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Solving 

the Powertrain 



10th Schaeffler Symposium 
April 3/4,2014 


4^ Springer Vieweg 


OPEN 





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10 th Schaeffler Symposium 

Welcome 


M ^ FAB 


SCHAEFFLER 


Schaeffler Technologies GmbH & Co. KG 

Herzogenaurach 

Germany 


© The Editor(s) (if applicable) and the Author(s) 2014. The book is published with open access at 
SpringerLink. com 

Open Access. This book is distributed under the terms of the Creative Commons Attribution 
Noncommercial License which permits any noncommercial use, distribution, and reproduction in 
any medium, provided the original author(s) and source are credited. 

All commercial rights are reserved by the Publisher, whether the whole or part of the material is 
concerned, specifically the rights of translation, reprinting, re-use of illustrations, recitation, bro¬ 
adcasting, reproduction on microfilms or in any other way, and storage in data banks. Duplication 
of this publication or parts thereof is permitted only under the provisions of the Copyright Law of 
the Publisher’s location, in its current version, and permission for commercial use must always be 
obtained from Springer. Permissions for commercial use may be obtained through Rights Link at 
the Copyright Clearance Center. Violations are liable to prosecution under the respective Copyright 
Law. 

The use of general descriptive names, registered names, trademarks, etc. in this publication does not 
imply , even in the absence of a specific statement, that such names are exempt from the relevant 
protective laws and regulations and therefore free for general use. 

DOI 10.1007/978-3-658-06430-3 

Library of Congress Control Number 2014946900 


SCHAEFFLER 

ujl FAG 


Publisher: 

Schaeffler Technologies GmbH & Co. KG 

Industriestrasse 1 - 3 

91074 Herzogenaurach • Germany 

Telephone +49 9132 82-0 

Telefax +49 9132 4950 

www.schaeffler.com 

Editor: Dr. Wolfgang Reik; 

Redaktionsburo delta eta, Frankfurt, Germany 

Layout: Nicole Daniel; Alexander Brand 

Copy editor: Heike Pinther; Jutta Dietz 

Printed by: Wunsch Druck GmbH, Neumarkt, Germany 

The technical papers found in this issue are available 

in electronic format in the libraby 

on our homepage (search criterion: Symposium). 

Reproduction in whole or part 
without our authorization is prohibited. 




3 


Contents 


Foreword 

4 

Overview of Technical Papers 

6 

Concept Vehicles 

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5 


Foreword 


There are plenty of technical ideas for the 
drive systems to power future genera¬ 
tions of vehicles. But which of these is the 
right solution for which application? How 
will the markets change in the future? 
What basic innovations can we use to 
make engines and transmissions even 
more efficient? 

At the 10th Schaeffler Symposium, 2014, 
we will venture a glimpse into the distant 
future - beyond 2020, when the most 
world's most stringent fuel consumption 
standards come into force in Europe. In the 
tradition of the Symposium, we will, on the 
one hand, be presenting our evolutionary 
technologies, which can make a significant 
contribution to optimizing the drive system. 
On the other hand, we will be discussing 
our radical innovations: Hybrid concepts 



for 48-volt on-board electric systems or a 
transmission concept with electric power 
splitting as well as electric wheel hub 
drives for passenger cars. 

We are convinced that the challenges 
of the future can only be overcome if per¬ 
manent further development of conven¬ 
tional powertrains based on internal com¬ 
bustion engines, and the courage to realize 
new ideas for electrification go hand in 
hand. As a supplier, we focus more than 
ever on people's changing behavior with 
regard to mobility, because public accep¬ 
tance will ultimately decide if and which 
technologies become established on the 
market. 

In this spirit, we hope you will have exciting 
discussions about an exciting topic: The 
automotive drive systems of the future! 


Norbert Indlekofer Prof. Dr. Peter Pleus 


Member of the Executive Board 
CEO Automotive 


Member of the Executive Board 
CEO Automotive 





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7 


Technical Papers 


1 

Mobility of the Future 

8 

2 

Powertrain Systems 

24 

3 

Developing Combustion Engines 

42 

4 

Manual Gearbox 

56 

5 

Centrifugal Pendulum-type Absorber 

78 

6 

Clutch Release Systems 

94 

7 

Clutch Systems 

112 

8 

Synchronisation Systems 

126 

9 

Simulation Engine Systems 

140 

10 

Camshaft Phasing Systems 

156 

11 

Valvetrain Systems 

172 

12 

Variable Valvetrain 

188 

13 

Rolling Bearings for Turbochargers 

202 

14 

48-volt Electric Axle 

212 

15 

Double Clutch Systems 

224 

16 

Dry Double Clutch 

230 

17 

Wet Double Clutch 

244 

18 

Planetary Transmission 1 

256 

19 

Planetary Transmission 2 

270 

20 

Torgue Converter 

280 

21 

Thermal Management 

302 

22 

Timing Drive Systems 

318 

23 

Customized Friction 

330 

24 

Start-Stop 

346 

25 

Transmission Actuators 

360 

26 

Differential Systems 

378 

27 

Chassis 

392 

28 

Range-Extender 

412 

29 

Hybrid Modules 

426 

30 

Wheel Hub Drives 

440 

31 

What Powertrains Could Learn from Each Other 

452 

32 

CAFE Demonstrator 

468 

33 

Belt Drive Systems 

480 

34 

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498 




































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9 


Individuality and Variety 

Paradigms of 
future mobility 


Prof. Dr. Peter Gutzmer 


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10 


Requirements 
for future mobility 


Mobility and climate protection 

Mobility is not only a basic human need, it 
also correlates closely with economic 
growth. This is not only true for passenger 
traffic (Figure 1) but also for commercial 
transport, particularly on the road. Ex¬ 
perts assume that there is a self-reinforc¬ 
ing effect between traffic and economic 
performance [1]. 

In contrast with total primary energy 
consumption, it has not been possible to 
disconnect traffic growth from economic 
growth. In spite of considerable savings in 
fuel consumption that have been achieved 
as a result of technical progress over the 
past few decades, the overall emissions of 


Gross domestic product 
GDP 1970-2010 



Source: Federal Office of Statistics and VDA 


Figure 1 Development of Germany’s gross 
domestic product and the number 
of passenger kilometers traveled 
annually for several years in the 
period from 1970 to 2010 


carbon dioxide related to traffic have con¬ 
tinued to increase. The political system re¬ 
acts to this by issuing ever stricter limits. 
This is a worldwide phenomenon in which 
limits converge, although with a time lag 
(Figure 2). 

In late 2013, the European Union made 
a commitment to the most stringent C0 2 
limit values worldwide [2]. According to 
this, a fleet limit value of 95 g/km will apply 
from 2020. This fleet limit value must be 
met initially by 95 % of the fleet and by 
100 % from January 1, 2021. Vehicles 
emitting less than 50 g/km may be count¬ 
ed multiple times for three years (2020- 
2022). The total effect from these so- 
called “super credits” is limited to 7.5 g/km 
for each fleet. 

This basically describes the primary 
task for future technical developments in 
motor vehicles. The challenge is to cover 
rising mobility requirements with fewer en¬ 
ergy resources, and particularly lower C0 2 
emissions. However, one consequence of 
the correlation between economic growth 
and mobility is that the greatest increase in 
the number of vehicles will be in the emerg¬ 
ing regions outside of the “old” industrial 
nations of the triad (EU, USA, Japan). The 
question here is whether technical solu¬ 
tions from Europe - the Schaeffler Group's 
home region - can be applied without any 
changes. 

What do customers want? - 
The Schaeffler mobility study 

The question of whether technologies can 
be transferred to another region is often re¬ 
duced to the issue of costs. This is a one¬ 
sided view that bears the risk of losing 
sight of the customer and the customer’s 
needs. For this reason, Schaeffler has de¬ 
cided to use a comprehensive approach 
for working out the development of future 
market scenarios. A recently completed 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_l, © The Author(s) 2014 





Mobility of the Future 


1 


11 



Years 

••• Proposed targets 

— Enacted targets USA: incl. light duty vehicles 


Source: ICCT 8/2013 


C0 2 limits for major passenger car markets 


Figure 2 

mobility study was based on a three-step 
method: 

1. Prepare mobility patterns for 12 selected 
world regions 

2. Cluster the patterns in a matrix 

3. Work out four in-depth scenarios for 
future mobility. 


Step 1: Mobility patterns 

The first step involved the preparation of 
twelve detailed mobility patterns for se¬ 
lected world regions during several work¬ 
shops. These patterns not only serve to 
analyze the current situation, they also ex¬ 
trapolate into 2025. Professional input 
was provided by in-house experts and 
from sources outside the company. Ex¬ 
cerpts from four of these analyses are 
presented below as examples: 

With an average household income of 
52,900 euros (2011), there is no doubt that 
the borough of Manhattan in New York 


can be called affluent. Its high traffic den¬ 
sity results in permanently congested 
roads, particularly during the day. Inhabit¬ 
ants are very willing to use public trans¬ 
portation because their priority is to mini¬ 
mize travel time. However, the capacity of 
public transportation is also limited. At the 
same time, the city needs to reduce its 
noise and pollutant emissions. Approach¬ 
es to solving these problems not only in¬ 
clude the expansion of public transporta¬ 
tion and increased use of bicycles, but 
also the introduction of small, agile elec¬ 
tric vehicles to maximize traffic area utili¬ 
zation. As part of its “PlaNYC” sustain¬ 
ability initiative, the city plans to set up a 
dense network of charging stations for 
electric vehicles. The plan also includes 
the addition of electric vehicles to the mu¬ 
nicipal vehicle fleets. 

The German state of Mecklenburg-Vor¬ 
pommern is quite the opposite. The state’s 
population of 1.6 million is similar to that of 
Manhattan but it is 260 times larger. Not 












12 


only is population density much lower, but 
also the average income of 22,884 euros 
per capita (2011). Outside the towns, public 
transportation is scarce due to low demand. 
As a result, the majority of passenger traffic 
consists of - for the most part used - cars. 
Since the population is also getting older, 
a further increase in mobile services is to 
be expected. A mobile medical service is 
already being tested in Mecklenburg- 
Vorpommern. 

Things are very different in the sprawl¬ 
ing city of Medellin in Colombia, which 
has a population of 2.7 million. With more 
than 7,000 inhabitants per square kilo¬ 
meter, population density is very high, 
and a large number of the poorest live in 
unofficial shanty towns (Favela) on the 
outskirts of the city. Most people use 
“paratransit” to get downtown. This 
means privately operated vans or large 
taxis without fixed routes or stops. Ex¬ 
panding public transportation and imple¬ 
menting stricter emissions standards for 
vehicles could provide relief for the 
smog-filled downtown area. In addition, a 
very unusual idea has been put into prac¬ 
tice. Medellin has integrated two cable 
car tracks into its regular public transpor¬ 
tation system that serve the shanty towns 
on the hills surrounding the city and can 
transport 3,000 people per hour. 

Bangkok has been pursuing yet another 
approach. This metropolitan area, which 
has a population of more than 12 million, 
has achieved a level of wealth that is con¬ 
siderable for a developing country. Annual 
household income is around 9,600 euros 
(2007), more than twice that of Medellin. 
Streets are gridlocked, and Bangkok’s in¬ 
habitants are very willing to use public 
transportation to get to work. However, 
the widely used buses sit in traffic along 
with the passenger cars. Expanding rail 
traffic would be time-consuming and ex¬ 
pensive, which is why Bangkok has been 
depending on a “bus rapid transport” sys¬ 


tem. This is made up of urban bus lines 
that use their own tracks separate from all 
other traffic and metro-like stops. At 
18,000 people per day on the system’s 
first line that covers 16.5 km, transport 
capacity is very high, and costs are twen¬ 
ty times lower than those of an elevated 
train. 

These mobility patterns primarily 
prove one thing: There is no single an¬ 
swer to the question of how to manage 
ever-increasing traffic volumes. Instead, 
there is a variety of answers that give 
consideration to issues ranging from lo¬ 
cal conditions to topography. It can also 
be seen that, at least in urban areas, lo¬ 
cal authorities are keen on finding solu¬ 
tions and have identified mobility as a 
factor in the international competition 
over geographic locations. 


Step 2: Clustering 

In the second step, we looked for a mas¬ 
ter pattern behind the various patterns. 
Regional mobility patterns were assigned 



Countryside City 
Level of urbanization 


High 


(/> <D 

52 I 


Low 


Industrial 
country 
Emerging vO ^ 
country 




Figure 3 Matrix for the categorization of 
mobility patterns 


Mobility of the Future 


1 


13 


to a three-dimensional matrix that includ¬ 
ed the dimensions of level of urbaniza¬ 
tion, purchasing power of users, and 
economic development level of each re¬ 
gion (Figure 3). 

It can be seen that all of the analyzed 
brands can be clearly assigned to one of 
the cubes in the 3-dimensional matrix. 
Manhattan meets the criteria for “City - 
Industrial Country - High Purchasing 
Power,” while Mecklenburg-Vorpommern 
can be categorized as “Countryside - In¬ 
dustrial Country - Low Purchasing Pow¬ 
er.” This clustering is important for the 
transferability of solutions from one region 
to another. 


Step 3: In-depth scenarios 

In the third and final step, we worked out 
four in-depth scenarios that Schaeffler be¬ 
lieves will determine future mobility. These 
are: 

Consideration of the entire energy 
chain 

Future mobility solutions will no longer 
consist of isolated measures but incor¬ 
porate the C0 2 footprint of the entire en¬ 
ergy chain. Here, special consideration 
must be given to the generation of elec¬ 
tricity for electric cars and to the genera¬ 
tion of hydrogen for fuel cell vehicles. In 
addition, storage also plays an important 
role in an energy supply that is primarily 
based on fluctuating renewable energies. 
Regardless of whether it is the methana- 
tion of hydrogen or electric cars as part 
of a smart grid - mobility will be increas¬ 
ingly regarded as part of an energy sys¬ 
tem. 

New mobility schemes for cities 

Intermodal traffic with seamless switch¬ 
ing from one form of transportation to 
another will be a matter of course in the 


cities of the future. For the continued de¬ 
velopment of motor vehicles, this means 
that it must fit seamlessly into the urban 
traffic network. In addition, the majority 
of the population in many fast-growing 
cities outside of the established industrial 
countries will develop a pragmatic atti¬ 
tude towards their own mobility and 
choose the most time-saving and cost- 
efficient option. 

Resource-efficient inter-urban 
mobility 

For a growing portion of the world’s popu¬ 
lation, it is becoming important to move 
between urban economic centers in a 
time-saving manner. Resource efficiency 
will increasingly become an essential 
characteristic for all carriers, regardless of 
whether they are airplanes, high-speed 
trains, or cars. At the same time, the auto¬ 
mation of inter-urban traffic continues, 
which also applies to automobiles (auton¬ 
omous, automated, or piloted driving), not 
forgetting the integration into communica¬ 
tion networks. 

Environmentally friendly drives 

Vehicles’ drives are one of the major fac¬ 
tors that determine the energy efficiency 
and environmental compatibility of mobil¬ 
ity. That is why the development of ener¬ 
gy-efficient drives will continue to take top 
priority. This includes the optimization of 
existing drives as well as the introduction 
of entirely new systems. The goal of re¬ 
ducing C0 2 and pollutant emissions - or 
of someday eliminating them entirely - not 
only extends to the use of a vehicle but to 
its entire lifecycle, particularly its produc¬ 
tion. 


14 


Energy efficiency as a 
driving force behind drive 
development 


Most experts would agree that so-called 
“conventional” powertrains - consisting of 
an internal combustion engine and a trans¬ 
mission with a high ratio spread - will domi¬ 
nate most of the world’s private transport. 
The market shares that electric drives and 
hybrid drives may be able to gain over the 
next few years vary by region and political 
provisions. Figure 4 shows a forecast by IHS, 
a renowned market research company. 

The market data show that an effective 
strategy for the reduction of C0 2 emissions 
from private transportation must prioritize 
increased efficiency in conventional pow¬ 


ertrains based on internal combustion en¬ 
gines. Since diesel engines as efficiency 
drives will only gain large market shares in 
certain regions such as Europe, India, and 
South Korea, the optimization of the gaso¬ 
line engine (which was first produced in 
1877) remains the most important task in 
engine development. 


Muda! - minimizing power loss 

The starting point for the optimization of 
every process is an evaluation of the loss¬ 
es incurred - that is, an increase in effi¬ 
ciency. In production circles, this ap¬ 
proach is known as the Muda principle, 
going back to an engineer named Taiichi 
Ono who is considered to be the inventor 
of the Toyota production system. “Muda” 
simply means “avoid waste.” 



Gasoline Diesel Hybrid 


Electric vehicle 


China 


f North 
America 


Europe 


rSouth 

AmeriCc 


Korea 


Figure 4 Market shares of various drive systems in regions of the world for 2011, 2016 and 2020 in % 



Mobility of the Future 


1 


15 


When applied to vehicle drives, power loss¬ 
es that distinguish real engines from a ther¬ 
modynamically optimum process must be 
analyzed consistently and technical coun¬ 
termeasures must be taken. A good exam¬ 
ple here is the reduction of frictional power 
loss in the powertrain which, according to 
[3], can lower the fuel consumption of a 
mid-size vehicle with a gasoline engine by at 
least 3 %. 

One example of applied frictional power 
loss reduction is lightweight balancer shafts 
with rolling bearing supports. As part of the 
reduction of displacement and the number 
of cylinders, balancer shafts are increas¬ 
ingly used because they allow the quiet op¬ 
eration of small engines with a high specific 
performance that customers demand. The 
problem: Balancer shafts “eat up” part of 
the energy saved, both because of the nec¬ 
essary acceleration of their mass and the 
frictional power loss in their bearing sup¬ 
ports. Schaeffler has found a solution by 
creating a new balancer shaft with rolling 
bearing supports. The balancer shafts are 
optimized geometrically so that a mass re¬ 
duction of up to one kilogram can be 
achieved for a four-cylinder engine. In addi¬ 
tion, the rolling bearing supports of the 
shaft(s) have helped achieve a friction re¬ 
duction of up to 50 % (Figure 5). 



Engine speed in rpm 

— Plain bearing 

— Rolling bearing 


Figure 5 Minimization of frictional power loss 
through balancer shafts with rolling 
bearing supports 



Figure 6 Axial needle roller bearing supports 
of planet carriers 

Efficiency, which is already high, can also be 
increased in the transmission that is charac¬ 
terized by numerous rotary, load-transmitting 
parts. Substituting the plain bearing sup¬ 
ports with planetary gears for planet pinions 
with thrust needle roller bearings is one ex¬ 
ample here (Figure 6). In third gear, for in¬ 
stance, maximum frictional power loss is re¬ 
duced from 470 W to just 50 W. For a 
transmission with four planetary gear sets, 
this means a frictional power loss reduced by 
420 W in third gear, and thus a 90 % reduc¬ 
tion. Based on the simulation, consumption 
can be expected to be reduced by around 
0.5 % when substituting the thrust washers 
with thrust needle roller bearings in the NEDC. 

It is very obvious that the analysis of 
power losses can not be restricted to the 
engine and transmission unit. It is neces¬ 
sary to look at the entire powertrain, includ¬ 
ing the wheels. This will help identify other 
sources of loss, such as the differential and 
the wheel bearings. Over the past few years, 
significant progress has been achieved 
here, such as by replacing tapered roller 
bearings with tandem angular contact ball 
bearings in the rear axle differential. 




16 


Added value through increased 
variability 

As important as the reduction of me¬ 
chanical transmission losses in the pow¬ 
ertrain may be, this in itself will not result 
in a thermodynamic optimum. The losses 
that occur in an engine are also influ¬ 
enced significantly by the throttle losses 
that depend on the operating point. This 
is even more true for modern internal 
combustion engines as the valve opening 
times cannot be controlled on the basis 
of the maximum power output alone. In¬ 
stead, the raw emissions that depend on 
the combustion process have become 
an important design criterion. In terms of 
thermodynamics, it would be ideal to 
have an entirely free control of the gas 
exchange that is adjusted to the relevant 
operating point. This ideal situation could 
only be achieved by using electrome¬ 
chanical valves that are completely de¬ 
coupled from the crankshaft. However, 
there are numerous arguments against 
this solution - such as the fact that a 
software error might lead to the immedi¬ 
ate destruction of the engine. 

Systems for camshaft phasing adjust¬ 
ment permit an initial approach towards 


this solution. They allow the valve lift curve 
to be “moved,” i.e. the valves can be 
opened or closed earlier or later. The lift 
curve as such remains unchanged. The 
timing velocity is an essential quality crite¬ 
rion that is usually expressed as degrees of 
crank angle per second (°CA/s). The high¬ 
est adjustment speeds, as well as com¬ 
plete freedom for the valve opening times 
when the engine starts, are provided by 
electromechanical phasing units. Schaeffler 
will launch the volume production of this 
type of system for the first time in 2015. 
However, because electromechanical so¬ 
lutions will have an impact on costs, 
Schaeffler continues to develop its hydrau¬ 
lic phasing units. 

Valve lift can be varied - usually between 
two predefined points - by means of various 
technical solutions, such as switchable tap¬ 
pets. This creates the prerequisite for limiting 
throttle losses in low-load ranges. 

It is Schaeffler’s electrohydraulic Uni- 
Air valve train system, launched around 
four years ago and since produced for ap¬ 
proximately 400,000 engines, that pro¬ 
vides near-complete variability. It permits 
nearly arbitrary formation of the lift curve 
within a predefined maximum valve lift 
(Figure 7). 



Variable valve train 

Lift, timing, and duration 


Discrete (switchable) Continuous 



Figure 7 Variable lift curves through camshaft phasing units (left), switchable valve actuation (center) 
and the electrohydraulic UniAir valve train system (right) 








Mobility of the Future 


1 


17 


The UniAir system’s current applications 
are limited to the intake side, and the two 
intake valves are controlled simultane¬ 
ously via a hydraulic bridge. Even this so¬ 
lution permits fuel consumption to be re¬ 
duced by up to 15 %, compared to a 
naturally aspirated engine as the starting 
point. During the 2014 Schaeffler Sympo¬ 
sium, a variety of new functions will be 
shown that can be achieved with a refined 
UniAir. Examples include a system for 
varying the valve overlap by means of a 
two-stage actuating cam. If such func¬ 
tions are utilized consistently, additional 
C0 2 savings potential can be developed 

- incidentally, not only for gasoline en¬ 
gines but also for diesel engines and even 
for ship propulsion. 

More variability for lower C0 2 emissions 
is not just an issue for engines but also for 
transmissions and chassis. Here are some 
examples: 

- For transmissions, there is a definite 
trend towards higher ratio spreads and 
thus a higher number of gears. These 
transmissions permit the engine to be 
operated at operating points with low 
specific consumption as often as pos¬ 
sible. This development has conse¬ 
quences for conventional Schaeffler 
products such as clutches since the 
number of gearshifts increases along 
with the number of gears. 

- More variability can thus lead to less 
friction. What is new here is a switch- 
able wheel bearing for vehicles with 
high wheel and axle loads. It is a four- 
row angular contact ball bearing. 
When the car is driven in a straight 
line, load is applied only to the center 
rows of balls, and no load is applied to 
the external rows. When driven around 
curves, the external rows are engaged 
to support driving behavior in curves 
with the required high rigidity. Initial test 
results have shown an additional fric¬ 
tion reduction of more than 25 %. 


Intelligent electrification 


The stricter C0 2 regulations become 
(and the smaller the market share of die¬ 
sel engines), the sooner automobile 
manufacturers reach a point when the 
electrification of the drive becomes rele¬ 
vant. The degree to which emissions are 
reduced is highly dependent on the level 
of electrification that can essentially be 
described by the output of the electric 
motor and the energy content of the bat¬ 
tery. These parameters determine the 
functions that can be used to avoid the 
consumption of carbon fuel: 

- Turning off the engine when stopping 
(start/stop) or in coasting mode at high¬ 
er speeds 

- Moving the engine load point to point 
to mapping areas with low specific 
consumption (“boosting”) 

- Recuperating braking energy 

- Electric driving in low-load ranges in 
which an internal combustion engine is 
operated with a highly unfavorable effi¬ 
ciency factor 

- Using renewable energy for the drive 
provided that the battery can be 
charged externally 

Unfortunately, the necessary engineering 
and expense increase along with in¬ 
creasing electric power. This is particu¬ 
larly true for, but not limited to, the 
battery. It is thus a good idea touse a 
step-by-step procedure for electrification 
to keep mobility affordable (Figure 8). 
Schaeffler development activities for all 
stages of electrification have been con¬ 
centrated in its eMobility Systems Divi¬ 
sion since 2012. 



18 


Micro hybrid Mild hybrid Full hybrid Plug-in hybrid Electric car 


Electric driving 
in all operating 
conditions 


Functionality 


Start-stop 


Boosting, 

recuperation 


E-creeping, 

stop-and-go, 

e-sailing 


Electric 

driving 


Charging 


yes 


yes 


Elec, motor power 0.5... 8 kW 8 ... 20 kW 10... 50 kW 30... 125 kW 30... 125 kW 


Voltage 


Electrical range 


C0 2 saving 
E-Wheel Drive 

Electric axle 

Hybrid module 
Start-stop 


12 ...48 V 


4 ... 6 % 


48 ...280 V 48 ...400 V 200 ...400 V 200 ...400 V 


0.1 ...5 km 


10 ...50 km 


> 75 km 


12... 16% 


15 ...25% 


>50% up to 100% 



Figure 8 Stages of electrification 

Affordable and efficient: 

The 48-volt on-board electric system 
as an opportunity 

Until recently, the hybridization of a vehicle 
meant adding a high-voltage level to the 
conventional 12-volt on-board electric sys¬ 
tem. In today’s volume-produced hybrid ve¬ 
hicles, voltages of up to 300 or 400 volts are 
generated and in some prototypes up to 
700 volts have been implemented in order 
to make the construction of the electric 
units as compact as possible. 

Driven by active chassis with their typi¬ 
cally brief power peaks, a few automobile 
manufacturers introduce a 48-volt on¬ 
board electric subsystem. This is a great 
opportunity for the drive, as electric trac¬ 
tion motors with an output of up to 15 kW 
can be produced at this voltage level with 
moderate system costs. These reduced 
costs can be attributed in part to much 
lower safety requirements. No separate 
contact protection is required for the com¬ 


ponents of a 48-volt on-board electric 
system. In combination with a small lithi¬ 
um-ion battery (approx. 125 Wh), short 
distances can be driven at low speed 
using electric power only, such as when 
parking or in stop-and-go traffic. Func¬ 
tions such as boosting or recuperating 
with a much improved energy intake are 
also possible. 

Schaeffler has been working on two 
solutions for the technical implementation 
of the 48-volt hybrid drive that will be pre¬ 
sented in detail during the 2014 Sympo¬ 
sium: A 48-volt variation of the hybrid mod¬ 
ule integrated into the drive and an electric 
axle. 

Integrating the electric motor into an 
automatic transmission in place of the 
torque converter has proven to be a good 
solution in previous hybrid vehicles since 
no additional design space must be pro¬ 
vided this way. The same can be achieved 
with a 48-volt hybrid module. However, an 
additional challenge lies in the fact that, at 
least in Europe, the transmission’s level of 
























Mobility of the Future 


1 


19 



Figure 9 Impulse clutch with an integrated 

electric motor for combination with a 
manual transmission 

automation is low specifically in vehicle 
categories for which hybridization with a 
relatively inexpensive 48-volt approach 
would be attractive. That is why Schaeffler 
has developed several solutions for com¬ 
bining the hybrid module with a manual 
transmission that will be presented during 
the 2014 Symposium. The use of an im¬ 
pulse clutch appears to be particularly 
attractive (Figure 9). 

Here, the starter is eliminated, and the 
internal combustion engine is brought up 


to speed exclusively by closing the clutch, 
or it is started by the electric motor of the 
hybrid module. It is a transmission that 
can be shifted very quickly and must be 
able to transmit very high alternating 
torques of up to 1,500 Nm. In this case, 
the entire hybrid module, including the 
electric motor, is installed on the crank¬ 
shaft side. 

An attractive alternative for automobile 
manufacturers is the use of an electric 
axle on a 48-volt basis because here, the 
conventional part of the powertrain can 
remain completely “untouched” with the 
exception of the engine control system. A 
48-volt axle can be integrated into the 
powertrain using various configurations 
(Figure 10). The drive axle can be assisted 
in both front-wheel and rear-wheel drive 
vehicles. In addition, an electric drive for 
the rear axle can be installed in a front- 
wheel drive vehicle, a configuration some¬ 
times described as an “electric four-wheel 
drive." The electric drive force can also be 
distributed between the front and rear 
axle, although this means that two electric 


Front-wheel drive Rear-wheel drive 




eAWD H All-wheel drive 









c 

4^4 

i 


L 

“P^JO- 

L 






GEL 






i 

□ 0 



I~pe1 




HE 



— m- 



1 

— CP-" 



> 

r 



5 

r 


Traction 




Traction 



support 




support 



Figure 10 Vehicle topographies with electric axle drive 


























































































20 


motors and two power electronics units 
are required. 

The 48-volt hybrid with an electric axle 
will be presented in detail during the 2014 
Symposium. 

Sporty and dynamic: High-voltage 
hybrid technology 

In future, large vehicles and sports cars 
will increasingly be designed as plug-in 
hybrids to achieve particularly favorable 
standard fuel economy. This trend to¬ 
wards plug-in vehicles has resulted in a 
significant increase in the electric power 
required. Hybrid vehicles will be designed 
to complete the entire test cycle on elec¬ 
tric power. Consequently, one of the pri¬ 
mary development goals for the next gen¬ 
eration of the Schaeffler hybrid module 
has been to increase the power and 
torque density while also reducing the re¬ 
quired design space. At the same time, 
the torques of the internal combustion 
engines used in hybrid vehicles also in¬ 
crease. The second generation of Schaeffler’s 
hybrid module takes this market trend into 
account. The transfer of extremely high 
torques of up to 800 Nm is made possible 
by a patented system for splitting the 
power flow. The torque of the internal 
combustion engine is transferred to the 
transmission by both the closed discon¬ 
nect clutch and simultaneously via a one¬ 
way clutch. 

Some essential features of the high- 
voltage variation of the electric axle have 
also been developed further over the 
past four years. The third generation, 
currently being tested, has been adjust¬ 
ed to the topology of a plug-in hybrid ve¬ 
hicle with a front-mounted engine and 
front-wheel drive. The drive unit contin¬ 
ues to be designed for coaxial installation 
in the rear axle. Water-cooled, hybrid- 
design electric motors (permanent-mag¬ 


net synchronous motors with a high level 
of reluctance) are used. These automo¬ 
bile-specific requirements are in contrast 
to the industrial motors used in the first 
generation. 

The transmission still has a planetary 
design but now has two transmission lev¬ 
els. With an increased power density, the 
transmission has a modular design that 
permits the traction and active torque dis¬ 
tribution (torque vectoring) to be offered 
as separate functions. 

Urban and flexible: Drives for electric 
vehicles 

As described in the first section, large cities 
with a high population density and great af¬ 
fluence will increasingly see electric vehicles 
as part of an intermodal traffic mix. Most of 
those vehicles will initially be model varia¬ 
tions of series in which conventional power- 
trains are dominant. Therefore, most elec¬ 
tric vehicles are currently equipped with a 
center drive. 

As market penetration increases, a 
larger number of battery-electric vehicles 
will become available that have been de¬ 
veloped specifically for the requirements 
of urban traffic. Schaeffler believes that a 
wheel hub drive is the best solution for 
these vehicles. Since there is no “engine 
compartment”, this permits the design of 
completely new body types that offer 
very good utilization of the available 
space - an important requirement for 
traffic in urban areas that are congested 
anyway. In addition, drive shafts are no 
longer required, which permits the wheel 
angle to be increased. From the custom¬ 
er’s point of view, this results in much 
better maneuverability. 

For customers, this makes cars more 
fun to drive as well as making them safer, 
since the control quality of the drive is 
above that of center drives due to its di- 



Mobility of the Future 


1 


21 


New vehicle 
concepts 



Increased 

maneuverability 



Increased 

safety 


* 


* 


/ - 


Increased 

agility 




Power 

electronics 


Liquid cooling 


Electric motor 
(internal rotor) 


Wheel bearing 


Brake 


Figure 11 Wheel hub drive with integrated electronic system - current development status 


rect transmission - without a transmis¬ 
sion and side shafts. These conventional 
goals of automobile development will de¬ 
cide customer acceptance of small city 
automobiles. Reason alone - such as a 
small traffic area and a good carbon 
footprint - will not make electric vehicles 
marketable. 

Based on this motivation, Schaeffler 
has been developing wheel hub drives 
since 2007. In cooperation with the Ford 
research center in Aachen, the current 
development status (Figure 11) has been 
installed in a Ford Fiesta that serves as a 
test vehicle. The total vehicle weight has 
not increased when compared to an identi¬ 
cal type of vehicle with a diesel engine 
(1,290 kg when empty). This includes a 
lithium-ion battery with a nominal capacity 
of 16.2 kWh. 

This test vehicle has been used for 
various driving dynamics tests on the test 
site. These tests have shown that, up to 
speeds of 130 km per hour, the prototype 
is at least equivalent to a volume pro¬ 
duced vehicle that was also driven. Ma¬ 
neuvers that utilized the potential of torque 
vectoring even yielded some significant 


performance increases. During a stan¬ 
dardized swerving-stability test with the 
traffic cones spaced 18 meters apart, the 
speed was increased by around 10 km 
per hour. 

Schaeffler has already been working 
on the next generation of wheel hub drives 
with Ford and Continental as well as with 
RWTH Aachen and the Regensburg poly¬ 
technic in the MEHREN research project 
(MEHREN stands for multiple-motor elec¬ 
tric vehicle with the highest possible 
space and energy efficiency and uncom¬ 
promising driving safety). The focus of the 
project is on implementing a new software 
architecture specifically designed for wheel 
hub drives. In addition, the MEHREN project 
is intended to show for the first time what 
kind of potential there is for new vehicle 
architectures if wheel hub drives are used 
as a standard drive to begin with. Com¬ 
pletion of a virtual prototype is expected 
for 2015. 








22 


Summary and outlook 


Mobility solutions for the future will be 
customized for specific applications more 
than ever before. As a consequence, the 
development of vehicle drives is an es¬ 
sential factor for energy efficiency in every 
mobility chain. Refined, highly efficient in¬ 
ternal combustion engines and transmis¬ 
sions work hand in hand with electric 
drives that are adjusted to the vehicle 
configuration but rely on a modular design 
system for core components. 

To be able to identify the right solution 
out of a wide variety of possible solutions, 
Schaeffler not only looks at technical po¬ 
tential but also at fundamental changes 
in markets and customer requirements. 
These requirements are transformed into 
ideas for solutions and finally technical in¬ 
novations by means of a well-structured 
process. This approach is true to the mot¬ 
to of Thomas Edison, whose “Menlo Park” 
laboratory was the first innovation factory: 
“I find out what the world needs. Then I go 
ahead and try to invent it.” 


Literature 


[1] Eichhorn, U.: Zukunft der Mobilitat - grun, 
sauber und vernetzt. cti Getriebesymposium, 
2013 

[2] Council of the European Union (eds.): Informal 
agreement on car C0 2 emissions reduction, 
press release, 2013 

[3] Ernst, C.; Eckstein, L.; Olchewski, I.: C0 2 - 
Reduzierungspotenziale bei Pkw bis 2020. 
Studie, Aachen, 2013 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



24 


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25 


Powertrain Systems of the Future 

Engine, transmission and damper systems for downspeeding, 
downsizing, and cylinder deactivation 


Dr.-lng. Hartmut Faust 


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26 


Introduction 


Besides hybridizing the powertrain, which is 
especially advantageous in city traffic, ef¬ 
forts must be made to improve the efficien¬ 
cy of conventional powertrains in order to 
reduce traffic-based C0 2 emissions. 

This will first require measures to directly 
reduce friction losses in internal combustion 
engines, transmissions, and chassis sys¬ 
tems, such as the use of friction-optimized 
bearing supports and seals as well as coat¬ 
ings to lower the friction coefficient. 

Furthermore, slippage losses in startup 
elements need to be reduced. Hydrody¬ 
namic torque converters with lock-up 
clutches are a notable example of this, as 
they can be engaged even at very low en¬ 
gine speeds by means of optimized damper 


systems. Double clutch systems with re¬ 
duced passive clutch drag torque losses of 
wet or - even better - dry running design 
are important contributions as well. 

The aim of this paper is also to report on 
improvements to the system as a whole, in 
which changes on the transmission side 
lead to an efficiency increase in the internal 
combustion engine. Examples of this in¬ 
clude transmissions with an increased 
spread of gear ratios, resulting in lower en¬ 
gine speeds even at higher travel speeds [1]. 
Optimized damper systems serve to further 
reduce and/or insulate torsional vibration 
excitation introduced into the entire pow¬ 
ertrain by cyclical combustion in the engine 
and facilitate downspeeding of drive sys¬ 
tems in order to reduce fuel consumption. 

At the same time, advanced damper 
systems permit the design of downsizing 
systems that reduce engine friction with a 



Figure 1 Samples from the product portfolio of the Schaeffler Group’s Transmission Systems 
Business Division designed to reduce losses and optimize comfort as well as NVH 
behavior 

Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 

DOI 10.1007/978-3-658-06430-3_2, © The Author(s) 2014 





















Powertrain Systems 


2 


27 


lower number of cylinders and substan¬ 
tially increased torsional vibration excita¬ 
tion without having strong NVH issues in 
the entire powertrain. Finally, a rolling cylin¬ 
der deactivation system is introduced that 
enables engines with three cylinders to run 
effectively on 1.5 cylinders (“RCD 1.5”). The 
measures taken on the engine and trans¬ 
mission system side to prevent excessive 
torsional vibrations along the entire pow¬ 
ertrain are described in detail. 


Reducing consumption 
by means inside the 
transmission 


An analysis of energy losses in the chain 
from well to wheel shows that the greatest 
percentage of energy losses occurs when 
the chemical energy bound up in fuel is con¬ 
verted to mechanical power at the crank¬ 
shaft. This is due to 
the high thermody¬ 
namic and friction 
losses in the internal 
combustion engine. 

In contrast, the 
power transmission 
efficiency is up to 
more than 90 %, 
depending on the 
transmission sys¬ 
tem and operating 
conditions. Never¬ 
theless, efforts to 
reduce this rather 
low proportion of 
the losses are valu¬ 
able as well, since 
such optimizing 
measures usually 
generate minimal 



additional costs relative to the increase in 
efficiency. Due to legislative regulations 
that - starting in 2020/2021 - will bring 
penalties of up to 95 euros per g/km in ex¬ 
cess of a C0 2 emission limit of 95 g/km in 
the EU, clear target values can now be de¬ 
rived with regard to the additional expendi¬ 
ture that is acceptable in order to increase 
efficiency. 

In presentations at the 10 th Schaeffler 
Symposium in 2014, many solutions for re¬ 
ducing C0 2 emissions will be introduced in 
detail. Figure 1 provides an overview of the 
product portfolio. 

In planetary automatic transmissions, 
plain bearing supports are being increas¬ 
ingly replaced by rolling bearing supports. 
Needle roller bearings are very frequently 
used for this application and in the case of 
planet gear bearing supports are subjected 
to centripetal acceleration. In the most re¬ 
cent nine-speed automatic transmissions, 
both for inline and FWD arrangements, val¬ 
ues up to 7,200 g must be taken into con¬ 
sideration and made sustainable by means 
of a suitable design (Figure 2). 


,n; 


a> c . 3,500 g 
o. o ' 


max. 

max. 7,200 g 
max. 6,000 g 
4,700 g 




6 gear 


9 gear 


Figure 2 


Centripetal acceleration values in the planet gear bearing 
supports of automatic transmissions and a newly developed 
axial needle roller bearing support for planet gears with a high 
relative speed 





28 


For the CVT, the advantages of the LuK 
chain with low-friction rocker joints com¬ 
pared to other CVT linking elements [2, 3] 
are being increasingly implemented on the 
market with an improved fuel consumption 
of up to 4 %. Starting with applications that 
have a high torque of 400 Nm, chains with 
smaller pitch lengths are now being used as 
well. Besides the volume-produced 08 and 
07 chain types, the smaller 06 and 05 types 
are being developed in order to make use of 
the robustness and efficiency advantages in 
the lower torque and vehicle class range 
also. 


Startup elements 


Hydrodynamic torque converters 

Along with optimizing the hydrodynamic cir¬ 
cuit in order to keep losses to a minimum 
even in open converter operation, the hy¬ 
drodynamic torque converters provided for 
automatic transmissions take the following 
key developmental aspects into account: 

- High-capacity torsional dampers, in¬ 
cluding centrifugal pendulum-type ab¬ 
sorbers running in oil that facilitate early 
lock-up even at very low engine speeds 
and 

- Reduction of the rotating masses being 
accelerated. 

Great progress is being made with the new 
development referred to as iTC with its in¬ 
novative integration of the lock-up clutch 
into the turbine wheel [4] (Figure 3). 


A broad portfolio of startup elements is pro¬ 
duced under the Schaeffler LuK brand - 
from a dry clutch for manual transmissions 
and torque converters to double clutch sys¬ 
tems with a wet or dry design. 


Double clutch systems and their 
actuators 

For double clutch system solutions [5, 6], 
which are gaining an ever greater share 



Figure 3 Innovative iTC with lock-up clutch integrated into the turbine wheel 






































Powertrain Systems 


2 


29 


180 Nm “dry” 


370 Nm “wet” 



■ Parts supplied by Schaeffler 


Figure 4 Dry and wet running double clutch systems, including electrically power on demand 
operated clutch and transmission actuators from Schaeffler for hybrid transmissions 


of the market, Schaeffler’s LuK brand 
has been offering dry double clutch sys¬ 
tems since the end of 2007. In contrast to 
wet double clutches, they have the ad¬ 
vantage of not causing fluid-induced 
drag losses in the passive clutch, which 
account for approx. 2 % fuel consump¬ 
tion and C0 2 emission advantages in the 
NEDC. In the meantime, volume-pro¬ 
duced dry double clutches have been 
delivered to five international OEMs and 
transmission manufacturers, even for 
hybridized versions (Figure 4). 

The range of applications of dry double 
clutch systems currently includes engine 
torques of up to 250 Nm. The main objec¬ 
tive of current development work is to con¬ 
tinue optimizing comfort features in order to 
meet increasing demands and the wide 
range of usage profiles - including for hy¬ 
bridized powertrains. 

After Schaeffler had already been in¬ 
volved in the initial basic development of 
wet multi-disk clutches in the 300 Nm 
range, volume production of the first wet 
double clutches from Schaeffler’s LuK 
brand started in 2013 (Figure 4 right). 


In many applications, LuK not only offers 
double clutches, but also the clutch actua¬ 
tion system with optimized auxiliary energy 
consumption. For example, the lever actua¬ 
tor made it possible to pursue the power- 
on-demand principle so that the clutch can 
be actuated with small electric BLCD mo¬ 
tors and the electrical power consumption 
is under 20 W during practical driving oper¬ 
ation including electromechanical gear ac¬ 
tuation [7]. 

Moreover, volume production has be¬ 
gun for a new electrically operated hydro¬ 
static clutch actuator (HCA). The HCA was 
developed in a modular design approach so 
that it could be used for actuating both dry 
and wet double clutches in conjunction with 
engagement bearings. 

At the same time, volume production of a 
new kind of gearshift actuator was launched, 
which uses the active interlock concept to 
actuate all of the gears of the hybridized dou¬ 
ble clutch transmission with the help of two 
electric motors. This actuator was also de¬ 
veloped with a modular design so that it can 
be used in both dry and wet double clutch 
transmissions (Figure 4 left and right). 


30 


Damper systems for 
torsional vibration 
isolation 


Trends in engine development place high 
requirements on damper systems: 

- Downsizing to reduce internal engine 
losses resulting in higher torsional vi¬ 
bration excitation due to lower num¬ 
bers of cylinders coupled with lower 
excitation frequencies 

- Higher turbocharging pressures with a 
corresponding torque increase and 
higher peak pressures, leading to in¬ 
creased excitation amplitudes 

- Downspeeding with high torques even at 
very low engine speeds thanks to opti¬ 
mized turbocharging concepts, which 
leads to even lower excitation frequen¬ 
cies coupled with very high amplitudes. 

The developmental history of damper 
systems extends from the transition from 


torsionally damped clutch disks to the 
dual mass flywheel with an extremely low 
first natural frequency and corresponding 
isolation of all higher excitation frequen¬ 
cies to the introduction of the centrifugal 
pendulum-type absorber (Figure 5). 

The centrifugal pendulum-type absorb¬ 
er is a kind of vibration absorber, whose 
frequency is inherently regulated by the en¬ 
gine speed frequency due to the centrifu¬ 
gal effect so that the damping effect can 
be utilized for all speeds according to the 
main engine vibration order. Due to the po¬ 
sitioning of the centrifugal pendulum-type 
absorber (CPA) on the secondary side of 
the dual mass flywheel (DMF), it was pos¬ 
sible with a small mass to achieve a signifi¬ 
cant additional reduction of the engine ex¬ 
citation on the transmission input shaft, 
which was already insulated by the DMF. 
This is used for both manual transmissions 
(MT) and double clutch transmissions 
(DCT). It has not been needed in previous 
applications of dry double clutch transmis¬ 
sions, since the required thermal masses of 


Full-load 


Down- 



Higher 

sensitivity 


DMF + 


Torsional 

damper 




Dual-mass 


1985 flywheel (DMF) : 



Prim. Sec. Vehicle 
Trans. — 


centrifugal ^ A ^ n^n 

2008 pendulum 

absorber WiBt'd? ^ — 


Figure 5 History of damping system development 












Powertrain Systems 


2 


31 


CPA in DMF 
for MT & DCT 


1,000 1,500 2,000 


the pressure plates 
already provide suf¬ 
ficient isolation for 
torsional vibrations 
with conventional 
dual mass fly¬ 
wheels. It has been 
possible to use the 
centrifugal pendu¬ 
lum-type absorber 
even in torque con¬ 
verter dampers 
(Figure 6). 

When used in 
torque converters, it 
is important to con¬ 
sider here that the 
centrifugal pendu¬ 
lum-type absorber 
is immersed in oil, 
meaning that corre¬ 
sponding adjust¬ 
ments of the char¬ 
acteristic curve must 
be calculated by 
means of simula¬ 
tions and measure¬ 
ments on the com¬ 
ponent test stand 
and in the vehicle in 
order to arrive at 

optimum operational results. By using the 
centrifugal pendulum-type absorber, it is 
possible to close the lock-up clutch soon¬ 
er, for one thing - at speeds even below 
1,000 rpm - and, for another, to avoid loss- 
inducing acoustic micro-slip. Besides sav¬ 
ing on consumption, this also achieves a 
stronger connection in the entire powertrain 
with a better dynamic sensation. 


Damper systems for 
cylinder deactivation 

The deactivation of cylinders in internal 
combustion engines running under partial 


CPA in Torque Converter 
for AT 




— Standard 

— with CPA 


1,000 1,500 2,000 


Speed in rpm 



Figure 6 


Use and effect of the centrifugal pendulum-type absorber in 
dual mass flywheels for manual and double clutch transmis¬ 
sions as well as in torque converters 


load is increasingly being introduced for re¬ 
ducing fuel consumption and C0 2 emis¬ 
sions. This leads to the requirement for the 
damper system to ensure good NVH qual¬ 
ity when the engine is operating both on all 
cylinders and a partial number of cylinders. 
The easiest solution is still to manage a V8 
engine running on four-cylinders. Depend¬ 
ing on the application, a conventional 
damper can be designed for when the en¬ 
gine is operating on all cylinders and the 
additional centrifugal pendulum-type ab¬ 
sorber designed for cylinder deactivation 
operation only so that good torsional vi¬ 
bration behavior can be ensured in both 
cases. In a four-cylinder engine with the 
















32 


CPA for CPA for 

4-cylinder mode 8- and 4-cylinder mode 



■ 4-cylinder CPA 

■ 8-cylinder CPA 


cylinder engines are resulting in increased 
requirements, both when operating the 
engine on all cylinders and a partial num¬ 
ber of cylinders. Solutions are being de¬ 
veloped that actually incorporate two dif¬ 
ferent centrifugal pendulum-type absorber 
systems in order to optimize both operat¬ 
ing modes independently of each other 
(Figure 7). To do so, one pair of pendu¬ 
lum-type absorbers is calibrated for op¬ 
eration of the engine on all cylinders and 
the other for operation on a partial num¬ 
ber of cylinders with half of the primary 
order of excitation. 


Figure 7 Centrifugal pendulum-type absorber 
combination matched for operation 
of the engine on all cylinders and 
with cylinder deactivation 

two center cylinders deactivated, it has 
been sufficient to implement an adequate 
damper solution by optimizing a two-stage 
curve for the dual mass flywheel due to the 
limited torque range in two-cylinder opera¬ 
tion. 

However, new applications with very 
high nominal torques, both in V8 and four- 

Engine 



Transmission 



— 3-cylinder 
-CDA2/3 


Speed in rpm 


Figure 8 Torsional vibration excitation for conventional static cylinder 
deactivation with two of the three cylinders active (CDA 2/3) 


New kinds of rolling 
cylinder deactivation for 
the “1.5-cylinder engine" 


If additional C0 2 reduction must be 
achieved by means of cylinder deactiva¬ 
tion for three-cylinder engines as well, this 
raises the question 
as to whether this 
can be attained 
through simple 
static cylinder de¬ 
activation. Tor¬ 
sional vibration 
simulations indi¬ 
cate large excita¬ 
tion amplitudes, 
however (Figure 8). 

What is more, 
the order analysis 
shows that excita¬ 
tion is mainly char¬ 
acterized by a very 
low 0.5 th funda¬ 
mental order (Fig¬ 
ure 9). This can 
hardly be brought 














Powertrain Systems 


2 


33 



Transmission 



1.5 th order 2.0 nd order 


Figure 9 Order analysis with conventional static cylinder deactivation 
CDA 2/3 


to a torsional vibra¬ 
tion level that is ac¬ 
ceptable for the 
powertrain with the 
clamper designs of 
today. 

Further reflec¬ 
tions on the physi¬ 
cal and mathemati¬ 
cal background of 
the origin of excita¬ 
tion orders have 
led to the sugges¬ 
tion of designing 
rolling cylinder de¬ 
activation in three- 
cylinder engines, 
ultimately leading 
to “1.5-cylinder op¬ 
eration” (Figure 10). 

The basic idea is 
that the time signal of excitation recurs al¬ 
ready after two cylinder operating cycles 
have elapsed if there is alternation be¬ 
tween the active and inactive cylinder. The 
frequency spectrum of excitation is there¬ 
fore determined by a fundamental fre¬ 
quency resulting from the inverse of the 
duration of only two consecutive cylinder 


operating cycles, and their higher har¬ 
monics. The periodic recurrence comes 
after just 2/3 of a camshaft revolution and 
not only after a complete revolution, as 
would be the case with static deactivation 
of a fixed cylinder. 

The fundamental frequency of the exci¬ 
tation function is 3/2, or 1.5 times the cam- 



1 

II 

III 

1 

II 

III 

| 

II 

III 

Engine operating 
on all cylinders R3 

/ 

o 

/ 

/ 

Ca 

/ 

/ 

/ 

/ 

/ 

/ 

Ca 

f R3 = 1/T R3 ~ 1.5 th order 

T R3 


Static cylinder 
deactivation CDA 2/3 

/ / / 
*0a O' *0a 

/ / / 

O Cj 

/ / / 
€03 O *0a 

fcDA = 1 /T cda - 0.5 th order 

T CDA 



Rolling cylinder 
deactivation RCD 1.5 

/ / 

Ca O 

/ / 

Ca O 

/ / 

Ca O 

/ / 
Ca O 

/ 

fpcD = 1 /Trcd - 0.75 th order 

T rcd 



Figure 10 Principle of rolling cylinder deactivation “RCD 1.5” with 1.5 of the three cylinders active 







































34 



Speed in rpm 

- 0.75 th order - 1.5 th order 

- 2.25 th order - 3.0 rd order 


Figure 11 Order analysis for ROD 1.5 operation with a 0.75 th fundamental 
order without centrifugal pendulum-type absorbers 


shaft speed and 
thus the 0.75 th or¬ 
der of the crank¬ 
shaft frequency 
(Figure 11). It is 
plausible that the 
alternating opera¬ 
tion of active and 
inactive cylinders in 
three-cylinder en¬ 
gines results in 
1.5-cylinder opera¬ 
tion, generating a 
0.75 th fundamental 
order for the four- 
stroke cycle princi¬ 
ple. 

The rolling cyl¬ 
inder deactivation 
“RCD 1.5” suggest¬ 
ed here with 1.5 
rolling active cylin¬ 
ders out of three cylinders therefore offers 
the following basic advantages over static 
cylinder deactivation with two fixed active 
cylinders out of three cylinders (CDA 2/3): 

- Fundamental excitation frequency of 
the 0.75 th order instead of the practi¬ 
cally uncontrollable low-frequency 0.5 th 
order, with all excitation frequencies 
50 % higher - the main objective of this 
development; 

- Even higher reduction in fuel consump¬ 
tion due to only 1.5 instead of two ac¬ 
tive cylinders. 

As a result of further tests, it is possible to 
provide the following advantages over static 
cylinder deactivation as well: 

- No oil suction due to a vacuum, since 
each deactivated cylinder is actively 
fired during the next camshaft revolu¬ 
tion, and thus there are no prolonged 
vacuum phases in the cylinder. 

- This also prevents the deactivated cyl¬ 
inder from cooling down, thereby re¬ 
ducing heat-related cylinder distortion 
during deactivation operation. 


- Since no cylinders are deactivated for 
prolonged periods with the RCD 1.5 
concept, fewer warmup measures are 
needed than for the static cylinder de¬ 
activation concept. For this reason, it is 
possible to drive in RCD 1.5 mode even 
directly after a cold start, which leads 
to another improvement in fuel con¬ 
sumption compared to static cylinder 
deactivation. 

Optimizing cylinder charging in 
deactivation operation 

At this point, one might ask how and with 
what charges the deactivated cylinders 
should be operated. With current cylin¬ 
der deactivation systems, fresh air is 
generally locked into the deactivated cyl¬ 
inder, where it is compressed and pas¬ 
sively expanded without combustion. In 
principle, the options of “exhaust gas in 
the cylinder” or “nearly no gas in the cyl¬ 
inder” are also open for discussion. A de- 














Powertrain Systems 


2 


35 


activated cylinder compresses and ex¬ 
pands twice without ignition and 
combustion during one revolution of the 
camshaft, while an active cylinder in four- 
stroke operation only compresses and 
expands once, using the second half of 
the camshaft’s revolution to exchange 
the gas. Excitation therefore originates 
from a deactivated cylinder twice per 
camshaft revolution and only once from 
an active cylinder. 

Consideration of the three options for 
potential cylinder charging leads to the fol¬ 
lowing results for RCD 1.5: 

- Variant 1, leaving the exhaust in the cyl¬ 
inder: 

Here, relatively high working pressures 
occur analogous to the pressure of the 
residual gas, which is unfavorable with 
respect to thermodynamic process and 
friction losses. Moreover, the torsional 
vibration excitation in the 0.75 th order is 
unacceptable due to the high exciting 
cylinder pressures. 


- Variant 2, fresh air in the cylinder: 

The disadvantage here are the losses 
due to working pressures. In addition, 
excitation still partly produces the 0.75 th 
fundamental order here due to the ad¬ 
ditional second “dummy” compression 
in asynchronous phasing relative to the 
omitted ignition. 

- Variant 3, almost no gas in the cylinder: 
After expelling the last combustion gas 
from the previous stroke, the intake and 
exhaust valves remain closed so that 
the piston completes two intake strokes 
against a vacuum, after which compres¬ 
sion occurs with a large portion of the 
compression energy being recuperated. 
The second time that the piston returns 
to TDC, the intake valves are then re¬ 
opened so that the normal intake, com¬ 
pression, ignition, and exhaust opera¬ 
tion is restored. 

Simulations of torsional vibration excita¬ 
tion based on the cylinder pressure curves 
do not indicate the presence of any dis- 


Fired 

cylinders 


= £ 


Non-fired 

cylinders 


I £ 

& T 


0° 360° 720° 1,080° 1,440° 1,800° 2,160° 2,520° 

Angle in °KW 



Intake 

■ Compression 

■ Ignition & expansion 
Exhaust 


RCD downwards (intake) — Cylinder I 

■ RCD upwards (compression) — Cylinder II 

— Cylinder III 


Figure 12 Formation of alternating torques of cylinder deactivation operation in three-cylinder engines 
in the variant with relatively high exhaust gas pressure in the cylinder 

















36 


turbing low-frequency 0.5 th order; instead, 
the lowest occurring order is the 0.75 th , as 
expected. The excitation amplitude is 
smaller than with the first two cylinder 
charge options and basically stems from 
the lack of ignition and to a lesser degree 
from the dummy intake strokes completed 
against a vacuum with subsequent re¬ 
compression. Advantageous here is the 
fact that relatively low pressures are in¬ 
volved, so that the friction losses in the 
deactivated cylinders are small, thereby 
achieving a considerable reduction in fuel 
consumption. Since the deactivated cylin¬ 
der is fired normally on the next camshaft 
revolution, no oil is sucked in despite the 
short vacuum phase. 

Implementing the RCD concept with 
various numbers of cylinders 

The outcome that must be kept firmly in 
mind is that the RCD 1.5 concept in con¬ 
junction with nearly no cylinder charge 
attained the best results with respect to 
both a reduction in fuel consumption as 


well as torsional vibration excitation. In 
essence, 1.5-cylinder operation was re¬ 
alized with a three-cylinder engine. The 
cycles of the individual strokes and the 
RCD strokes contained in them are por¬ 
trayed in Figure 13. 

Using the same principles, a five-cylin¬ 
der engine can effectively be operated as a 
2.5-cylinder engine with RCD 2.5 in cylinder 
deactivation operation. Fundamental excita¬ 
tion then occurs in a 1.25 th order, which can 
be controlled by means of relevant damper 
systems. 

Rolling cylinder deactivation can also 
be implemented in engines with an even 
number of cylinders. For example, de¬ 
pending on the power required, a four- 
cylinder engine can either run as RCD 
1.33 or as RCD 2.66 along with normal 
static deactivation CDA 2/4. A 0.66 th fun¬ 
damental order is produced, however, in 
the first two cases that is hard to control 
due to the fundamental period duration 
according to the sequence of three of the 
four cylinders up to the periodic recur¬ 
rence of the sequence. 


3-cylinder mode 



Angle °KW 


Intake ■ Ignition & expansion ■ RCD downwards (intake) 

■ Compression Exhaust ■ RCD upwards (compression) 

Figure 13 Comparison of the stroke cycles in a three-cylinder engine operating on all cylinders and in 
RCD 1.5 operation 






















Powertrain Systems 


2 


37 


The valve control required for RCD oper¬ 
ation, i.e. the deactivation of intake and 
exhaust valves of each cylinder being de¬ 
activated during a camshaft revolution, 
can be implemented so as to be com¬ 
pletely variable with the Schaeffler UniAir 
system for electro-hydraulic valve actua¬ 
tion [8]. 

As a rule, the intake and exhaust 
valves can be deactivated by means of 
switching mechanisms as well [9]. Op¬ 
tions include switchable tappets, finger 
followers, pivot elements, and - with cer¬ 
tain limitations - even the principle of cam 
shifting. These types of components are 
currently used for valve switching, and 
are capable of switching within parts of a 
camshaft revolution. In order to be used 
with RCD 1.5 and the considerably great¬ 
er number of switching cycles involved, 
further development would be required, 
since switching would have to occur after 
each camshaft revolution. 


Torsional vibration 
damper development for 
RCD 1.5 


The 0.75 th fundamental order occurring in 
RCD 1.5 operation places heavy demands 
on the torsional damper system. Figure 14 
shows a design solution in connection with 
dry double clutches - the result of DMF op¬ 
timizations and a centrifugal pendulum-type 
absorber designed for the 0.75 th order. Due 
to the advantage of the overall length of 
three-cylinder engines as compared to four- 
cylinder engines in identical vehicles, it was 
possible here to choose a design for which 
the arc spring damper and the centrifugal 
pendulum-type absorber masses are both 
arranged axially one behind the other on 
large effective radii. 




Figure 14 DMF design with a centrifugal pendulum-type absorber for the 0.75 th order for RCD 1.5 
rolling cylinder deactivation in three-cylinder engines 














38 



Transmission 



Speed in rpm 

- 0.75 th order 1.5 th order 

— 2.25 th order - 3.0 rd order 


Figure 15 Order analysis of ROD 1.5 operation with a centrifugal 
pendulum-type absorber 


The resulting order 
analysis of the simu¬ 
lations shows how 
the excited 0.75 th or¬ 
der is reduced by 
the matched centrif¬ 
ugal pendulum-type 
absorber to the very 
low amplitudes on 
the transmission in¬ 
put (Figure 15). 

Figure 16 depicts 
the behavioral com¬ 
parison of a three- 
cylinder engine run¬ 
ning operating on all 
cylinders and under 
full load as well as in 
cylinder deactivation 
operation according 
to the RCD 1.5 prin¬ 
ciple at its highest 
operating load, which is set at 70 % of the 
theoretically highest producible half-engine 
torque. It is evident that practically the same 


speed amplitude occurs under such condi¬ 
tions at the transmission input in RCD 1.5 op¬ 
eration as when the engine is operating on all 
cylinders. The 
means for this is the 
centrifugal pendu¬ 
lum-type absorber 
with a total mass of 
approx. 1 kg that has 
been optimally 
matched for the oc¬ 
curring 0.75 th order. 

In addition, a 
centrifugal pendu¬ 
lum-type absorber 
approx. 800 g larger 
was designed for 
manual transmis¬ 
sions for which the 
secondary moment 
of inertia of the mass 
is less than with the 
dry double clutch, 
which has a thermal 
mass that is practi- 


Engine 



Operating range RCD 1.5 



800 1,200 1,600 2,000 2,400 2,800 3,200 3,600 4,000 

- 3-cylinder Speed in rpm 

— 1.5-cylinder 

Figure 16 Comparison of torsional vibrations in the powertrain in a three- 
cylinder engine operating on all cylinders and for rolling cylinder 
deactivation in RCD 1.5 operation with a dry double clutch 































Powertrain Systems 


2 


39 


cally used twice (Fig¬ 
ure 17). 

In this way, the 
goal of implementing 
cylinder deactivation 
operation in three- 
cylinder engines with 
acceptable torsional 
vibration behavior in 
the powertrain was 
achieved, both with 
a dry double clutch 
and for manual 
transmissions. In 
RCD 1.5 operation, 
this can in effect be 
managed with only 
1.5 active cylinders 
to reduce fuel con¬ 
sumption and C0 2 
emissions. 




— 3-cylinder Speed in rpm 

— 1.5-cylinder 

Figure 17 Comparison of torsional vibrations in the powertrain in a three- 
cylinder engine operating on all cylinders and for rolling cylinder 
deactivation in RCD 1.5 operation with a single clutch for manual 
transmissions with a larger centrifugal pendulum-type absorber 


Summary 


This article describes measures for reduc¬ 
ing fuel consumption and C0 2 emissions in 
motor vehicles to the extent that they are 
primarily influenced by transmission sys¬ 
tems: 

- Direct friction reduction in the transmis¬ 
sion through optimized bearing sup¬ 
ports 

- Wet and dry double clutches with re¬ 
duced drag torque 

- Transmission designs with a large 
spread of gear ratios 

- Optimized damper technology for achiev¬ 
ing downsizing and high turbocharging 
pressures, along with downspeeding for 
reducing losses in combustion engines. 

Such drive trends are related to an increase 
in torsional vibration excitation from the in¬ 
ternal combustion engine into the pow¬ 


ertrain. Finally, a new approach is intro¬ 
duced for implementing RCD 1.5 rolling 
cylinder deactivation for three-cylinder en¬ 
gines to attain 1.5-cylinder operation. The 
basic characteristics are: 

- Sophisticated rolling cylinder deactiva¬ 
tion in order to increase the fundamen¬ 
tal frequency of the excitation spectrum 
from the 0.5 th order with static cylinder 
deactivation to the much more control¬ 
lable 0.75 th order with rolling cylinder 
deactivation 

- Optimized cylinder charge setting to re¬ 
duce the excitation amplitude. 

The resulting torsional vibration excitation is 
controlled by the innovative damper technol¬ 
ogy developed by Schaeffler, which entails a 
dual mass flywheel with an optimized curve, 
the use of centrifugal pendulum-type ab¬ 
sorbers on the secondary DMF mass that 
are matched to the occurring 0.75 th main ex¬ 
citation order, and an additional damped 
clutch disk if needed. Similarly, it is possible 
















40 


to implement RCD 2.5 operation, which is 
advantageous for five-cylinder engines. 

This approach can be implemented for 
applications with manual transmissions (MT), 
automated manual transmissions (AMT), dou¬ 
ble clutch transmissions (DCT) with a dry or 
wet double clutch, and also for planetary au¬ 
tomatic transmissions or CVTs with convert¬ 
ers that have dampers equipped with added 
centrifugal pendulum-type absorbers. 


Literature 


[1] Faust, H.: Requirements for Transmission 
Benchmarking. (FWD Automatic Transmis¬ 
sions). GETRAG Drivetrain Forum, Symposium, 
2012 

[2] Nowatschin, K.; Fleischmann, H.-P.; Gleich, T.; 
Franzen, P.; Ftommes, G.; Faust, H.; Fried¬ 
mann, O.; Wild, H.: multitronic - The New 
Automatic Transmission from Audi. Part 1: ATZ 
worldwide 102, 2000, no. 7/8, pp. 25-27. Part 2: 
ATZ worldwide 102, 2000, no 9, pp. 29-31 


[3] Teubert, A.: CVT - The Transmission Concept of 
the Future. 10 th Schaeffler Symposium, 2014 

[4] Lindemann, R: iTC - Innovative Solutions for 
the Converter Pave the Way into the Future. 

10 th Schaeffler Symposium, 2014 

[5] Faust, H.; Steinberg, I.: Die neuen GETRAG 
PowerShift-Getriebe 6DCT450 & 6DCT470. 
VDI-Berichte no. 2029. Dusseldorf: VDI-Verlag 
2008, pp. 69-90. 2008 VDI Conference on 
Vehicle Transmissions, 2008 

[6] Faust, H.; Ruhle, G.; Herdle, L.: Optimization of 
Driving Fun and Reduction of C0 2 Emissions 
with the New GETRAG PowerShift Trans¬ 
missions. 3 rd International CTI Symposium 
Automotive Transmissions. Detroit/USA, 2009, 
Session A2, pp. 1-20 

[7] Faust, H.; Bunder, C.; DeVincent, E.: Dual 
Clutch Transmission with Dry Clutch and 
Electro-mechanical Actuation. ATZ worldwide 
112, 2010, no. 4 

[8] Scheidt, M.: Pure Efficiency. Developing 
Combustion Engines from the Perspective of a 
Supplier. 10 th Schaeffler Symposium, 2014 

[9] Ihlemann, A.: Cylinder Deactivation. Something 
of a Niche or a Technology for the Future? 

10 th Schaeffler Symposium, 2014 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



42 


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43 


Pure Efficiency 

Developing combustion engines 
from the perspective of a supplier 


Dr. Martin Scheidt 
Matthias Lang 


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44 


Efficiency as the primary 
development objective 


At the end of 2013, the European Union 
agreed new C0 2 limits. As of 2020, these 
specify fleet emission values of 95 grams 
of C0 2 . This figure corresponds to a con¬ 
sumption of approx. 3.6 1/100 km for die¬ 
sel vehicles and 4.1 1/100 km for gasoline- 
operated vehicles. These limits will be the 
most stringent in place anywhere in the 
world. It is expected that it will only be 
possible for premium vehicle manufactur¬ 
ers (with a virtually identical mix of vehi¬ 
cles) to achieve this limit value by partially 
electrifying large and heavy vehicles. The 
plug-in hybrid drive is set to play a signifi¬ 
cant role in electrification, as it is favored 
by applicable legislation. 

Despite increasing electrification, engi¬ 
neers across the entire automotive industry 
will focus on optimizing the combustion en¬ 
gine for many years to come and for a num¬ 
ber of reasons. The most important reason 
is the tremendous growth trajectory that the 
global automotive 
industry can expect 
over the coming 
years. Increasing 
prosperity means 
that the number of 
newly registered 
passenger cars and 
light commercial 
vehicles will grow to 
around 105 million 
units by 2020, 
which corresponds 
to a growth of 40 % 
compared to 2012 
[1]. Emerging econ¬ 
omies, as well as 
newly industrialized 
countries such as 


Brazil, Russia, China and India, will see the 
majority of this growth. However, there are 
many first-time car buyers in these coun¬ 
tries who cannot afford the costs associat¬ 
ed with drive electrification. Therefore in 
these kinds of markets, automotive manu¬ 
facturers that use efficient combustion en¬ 
gines to shift electrical drive components 
into heavy vehicles as far as is possible will 
be especially successful. 

The second key reason, this time for 
the developed markets such as in Europe 
and the United States, is the expectation 
of car buyers for standard consumption 
figures to be approximately achieved in 
real life. For plug-in hybrids, this is partic¬ 
ularly the case if the distances travelled 
far exceed the electric range and the ve¬ 
hicle must bear the additional weight of 
the electrical drive components and the 
battery. For instance, Volkswagen has an¬ 
nounced that the plug-in version of the 
Golf to be introduced in 2014 will be 250 kg 
heavier than its comparable gasoline en¬ 
gine version. An efficient combustion en¬ 
gine with a high weight-to-power ratio can 
help to fulfill the expectations of end cus¬ 
tomers in this regard. 

4 


E 

XL 

O) 

c 

6 

o 


160 


140- 



EU Limit Value Curve 



Figure 1 


C0 2 fleet consumption of vehicles sold in the EU 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_3, © The Author(s) 2014 







Developing Combustion Engines 


3 


45 



Figure 2 Typical power losses on the efficiency chain from tank to wheel 


Finally, it should be noted that the Europe¬ 
an C0 2 limits are particularly strict, but 
international legislation aiming for similar 
values has experienced some setbacks 
(Figure 2 in [2]). From the point of view of a 
European supplier, there is the possibility 
of bringing consumption-reducing tech¬ 
nologies onto the domestic market at an 
early stage and thus gaining a competitive 
advantage on a global scale. 

Both EU limit values of 130 g C0 2 /km 
for 2015 or 95 g C0 2 /km relate to a vehicle 
weight of 1372 kilograms; limit values for 
vehicles of different weights are calculat¬ 
ed using the straight-line method and a 
weighting factor. As Figure 1 shows, no 
manufacturer in the EU currently meets 
the limit value for 2015; the limit values are 
currently only met by segments of some 
manufacturers’ fleets. In addition, the 
weighting factor is reduced for the 2020 
target, which represents a huge disadvan¬ 
tage for manufacturers of heavier vehi¬ 
cles. On the whole, it is apparent that all 
manufacturers will need to put great effort 
into boosting the efficiency of their vehicle 
fleets. 


Approach to improving efficiency 

The efficiency of combustion engines can 
only be increased if the actual engine comes 
as close as technically possible to the at¬ 
tainable thermodynamic optimum. There¬ 
fore, the engineers’ first priority must be to 
focus on losses that occur in actual en¬ 
gines. Losses for a typical cycle, in which 
the vehicle’s aerodynamics are not taken 
into account, can be seen in Figure 2. 

In modern powertrain concepts, the en¬ 
gine’s special operating states also play a 
vital role when it comes to standard and ac¬ 
tual fuel consumption. This applies to tran¬ 
sient states, such as increased accelera¬ 
tion. During acceleration, engines with high 
weight-to-power ratios and low displace¬ 
ment exhibit relatively high deviations from 
their optimum operating point. One of the 
reasons for this is “full-load enrichment”, 
which often needs to be applied at relatively 
low speeds to avoid knocking combustion 
and protect exhaust components from ex¬ 
cessive temperatures. Other operating 
states that should be considered in particu¬ 
lar include engine warm-up after a cold 





46 


start, and more and more frequently situa¬ 
tions in which an engine is partially shut 
down (cylinder deactivation) or even com¬ 
pletely stopped (start-stop/coasting). 

Furthermore, the efficiency of modern 
engines must not be detrimental to the en¬ 
gine-out emissions of exhaust pollutants 
nor should it result in reduced comfort for 
end customers. 

Potential for efficiency 
enhancement 


There are two ways of minimizing the losses 
that occur within combustion engines, and 
they must be initiated simultaneously: first, 
increase the actual combustion efficiency 
and, second, minimize losses, especially 
friction and pumping losses. 




Reducing pumping losses 

Pumping losses depend heavily on how 
much the engine must be throttled at a spe¬ 
cific operating point. Or put another way: 
how often can operating points with low 
throttling, i.e. high load at low speed, actu¬ 
ally be achieved through the transmission 
curves. 

The combination of direct injection 
and exhaust-gas turbo charging to en¬ 
able this kind of operation has become 
established on the market. It results in 
high specific output that can be used for 
reducing eingine displacement (downsiz¬ 
ing). Engines of this type tend to be oper¬ 
ated more frequently at dethrottled map 
points. Cylinder deactivation has a simi¬ 
lar effect, also resulting in a higher indi¬ 
cated mean effective pressure in the cyl¬ 
inders still running — and thus resulting 
in dethrottling. 

Extensive dethrottling can be achieved 
by closing the intake valves early (EIVC) or 




Figure 3 Dethrottling by changing intake valve timing 
















Developing Combustion Engines 


3 


47 



Hydraulic 

Electric 


Two-Step Three-Step 

Electric Shifting Cam 

Tappet 

Pivot Element 
Finger Follower 
Shifting Cam 
Roller Lifter 


Electro-magnetic 

Mechanical 
e.g. Valvetronic 

Electro-hydraulic 

UniAir 



Figure 4 Different types of variability in the valve train 


late (LIVC). Both methods reduce the effec¬ 
tive compression ratio and are also known 
as Miller or Atkinson cycles (Figure 3). With 
the valve opening times thus modified, four- 
stroke engines experience lower pumping 
losses but suffer the challenge of reduced 
combustibility. This effect can be counter¬ 
acted by increasing the charge motion in 
the combustion chamber, thereby enabling 
improved mixture formation and more effi¬ 
cient combustion. 

Ideally, to achieve complete dethrottling, 
it would be possible to freely select the 
opening and closing times as well as the 
valve lift for all operating states. 

Today camshaft phasers, which only 
allow partial dethrottling, have already be¬ 
come established on the market. Ele¬ 
ments in the valve train for deactivating 
cylinders continue to be used. The first 
applications of mechanical and electro- 
hydraulic fully variable valve trains are now 
available (Figure 4). 


Camshaft phase control 

Camshaft phasers are manufactured in 
large quantities. Hydraulic systems have 
taken hold, and electromechanical sys¬ 
tems are being developed at the same 
time. The latter provide optimum adjusting 
speed and variability (Figure 5). However, 
electromechanical systems are also more 
costly. With this in mind, Schaeffler is not 
only working towards the start of produc¬ 
tion for electromechanical cam phasing 
systems, planned for 2015; we are also 
continuously optimizing the performance 
of hydraulic systems. 

The adjusting speed of hydraulic cam¬ 
shaft phasing units is largely determined 
by the performance of the oil circuit. En¬ 
gine oil pressure has been consistently 
lowered over the past few years to reduce 
the power consumption of engine oil 
pumps. Low oil pressure is a challenging 
constraint when it comes to designing new 
and developing existing camshaft phasers. 
This is because the lower the oil pressure, 
the less energy is available to adjust the 
camshaft. 









48 


Hydraulic VCP 
with cartridge valve 



1,000 


2,000 3,000 

Engine speed in rpm 


4,000 


Figure 5 Adjusting speed of hydraulic and electrical camshaft phasers 


Schaeffler is therefore showcasing a phaser 
with a secondary oil reservoir for the first 
time: the additional oil reservoir is located in 
additional bores in the camshaft phaser ro¬ 
tor — in other words, right next to the oil 
chambers that trigger phasing when they 
are filled. This tank is not pressurized, it im¬ 
proves adjustment speed by providing vol¬ 
ume that does not have to be supplied by 
the oil pump [3]. 


Switching elements 

Another way to increasing valve train vari¬ 
ability is provided by switching elements 
that vary the lift of individual valves. These 


kinds of systems are aimed, in particular, 
at cylinder deactivation, partial dethrottling 
and internal exhaust gas recirculation, 
and are available in a range of designs (Fig¬ 
ure 6): 

- The simplest example merely involves 
shutting down individual valves — and 
therefore also the cylinders — via a 
switched pivot element; these types of 
elements have been used successfully 
on the market for several years. 

- Switchable finger followers or bucket tap¬ 
pets are also used for two-stage lift switch¬ 
ing and therefore for partial dethrottling. By 
using a sliding cam system, it is even pos¬ 
sible to vary the valve stroke in three 
stages. 3 step cam shifting systems either 












Developing Combustion Engines 


3 


49 


combine cylinder deactivation with switch¬ 
ing between two discrete strokes, or allow 
switching between three strokes. Schaef¬ 
fler is developing a mechanical solution for 
3 step switching, designed to be robust 
enough to meet all standard requirements 
regarding valve train service life. 

- Using a switchable finger follower, a 
second valve stroke can be performed 
outside of the specified first stroke con¬ 
tour. This enables internal exhaust gas 
recirculation to be performed by either 
pushing the exhaust gas back into the 
intake manifold or by re-breathing ex¬ 
haust gas by opening the exhaust valve 
a second time during the intake phase. 
Schaeffler has adapted a system of this 
kind for a Japanese diesel engine. 


Fully variable valve train 

Electromechanical or electrohydraulic fully 
variable valve train systems offer a high de¬ 
gree of variability, the latter are already in 


volume production. The electrohydraulic 
systems are still driven by the camshaft. 
Electromagnetic systems without a cam¬ 
shaft have been the subject of research for 
some time, but they have yet to be intro¬ 
duced which is not only attributable to the 
demanding electrical power requirements. 
The camshaft also acts as a safety element, 
preventing faulty actuations and thus the 
valve and piston coming into contact. 

In 2009, Schaeffler started volume pro¬ 
duction of the UniAir electrohydraulic valve 
train system. This Schaeffler system in¬ 
cludes: 

- The electrohydraulic actuator module 

- The software required to control valve 
timing; this software is integrated in the 
customer’s engine control unit 

- A calibration data set for the relevant 
application 

Since 2009, this system has been adapted 
for various production engines with capaci¬ 
ties between 0.9 and 2.4 I, and delivered in 
high volumes to customers in Europe and 
North and South America. 


Switchable 
pivot element 


Switchable 

tappet 


Electro-Hydraulic Actuated 

Electro-Mechanical Actuated 
(Enlarged Temperature Range) 




© 





Cam shifting 
system 






Valve Deactivation 
(1 Valve per Cyl.) 

M. 



s 

s 

Cylinder Deactivation 
(All Valves per Cyl.) 

M. 



s 

s 

Internal EGR 
(Retain) 

M. 


s 



Internal EGR 
(Re-breath) 

Ad 





Crossing of 

Valve Events 




2-Step 

Nk 





3-Step 

NL 




Figure 6 Switching systems for varying valve lift 

























50 



Figure 7 Valve lift curves in different engine map ranges 


UniAir not only enables continuously vari¬ 
able setting of the valve lift; it also enables 
largely free configuration of the valve lift 
event within the maximum contour speci¬ 
fied by the camshaft envelope. In this way, 
dethrottling is possible within broad en¬ 
gine map ranges (Figure 7). It results in a 
fuel consumption reduction of up to 15 % 
in the New European Driving Cycle 
(NEDC). 

Future generations of the UniAir system 
will feature new functions which will be pre- 



Crankshaft in ° 

Figure 8 Possible lift curve with individual 
valve control 


sented in more detail at the Schaeffler Sym¬ 
posium 2014 [4]. One noteworthy function is 
individual control of two intake valves. This 
kind of activation enables a specific charge 
motion to be generated (especially at low 
loads), thereby significantly increasing com¬ 
bustion efficiency. Figure 8 shows asym¬ 
metric valve lift curves, as enabled by indi¬ 
vidual control. 

From Schaeffler’s standpoint, the free¬ 
dom in combustion process design af¬ 
forded by the UniAir system can be ap¬ 
plied to all vehicle segments. Low-cost 
engines with a small number of cylinders 
can benefit from increased torque, while 
simultaneously lowering specific fuel con¬ 
sumption. In this vehicle segment the 
cost/benefit ratio is far superior to other 
measures, such as adding exhaust gas 
turbo charging and direct injection. Large 
engines benefit especially from dethrot¬ 
tling in the part load range. New functions 
can also support future combustion pro¬ 
cesses that can exploit the benefits of the 
system’s extremely fast actuating mecha¬ 
nism. 

























































Developing Combustion Engines 


3 


51 


Reducing friction losses 

Reducing friction lossesr has always been a 
crucial development objective in engine de¬ 
sign. In the past, focus was placed on inter¬ 
nal friction in the cylinder, particularly friction 
between the piston/piston ring cylinder pair¬ 
ing. On account of increasingly stringent 
C0 2 legislation, all other sources of loss are 
now also being studied. This applies in par¬ 
ticular to 

- Crankshaft 

- Valve train 

- Balancer shafts 

- Camshaft and auxiliary equipment 
drives 

- Losses caused by operating the cool¬ 
ant and the oil pump 

In total, these friction values account for 
about 50 % of the friction losses of an aver¬ 
age combustion engine (Figure 9). In addi¬ 
tion, the engine heat-up process becomes 
more important due to the relationship be¬ 
tween friction and oil temperature. This 
power loss has a direct impact on standard 


100 
90 
80 
^ 70 
■E 60 
.2 50 


30 
20 
10 
0 

1,000 2,000 3,000 4,000 5,000 

Engine speed in rpm 

Figure 9 Typical power loss values of 

individual causes of friction over 
engine speed for a petrol engine 

consumption due to the cold start section in 
the New European Driving Cycle (NEDC). 

The valve train is responsible for a par¬ 
ticularly high proportion of friction losses 
that occurs at low engine speeds. Over the 
past 20 years, great progress has been 
made in this area by means of tribological 



Piston ring 


Piston 



Connecting rod 


Crankshaft bearing 


Tappet 

with hydraulic valve lash adjustment 



Figure 10 Comparison of frictional power for different valve train types 




















52 


optimization of bucket tappets; the friction 
mean effective pressure has been reduced 
by around 50 % (Figure 10). At the same 
time, roller finger followers for valve control 
have become established — they link hy¬ 
draulic valve clearance compensators with 
inherently low friction. 

It is increasingly common for modern 
engines — both gasoline and diesel — with 
high specific power ratings and few cylin¬ 
ders to be fitted with balance shafts. The 
friction on the shaft bearing is particularly 
relevant due to its high speed (double 
crankshaft speed in a four-cylinder engine). 
Switching to a roller bearing arrangement 
while simultaneously designing lighter com¬ 
ponents (Figure 5 in [2]) can decrease a ve¬ 
hicle’s C0 2 emissions by up to 2 %. In a 
four-cylinder engine, this kind of solution can 
reduce the weight by approximately 0.5 kg 
per shaft/1 kg per system. 

Significantly lower friction losses can 
also be achieved by supporting the cam¬ 
shaft on roller bearings (Figure 11). However, 
if this approach is taken, it is essential to 
consider the assembly concept for the cyl¬ 
inder head. 

The key goal for auxiliary equipment drives 
is ensuring seamless functionality over the 
service life. Transferring ever-increasing 



— Plain bearing 60 °C 
— Plain bearing 100 °C 
— Rolling bearing 60 °C 
— Rolling bearing 100 °C 

Figure 11 Drive torques for camshafts with 
plain bearings and roller bearings 


torques and power ratings results in higher 
preloads in the belt drive, resulting in in¬ 
creased power loss. At the same time, dy¬ 
namic amplitudes in the belt drive are increas¬ 
ing as engines have fewer cylinders, and 
higher mean effective pressures; this results in 
high rotational irregularities. Innovative belt 
tensioners and crankshaft decouplers devel¬ 
oped by Schaeffler are able to transfer the in¬ 
creased torque reliably while simultaneously 
minimizing any power loss [5]. 

More dynamics, fewer 
losses - special operating 
states 


Optimizing steady state engine map points 
alone is not an effective way of improving 
the overall combustion engine. On the one 
hand, future consumption test cycles will 
have higher dynamic content; on the other 
hand, hybrid systems, in which there is no 
clear correlation between driving conditions 
and engine operating points, are being used 
more commonly. 


Acceleration 

The dynamic response characteristic of en¬ 
gines with a high degree of supercharging 
can be specifically enhanced by setting a 
positive scavenging gradient. Extremely 
rapid actuation of the camshaft phaser is 
desirable to quickly start adjusting the valve 
timing as required . 

Electromechanical phasers allow ex¬ 
tremely high adjusting speeds of more than 
250 °KW/s [6]. They also provide greater ri¬ 
gidity when torque is applied between the 
drive wheel and camshaft, thereby achiev¬ 
ing optimum adjusting accuracy. 






Developing Combustion Engines 


3 


53 


In addition, electric cam phasing is the only 
option that allows valve timing to be select¬ 
ed as required when starting the engine. By 
selecting the valve timing, the engine can be 
started with minimal compression, which 
results in a low-vibration start and requires 
considerably less starter power. Electrome¬ 
chanical phasers are largely unaffected by 
temperature, while hydraulically actuated 
systems only provide useful adjusting 
speeds at ambient temperatures of +7 °C to 
+20 °C, depending on the design. 

However, this high performance level 
goes hand in hand with increased cost. 
Schaeffler will put this kind of system into 
volume production for the first time in 2015. 
The crank angle adjustment range will be up 
to 95° in this new system. It is designed to fit 
to the series engine cylinder head with only 
small changes. 

Furthermore, it is of course important to 
bring the turbo charger up to maximum 
speed as quickly as possible when acceler¬ 
ating under a full load. Two-stage turbo 
charger systems are increasingly being 
used for this purpose. In these systems, the 
first supercharger is relatively small and has 
correspondingly low inertia. The use of roll¬ 
ing bearings for turbo chargers results in 
significantly lower frictional losses [7] and 
thus shorter response times. The reduction 
is so great that the charger could be made 
larger and yet retain the same response 
characteristics. For certain engine power 
ratings, a second turbocharger is therefore 
no longer required and considerable cost 
can be avoided. 


Engine warm up 

The high thermodynamic efficiency of mod¬ 
ern engines also has its disadvantages: sig¬ 
nificantly less waste heat is produced, which 
is however needed to heat the engine, 
transmission and, depending on weather 
conditions, the vehicle interior. At the same 


time, the test cycles for determining C0 2 
and exhaust emissions demand a cold 
start. To distribute the initial heat produced 
in an optimum manner, regarding passen¬ 
ger comfort and emissions, Schaeffler has 
introduced a thermomanagement module 
(Figure 12). 

In the engine warm up phase, the mod¬ 
ule can completely shut off the coolant en¬ 
tering the engine or set a minimum volume 
flow. When the engine is at operating tem¬ 
perature, the coolant temperature can be 
regulated quickly to various temperature 
levels, depending on load requirements and 
external conditions. The component has 
two coupled rotary slide valves that use a 
single drive. The first volume engine 
equipped with a Schaeffler multifunctional 
cooling water controller is the 1.8-1 TFSI 
engine manufactured by Audi (four-cylinder 
in-line engine, third generation). The module 
heats up the coolant at a rate that is up to 
30 % faster than the predecessor engine 
which has a wax-type thermostat. In fact, 
the time required to achieve target oil tem¬ 
perature is reduced by 50 % [8]. 



Figure 12 Structure of the Schaeffler 

thermomanagement module with 
integrated water pump 


54 


Compact designs for smaller engines and 
vehicles and further development of func¬ 
tional integration are the focus for future ap¬ 
plications [8]. Development includes a multi¬ 
functional module with separate circuits for 
the engine block and cylinder head (split 
cooling). Schaeffler estimates it is possible 
to save up to 4 g of C0 2 per kilometer with 
skilful application of a thermomanagement 
module. A controllable water pump is a par¬ 
ticularly good solution for commercial vehi¬ 
cles whose cooling systems are designed 
for hill climbs, and thus allow power reduc¬ 
tion when driving on level ground. 


Engine switch off 

Naturally, an engine has the lowest fuel con¬ 
sumption when it is not in operation, which 
is why modern vehicles increasingly switch 
off the engine not only during test cycles but 
also in real traffic situations. The expecta¬ 
tion for 2016 is that two thirds of all new ve¬ 
hicles sold in Europe will feature start/stop 
systems; from 2019, they will be standard 
for conventionally powered vehicles in most 
segments [9]. NEDC consumption can be 
reduced by up to 4.5 %. In the future cycle 
“Worldwide Harmonized Light Vehicles Test 
Procedure” (WLTP), the percentage of en¬ 
gine downtime decreases from 23 % to ap¬ 
prox. 13 %. This means that using start/ 
stop does not achieve the same level of re¬ 
duction. However, the WLTP is more dy¬ 
namic overall, so that vehicle coasting func¬ 
tions gain in importance. 

At its simplest, coasting - or better the 
restart of the engine at the end of the coast¬ 
ing phase -can be achieved using a belt- 
driven starter generator. The development 
target is to be able to switch from one oper¬ 
ating state to the other with the change be¬ 
ing barely perceptible or even imperceptible 
for the driver. However, compared to con¬ 
ventional belt drives, high torque spikes oc¬ 
cur that make new belt tensioner concepts 



Figure 13 Schaeffler decoupling tensioner 

necessary. The wide variety of possible 
concepts range from using twin mechanical 
tensioners to a hydraulically actuated ten¬ 
sioner. Schaeffler’s preferred option is a de¬ 
coupling tensioner installed on the genera¬ 
tor (Figure 13). 

The function of this new tensioner is ex¬ 
plained in detail in [5]. 


Outlook 


Using the technologies outlined in this arti¬ 
cle, the efficiency of today's already very 
economical combustion engines can be 
significantly improved. Schaeffler estimates 
the entire remaining potential for increasing 
efficiency of current volume engines to be 
20 % for petrol engines and 10 % for diesel 
engines. However, parts of this potential 
have already been implemented in engines 
now appearing on the market. 

Furthermore, consistent development of 
the combustion engine will yield additional 
potential, even if existing ideas cannot yet 



Developing Combustion Engines 


3 


55 


be covered by technology which is ready for 
volume production: 

- Complete omission of full-load enrich¬ 
ment even on gas-operated engines with 
a specific output of 100 kW/l or more. In 
addition to fuel savings, this would also 
reduce engine-out particle emissions 

- Replacing plain bearings with roller 
bearings in the crankshaft drive. The 
fundamental technical feasibility of 
this application has already been 
proven, even if an acoustically satis¬ 
factory solution has yet to be found. 
Studies conducted by Schaeffler 
have identified potential C0 2 savings 
of around 3 % 

- Cylinder deactivation that shuts down 
each cylinder in turn instead of always 
shutting down the same cylinder [10]; 
this prevents individual cylinders from 
cooling down 

Schaeffler’s viewpoint is that engine and 
transmission development must be even 
more closely coordinated in future to fully 
exploit this potential. After all, efficient 
drives will only be a success on the mar¬ 
ket if they meet customer expectations for 
acoustics and vibration. In turn, the de¬ 
gree by which overall C0 2 emissions from 
road transport can be reduced depends 
solely on how quickly efficient drives be¬ 
come the norm. The developers of com¬ 
bustion engines and transmissions must 
overcome this major challenge by working 
together. 


Literature 


[1] Gutzmer, P.; Hosenfeldt, T.: Marktgerechte Reibungs- 
optimierung. ATZ 115, 2013, no. 11, pp. 876ff. 

[2] Gutzmer, P.: Individuality and versatility: future mo¬ 
bility paradigms. 10 th Schaeffler Symposium, 2014 

[3] Dietz, J.; Busse, M.; Raecklebe, S.: Smart Phas¬ 
ing: Needs-based concepts for camshaft phas¬ 
ing systems. 10 th Schaeffler Symposium, 2014 

[4] Haas, M.; Piecyk, T.: Get Ready for the Com¬ 
bustion Strategies of Tomorrow. 10 th Schaeffler 
Symposium, 2014 

[5] Stuffer, A., et al.: Introduction of 48V Belt Drive 
Systems: New tensioner and decoupler solu¬ 
tions for belt-driven mild hybrid systems. 

10 th Schaeffler Symposium, 2014 

[6] StrauB, A., et. al: Quo vadis, hydraulic camshaft 
phasing unit? 9 th Schaeffler Symposium, 2010 

[7] Kropp, M., et al.: Optimising the transient be¬ 
haviour of turbo chargers, using the example of 
a PC diesel engine. 13 th International Stuttgart 
Symposium, 2013 

[8] Weiss, M.: Hot and Cold: Schaeffler thermo¬ 
management for up to 4 % C0 2 reduction. 

10 th Schaeffler Symposium, 2014 

[9] Kirchner, E.; Eckl, T.: The Long Path from Dis¬ 
comfort to Customer Acceptance - Start-stop 
systems yesterday, today and tomorrow. 

10 th Schaeffler Symposium, 2014 

[10] Faust, H.: Powertrain Systems of the Future: 
Engine, transmission and damper systems for 
downspeeding, downsizing, and cylinder deac¬ 
tivation. 10 th Schaeffler Symposium, 2014 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



56 


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57 


Mission C0 2 Reduction 

The future of the 
manual transmission 


Jurgen Kroll 
Markus Hausner 
Roland Seebacher 



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58 


Introduction 


The internal combustion engine will continue 
to be the dominating force behind individual 
mobility for some time to come. The biggest 
challenge in this context, however, revolves 
around lowering fuel consumption in line with 
ever more stringent legal requirements while 
at the same time maintaining driving comfort 
and pleasure. All aspects of the engine and 
transmission must be revisited with equal at¬ 
tention, whereby driving strategies that mini¬ 
mize consumption are key to achieving des¬ 
ignated performance targets. To improve on 
these aspects, the transmission must be 
further automated and coupled with electrifi¬ 
cation measures. The conventional manual 
transmission is therefore coming under pres¬ 
sure and runs the risk of being „overrun“ by 
other designs at least in the developed mar¬ 
kets. On the other hand, manual transmis¬ 
sions remain attractive for cost reasons and 
may continue to play a key role in the future if 
a way is found to develop systems that also 
enable „sailing“ and other efficient drive 
modes to be achieved in vehicles equipped 
with a standard transmission. 

Adopting a partially automated setup for 
the manual transmission would also open 
the door to integrating comfort, conve¬ 
nience, and safety-oriented functions with¬ 
out additional cost. Fuel consumption could 
then be further reduced by opting for longer 
gear ratios, for example. Misuse, or abuse 
of the clutch, causing it to overheat, can be 
reliably prevented thanks to the partially au¬ 
tomated setup. 

The end result - “extreme” downspeed¬ 
ing - has disadvantages, however, especially 
when it comes to future engines, where few 
cylinders and/or feature cylinder deactivation 
will be widely used. In order to realize the 
comfort and convenience expected by end 
customers, ever better systems for isolating, 
or dampening, vibrations, must be devel¬ 


oped. Although the centrifugal pendulum ab¬ 
sorber (CPA) developed by LuK also offers 
good potential for the coming years, in the 
long term, even more capable systems will 
need to be integrated. 

Initial situation - 
Manual transmissions 
under pressure 


In addition to the effort expended to further 
reduce the consumption of the engine itself, 
equal focus must be placed on developing a 
transmission that optimizes the efficiency of 
the entire powertrain. The manual transmis¬ 
sion is initially positioned quite well in this re¬ 
gard, since it offers a high level of operating 
efficiency. Additional, conventional improve¬ 
ment measures, such as reducing frictional 
loss and increasing the number of gears and 
gear ratio spread, are limited in their potential, 
however. The transmission can therefore only 
play a much more effective role if it enables 
the internal combustion engine to operate un¬ 
der conditions that allow it to burn as little fuel 
as possible. In terms of today’s engines, this 
translates to low operating speeds or deacti¬ 
vation of the engine as soon as the driver’s 
power requirement makes this possible. It 
goes without saying that a manual transmis¬ 
sion does not offer the ideal setup for tapping 
this potential and is the reason why it is receiv¬ 
ing more and is increasingly under pressure. 
Apart from visual shift point recommendations, 
it is not possible to implement any other, more 
sophisticated, fuel-saving shift strategies. In ad¬ 
dition, hybrid and advanced start/stop functions 
require a specific, baseline level of automation. 

Viewed from this perspective, automation 
is no longer only driven by the needs and 
wants of buyers looking for greater comfort 
and convenience, but is absolutely necessary 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3 4, © The Author(s) 2014 



Manual Gearbox 


4 


59 



■ MT DCT ■ AMT 

■ AT ■ CVT 


Figure 1 Global vehicle production based on transmission technologies (source: CSM, Aug. 2013) 


in several vehicle categories in order to com¬ 
ply with tomorrow’s C0 2 limits and avoid ex¬ 
pensive penalty payments. Vehicles currently 
permitted to expel 135 g/km will only be al¬ 
lowed to produce in Europe 130 g/km in 2015, 
and in 2020, this limit will drop to 95 g/km. 

Against this backdrop, the manual 
transmission isn’t out of the game yet, as 
you might think, since current estimates 
point in the opposite direction. The manual 
transmission still enjoys the highest share of 
the market, especially in the entry vehicle 
segments in the BRIC nations and in Eu¬ 
rope (Figure 1). 

If this predominant market position is to be 
maintained in the future, the manual transmis¬ 
sion will have to be upgraded. While emphasis 
needs to be placed on exploiting the potential 
available for reducing fuel consumption, as¬ 
pects pertaining to convenience and comfort, 
such as launch or stop-and-go assist manag¬ 
ing traffic jams, cannot be overlooked. 


New opportunities for the 
manual transmission 


Analyzing or assessing potential areas in which 
consumption can be reduced is best facilitated 
by conducting tests in line with the established 
driving cycles to pinpoint in which phases cer¬ 
tain measures can offer beneficial results. The 
stop rate of 20 % in the New European Driving 
Cycle (NEDC), for instance, led to the wide¬ 
spread implementation ofstart/stop systems in 
Europe, which can reduce overall fuel con¬ 
sumption in the range of 5 %. The logical en¬ 
hancement of this technology is to switch the 
engine off during normal driving, which in turn 
means that it has to be mechanically decou¬ 
pled from the rest of the powertrain. This is 
what is known as “sailing” and theoretically is 
always a practical mode to be in when vehicle 
deceleration forces lie between those of driving 














60 



■ Basis ■ Sailing (ICE on) Sailing G 4/5/6 (ICE off) 
Start-stop Sailing (ICE off) ICE = Internal combustion engine 


Figure 2 Consumption benefits of start/stop systems and sailing across 
different driving cycles 


resistance and engine braking torque. Current 
NEDC do not incorporate these phases, which 
is why the sailing function does not bring about 
any concrete benefits when comparing posted 
fuel economy numbers. This will not be the 
case when the WTLP (Worldwide Harmonized 
Light Duty Test Procedure) takes effect, how¬ 
ever. Internally conducted consumption simu¬ 
lations with a 2.0-liter diesel engine (Figure 2) 
show that a reduction in fuel consumption of 
more than 6 % is possible when a sailing strat¬ 
egy is incorporated. Even when sailing is only 
used in higher gears (4/5/6), it is possible to re¬ 
duce consumption by approximately 4 %. This 
is counteracted by the decreased benefits of 
modern start/stop systems under WLTP con¬ 
ditions, however, which perform more than 
50 % worse due to the lower stop rate. 

The sailing function currently can only be 


combined with an 
automatic transmis¬ 
sion and has already 
reached volume pro¬ 
duction for several 
models. The trans¬ 
mission itself does 
not need to be fully 
automated, howev¬ 
er, and an automat¬ 
ed clutch to discon¬ 
nect the engine from 
the transmission 
theoretically could 
be sufficient enough. 
Unlike vehicles with 
an automatic trans¬ 
mission, their manually shifted counterparts 
are required to hold a certain gear in a defined 
speed range under cycled testing. If an engine 
is to also operate efficiently at low speeds, the 
gear ratios provided must be adapted accord¬ 
ingly. The potential here should not be under¬ 
estimated, since a 10 % drop in engine speed 
reduces consumption by 7 % when traveling at 
a constant 70 km/h in fifth gear (example simu¬ 
lation with a 2.0-liter diesel engine); under 
NEDC and WLTP conditions, approximately 
5.6 % and 2.5 % less fuel is consumed, re¬ 
spectively. Start-off performance would suffer 
somewhat, however, as comfort levels de¬ 
crease and clutch wear increases. An auto¬ 
mated clutch could provide the answer here, 
too, however, by resolving this inherent conflict. 
The higher operative requirements could be 
compensated for with automated or assisted 



Figure 3 Motivation for clutch automation 















Manual Gearbox 


4 


61 


launch procedures, for example, and additional 
safety and reassurance could be provided by 
incorporating a strategy that prevents exces¬ 
sive heat to the clutch. 

Combining sailing with lower engine speeds 
can theoretically reduce consumption by 5 to 
10 %, depending on the driving cycle. Integrat¬ 
ing an automated clutch assembly would open 
up even more possibilities (Figure 3).The higher 
level of automation associated with this is per¬ 
fect for setting the stage to transition to a hybrid¬ 
ized manual transmission. Coupled with an ad¬ 
ditional electric drive, such as an electric 48-volt 
driven axle, it also would be possible to offer 
functions like electric launch and creeping in a 
special stop-and-go mode. Driving at constant 
speeds could likewise take place without the 
assistance of the internal combustion engine 
(electric sailing), and during braking, the effec¬ 
tiveness of an energy recovery system could be 
increased by the drag loss of the internal com¬ 
bustion engine. Internal calculations have shown 
that the total reduction in fuel consumption 
when all measures are combined can exceed 
20 % under cycled testing conditions [1]. 

Increased comfort and convenience 
represent an additional aspect that comple¬ 
ments the lower levels of consumption. In 
an automated stop-and-go mode, the driver 
could take his left foot off of the clutch ped¬ 
al, making it much easier to drive in con¬ 
gested traffic while at the same time mini¬ 
mizing wear and tear on the clutch. 



Figure 4 ECM at market launch (volume 

production) and in a concept vehicle 

Automation of manual 
transmission - Old friends 
for the 21 st Century 


The electronic clutch management system 
(ECM, Figure 4) developed by LuK, which al¬ 
lows the driver to shift without having to en¬ 
gage the clutch, was launched in 1993 [2, 3]. 
What started out as a great idea did not win 
over end customers, however. Vehicles 
equipped with an ECM were well received by 
only a few people and are no longer on the 
market. One of the reasons why acceptance 
was so low presumably has to do with the fact 
that when a vehicle comes only with an ac¬ 
celerator and a brake pedal (i.e. no clutch 
pedal), it very much resembles a vehicle with 



Figure 5 Automated manual transmission 




























62 


a conventional automatic transmission, and 
the assumption is made that an ECM should 
behave in this manner, which it cannot due to 
its different design. 

The automated manual transmission 
(AMT, Figure 5) also debuted in volume-pro¬ 
duced vehicles around this time and compet¬ 
ed directly with the ECM. Today, even this 
technology has not been able to win over cus¬ 
tomers and is currently offered on selected 
models only. This lack of acceptance can be 
attributed to the noticeable interruption in 
tractive power, which puts the AMT at an im¬ 
mediate disadvantage to the automatic trans¬ 
mission when it comes to comfort. The global 
market share for vehicles equipped with an 
AMT is under 1 %, making this type of trans¬ 
mission by far the one with the lowest unit 
quantities when viewed in the context of the 
other transmission technologies available. 

It therefore almost goes without saying 
that previous attempts to automate the 
manual transmission have been less than 
fruitful, as the unit did not impress drivers 
enough in terms of enjoyment or comfort. 
Today, however, new opportunities have 
presented themselves. The ECM and the 
AMT both provide a solid basis to facilitate 


the aforementioned operative strategies for 
reducing consumption. 

There are other ways to automate the 
manual transmission, however, without hav¬ 
ing to forego the clutch pedal. 

Clutch by wire - 
Intelligent clutch 


One well-known concept is the clutch-by¬ 
wire (CbW) design. For the driver, this trans¬ 
mission very much resembles a convention¬ 
al manual transmission because three 
pedals are provided and there is no immedi¬ 
ate sense of automation involved. Automa¬ 
tion is, in fact, working “behind the scenes”, 
since actuating the clutch pedal merely 
serves to communicate the driver’s inten¬ 
tion, which is detected by a position sensor. 
The clutch is actually operated by an actua¬ 
tor assembly. As the name “by wire” no 
doubt reveals, this system does not have a 
hydraulic or mechanical connection that 
links the clutch with the clutch pedal. 



Clutch 


Actuator pos. 


Clutch pedal 





Figure 6 Design and components of the clutch-by-wire (CbW) system 



























Manual Gearbox 


4 


63 


Spindle drive Position sensor Rotor 



Pressure connection Angle sensor 

Figure 7 Hydrostatic clutch actuator - HCA 


LuK has already presented the technology 
several times as a way to bring the manual 
transmission up to date, with design work 
focusing on improving comfort levels with 
regard to using the clutch, accelerating from 
a stop, and improving NVH behavior. The 
inherent problem with this approach, how¬ 
ever, was that the functions offered did not 
lead to a favorable cost-benefit ratio. The 
concept was then no longer pursued from 
the original design perspective and has 
never entered volume production. 

Figure 6 depicts the architecture of a 
clutch-by-wire system. The input data re¬ 
quired by the clutch control unit comprises 
information about the vehicle (CAN) and the 
driver’s intent (pedal position) as well as ad¬ 
ditional parameters such as transmission 
speed, which are provided by on-board 
sensors. Predefined strategies then deter¬ 
mine the target clutch torque on this basis, 
and the system can correct driver inputs as 
required. For example, if the driver inadver¬ 
tently misuses the clutch or does not coor¬ 
dinate it properly with the gas pedal which 
can cause the engine to stall, the system is 
clever enough to override the driver’s com¬ 
mands. 


In this arrangement, the physical release force 
of the clutch no longer acts on the pedal, 
which means that this must be emulated to 
provide for a realistic experience. Schaeffler 
has addressed this need by developing a new 
product that appeals from a cost and installa¬ 
tion perspective. The result is a very compact 
force emulator that replaces the conventional 
hydraulic master cylinder while mirroring its 
dimensions (refer to [4] for details). 

The hydraulic clutch actuator (HCA, Figure 7), 
also developed by Schaeffler, can likewise be 
fitted to actuate the clutch assembly and is 
described in detail in [5]. This actuator tech¬ 
nology was designed specifically for hydrauli¬ 
cally actuated clutches as found in automated 
transmissions and is now being used in vol¬ 
ume production double clutch transmissions. 

The inherent benefit of the HCA lies in its 
universal adaptability. Not only can it be ac¬ 
commodated without having to make major 
modifications to the vehicle; it can also actuate 
and control a CSC as well as a semi-hydraulic 
slave cylinder. The latter may not represent the 
best configuration, however. The internal axial 
stroke drives a hydrostatic system that, in turn, 
produces an axial stroke on the release lever of 
the clutch. It is therefore practical to actuate the 




64 



Electromechanical actuator for CbW - Compact and performance oriented 


Figure 8 

release lever directly instead of indirectly, by 
means of hydraulics. This has prompted 
Schaeffler to develop a compact, perfor¬ 
mance-oriented solution (Figure 8). The design 
objective is to replace the semi-hydraulic cylin¬ 
der with an electromechanical actuator without 
having to make substantial modifications to the 
transmission, since this makes it possible to 
add an automated clutch to an existing trans¬ 
mission with minimal additional cost. 

In an effort to enhance flexibility still fur¬ 
ther, Schaeffler has taken an additional step 
by developing a modular actuator system 


Mechanical module 



Hydraulic module 


Figure 9 Modular actuator concept for 
maximum flexibility 


that allows the same base actuator to be 
used in all applications (Figure 9). This ac¬ 
tuator houses all electronics, including the 
sensors, electric motor, and a special spin¬ 
dle drive for manual clutches (self-locking in 
the closing direction). Depending on the 
constraints of the application, the base ac¬ 
tuator is mated to a mechanical or hydraulic 
module, which also serves as the connec¬ 
tion point to the transmission. Development 
and system costs are minimized as a result, 
which is absolutely required if these sys¬ 
tems are to be offered in conjunction with 
price-sensitive manual transmissions. 

An additional description of this system 
and current developments in actuator tech¬ 
nology as pursued by Schaeffler can be 
found in [6]. 

The design requirements for the actua¬ 
tor are comparably high with respect to the 
aforementioned possibilities for automating 
the manual transmission. The ECM and 
CbW in particular require a pronounced dy¬ 
namic response to also enable fast gear¬ 
shifts. If progress is made to considerably 
reduce these requirements, costs can be 
lowered further. With this in mind, Schaeffler 
has taken a new direction whereby the 
clutch is no longer operated by an actuator 
every time. 



Manual Gearbox 


4 


65 


t 


o. 

E 

o 



Impulse start 
Launch/ 
modulation 

Stall Torque demand 
protection 


Sailing^ 

Start-stop ifo] 


Torque 

tracking 


Torque 

limitation 



Anti-judder 

control 



Requirements 


Basic functions 
(opening/closing) 




Temporary modulation 
with low dynamics 




Permanent 

modulation 


T 


Permanent 
modulation with 
high dynamics 


Figure 10 Actuator requirements versus functions 

MT plus - Partially 
automated alternative 


The underlying idea is to arrange an actuator 
in parallel with the release system to consid¬ 
erably reduce the actuator performance or 
capacity required. Consideration must also 
be given to the functions that can still be ex¬ 
ecuted, however, and whether the remaining 
added value can justify an automated setup. 


Figure 10 provides a rough estimate or out¬ 
line in this context by assessing several func¬ 
tions based on dynamic performance and 
application times as pertinent evaluation 
criteria. The highest requirements relate to 
functions for reducing vibrations. The re¬ 
quirements for accelerating from a stop and 
sailing are small by comparison as they do 
not require high dynamic response or ongo¬ 
ing clutch modulation. 

According to this estimation, a smaller 
actuator would already offer sufficient po¬ 
tential for upgrading a manual transmission 



Figure 11 Basic concept of MTplus partial automation with OR logic 

























66 


Reservoir connection 


Position sensor 
Hydraulic cylinder 
Piston rod 



Clutch pedal Spindle drive E|ectrj ’ ca| 

connection connection 


Figure 12 Example of an active master cylinder 
(OR logic) 


and make it possible to include the func¬ 
tions mentioned above for reducing con¬ 
sumption. 

The challenge is to find a suitable actua¬ 
tor concept that allows a clutch to be actu¬ 
ated conventionally and automatically. 
Steps must also be taken to ensure that the 
actuator does not interfere with foot-actuat¬ 
ed operation and that the driver always has 
complete control over the vehicle. 

Detailed concept studies were conduct¬ 
ed to find solutions for this application sce¬ 
nario. The basic concept devised is shown 


in Figure 11 and has two defining character¬ 
istics: 1) At no time when the actuator is ac¬ 
tuated does this translate into the clutch 
pedal being moved and 2) the release posi¬ 
tion of the clutch is well defined by OR logic. 
This, in turn, ensures that the driver’s intent 
is highly prioritized at all times. 

The sketch provided in Figure 11 char¬ 
acterizes an active master cylinder in prin¬ 
ciple, with a structural design shown in Fig¬ 
ure 12. The electric motor with spindle drive 
is arranged next to the master cylinder. The 
connections linking the pedal and spindle 
drive to the piston rod allow only one force 
to be transmitted in the disengaging direc¬ 
tion, which correlates with the OR logic. 

An active master cylinder has noticeable 
drawbacks, however, including a greater 
risk of noise being transmitted by the elec¬ 
tric motor to the interior, additional installa¬ 
tion space required in the already cramped 
area surrounding the cylinder, and little to 
no universal adaptability. This type of actua¬ 
tor would have to be modified or redesigned 
in many cases for different application sce¬ 
narios, which does not make it very attrac¬ 
tive from a cost standpoint. The same holds 
true for the majority of installation arrange¬ 
ments near the slave cylinder, which like¬ 
wise lead to moderate results. 


Master cylinder connection 



Electrical 
connection 
Electric motor 


Reservoir connection 


Piston 1 


Carrier ring 


Slave cylinder connection Piston 2 


Figure 13 Actuator variant for MTplus with two intermediate pistons 


Manual Gearbox 


4 


67 


Active secondary piston 
Master cylinder connection 



. Electrical 

Spindle drive / connection 

Electric motor 

Slave cylinder connection 

Figure 14 Alternative intermediate piston variant without additional loss 
encountered during foot-operated actuation 


Integrating the ac¬ 
tuator in the hydrau¬ 
lic pressure line, on 
the other hand, is 
much more favor¬ 
able with respect to 
installation space 
and adaptability. In 
this setup, the ac¬ 
tuator unit is posi¬ 
tioned where it can 
be physically ac¬ 
commodated and is 
connected to the 
hydraulic line. A di¬ 
rect transfer fom the design shown in 
Figure 11 leads to an intermediate cylin¬ 
der with two pistons which divide the hy¬ 
draulic system (Figure 13). During auto¬ 
mated actuation, piston 2 is driven 
directly by the actuator, while piston 1 
remains stationary. 

During manual, foot-operated actua¬ 
tion, piston 1 drives piston 2 by way of the 
carrier ring, which in turn leads to two 
drawbacks: 1) The seals produce addition¬ 
al friction and 2) the “sniffing” function re¬ 
quired of the piston 2 cylinder further mini¬ 
mizes travel. 

To counteract these drawbacks, design 
work is being carried out on an alternative 
variant that does not call for the release sys¬ 
tem to be permanently split into two sepa¬ 
rate parts (Figure 14). The result is a direct 
fluid path extending from the master to the 
slave cylinder (blue arrow) during foot-oper¬ 
ated actuation, with minimal additional loss 
encountered. In automated mode, the ac¬ 
tive intermediate piston blocks the inlet ac¬ 
cess point of the master cylinder and as¬ 
sumes actuation of the clutch. Another 
problem area that needs to be addressed 
for this concept is ensuring a smooth transi¬ 
tion when a driver override input is received. 
To this end, different valve and reservoir ar¬ 
rangements are currently being investigated 
(not shown in Figure 14). 


System comparison - 
Limitless possibilities 


The previous sections discuss a number of 
possibilities for automating the clutch used 
in a manual transmission. Figure 15 com¬ 
pares each of these variants side by side. 
The most consequent variant is the ECM, 
which does away with the clutch pedal and 
only senses driver inputs through the gear 
selector. The CbW offers similar possibilities 
at comparable cost. Although the driver 
must engage the clutch, all direct actuations 
of the clutch are executed by an actuator as 
is the case with the ECM. 

The new MTplus concept was devised 
to offer a cost-effective alternative with a re¬ 
duced functional scope by partially auto¬ 
mating the clutch assembly. Unlike the ECM 
and CbW, the clutch is only automated 
when accelerating from a stop in gears 1, 2, 
and R; when the driver shifts to higher 
gears, the clutch is operated manually only. 
The design challenges specific to this con¬ 
cept are to provide for good operability 
while optimally coordinating actuator and 
foot-operated actuation inputs. Further 
analysis will be conducted in a trial test us¬ 
ing a demonstrator. The following benefits 



68 


ECM 


CbW 


MT plus 



• 2-pedal design 

• Fully automated clutch 

• Actuator with high dynamics 

• Gearshift intention detection 

• Gear recognition sensor 

• Automated actuation for launch 
and shifting 


• 3-pedal design 

• Fully automated clutch 

• Actuator with high dynamics 

• Pedal force emulator 

• Gear recognition sensor 

• Automated actuation for launch 
and shifting 


• 3-pedal design 

• Partially automated clutch 

• Actuator with reduced dynamics 

• Hydraulic connection between 
pedal and clutch 

• Gear recognition sensor 

• Automated actuation for launch 
only (gears 1 st , (2 nd ), reverse) 


Figure 15 Variations of clutch automation for manual transmissions 


are achieved in comparison to an ECM or 
CbW: 

- Lower cost thanks to reduced actuator 
requirements (dynamic response and 
application times) 

- Mechanical override capability (re¬ 
duced functional safety requirements) 


- No possibility of a breakdown should 
the actuator system fail 
All three systems offer comprehensive func¬ 
tionality (Figure 16). This especially applies to 
the options available for reducing consump¬ 
tion, which are supported by each system. The 
sailing and other functions offered make the 


f 


o 


LLl 


I 


£ 

n 

O 


f 


Concept 


Driving without 
clutch pedal 





Optimal 
pedal force 





Microslip 



Traction 

control 


Anti-judder 

control 



Electric clutch 
during gear shift 

Impact 

protection 

Tip-in/back-out 

damping 


Longer drive 
ratios 





Sailing 





Regeneration 


Stall prevention 



Electric 

launch/creeping 

Automated 

launch 

Resonance 

drive 


Collision 

protection 

Hybrid 

capability 

Autonomous 

driving 

Clutch 

protection 

Pedal force 
assistance 

Emergency 

braking 

C0 2 potential 

Assistance 

Protection 

Comfort 

Safety 


Figure 16 Functions afforded by clutch automation 
























































Manual Gearbox 


4 


69 


manual transmission much more hybrid friend¬ 
ly from an overall design perspective. A wide 
variety of technical features and options also 
improves comfort and durability and can even 
be extended to include assistance systems. 

Looking optimistically 
into the future 


The trend toward greater levels of automation 
and electrification to reduce fleet consump¬ 
tion also requires solutions for the manual 
transmission. Schaeffler is dedicated to find¬ 
ing these solutions by promoting technical 
developments for automating the clutch. In 
the process, the effects on the overall pow¬ 
ertrain cannot be overlooked. For example, 
further reducing consumption by adding lon¬ 
ger gear ratios leads to increased engine exci¬ 
tations as a result of lower operating speeds, 
which in turn necessitate better operative 
characteristics of the torsion dampers. 



Improving the efficiency of 
the powertrain and the 
challenges to be overcome 


The previous section already discussed the 
importance of shifting the operating point of 
an engine to lower operating speeds 
(downspeeding) in order to significantly re¬ 
duce fuel consumption. For example, when 
the mean operating speeds of a current 
2.0-liter diesel engine are reduced by 10 %, 
it is possible to consume 5.6 % less fuel un¬ 
der NEDC testing conditions. This potential 
can only be tapped, however, if doing so 
does not lead to any drawbacks in driving 
dynamics or comfort. Thus, to ensure that 
these driving dynamics remain fairly consis¬ 
tent and comparable, the same output must 
be achieved when the engine operates at a 
speed that is 10 % lower, which is why max¬ 
imum torque must also be increased by ap¬ 
proximately 10 % (Figure 17). 


Consumption 



■ Today's engine 

■ 10 % operating point shifting 

■ Extreme downspeeding 


Figure 17 Operating point shifting and potential reduction in consumption with downspeeding 












70 


In addition, it is foreseeable that usable 
speeds will be expanded much further 
down in the rev range. Some engines in the 
future will even reach their peak torque at 
below 1,000 rpm! Compared to today’s en¬ 
gines, this will allow these power units to 
theoretically reduce their consumption by 
11 % under NEDC testing conditions. 

Such engine developments ultimately 
lead to considerably higher vibrations from 
the powertrain. This initially becomes evi¬ 
dent in the rotational irregularity that in¬ 
creases proportionately to an increase in 
torque or a drop in engine speed. Adding to 
this is the fact that as engine speed goes 
down, the excitation frequency becomes 
more closely aligned with the natural fre¬ 
quency of the rest of the powertrain. 

Figure 18 summarizes the effects on the 
rotational irregularity in the powertrain. Relative 
to a current engine (green line), the oscillation 
range at the transmission input doubles for the 
same damper technology when engine speed 
is reduced by 10 % (blue line). This marks the 
starting point at which target comfort levels can 
no longer be attained. Some drivers would 


even intentionally avoid low engine speeds for 
this reason and thereby not profit from the 
lower fuel consumption otherwise possible. 

Further downspeeding amplifies the situa¬ 
tion disproportionately (red line). When maxi¬ 
mum torque is available below 1,000 rpm, the 
comfort target at this speed is undershot by 
more than 600 %. In order to achieve an ac¬ 
ceptable comfort level with these engines, 
performance-oriented damper systems must 
be fitted and are critical to ensuring that the 
consumption benefits afforded by downspeed¬ 
ing can, in fact, be realized. 

Vibration isolation - 
State of the art 


Some 20 years ago, the requirements 
placed on damper technology dramatically 
rose as a result of the direct-injected diesel 
engines then offered for passenger cars 
(Figure 19). 



Time 

-- Target —Engine = Transmission 

Figure 18 Rotational irregularity at the engine and transmission input for current and future engines 



















Manual Gearbox 


4 


71 


Full-load 


Down- 



Yesterday 


Higher 

sensitivity 


Torsional 

damper 




Dual-mass 


1985 flywheel (DM F) 



9 


DMF + 
centrifugal 


Tomorrow 


2008 pendulum 
absorber 



Prim. Sec. 

_ Trans. 


Vehicle 


urs\ M UM p 
in^-o—y 


Figure 19 Dramatic increase in performance requirements for vibration-dampening systems 


This shift in engine technology presented 
the developers of these systems with en¬ 
tirely new challenges. The resulting rota¬ 
tional irregularity could not be sufficiently 
counteracted using the available torsion- 
damped clutch disks. Although the princi¬ 
ple of the low-pass filter was known, it was 
not regarded as being technically feasible 
until the dual-mass flywheel (DMS) was in¬ 
troduced in passenger-car applications. 
By leveraging its comprehensive knowl¬ 
edge of the operating principles of passive 
damping systems, LuK systematically 
started investigating the underlying corre¬ 
lations early on and was consequently able 
to offer a compatible solution that met the 
emerging challenges in good time. Many 
years of know-how in metalworking then 
finally led to a robust product. 

In the years that have passed, specific 
torque outputs have more than doubled in 
comparison to the first turbocharged, di¬ 
rect-injection diesel engines. The resulting 
effect is that even today, some engines 


experience torsional vibrations that can¬ 
not be counteracted with a DMS alone. 
The answer to these increased require¬ 
ments is the centrifugal pendulum ab¬ 
sorber (CPA), which is a damper assembly 
that introduces additional mass external 
to the power flow. The dual-mass flywheel 
and centrifugal pendulum absorber have 
been continually refined and advanced 
and will meet the requirements associated 
with the upcoming evolutionary stages set 
for the current generation of engines [7]. 

The next engine generation, however, 
which is currently under development, 
will call for vibration isolation measures 
that are even more capable, which is why 
Schaeffler is not only investigating the 
possibilities and constraints of today’s 
technology, but is also looking at alterna¬ 
tive solutions. 










72 


Alternative solutions - 
Options and the operating 
principles that define them 


Before implementation concepts are consid¬ 
ered at product level, the operating principles 
that govern them must be thoroughly evalu¬ 
ated with respect to future requirements. It is 
in this context that the method that uses 
simple, linearized models to investigate the 
relative operating principles has proven suc¬ 
cessful. Not only the technical potential of the 
different approaches must be factored into 
the overall assessment, however, but also 
their cost-benefit ratio, whereby the objective 
must always be to find approaches that offer 
equal, uniform performance across an en¬ 
gine’s entire operating speed range. Im¬ 
provements made at very low engine speeds 
are not optimal if they compromise the prog¬ 
ress already achieved in the mid and high¬ 
speed ranges. In addition, only those solu¬ 
tions that comply with the restrictions for 
installation space and weight and are just as 
robust as current systems when it comes to 
friction, wear, and manufacturing tolerances 
are promising candidates. 

The following systems will be investigat¬ 
ed to determine whether (and under which 
conditions) their physical potential is capa¬ 
ble of isolating the torsional vibrations of a 
motor that utilizes an extreme downspeed¬ 
ing concept so that comfortable driving is 
possible from 800 rpm. 

Spring-mass system - Principle of 
the dual-mass flywheel 

The basic operating principle of this ar¬ 
rangement is that two masses connected to 
each other by a spring-damping system os¬ 
cillate against one another. In terms of the 
operating range used today and the excita- 


Stiffness /17 



Engine speed n Eng in rpm 
— Target — Basis 

— Mass x 3.5 or Stiffness /17 

Figure 20 Isolation capacity and limitations of 
the spring-mass system 

tions that are encountered as a result, 
dampers demonstrate overcritical perfor¬ 
mance, and provide better isolation as fre¬ 
quency increases. When frequency drops, 
the resonance frequency is more closely 
aligned with these excitations and torsional 
vibrations become more prevalent. 

Theoretically, it is also possible to use a 
spring-mass system to reach the required 
target even in extreme downspeeding sce¬ 
narios. This, however, would require the 
mass to be increased by a factor of 3.5 or 
the spring rate to be reduced by a factor of 
17 compared to the base construction. 
Neither is realistic. Arguments not in favor 
of increasing mass are the increased in¬ 
stallation space required, added weight, 
and worse driving dynamics. Reducing the 
spring rate by an extreme amount is also 
not possible as a result of the installation 
space problem and the compromised driv¬ 
ing experience that would result. 












Manual Gearbox 


4 


73 


Anti-resonance - Principle of 
interference 

The following describes two concepts for 
generating anti-resonance: The spring-mass 
absorber and the summation damper. 
Although both concepts use a different 
operating principle, they produce similar 
results under the same conditions. 

The spring-mass absorber 

The spring-mass absorber is based on a 
second spring-mass system. When this 
system is excited at its resonance frequen¬ 
cy, an opposing oscillation is generated that 
ideally completely cancels out the original 
excitation. With a conventional absorber 
connected via a spring, this effect occurs at 
exactly one frequency - the resonance fre¬ 
quency of the absorber. The drawback is an 
additional resonance point above the ab¬ 
sorber resonance frequency. 

A conventional absorber is therefore not 
a suitable means of reducing torsional vi¬ 
brations in the powertrain. What is required 
is a absorber whose dampening frequency 
corresponds to the ignition frequency of the 
engine at all times. This property is fulfilled 
by the centrifugal pendulum absorber (Fig¬ 


Absorber 

Basis 


^y ^y 



— Target — Basis 

— Absorber 


Figure 21 Principle and isolation effect of a 
conventional absorber 

ure 22), which restoring force is dominated 
by the centrifugal force of the absorber 
mass. Since the centrifugal force changes 
quadrically in relation to the engine speed, 





Figure 22 Centrifugal pendulum absorber (CPA) as a speed-dependent absorber 
















74 


Jeff 

J(Sec+CPA) 



Figure 23 Equivalent effective mass inertia of a 
centrifugal pendulum absorber 

the centrifugal pendulum absorber has a 
absorber frequency that is proportionate to 
this speed. This is the ideal property or at¬ 
tribute for reducing torsional vibrations in 
the powertrain, since a fixed excitation or¬ 
der can be dampened. 

Figure 23 shows how effective the mass 
of a centrifugal pendulum absorber is. The 
graph depicts, in relation to the engine 
speed, by what factor the secondary mass 
would have to be increased for similar per¬ 
formance - e.g. by a factor of 3 at a speed 
of 1,000 rpm or a factor of 9 at 1,500 rpm. 

With a CPA, a vibration isolation figure of 
100 % could theoretically be achieved up 
on a defined frequency. In demonstrator ve¬ 
hicles, a decoupling performance rating of 
up to 99 % was already demonstrated in 
conjunction with a DMF. This, in turn, makes 
it easy to meet the requirements of today’s 
engines and their upcoming evolution stag¬ 
es. Current systems are even capable of 
fulfilling the requirements of two-cylinder 
engines. The potential offered by the CPA is 
described in an additional article in this 
book [7]. 

When engine speed drops, the centrifu¬ 
gal pendulum absorber must absorb more 
energy. The ability of this pendulum to re¬ 
spond depends on the mass involved and 
the vibration angle, whereby the latter is in¬ 


trinsic. The mass of the pendulum can also 
only be increased to a certain extent due to 
the installation space available. 

Whether the CPA can produce a vibration 
isolation that is also compatible with the ex¬ 
citations of the next generation of engines is 
not entirely clear at present. Recent im¬ 
provements made to the system support 
this working hypothesis, however. 

Nevertheless, Schaeffler also continues 
to search for alternative approaches. Using 
mass intelligently is the key to implementing 
future solutions. 


The summation damper 

Another way of dampening vibrations with 
anti-resonance is to add two vibration paths 
together. Figure 24 charts this principle. Vi¬ 
brations are transferred via a spring-mass 
system along the one path and directly to a 
lever on the other. The pivot point of the le¬ 
ver (summation unit) is void offeree and mo¬ 
tion from a dynamic vibration perspective. 



Engine speed n Eng in rpm 


Target — Basis 

— Summation damper 


Figure 24 Principle and isolation effect of a 
summation damper 



















Manual Gearbox 


4 


75 



M 


hAAAAAAAAAr 

- 41 =° 


1 .i Mi-lF 

2n ^ J j 




AM 

Aq> 


f A = Anti-resonance frequency c 0 = Effective stiffness 
Figure 25 Variations in spring arrangement for the summation damper 


mmm 

-0 


As in the case with a conventional absorber, 
a summation clamper can also decouple 
100 % of vibrations but only for a single fre¬ 
quency. The summation damper therefore 
has an advantage over the absorber in that 
no additional natural frequency is generat¬ 
ed. Unwanted vibrations above and below 
the anti-resonance frequency remain pres¬ 
ent, however. 

The frequency to be isolated, or target¬ 
ed, can theoretically be selected as re¬ 
quired. When coordinating the system, the 
summation damper provides one additional 
parameter not available with the conven¬ 
tional absorber-the lever ratio in addition to 
the spring rate and the rotary mass (J). An¬ 
other benefit is that the system can also be 
configured so that a dampening effect is 
achieved on the primary side (engine side). 

Further arrangements are possible in 
addition to the summation damper charac¬ 
terized in Figure 24. For example, the spring 
can be positioned at any point required (Fig¬ 
ure 25). Comprehensive testing has re¬ 
vealed that the same basic laws and princi¬ 
ples apply irrespective of the positional 
arrangement of the spring. The anti-reso¬ 
nance frequency can even be calculated for 
all concepts using a single formula. Assum¬ 
ing that the lever ratio, spring capacity, and 
mass J do not change, not only is the same 
anti-resonance frequency yielded for all of 


the concepts, but also an identical transfer 
response. 

When the transfer response for design 
concepts with different anti-resonance 
points is considered, the typical properties 
of a summation damper become apparent. 
Anti-resonance frequencies can theoreti¬ 
cally be shifted to any low engine speed. 
Doing this, however, not only reduces the 
absorbtion width, but also the isolating 
properties above the anti-resonance fre¬ 
quency (Figure 26). This, in turn, means that 
a summation damper configured for very 
low anti-resonance responds sensitively to 
fluctuating parameters. A satisfactory solu¬ 
tion can only be achieved if at least one of 
the three relevant parameters is variable 
with respect to engine speed. 

In a direct comparison, the summation 
damper has a slightly higher theoretical 
potential for dampening vibrations than the 
conventional damper (Figure 27). Having 
said this, the advent of the centrifugal pen¬ 
dulum absorber has already provided a so¬ 
lution for realizing a variable-speed damp¬ 
er and is currently being used in volume 
production applications. Variable-speed 
summation dampers, on the other hand, 
have yet to be integrated. 

































Summation damper Absorber Amplitude 


76 



— Summation damper • Anti-resonance frequency 

ire 26 Influence of the anti-resonance frequency on the absorbtion width 


Constant parameters 


Speed-dependent parameters 





Figure 27 Evolution of the absorber and summation damper 


























Manual Gearbox 


4 


77 


Summary 


In the race to achieve global C0 2 targets, au¬ 
tomatic transmissions have clearly taken an 
early lead as they allow engineers to develop 
fuel-saving strategies by decoupling the en¬ 
gine from the transmission. The manual trans¬ 
mission also offers certain benefits, however, 
including reliability, durability, and a low price, 
the latter of which continues to appeal to buy¬ 
ers of small vehicles in particular. The logical 
next step of advancing the technology of the 
proven manual transmission must therefore 
focus on automating the clutch so that the 
driving strategies explored here can also be 
implemented in vehicles with manual trans¬ 
missions. In addition to offering technical solu¬ 
tions that have already been developed (ECM, 
CbW), Schaeffler is working on systems that, 
when scaled down in scope, largely maintain 
the price advantage that a manual transmis¬ 
sion has over its automatic counterpart. 

Automated clutches are not only capable 
of decoupling the engine from the rest of the 
powertrain, but also actively support and fa¬ 
cilitate many other comfort and protective 
functions. Automating acceleration from a 
stop, for example, can prevent the clutch from 
being overloaded or misused, which in turn 
allows the powertrain to be configured differ¬ 
ently so that longer gear ratios can be imple¬ 
mented to further reduce fuel consumption. 

The operating point of the internal com¬ 
bustion engine then shifts to lower speeds 
and specific torque is increased. Both mea¬ 
sures lead to more pronounced rotational ir¬ 
regularity, however. The resulting higher de¬ 
sign requirements for mechanisms that isolate 
frequencies will nevertheless be reliably met 
by current technology as it is incorporated into 
today’s engines and those targeted for the 
next evolution stage. The problem revolves 
around the next generation of engines, which 
will require even more capable systems. Al¬ 
though the technology offered by the dual¬ 


mass flywheel in conjunction with a centrifu¬ 
gal pendulum absorber is a prime candidate, 
the summation damper is also worth consid¬ 
ering if a way can be found to extend its high 
potential at low operating speeds to mid¬ 
range and higher speeds. Schaeffler contin¬ 
ues to investigate both concepts with a great 
deal of interest. The key to developing a more 
responsive summation damper lies in the abil¬ 
ity to vary one relevant parameter with respect 
to engine speed. A solution that is robust, af¬ 
fordable, and can be deployed on a large 
scale has not yet crystallized, however. 


Literature 


[1] Gutzmer, P.: Individuality and Variety. 

10 th Schaeffler Symposium, 2014 

[2] Kremmling, B.; Fischer, R.: The Automated 
Clutch. 5 th LuK Symposium, 1994 

[3] Fischer, R.; Berger, R.: Automation of Manual 
Transmissions. 6 th LuK Symposium, 1998 

[4] Welter, R.; Herrmann, T.; Honselmann, S.; 

Keller, J.: Clutch Release Systems for the 
Future. 10 th Schaeffler Symposium, 2014 

[5] Muller, M.; Kneissler, M.; Gramann, M.; Esly, N.; 
Daikeler, R.; Agner, LComponents for Double 
Clutch Transmissions. 9 th Schaeffler 
Symposium, 2010 

[6] Muller, B.; Rathke, G.; Grethel, M.; Man, L.: 
Transmission Actuators. 10 th Schaeffler 
Symposium, 2014 

[7] Kooy, A.; The Evolution of the Centrifugal Pen¬ 
dulum Absorber. 10 th Schaeffler Symposium, 
2014 

[8] Fidlin, A.; Seebacher, R.: DMF Simulation Tech¬ 
niques. 8 th LuK Symposium, 2006 

[9] Kroll, J.; Kooy, A.; Seebacher, R.: Torsional 
Vibration Damping for Future Engines. 

9 th Schaeffler Symposium, 2010 

[10] Reik, W.; Fidlin, A.; Seebacher, R.: Gute 
Schwingungen - bose Schwingungen. VDI- 
Fachtagung Schwingungen in Antrieben, 2009 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 




78 


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79 


Isolation is the Key 

The evolution of the centrifugal 
pendulum-type absorber not only for DMF 


Dr. Ad Kooy 


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80 


Introduction 


A key task that has concerned the automo¬ 
tive industry in recent years has been to re¬ 
duce consumption. One effective measure 
for achieving this goal is to exploit even 
lower engine speeds for driving. Torque is 
increased to achieve this without losing 
power. Doing so allows the engine to run 
only very slightly above idle speed and 
therefore in an extremely consumption- 
efficient range. One challenge is to achieve 
adequate powertrain isolation even for 
these low engine speeds and thus provide 
drivers with their usual level of comfort. 

Figure 1 [1] shows that the dual mass 
flywheel (DMF) is a factor in achieving this 
goal, particularly in connection with the 
centrifugal pendulum-type absorber. While 
twin-cylinder engines have yet been un¬ 
able to reach the projected fuel savings for 
day-to-day use, the increasing numbers of 
three-cylinder engines have achieved low¬ 
er consumption figures. However, lower 
consumption places stricter demands on 
vibration isolation. The secondary-side 
centrifugal pendulum-type absorber (CPA) 



■ Conventional system ■ DMF ■ 


was introduced as a concept in conjunc¬ 
tion with the DMF as early as 2002 [2], and 
successfully went into series production a 
few years later. The simple physical princi¬ 
ple, modular design and extremely good 
isolation have led to increasing acceptance 
and proliferation not only in the DMF, but 
also in other damping concepts such as 
torque converters and clutch discs. There 
have also been huge improvements in how 
the centrifugal pendulum-type absorber 
works thanks to far-reaching understand¬ 
ing of the centrifugal pendulum-type ab¬ 
sorber; more detailed information is pro¬ 
vided about this below. 

As the DMF must also be optimised for 
other operating points, such as startup, or 
optimised for so called impacts - very high 
torque peaks when bottoming out the arc 
springs - compromises must be made. 
These compromises also have an indirect 
influence on isolation in drive mode. We will 
be using examples of impacts that affect 
DMFs when stalling the vehicle and demon¬ 
strating methods of preventing stress of this 
kind and making DMFs more robust. These 
result in greater freedom for optimising tor¬ 
sion isolation and so improving driving com¬ 
fort. 



DMF with CPA 


Modified value for NEDC 
fuel consumption 



Figure 1 Fuel economy potential with DMF and DMF with CPA 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_5, © The Author(s) 2014 














Centrifugal Pendulum-type Absorber 


5 


81 


First generation 


Second generation 



Layout of end stop dampers on first and second-generation CPAs 


Figure 2 

Development of DMF 
centrifugal pendulum-type 
absorbers 


To date, one million centrifugal pendulum- 
type absorbers have been produced for six- 
cylinder, four-cylinder and three-cylinder 
engines, and the concept has been contin¬ 
ually developed. Prototypes show that the 
technology could also be employed in twin- 
cylinder engines. 

The secondary-side arrangement of the 
centrifugal pendulum-type absorber makes 
the arc spring damper, which provides pre¬ 
isolation, especially important. Taking en¬ 
gine torque development into consideration 
largely automated simulation programs run 
through hundreds of variations evaluating 
start and drive to find the optimum combi¬ 
nation of arc spring and CPA for a vehicle 
application. Of course, this requires vehicle 
parameters of adequate quality which are 
not always available during the early stages 
of development in which design takes place. 
This is where LuKs wealth of experience re¬ 
ally comes into its own, as it allows us to 
complete missing data in a meaningful 
manner. However, should corrections be re¬ 
quired subsequently during to vehicle test¬ 


ing, simulations of this kind can quickly be 
repeated. Interaction with other critical op¬ 
erating points can also be integrated , such 
as stalling the engine along the critical de¬ 
tails of engine timing management. 

It is easy to calculate the natural fre¬ 
quency of a thread pendulum, in other 
words a point mass moving on a circular 
path, if the angle is small. However, this ap¬ 
proach is inadequate for centrifugal pen¬ 
dulum-type absorbers. The path curvature 
must be more pronounced to maintain a 
constant order (natural frequency to speed 
frequency ratio) independently of the mag¬ 
nitude of the angle. This approach is the 
only way to achieve optimum isolation over 
the whole engine speed for partial throttle 
as well as for wide-open throttle. Special 
attention must be given to the rpm range 
slightly above idle speed. On account of 
the low centrifugal forces in this range, the 
CPA needs as large a vibration angle as 
possible to store sufficient vibration ener¬ 
gy. High engine torques exacerbate the 
situation. Therefore, the goal is to maxi¬ 
mise this angle along with the pendulum 
inertia. For this reason, the three circular 
end stop dampers previously present on 
first-generation centrifugal pendulum-type 
absorbers have been combined into a V- 
shaped end stop damper on an additional 
intermediate mass in second-generation 
absorbers (Figure 2). 






82 


This eliminates the need for the bean¬ 
shaped holes in the flange required for the 
circular end stop dampers and creates ad¬ 
ditional space for greater vibration angles 
or heavier pendulums. The added interme¬ 
diate mass lies relatively far towards the 
outer edge in radial terms, thereby improv¬ 
ing isolation in the low speed range through 
increased inertia. A number of other opti¬ 
misations, such as optimising the arc 
spring damper with the centrifugal pendu¬ 
lum-type absorber as a system, smoother 
pitch surfaces and optimised paths, have 
together resulted in a significant perfor¬ 
mance boost, especially at low engine 
speeds (Figure 3). 

The example of a four-cylinder diesel 
engine shows that when using the first gen¬ 
eration absorber an increasing of the engine 
torque from 360 to 450 Nm leads to a clear 
deterioration in isolation. In contrast, when 
the second generation is used, a torque in¬ 


crease can be handled without loss of com¬ 
fort. For three-cylinder engines acceptable 
values of 500 rad/s 2 from about 1,000 rpm 
are already achieved (in this example, a die¬ 
sel engine with 270 Nm). However, these 
values can still be significantly reduced: If 
the entire clutch system — i.e. DMF with 
centrifugal pendulum-type absorber and 
clutch — is designed according to an en¬ 
tirely new layout, (third generation), it is pos¬ 
sible to achieve angular acceleration ampli¬ 
tudes of below 200 rad/s 2 from 800 rpm 
upwards and without requiring any further 
space. The rigidity of drive shafts, in partic¬ 
ular, must be incorporated into this concept. 
If rigidity changes, it results in a completely 
new design. It makes close coordination 
with the vehicle manufacturer’s develop¬ 
ment process essential. 

The considerations mentioned above 
relate to a centrifugal pendulum-type ab¬ 
sorber integrated below the arc spring 



Four-cylinder — 450 Nm — generation 1 Three-cylinder — 270 Nm — generation 1 

— 450 Nm — generation 2 — 270 Nm — generation 3 

— 360 Nm — generation 1 

Figure 3 Comparing DMF isolation with various CPA generations 
in three-cylinder and four-cylinder engines 









Centrifugal Pendulum-type Absorber 


5 


83 



Figure 4 Kinematic simulation of pendulum 
motion at 150 rpm 


damper. As already shown in [1], a centrifu¬ 
gal pendulum-type absorber can also be 
arranged next to the arc spring, i.e. radially 
further towards the edge, if sufficient space 
is available; this improves isolation even 
further, where necessary. For engines with¬ 
out cylinder deacti¬ 
vation, as commonly 
used in series pro¬ 
duction, it is there¬ 
fore possible to 
achieve adequate 
isolation using a 
centrifugal pendu¬ 
lum-type absorber. 

Should a CPA of this 
kind provide isola¬ 
tion better than that 
required, costs can 
be reduced by omit¬ 
ting two of the four 
pendulum masses. 


Path wear is not expected due to the fact 
that the pendulum only has a rolling motion. 
However rattling noises may occur when 
switching the engine off: As soon as the en¬ 
gine speed drops below approx. 200 rpm, 
the pendulum’s centrifugal force drops be¬ 
low the force of gravity. It falls a few millime¬ 
tres within the designed degree of freedom 
until it strikes the bolts on the flange. In or¬ 
der to better understand this process, kine¬ 
matic simulations have been carried out 
and compared using high-speed record¬ 
ings (Figure 4). 

The simulation demonstrates how the 
two rollers strike differently; the precise ar¬ 
rangement of the damper and rollers and 
the clearance between the roller compo¬ 
nents have an effect on these striking pat¬ 
terns. These parameters must be precisely 
analysed and optimised. In addition to these 
kinds of optimisations, ways of preventing 
stopping noises have also been investigat¬ 
ed. One option is to arrange circular end 
stop dampers at the end of the pendulum 
(Figure 5). 

This causes the pendulums to strike 
each other after a short fall, and a part of 
the kinetic impact energy stored in the pen¬ 
dulum system is neutralised without any 
noise occurring. The rollers striking on the 



Figure 5 CPA on outer edge with end stop dampers between the pendulums 














84 


flange only cause slight noise. This concept 
works well with a closed throttle valve, as 
pure torsional acceleration is low in this 
state. However, much higher torsional ac¬ 
celeration occurs when stopping if the 
throttle valve is to remain open, for instance 
to enable cylinders to charge correctly to 
enable quick automatic start-up. The result 
is that all pendulums have a virtually syn¬ 
chronised torsional motion, thereby render¬ 
ing the rubber stops on the end of the pen¬ 
dulum ineffective; their job is then assumed 
by the central V-shaped end stop damper. 

Centrifugal pendulum- 
type absorber mounted on 
the clutch disc 


The success of the DMF is due to the fact 
that hypercritical operation is largely pos¬ 
sible, compared to torsion-damped clutch 
discs. The result is an enormous increase 


in isolation, as already shown in an example 
in [3]. Also discussed in [3] was the option 
of arranging a centrifugal pendulum-type 
absorber on the clutch disc - positioned 
on the gearbox input shaft for simulation 
purposes. Based on the knowledge of 
pendulum path design, permissible mass 
moments of inertia and tolerances permit¬ 
ted in series production available at that 
time, a viable solution was not within 
reach. Today, our in-depth expertise con¬ 
cerning the design of centrifugal pendu¬ 
lum-type absorbers coupled with new 
ideas on the reduction of clutch disc mass 
inertia means this approach can be imple¬ 
mented (Figure 6). 

For clutch discs with a single pendulum 
system, it comprises of two or three pendu¬ 
lums and is calibrated to the main excita¬ 
tion, i.e. order 1.5 for a three-cylinder en¬ 
gine. Clutch discs with double pendulum 
systems have two additional auxiliary pen¬ 
dulums, calibrated to double the main exci¬ 
tation frequency. In both designs, the pen¬ 
dulums are arranged next to the damper. 
During development, a particular aim was to 
keep the extra clutch disc inertia caused by 



Clutch disc 
with damper 


With additional 
single pendulum 



Pendulum I 


Pendulum II 


With additional double 
pendulum and 
two-flange design 


Flange 


Figure 6 Clutch discs without a pendulum, with a single pendulum and with a double pendulum 
system 



Centrifugal Pendulum-type Absorber 


5 


85 


the pendulums to a minimum, so that gear 
synchronisation was not overloaded. There¬ 
fore, the pendulums needed to be particu¬ 
larly effective despite their low mass. As the 
effect of a pendulum is mainly determined 
by the product of mass and vibration angle, 
the vibration angles consequently had to be 
hugely enlarged. 

Initial designs for the first generation 
used three pendulums. In the optimised, 
second-generation version, two pendulums 
with secondary spring masses were used 
for clutch discs with a single pendulum sys¬ 
tem (Figure 7). 

The additional intermediate mass was 
introduced along the same lines as the DMF 
(Figure 2): Therefore, more mass can be ar¬ 
ranged on the outer edge in radial terms. 
But the most important innovation concerns 
the two roller paths of each pendulum. The 
paths are now no longer identical and are 


now skewed relative to one another instead 
of merely displaced. This is reflected in the 
skewed arrangement of the bean-shaped 
holes for the rollers in the pendulums, as a 
comparison of the first and second genera¬ 
tions shows. This arrangement causes the 
pendulum to execute a rotation in addition 
to oscillation. The sketch in Figure 7 illus¬ 
trates this principle: During movement, the 
end of the pendulum is guided radially in¬ 
wards while the other end simultaneously 
moves radially outwards. This arrangement 
has become known as a trapezoidal pendu¬ 
lum, while the first generation is called a 
parallel pendulum. 

Thanks to their trapezoidal oscillation, 
the pendulums need less space meaning 
that considerably larger pendulum vibration 
angles can be achieved. Additional rotation¬ 
al energy is also stored when turning, so 
better use is made of the pendulum mass. 


First generation 
single pendulum 



a) Flange 

b) Main pendulum 

c) Secondary pendulum 

d) Intermediate mass 

e) End stop damper 

f) Pressure spring 

g) Roller 


Parallel pendulum 
(first generation) 



Second-generation 
single pendulum 



Second-generation 
double pendulum 



Figure 7 CPA for clutch discs 






86 


This effect can also be utilised on the DMF, 
but it is not so effective there due to the 
mounting space available. 

Although the pendulum masses are 
lower than those of the DMF, undesirable 
knocking noises may occur when stopping 
if the bell housings are sensitive or open. 
The spring bracing of second-generation 
pendulum masses (Figure 7) also helps 
combat this problem. The preloaded 
springs can be designed to be especially 
soft thanks to the reduced pendulum mass¬ 
es. This is important because the spring 
forces are not speed-dependent and do not 
follow the principle of the centrifugal pendu¬ 
lum-type absorber. An angular correction of 
path geometry minimises this effect. 

Figure 8 shows a comparison of a DMF 
with a single mass flywheel with CPA on the 
clutch disc and a torsion-dampened clutch 
disc using the example of a four-cylinder en¬ 
gine. The single mass flywheel with a CPA 
on the torsion-dampened clutch disc takes 


the middle position with regard to isolation 
of the torsional vibrations from the gearbox. 
On the engine side it even leads to smaller 
irregularities than a DMF, resulting in a lower 
load on the belt drive. This configuration 
proves its worth for three-cylinder engines 
in conjunction with soft drive shafts. How¬ 
ever, when combined with rigid shafts, we 
have the problem that the third order comes 
through very dominantly in the overall ampli¬ 
tude of gearbox acceleration (Figure 9). The 
figure shows the total amplitude in which 
both orders arrive. 

To dampen the third order, an additional 
pendulum system calibrated to this order 
has to be added; in other words, a double 
pendulum system is required. Figure 6 and 7 
show the layout of both pendulums on the 
clutch disc. It goes without saying that only 
smaller pendulum masses are possible due 
to space constraints, but this is compen¬ 
sated in part by a dual-flange design. In this 
design, the pendulum is situated between 



Transmission 



1 - 1 - 1 - 1 — 

1,000 1,500 2,000 2,500 


Speed in rpm 


— Torsion-damped 
clutch disc 


— Torsion-damped — DMF 
clutch disc + CPA 


Eng. Trans. Veh. 


Eng. Trans. Veh. 


Sec. + 

Eng. Trans. Veh. 

1 1n iiiiimm 1 1 

i -|j wvvvvvnvvi 


1 FI uiimini 1 

1 |—— m fffffffffffi 


rwi_ l~l 

Li 


Figure 8 Comparing three damping concepts based on isolation of a four-cylinder engine in 6 th gear 














Centrifugal Pendulum-type Absorber 


5 


87 



— Torsion-damped — Torsion-damped — Torsion-damped — DMF — DMF + CPA 

clutch disc clutch disc clutch disc 

+ CPA + 2x CPA 


Eng. Trans. Veh. 


Eng. Trans. Veh. 

Eng. Trans. Veh. 

Sec. + 

Eng. Trans. Veh. 


Sec. + 

Enq. Trans. Veh. 

1 FI iAiimmi 1 

1 i”””i 1 tlWHWWrl 


1 1__ T1 Miimiiu 1 

1 L_FI umiiiiM 1 I 

LTXTJ 

rwi_n 


rwi_ 

HTL 



Figure 9 Comparing five damper concepts based on isolation of a three-cylinder engine with rigid 
side shafts in 6 th gear 


two flanges. In the contrary, on a DMF it is 
usual practice for two pendulums to be ar¬ 
ranged around a central flange (Figure 2). 
This new design principle omits the con¬ 
nection elements of the sub-pendulums, 
which weaken the flange. As a result, larg¬ 
er pendulum vibration angles can be inte¬ 
grated. The achievable isolation reveals 
astonishing results: in 6 th gear, isolation 
below 1,300 rpm is even better than with a 
DMF. Fiowever, if the DMF is combined 
with a centrifugal pendulum-type absorb¬ 
er, it is once again clearly the superior 
combination. 

In order not to place additional stress on 
gear synchronisation, the entire mass inertia 
must not be significantly greater than for a 


normal torsion-dampened clutch disc, de¬ 
spite the CPA. This is achieved by reducing 
the mass of all individual parts affected. De¬ 
tailed comments about mass reduction of 
this kind can be found in another article [4]. In 
conjunction with a CPA, the actual torsional 
damper in the clutch disc is dampened to a 
lesser extent which benefits isolation at 
higher engine speeds. Another significant 
benefit is that the centrifugal pendulum-type 
absorber aids isolation in the creeping range, 
i.e. the low torque range. This allows the 
creeping stage to be designed for steeper 
rates and higher torques. In this way, creep¬ 
ing rattle can be largely prevented. 

The introduction of clutch discs with 
centrifugal pendulum-type absorbers pro- 



















88 


vides additional damping solutions, de¬ 
pending on vehicle configuration and the 
required isolation level. Simulations help 
when it comes to selecting the optimum 
damping parameters whilst taking complex 
boundary conditions into account. They can 
be used to implement a solution halfway 
between a DMF and a torsion-damped 
clutch disc both in terms of isolation and 
costs and long-awaited by the automotive 
industry. 


Centrifugal pendulum-type absorbers 
for trucks 



Engine: 2,400 Nm 
Transm.: 12 gears 
Weight: 401 


CPA of 
6 kg reduces 
vibrations: 


Engine side by 30 % 
Transmission side 46 % 


In comparison to passenger cars, signifi¬ 
cant damping of a truck gearbox requires 
considerably higher inertia of the centrifugal 
pendulum-type absorber on the clutch disc. 
However, this higher inertia leads to an un¬ 
acceptable reduction of synchronisation 
service life, which at 1,000,000 km is well 
above the requirements for passenger cars. 
For this reason, other ways of improving iso¬ 
lation have been explored: The CPA was ar¬ 
ranged on the single mass flywheel (Figure 10). 
It can be detached for easy maintenance. 


Figure 10 CPA on a truck single mass flywheel 

The solid pendulums, which weigh around 
6 kg, reduce engine vibrations by 30 % for a 
typical six-cylinder engine at 2,400 Nm, and 
reduce gear vibrations by 46 %. The latter 
directly improves the gearbox service life, 
as it is restricted if the vehicle is often driven 
at low engine speeds. In contrast, using a 
single mass flywheel with CPA can reduce 
engine speed without compromising ser¬ 
vice life when only low to medium engine 
torques are used, as is often the case (Fig¬ 
ure 11). Fuel consumption is reduced by 5 %, 
which represents a 
competitive edge for 
end customers that 
should not be un¬ 
derestimated. The 
service life of the 
belt drive also ben¬ 
efits from reduced 
engine vibrations, 
thereby allowing 
this drive to be more 
simply constructed 
or service intervals 
to be extended. 



500 750 1,000 1,250 1,500 1,750 

Speed in rpm 

- Without CPA 

- With CPA 


Figure 11 Fuel consumption savings in a truck thanks to a single mass 
flywheel with CPA 



















Centrifugal Pendulum-type Absorber 


5 


89 



Figure 12 Classifying impacts 

Reducing impacts 


The principle of a DMF (without centrifugal 
pendulum-type absorber) is ultimately based 
on shifting the resonance speed of the pow¬ 
ertrain from the driveable range into ranges 
well below idle speed. By shifting this speed, 
hypercritical driving is possible throughout the 
entire speed range with the resulting excellent 
isolation. Even in the early days of DMF devel¬ 
opment, it became clear that driving situations 
below idle speed, such as that occur when 
stalling a vehicle, lead to large vibration angles 
and the DMF can strike the end stops (im¬ 
pact). The energetic transition of high kinetic 
energy in the relatively rigid end stop results in 
torques that can be up to 40 times the engine 
torque. Impacts can also occur at other oper¬ 
ating points, however not usually at this level 
or with this regularity (Figure 12). 

Many ideas for reducing impacts have 
already been developed and implemented. 
The majority of them actually contradict the 
primary task of the damper system, i.e. iso¬ 
lation, by requiring additional mounting 
space (such as a slipping clutch in the 
flange) or using thicker (more robust) spring 
wires (damping arc springs). The following 
describes one approach using software 
and one using hardware; these approaches 
dramatically cut the severity and regularity 
of these kinds of impacts. 


Influence of the engine control unit 
when stalling 

Figure 13 shows a typical stalling mea¬ 
surement of a three-cylinder diesel engine 
plotted in an engine speed/time diagram. 
It can be seen that powerful impacts are 
caused by the extreme difference in 
speed between the primary and second¬ 
ary side. To aid understanding, this dia¬ 
gram is converted into speed squared (n 2 ) 
over crankshaft angle; this is because 
combustion causes an injection quantity 
to be turned into kinetic energy, which, in 
turn, is proportional to n 2 . Thus, cyclic en¬ 
gine irregularities for the same injection 
quantity are shown as the same ampli¬ 
tudes regardless of engine speed. It is ap¬ 
propriate to use the crankshaft angle, as 
the ignitions occur at equidistant intervals 
in the diagram. Very high impacts occur 
when the engine stops at TDC or when 
the engine does not reach TDC at all due 
to the retroactive effect of the secondary 
flywheel. In the latter case, reverse com¬ 
bustions are produced with extreme im¬ 
pacts. The aim must be to anticipate this 
situation and disable the injection process 
in time. Until now, fixed speed limits have 
been implemented in the control system 
to disable the process, but Figure 13 
shows that it is advisable to use an addi¬ 
tional gradient-based limit. If a straight line 
is drawn in this diagram through the two 




























90 


previous ignition or injection points just 
before TDC, as shown in Figure 13, it is 
immediately apparent that the engine will 
stop at approx. 0 rpm when it reaches the 
next TDC. This causes the high impacts 
afterwards. 

The last ignition or injection were there¬ 
fore not only useless - the engine was at 
a standstill afterwards - they also dam¬ 
aged the DMF and it would have been bet¬ 
ter for them to have been disabled by the 
engine control unit. These types of prob¬ 
lems can now be identified early on in the 
project using simulations. By them, it is 
apparent that even a small difference of 
10 ms in the engagement time can cause a 


tremendous difference in the impact level 
(Figure 14). 

This finding also matches the large vari¬ 
ations observed time and again in road 
tests. Statistical analysis is therefore essen¬ 
tial, and can be conducted by means of 
simulations using a well-calibrated model 
(Figure 15). 

These simulations then form the basis 
for estimating field quality. During this pro¬ 
cess, the behaviour of multiple drivers is 
calculated using Monte Carlo methods (roll¬ 
ing the dice for impact levels) in conjunction 
with the S/N curves of the arc spring and 
the regularity of occurrence. It is possible to 
evaluate the software using the simulation 


— Engine — Transmissions 

— Secondary 



d> 

3 

O" 

0.0- 

k. 

o 

+•> 

6,000 - 

c 

"C 

Q_ _ 

3,000 - 

Arc s| 
in Nrr 

o - 


An = 770 rpm 




Impact 


Point of 
injection 


Impact 
similar to 
.5 s 


5 10 

Crankshaft revolutions 


Figure 13 Impact when stalling 














Centrifugal Pendulum-type Absorber 


5 


91 





Figure 14 Influence of clutch engagement time on the impact level in stalling simulations 


by integrating the software parameters. It is 
important to trigger necessary software ad¬ 
justments early on in the project, preferably 
at the start of the project, as testing of soft¬ 
ware changes is extremely time-consum¬ 
ing. The engine control unit should also pre¬ 
vent the engine being restarted by the 
continuing motion of the vehicle after stall¬ 


ing as this causes speed ratios and impacts 
that are difficult to control. 


The High Capacity spring 

It is difficult to develop an active engine 
control unit strategy that can prevent im¬ 
pacts entirely for all 
operating condi¬ 
tions and combina¬ 
tions of parame¬ 
ters. Therefore, the 
remaining impacts 
must be intercept¬ 
ed by an increased 
robustness of the 
DMF. This is where 
the High Capacity 
spring (HC spring) 
can play a vital role 
(Figure 16). 



Figure 15 Cumulative frequency of impacts during stalling and illustration 
of the Monte Carlo method 


























92 


High capacity arc spring: 

• Increased distance between coils 

• Can store up to 50 % more energy 

• Helps to prevent deformations 

• Similar wire thickness 

• Same tension when engine torque applied 

• Same level of insulation 

• Spring rate only slightly higher 

• No significant load losses due to setting 
under approx. 8,000 Nm 



Figure 16 High capacity spring (HC spring) 

The basic idea is to considerably increase 
the torque capacity of the arc spring and 
therefore absorb approx. 30 % to 50 % 
more energy in the characteristic curve, 
without hitting the end stop. Figure 17 
shows the end of a start-up procedure, in 
which high clutch torque results in the 
damper striking the end stop. 

The higher torque capacity of the HC 
spring is achieved by an increased distance 
between the coils and largely absorbs the 
high clutch torque. Wire thickness is kept ap¬ 
proximately the same, so that the stress ex¬ 
erted on the springs by engine torque, and 
thus the service life, remains unchanged. As 
the distance between the coils increases as 
a consequence of the concept, fewer coils 
can be accommodated in the same space. 
The nominal spring rate therefore increases 
slightly. This affects starting behaviour to a 
small extent, but not drive characteristics. 
This is because the rear coils are disabled in 
drive mode as a result of the friction caused 
by centrifugal force. The shortened spring 
consequently has absolutely no effect on re¬ 
ducing the number of active coils. 

Fatigue strength is not an issue for small 
impacts as impacts are relatively rare — 
typically fewer than 1,000 load cycles over 
the vehicle’s service life. The determined 


service life for small impacts is more than an 
order of magnitude greater. However, if 
higher impacts does happen to act on the 
HC spring despite its capability of absorb¬ 
ing energy, flattened coils can absorb the 
difference without serious crushing. LuK 
has used flattened coils successfully on 
standard springs for quite some time now. 
As HC springs have a significantly higher 
torque capacity than standard springs, set 
HC springs can still safely absorb the en¬ 
gine torque. Overall, HC springs yield huge 
benefits for the DMF in terms of robustness 
without compromising torsion isolation. 


Summary 


The evolution of the centrifugal pendulum- 
type absorber in conjunction with overall 
damper tuning improved the isolation 
achieved by DMFs to such an extent that it 
can also cope with higher engine torques 
and cover today’s three-cylinder and even 
twin-cylinder engines . Furthermore, they 
still have further potential, as regard to isola¬ 
tion, for dealing with the expected further 




Centrifugal Pendulum-type Absorber 


5 


93 






— Engine —Transmission —Secondary 
Figure 17 Influence of the HC spring when driving off 


increase of engine torque from idle speed 
upwards. However, close interaction be¬ 
tween powertrain design and damper con¬ 
cept is absolutely essential if this potential is 
to be achieved. 

Locating the centrifugal pendulum-type 
absorber on the clutch disc succeeded in 
providing a long-awaited solution halfway 
between a simply damped clutch disc and a 
DMF. For trucks, arranging the CPA on the 
single mass flywheel also leads to reduced 
strain on the gearbox and the belt drive. Im¬ 
pact situations can be managed through 
early optimisation of the engine control unit 
and the use of High Capacity springs. No 
additional protective measures must then 
be implemented in the DMF; the system 
comprising DMF and centrifugal pendulum- 


type absorber can be designed specifically 
for maximum isolation. 


Literature 


[1] Kroll, J.; Kooy, A.; Seebacher, R.: Land in 
sight? 9 th Schaeffler Symposium, 2010, 

[2] Kooy, A.; Gillmann, A.; Jaeckel, J.; Bosse, M.: DMF 
- Nothing new? 7 th LuK Symposium, 2002 

[3] Reik, W.: Torsional vibration isolation in the 
powertrain. 4 th LuK Symposium, 1990 

[4] Schneider, M. et al.: The Clutch Comfort Portfo¬ 
lio: From a supplier’s product to an equipment 
criterion. 10 th Schaeffler Symposium, 2014 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 

















94 


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95 


Clutch Release Systems 

From system know-how to a 
successful volume produced product 


Roland Welter 
Tim Herrmann 
Sebastian Honselmann 
Jeremy Keller 


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96 


Introduction 


More than 100 years after the invention of 
the automobile, it seems as if the technol¬ 
ogy of clutch release systems, is a mature 
one, without the necessity of changes. 
However, even in this seemingly evolved 
family of products, the innovation dynamic 
remains high. Current developments aim 
to further increase robustness, replace 
existing materials with polymer materials, 
and integrate sensors in the master cylin¬ 
der. 

Master cylinders with integrated sen¬ 
sors have only been used in a few cases 
in the past. The proliferation of systems 
such as start/stop or the electronic park¬ 
ing brake is now leading some car manu¬ 
facturers to consider such sensors in the 
master cylinder as obligatory. The sen¬ 
sors make it possible to measure the trav¬ 
el on the clutch pedal and thus determine 
the driver’s intent. 

Materials too are evolving. While for 
decades cast iron or aluminum alloys 
were dominant, in new applications, 
master cylinders, pipes and slave cylin¬ 
ders are almost always made of plastic. 
Initial problems with the use of polymer 
materials, such as master cylinder 
squeaking, high adhesive friction and 
volume expansion, have since been re¬ 
solved. The technologies necessary for 
the use of plastics have been constantly 
refined and are now solid and economi¬ 
cal. Even in double clutch systems, 
which have higher, continuous loads, 
plastic cylinders are gradually becoming 
established. Current developments are 
focused on using plastic in the pedal 
box. 

Ultimately, the robustness require¬ 
ments for the components used in clutch 
operation have risen significantly. Even 
just a few years ago, one million cycles 


was the going operating load specifica¬ 
tion for release systems. Now, it is not 
unusual to require two to three million 
cycles - accompanied by increased re¬ 
quirements regarding the ambient condi¬ 
tions of temperature, water and contami¬ 
nant exposure. 


Clutch master cylinder 


Schaeffler’s LuK brand currently has three 
different types of plastic clutch master cyl¬ 
inders in the product line. The difference in 
the designs is in the seal configuration 
used. 

The clutch master cylinder that is used 
the most in volume production has a mov- 




P = Primary seal 
S = Secondary seal 



Figure 1 Types of clutch master cylinders 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_6, © The Author(s) 2014 




































Clutch Release Systems 


6 


97 



Primary Secondary 


Figure 2 New clutch master cylinder with a one-piece housing and seals mounted to 
the pistons 


ing piston and two stationary seals (Figure 1, 
top). This configuration allows the prima¬ 
ry seal to build up pressure toward the 
slave cylinder while the secondary seal 
retains the fluid without pressure in the 
reservoir. The advantages of this design 
include the fact that the pistons can be 
manufactured using a duroplastic materi¬ 
al. With this material, the squeaking en¬ 
countered with the usual seals made of 
synthetic ethylene propylene diene mono¬ 
mer (EPDM) rubber is effectively sup¬ 
pressed. The disadvantage to this design 
is that for installation reasons of the seals 
the housing must be made in two pieces 
and thus is comparatively more costly to 
set up. In addition, the entire pressure 
chamber expands radially during opera¬ 
tion, which results in a relatively high vol¬ 
ume expansion. 

The second design (Figure 1, center) 
uses a primary seal that moves with the 
piston and a stationary secondary seal 
in the housing. The volume expansion 
during operation is lower due to the 
smaller pressure space when the piston 
is pushed in. The disadvantages of this 
design include the fact that the noise 
problem has not thus far been satisfac¬ 


torily resolved, and that, in addition to 
the installation of the secondary seal, a 
costly, two-piece housing is necessary 
here as well. 

In new applications, LuK is therefore 
focusing on a third variation (Figure 1, be¬ 
low), which has a one-piece housing made 
of thermoplastic and uses seals that are 
mounted to the piston. Figure 2 shows an 
example of the technical design of one 
such clutch master cylinder. 

Seals made of EPDM are generally 
used, with the primary seal being protect¬ 
ed on the outer diameter against the cylin¬ 
der raceway by a shield made of unrein¬ 
forced polyamide in order to improve 
friction and wear. This type of measure is 
not needed for the secondary seal, which 
is not under pressure during operation. To 
prevent squeaking noises with critical 
brake fluids, seals made of special materi¬ 
als can be used with this master cylinder 
design. 

A well-thought-out test stand system is 
helpful in researching suitable seal and 
raceway material combinations for the 
specific brake fluid. In addition to stan¬ 
dardized noise measurements with com¬ 
plete master cylinders, there is a tribologi- 










98 



Time 


Figure 3 Tribological test stand for basic trials on seal friction and noise excitation 


cal test stand available for basic trials that 
was developed specifically for this purpose 
(Figure 3). 

Current trials on this Tribometer involve 
mounting and loading a flat specimen made 
of the raceway material with a seal material 
specimen, which is pushed along the flat 
specimen. In the process, the contact 
points can be flooded with brake fluid and 
maintained at a constant temperature. 
Measuring devices allow the friction load 
and frictional vibrations to be recorded. 
Suitable material combinations show two 
clear effects: The gradient of the friction co¬ 
efficient over the piston speed is small and 
there is no detectable frictional vibration. 
Figure 4 shows an example of the contact of 
polyamide with standard ethylene propyl¬ 
ene diene monomer rubber (EPDM) and liq¬ 
uid silicon rubber (LSR). 

In all of the trials conducted thus far, 
LSR has proven itself superior with regard 
to friction gradient, friction level and friction 
vibration behavior. There are, however, a 
few brake fluids on the market, particularly 
in Asia, that are not compatible with LSR. 
The goal for future development is to avoid 
this limitation. 


Advantages of the new cylinder design 
with moving seals include its low volume 
expansion. This is due to the design, since 
the highest pressure occurs with the pis¬ 
ton nearly pushed in, when the “breath¬ 
ing” cylinder surface is comparatively 
small. Due to the one-piece housing, stat¬ 
ic burst pressures of more than 200 bar 
were achieved with the new cylinder de¬ 
sign. 



— Seal standard EPDM 

- Seal LSR 


Figure 4 Friction load gradient for an EPDM 
and an LSR specimen against a 
raceway made of polyamide PA 66 
with fiberglass fill 



































Clutch Release Systems 


6 


99 



- 20 °C 

Figure 5 Volume expansion of the new master cylinder 


The cylinder can be equipped with addi¬ 
tional attachments if desired, such as a 
premounted bleeding pipe, which is most 
cost-effective designed as a plastic con¬ 
voluted tube. In contrast to the seal cus¬ 
tomarily used thus far for this type of 
convoluted tube, which has to be greased 
to attain acceptable mounting forces, a 
new type of self-locking seal is used. The 
seals have locking hooks on the adjacent 
side, which engage into a groove of the 
convoluted tube during mounting (Figure 6). 
Additional lubrication is not required. 
The associated problem of contamina- 



0 2 4 6 8 10 12 14 


Travel s in mm 

Figure 6 Ergonomically configured connec¬ 
tion between bleeding hose and fir 
tree connection 


tion is thus eliminated. The mounting 
forces are 40 N, maximum. The haptic 
indication of a successful mounting is a 
noticeable drop in the sliding force. Nor¬ 
mal fill pressures during vehicle assem¬ 
bly of up to 10 bar are endured without 
issue. 


Integrated sensor system 


There is currently an increased demand 
for master cylinders with travel and/or po¬ 
sition sensor systems. The travel sensors 
continuously measure the piston travel 
and thus replace the potentiometer on the 
pedal axle, whereas the position sensors 
generate a digital signal when passing 
through defined piston positions and thus 
take over the pedal switch function. Both 
measuring tasks can take place in one 
space-saving sensor on the clutch master 
cylinder. In addition, travel measurement 
on the clutch master cylinder is less de¬ 
pendent on tolerances and thus more ac¬ 
curate. The sensors used in the clutch 
master cylinder work exclusively without 
contact and thus cause no noise or wear. 
In the meantime, in addition to the familiar 
Hall sensors, magnet-free inductive sen- 




























100 



Figure 7 Master cylinder with integrated inductive travel and position 
sensor 


sors integrated into the housing are now 
also available in volume production, as the 
example in Figure 7 shows. 

The reason for the increased demand 
for sensor information from the clutch 
pedal is the need to make travel and po¬ 
sition information available to the engine 
controller or the control units for start/ 
stop or the electronic parking brake, 
which allow conclusions to be drawn re¬ 
garding the engagement status of the 
clutch as well as the driver’s intention. 
Common functions today are shown in 
Figure 8. 

Additional signal utilization is conceiv¬ 
able, but are is not yet used on a large 
scale. Examples include detecting operat¬ 
ing errors such as insufficient load release 
of the pedal when driving, too frequent 
slipping of the clutch or resetting a calcu¬ 
lation model for wear predictions. 

Hall switches, integrated Hall ICs for 
travel measurement as well as magnet-free 
inductive sensors are used in volume pro¬ 
duction. The use of intelligent sensors 
makes it possible to compensate for the 
tolerances related to production and 
mounting with a calibration after mounting 
on the clutch master cylinder. They also 


have a certain diag¬ 
nostic capability. 
Cable breaks, short 
circuits and internal 
sensor errors can 
be detected in this 
way and communi¬ 
cated to the con¬ 
troller. 

The require¬ 
ments for the func¬ 
tional safety of elec¬ 
tronic components 
in vehicles accord¬ 
ing to ISO 26262 
are met with design 
up to level “ASILC”. 
This is dependent, 
however, on the specific safety goals of the 
sensor-related vehicle functions, which 
must be specified by the automobile man¬ 
ufacturer. Usually multiple pieces of avail- 


Early detection of driver 
intention when disengaging 
at engine start 

Preventing engine start for 
unseparated clutch 



Automatic release of the 
electronic parking brake 


Controlled start up on hill 


Turn-off of speed controller 
for clutch operation 


Comfort increase from engine 
engagement (speed increase) 


Figure 8 Use of the sensor signal on the 

clutch master cylinder in the vehicle 



101 


Clutch Release Systems 6 


Hall series 
linear 




Hall Array 


3D Hall 
linear 


Blanks Lead 
Solution frame 




Smart Hall 1C 


Inductive 

linear and position 



Induktiv 

(magnet-free) 


Hall 

position 




Hall switch 


Sensors used in the clutch master cylinder 


Figure 9 

able sensor information are accessed in 
order to reach the safety goals at a vehicle 
level. This reduces the requirements on in¬ 
dividual sensors. 

One important advantage of the Hall 
sensors is their short axial installation 
length, which can be further reduced in 
the future. Thus, Hall array sensors use 
two Hall cells connected one behind the 
other and signal processing via a micro¬ 
controller. Highly integrated chips use 
multiple Hall elements, which make it 
possible to measure the magnetic field in 
multiple dimensions and derive travel in¬ 
formation. In miniaturized form, these 
sensors no longer have boards, but rath¬ 
er are all mounted directly on the lead 
frame together with the necessary cir¬ 
cuit. The price of the advantage of a 
small installation space is that additional 
circuits or custom solutions are not pos¬ 
sible. 


Development goals include reducing the 
mass of the magnets and minimizing the 
proportion of rare earths. While cylindrical 
magnets were used originally, LuK is in¬ 
creasingly switching to segment magnets 
and using an anti-rotation device for the pis¬ 
tons. In the Hall switch-point sensor, the 
magnet has now been reduced to a small 
cube. 

Despite these advances, efforts are 
being made in newer solutions to com¬ 
pletely eliminate the use of magnets in or¬ 
der to circumvent the price volatility for 
rare earths. One initial result of these ef¬ 
forts is a contact-free inductive travel sen¬ 
sor that uses a small aluminum ring as the 
measuring element. Higher precision can 
be achieved with this type of system than 
with a Hall sensor, and additional switch 
points can be derived from the signal of 
the integrated controller or from an addi¬ 
tional switch as needed. 


















102 




- Linear signal — Power supply 

- Position — Vehicle CAN 


ECM: Engine control unit EPB: electronic 
BCM: Chassis control unit parking brake 


Figure 10 Master cylinder signal processing in the vehicle; left: commonly used today; right: future 
concept 


The only disadvantage to the inductive 
solution is the comparatively large instal¬ 
lation length. The length of the coil sys¬ 
tem, depending on the design, can be up 
to 135% of the measured travel. This pos¬ 
es no problem for most applications. Nev¬ 
ertheless, LuK is working on shorter in¬ 
stallation solutions, but they are not yet 
production-ready. 

Highly integrated sensors emit a trav¬ 
el signal as well as position signals and 
provide this information to different con¬ 
trollers. Since the signal interfaces and 
the expected voltage levels are not uni¬ 
form, and the on-board electrical system 
is the only available power supply, the full 
potential of an intelligent sensor solution 
cannot be completely realized at this 
time. 


Since, in contrast to the linear travel sig¬ 
nals, the position signals are not available 
from active diagnostic functions, safety 
goals according to ISO 26262 are often 
not completely met at vehicle levels with 
position points derived from the linear 
travel signal. For applications with high 
safety requirements, LuK therefore rec¬ 
ommends using a travel sensor with two 
independent travel signals, which are pro¬ 
cessed by the respective controllers and 
which can be compared as needed for 
increased safety. A stabilized 5 V power 
supply is provided in this case by a con¬ 
troller and the sensor signals are prefera¬ 
bly provided as pulse-width modulation 
(PWM) or as digital signals (for example 
SENT). One advantage of this solution is 
that the information needed for different 





















Clutch Release Systems 


6 


103 


vehicle functions is derived directly from 
the linear signal and can be transmitted 
via CAN bus to the respective controllers. 
The first manufacturers are already plan¬ 
ning to use this concept. 

The travel measurement via integrat¬ 
ed sensor system can also be used for 
clutch-by-wire applications. To this end, 
a broad spectrum of more or less com¬ 
plex solution suggestions are being dis¬ 
cussed. LuK favors a solution in which 
only the conventional master cylinder in 
the pedal box is replaced by an element 
with similar installation space. This could 
consist of a housing with a piston rod 
and interior spring assembly. A spring 
with a linear characteristic curve is 
eclipsed by the spring and hysteresis ef¬ 
fect of a clamping element and ensures 
the usual plateau of the clutch operation. 
The sensor is on the outside of the hous¬ 
ing in this design, the same as with the 
conventional master cylinder. 


Clutch pipes and 
installation elements 


Pipes have the task of transferring the hy¬ 
draulic pressure safely and with as little 
friction and volume loss as possible. In 
addition, pipes are supposed to prevent 
engine vibrations propagating as far as 
the pedal box. Installation elements such 
as dampers and anti-vibration units are 
used for this. 

Pipes are currently made of steel/rub¬ 
ber or polyamide (PA 12 and PA 612) ma¬ 
terials. Currently, pipes made of plastic 
are becoming increasingly common be¬ 
cause of their low costs [1]. LuK now uses 
PA 610 for almost all pipes. This plastic is 
more than 60 % based on plant raw mate¬ 
rials. The global availability of prematerials 
is better than for PA 12 and PA 612. The 



— Disengage 
Engage 


Figure 11 Compact pedal load emulator with sensor for clutch-by-wire systems 











104 



Figure 12 High Pressure Pipes made of PA 610 
for clutch operation 


mechanical properties are almost the 
same as for PA 612 and the chemical 
compatibility is better. 

Plastic pipes in vibration-critical ap¬ 
plications (diesel engines and engines 
with few cylinders) mostly require the use 
of a filter to counter pedal vibration and 
interior noises. This filter traditionally op¬ 
erates like a soft added volume in the 
pipe. However, this regularly caused a 
conflict between good filter effect and 
low-loss direct operation. 

Due to the complexity of this conflict 
of interest, an optimal solution was very 
hard to come by in testing. Therefore, 
specialized simulation tools had to be 
used. In general, it is sufficient to calcu¬ 
late the transmission behavior of the pipe 
within the frequency range. For this pur¬ 
pose, LuK has the PipeSim program, 
which calculates the flow and vibration 
behavior in the pipe based on the numer¬ 
ical solution of the Navier-Stokes equa¬ 
tions. 


Calculation application 
Flow profile 



Figure 13 Simulation of the transmission behavior of clutch pipes and built-in 
vibration dampers 










Clutch Release Systems 


6 


105 


a> 



Frequency f in Hz 


Steel / rubber 
— Polyamide 


Slave cylinder Master cylinder 



— Without damper 
Without damper 

— Damper in position 2 


Figure 14 Example of a variation calculation using PipeSim 


PipeSim helped in carefully studying the vi¬ 
bration transmission up to the master cylin¬ 
der and identifying the best corrective mea¬ 
sures. This generally involves a vibration 
damper with appropriate tuning, an anti-vi¬ 
bration unit or a combination of the two at 
an optimal point along the pipe. The simula¬ 
tion also allows for early determination of 
the pipe routing, which is available even be¬ 
fore test vehicles. 

The following example shows the pro¬ 
cedure and the advantages of the simula¬ 
tion: The pipe is first divided into multiple 
segments based on its mechanical de¬ 
sign. An excitation is then specified as a 
frequency curve via the slave cylinder. 
Based on the transmission behavior of the 
line, this generates a corresponding pres¬ 
sure vibration in the master cylinder. The 
diagram of the results is shown in Figure 14 
on the left; the black line shows the curve 
for a steel/rubber pipe and the red line 
shows a typical PA pipe: The steel/rub¬ 
ber variant exhibits a resonance of the 
incoming vibrations at approx. 150 Hz. 
Problems with pedal vibration can be ex¬ 
pected there. The pressure curve over the 
frequency for the plastic pipe is largely 


lower, but there are also several potential 
resonances. In the example shown, only 
the resonances at approx. 250 Hz showed 
as unpleasant in the vehicle. This can be 
countered by installing a vibration damp¬ 
er, whose optimal placement can be cal¬ 
culated using PipeSim. 

The technology for the damper and 
anti-vibration unit could be improved con¬ 
siderably, to a certain extent as a side ef¬ 
fect of the simulation technology: These 
elements can be adjusted perfectly and 
individually to the respective application. 
They only show a minimal volume expan¬ 
sion which does not disrupt the pedal 
characteristic curve and are available as 
modules. The anti-vibration unit (AVU) in 
Figure 15 left is, from a hydraulic perspec¬ 
tive, a type of mutual automatic shut-off 
valve when the pedal is depressed, or a 
restrictor in case of light flow. It is used to 
counter low-frequency pedal vibrations up 
to approx. 150 Hz, which can be felt by the 
foot as vibration. The vibration damper in 
Figure 15, center, was based on a Helm¬ 
holtz damper in the gas dynamic. This in¬ 
volves a resilient capacity with a defined 
restriction as a cross connection to the 










106 


Anti-vibration unit 










— Without anti-vibration unit — Without damper 

— With anti-vibration unit — With damper 


— Without anti-vibration unit 
+ damper 

— With anti-vibration unit 
+damper 


Figure 15 Anti-vibration unit and modular system of vibration dampers 


pressure pipe. The effect is used more in 
the high frequency range and serves to 
counter interior noises. The volume expan¬ 
sion of the connected capacity as well as 
the length and the diameter of the restrictor 
determine the damper frequency and the 
bandwidth. The goal is to keep the volume 
expansion as low as possible in order to 
minimize release travel losses. Thus the 
damper is adjusted specifically for each 
application. A combination of anti-vibration 
unit and damper is shown in Figure 15, 
right. There, the damper is tuned so that 
the resonance, at approx. 550 Hz from Fig¬ 
ure 15, left, is corrected. Thus far, this ap¬ 
proach has been successful in practically 
all cases, even difficult problems, by using 
a combination of plastic pipe and corre¬ 
sponding filter. This is an argument for fur¬ 
ther substitution of steel/rubber pipe with 
cost effective plastic solutions. 


In addition to the vibration dampers, the 
installation of other elements is possible 
in the pipes. Examples include ventilation 
aids for long and non-continuously 
sloped pipe such as are needed for rear- 
wheel drives. For these types of installa¬ 
tion situations, a double pipe and two 
supply reservoirs have often been used 
thus far. Ventilation assistance makes 
this double design superfluous. The 
small hydraulic stage allows air bubbles 
to move only toward the master cylinder 
even if the line is partially tilted away from 
it. The air thus collects at the highest 
point of the ventilation aid, is mostly 
transported toward the master cylinder 
during engagement and can be dis¬ 
charged via ventilation holes. 















Clutch Release Systems 


6 


107 



MC: Master cylinder —► Direction of flow 
SC: Slave cylinder Air bubbles 

Figure 16 Ventilation aid for clutch that slope 
downward to the master cylinder 


Slave cylinder - 
plastic prevails 


Prior years show a clear trend worldwide 
toward concentric slave cylinders (CSC) 
and toward housings made of tempera¬ 
ture-resistant plastic for practically all pas¬ 
senger cars and light utility vehicles. The 
advantages of CSCs include its compact 
design, uniform bearing load and reason¬ 
able price in comparison to all of the other 
systems. Technical further developments 
allow for a continuous increase in reliability. 
Core topics in the development are protec¬ 
tion of the hydraulic system and the bear¬ 
ing from contamination, extended service 
life of the central seal up to three million 
cycles and constantly low friction. Almost 
100 % of all new CSCs for manual trans¬ 
missions are now built with plastic hous¬ 
ings. A detailed technical description is 
provided in [3]. 



Figure 17 Travel adjusted clutch (TAC) and cover-mounted release system (CMR) with smaller cover 
bearing on the end of the guide sleeve [2] 






108 



Figure 18 Cover-mounted release system for 
double clutch 

The cover-mounted release system (CMR) 
presented earlier is now running in initial ap¬ 
plications including in volume production 
and presents comfortable driving behavior 
with regard to pedal vibrations, slip and jud¬ 
der. Further developments of the CMR fo¬ 
cus on reducing the size of the cover bear¬ 
ing and a combination with the new clutch 
with travel-controlled wear adjustment 
(Travel Adjusted Clutch, TAC ). The small¬ 
er cover bearing should save money and 
space. The CMR with TAC is configured 
such that a conventional release cylinder or 
a CMR can be used with the same cover 
tool. This provides the customer maximum 
flexibility for volume production. 


The proven technology for the manual 
transmission is, to some extent, now being 
transferred to the double clutch transmis¬ 
sion with hydraulic actuation and expanded 
upon. Thus, CSCs for double clutch trans¬ 
missions now use the reinforced seal, per¬ 
manent lubrication to reduce friction and 
the CMR technology. Additional enhance¬ 
ments include: 

- Piston with universal joint for parallel 
lift-off of the clutch, 

- Drag torque secured to the housing 
via springs instead of a pre-load 
spring 

- Dimensioning of the engagement bear¬ 
ing for constant high loads. 


Pedal boxes - wallflowers 
with great potential 


Pedal boxes for the clutch operation are 
increasingly being used separately from 
the brake and driving pedal. This topic 
would therefore also be of interest to a 
clutch system manufacturer. Current ac¬ 
tivities at LuK involving pedal boxes in¬ 
clude a design simplified by integrating 
the sensor system, lightweight construc¬ 
tion by a direct use of plastic with an inte¬ 
grated master cylinder and, last but not 
least, the most ergonomic pedal charac¬ 
teristic curve possible. 

The pedal box structure is simplified 
considerably by the possibility of integrating 
the sensor system in the master cylinder. 
Thus far, three switches and a potentiome¬ 
ter with corresponding retainer, stops and 
cables have been used in an extreme case. 
When using a sensor in the master cylinder, 
these components can be completely omit¬ 
ted except for one cable. The measurement 
precision of the system is increased at the 
same time. 



















Clutch Release Systems 


6 


109 


State of the art 


LuK technology 




Figure 19 Simplified design of the pedal box by integrating the sensor system in the master cylinder 



From a material technology perspective, it is 
conceivable to manufacture the above new 
type of master cylinder as one part with the 
pedal box housing using plastic injection 
molding. This reduces the assembly ex¬ 


pense and increases the stiffness by elimi¬ 
nating the joints. Even the pedal would be 
manufactured from plastic for this type of 
solution. In the example shown, there are 
two possible joint points between the pedal 
and pedal box. As a 
result, two different 
ratios can be used 
in the same struc¬ 
ture. The spring is 
configured as a cy¬ 
lindrical coil spring 
and mounted in the 
middle, covered by 
the housing. The 
sensor is mounted 
on the side of the 
master cylinder or 
integrated in the 
structure. 

An ergonomi¬ 
cally perfect de¬ 
sign of the load- 
travel characteristic 
curve on the pedal 
is indispensable 
[4]. Various auto¬ 
motive manufac¬ 
turers are pursuing 


Figure 20 Pedal box for clutch operation made of plastic with integrated 
master cylinder and sensor 















110 




Figure 21 Pedal box with self-adjusting OCS for pedal effort limitation 


the goal of reducing tolerance-based 
load fluctuations in conventional systems 
in new condition. The idea behind this is 
to create a brand-specific pedal feeling. 
Since this is not sufficiently feasible due 
to tolerance limitations, LuK is posed 
with the task of studying an adjustment 
mechanism in the pedal box. Two adjust¬ 
ment mechanisms were considered for 
this: An adjustment of the pedal ratio as 
well as a preload of the over-center 
spring (OCS). 

Finally, the idea to make changes to 
the pedal transmission was proposed be¬ 
cause this also changes the travels on the 
pedal or on the release bearing. The ad¬ 
justment of the preload of the OCS offers 
an elegant option for influencing the load 
level. This makes a manual or automatic 
adjustment equally conceivable. The 
manual adjustment could, for example, be 
made by a simple setting screw on the 
pedal and a measurement of the pedal ef¬ 
fort in the vehicle could be taken. Since 
this type of step is not provided for in the 
vehicle assembly lines, LuK is focusing on 


the automatic setting. For this, the base 
point of the OCS is acted upon via a small 
hydraulic cylinder with the pressure from 
the release system. In new condition, the 
OCS is unloaded, and thus compensates 
very little. With maximum pedal effort or 
increasing pressure in the system, the 
spring is preloaded further until a balance 
is reached between spring load and pres¬ 
sure. A mechanism for engaging prevents 
the tension piston from resetting. 

In this way, the complete form of the 
characteristic curve is not adjusted to a 
set curve, although the height of the 
maximum load is. The form of the auto¬ 
matic adjustment shown has the side ef¬ 
fect that force increases in the operation 
can be prevented in part. In this way, 
wear adjustment is also achieved within a 
certain range. If a clutch repair is need¬ 
ed, the stop mechanism is triggered and 
the automatic adjustment starts over. 
Details on this mechanism are currently 
being developed; the target application is 
in vehicles with conventional clutches. 



Clutch Release Systems 


6 


111 


Summary Literature 


There are numerous starting points for 
innovation in what appears to be the ma¬ 
ture field of release systems. New sen¬ 
sors and pedal boxes with integrated 
master cylinder made of plastic promise 
numerous advantages for future volume- 
produced vehicles. In addition to an in¬ 
crease in functionality for customers, 
there is also a benefit from the lightweight 
design and savings in fuel consumption. 


[1] Welter, R.; Wolf, B.; Ineichen, L.: Leitungs- 
systeme fur die Kupplungsbetatigung. 

VDI Reports no. 2139, 2011, pp. 231-240 

[2] Welter, R.; Wittmann, C.; Hausner, M.; Kern, A.; 
Ortmann, S.: Deckelfester Zentralausrucker fur 
Kupplungen. VDI Reports No. 2206, 2013, pp. 67-79 

[3] Welter, R.; Lang, V.; Wolf, B.: Clutch Operation; 
9 th Schaeffler Symposium, 2010, pp. 61-74 

[4] Zink, M.; Hausner, M.; Welter, R.; Shead, R.: 
Clutch and Release System; 8 th LuK Sympo¬ 
sium, 2006, pp. 27-45 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 




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113 


The Clutch Comfort Portfolio 

From supplier’s product 
to equipment criterion 


Juergen Freitag 
Dr. Martin Haessler 
Steffen Lehmann 
Christoph Raber 
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Christoph Wittmann 


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114 


Introduction 


Buying a new automobile ranks as one of the 
most expensive single expenditures for pri¬ 
vate households. It therefore goes without 
saying that the emotions and expectations 
associated with such a purchase are quite 
high. These expectations are never fixed and 
rigid, however, but are constantly changing 
and evolving. Just take a look at history. At 
first, consumers were more than satisfied with 
such vehicles as the Messerschmitt Kabinen- 
roller, Opel Laubfrosch, and Goggomobil be¬ 
cause they enabled personal mobility. As the 
years went by and this newfounded feeling of 
excitement wore off, different consumer pri¬ 
orities emerged in the form of reliability, power 
and performance, and comfort. Once these 
requirements were met, people became more 
and more interested in safety, low fuel con¬ 
sumption, and equally low emissions while 


also expecting new developments and fea¬ 
tures in other areas. 

As technical developments frequently 
have a mutual influence on each other, inno¬ 
vations realized for one component more of¬ 
ten than not necessitate adaptations to other 
systems. The same applies to automotive 
clutch systems, which greatly facilitate driving 
comfort and convenience. Increased torque 
or ignition pressure in the engine, for example, 
leads to more pronounced axial vibrations 
along the crankshaft. To ensure that this in¬ 
herent tendency does not compromise the 
driving experience by inducing strong pedal 
vibrations, high pedal forces, or creating dis¬ 
turbing noise levels, the clutch systems in¬ 
stalled must be adapted accordingly. Figure 1 
shows a graph of the targeted areas, or sweet 
spots, targeted for achieving comfortable 
pedal forces and depicts a selection of differ¬ 
ent clutch designs that can be incorporated to 
approach these areas, depending on the 
amount of engine torque available. 


I6O-1 



80 -F -^-1 - 1 -^- 1 - 1 - 1 

100 200 300 400 500 600 

Max. engine torque in Nm 


Figure 1 Excerpt from the product portfolio of clutch pressure plates 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3 7, © The Author(s) 2014 





Clutch Systems 


7 


115 


Comfort 


A clutch pedal should not only be comfortable 
and convenient to operate, but also fulfill other 
design criteria such as complying with de¬ 
fined levels of vibration and manual shifting 
force when a gear is engaged as well as 
reliably withstanding extreme loads. Another 
requirement that is equally as important is en¬ 
vironmental compatibility. More recent devel¬ 
opments to this end include start-stop systems 
and clutches for hybrid and fully electric pow¬ 
ertrains. Extending beyond the range of tradi¬ 
tional performance applications, Schaeffler 
now also offers innovative designs for clutch 
assemblies used in motorcycles. 


Main spring 
adjustment 

Adjustment 

unit 

Ajustment 

ring 


Basis 



Release travel 



Comfort of actuation 

The current drive of automakers to leverage 
the concepts of downsizing, downspeed¬ 
ing, increased boost pressures, and trans¬ 
missions designed to reduce internal friction 
in an effort to minimize C0 2 emissions ne¬ 
cessitate a clutch system that is even more 
robust and resistant to axial vibrations ex¬ 
perienced along the crankshaft. In addition 
to realizing the stability required to harmo¬ 
nize the characteristic curve for pedal ap¬ 
plication forces, topics such as pedal vibra¬ 
tion, gear rattle, and judder are becoming 
increasingly important. 

By developing a clutch that utilizes trav¬ 
el-controlled wear adjustment (travel-ad- 
justed clutch, or TAC) [1] as an alternative to 
the classic, proven clutch design integrating 
a force-controlled wear adjustment mecha¬ 
nism (self-adjusting clutch, or SAC), it is 
possible to reduce pedal application forces 
despite higher engine torque outputs while 
at the same time catering to changed vehicle 
constraints. 

The TAC also facilitates a more flexible 
selection of plotted performance curve 


Figure 2 Clutch with travel-controlled wear 
adjustment and its optimization 
potential 

characteristics as well as lends itself to a 
high level of operational stability. Adding to 
this is the fact that the robust nature of TAC 
assemblies when it comes to resisting axial- 
based vibrations makes it possible to re¬ 
duce the torque transfer capabilities of the 
clutch to enable lower contact pressures 
and, in so doing, lower the release and ped¬ 
al pressures required for identical maximum 
torque ratings. 

Due to the pronounced active adjust¬ 
ment characteristics of the TAC, engi¬ 
neers can also enhance performance 
curves in relation to the overall system as 
it interacts with the pedal system or imple¬ 
ment additional componentry to further 
reduce pedal forces by up to 40 percent 
(Figure 2). 

As a result of the adjustment system 
used, the tongue height and operating range 
of the clutch remain constant. This, in turn, 
makes the TAC an ideal partner to combine 
with a cover fixed release system, or CFR [2]. 
Coordinating and harmonizing these two 








116 



Standard design 


Figure 3 


components allows 
NVH performance 
to be noticeably im¬ 
proved. To this end, 
the CFR eliminates 
pedal vibration and 
judder, since the 
clutch and clutch re¬ 
lease bearing are no 
longer braced by 
the transmission 
and can instead 
freely oscillate in the 
transmission bell 
housing to prevent 
the axial vibrations 
of the engine and 
integrated clutch 

system from transitioning to clutch modula¬ 
tion. 

The design configuration of the CFR as 
an easily adaptable ancillary component 
even makes it possible to use the TAC in 
conjunction with a conventional release sys¬ 
tem early in the development phase. Should 
it then be determined later on, right before 
the start of production (SOP), that undesir¬ 
able noises will have to be eliminated (as is 
often the case), minor adjustments can be 
made to the TAC to align it with the CFR. 
This introduces a whole new dimension of 
flexibility, which is further enhanced by the 
fact that the CFR is designed in modular 
design. As such, the CFR can also be used 
for different sizes of the TAC assembly 
(Figure 3). 

A servo-spring clutch can also be fit¬ 
ted as an alternative option for improving 
comfort levels. This development closes 
the gap between conventional and self- 
adjusting systems. Normally, the release 
force of a conventional clutch increases 
as the lining continues to degrade over 
time due to the characteristics of the dia¬ 
phragm spring. This effect is counteract¬ 
ed by the servo spring clutch as a result of 
an additional servo spring that overlays 


Adapter 



Integrated CFR 


Clutch with travel-controlled wear adjustment and cover fixed 
release system 


the characteristic curve of the diaphragm 
spring in such a way that a less pro¬ 
nounced difference in force is encoun¬ 
tered between the as-new and worn 
states. 


Diaphragm spring 
Servo-spring 



Release travel 

■ Wear without servo spring 

■ Wear with servo spring 

■ New condition 

Figure 4 Clutch with servo spring support 






Clutch Systems 


7 


117 


This, in turn, reduces the maximum level of 
release force as compared to a convention¬ 
al system that does not have an additional 
spring and minimizes the maximum pedal 
application force required by up to 20 per¬ 
cent across the entire service life of the as¬ 
sembly (Figure 4). Servo spring clutches are 
particularly well suited to applications in 
which a conventional clutch can no longer 
meet the target comfort requirements that a 
self-adjusting clutch system can more than 
fulfill. 


Comfort of launch 

In an effort to improve launch comfort, 
several passenger car powertrains have 
been realized with a judder damper inte¬ 
grated in the clutch disk since 2011. Char¬ 
acteristic for this product is not only the 
correct adjustment to the natural frequen¬ 
cy, but also a friction level between the 
damper mass and the mass to be damped 
that is directly proportionate to the twist 
angle. The result is that the oscillatory en¬ 
ergy that increases as a square of the fluc¬ 
tuations observed in transmission speed is 
optimally dampened [3]. 

In today’s series production versions, 
compression springs that target tangential 
forces coincide with the torsional rigidity of 
the judder damper. A ramping mechanism 
between the damper mass and a friction 
element generates the friction proportion¬ 
ate with the twist angle, while the force of a 
separate diaphragm spring as it contacts 
the ramps and axial support of the damper 
mass produces the corresponding friction¬ 
al torque. 

An alternative setup to this design 
would be to utilize the available tangential 
compression springs to produce this 
torque directly. In one such judder damper 
that has already entered its second gen¬ 
eration, the diaphragm spring is then no 
longer needed. As a result, fewer compo¬ 



1 st generation 2 nd generation 



Figure 5 First-generation (left) and second- 
generation (right) judder damper 

nents and less installation space are re¬ 
quired to provide the same level of func¬ 
tionality. In order for this to be possible, 
the ramping mechanism previously ar¬ 
ranged in parallel with the compression 
springs has been redesigned to connect 
them in series. When the damper mass is 
deflected against the clutch disk - with 
corresponding deflection of the compres¬ 
sion springs acting in the circumferential 
direction - the force associated with it 
produces an axial force by way of a 
wedge-shaped contact whose intensity is 
defined by the wedge angle. This axial 
force generates a frictional torque at the 
contacts of the axial support points that is 
proportionate to the torsional moment 
and either increases or decreases it, 
depending on the direction of motion. 
Figure 5 compares a first and second- 
generation judder damper. 























118 


Vibration isolation 

Disturbing noises are among the most fre¬ 
quent complaints made with respect to 
new vehicles. It is often difficult to localize 
these noises because they can have many 
culprits. In the case of the powertrain, for 
example, speed irregularities of the com¬ 
bustion engine can excite torsional vibra¬ 
tions. Resonance frequencies and low en¬ 
gine operating speeds in particular cause 
vibrational output to be perceived as 
bothersome. 

Torsional dampers in clutch disks con¬ 
nected to a rigid flywheel minimize the reso¬ 
nance of vibration amplitudes as a result of 
their friction-damping characteristics but at 
the same time can only isolate the vibrations 
experienced in different speed ranges to a 
limited extent. 



Engine speed 


H Engine 

■ Transmission with 

conventional clutch disk 
Transmission with centrifugal 
pendulum disk 

Figure 6 Clutch disk with centrifugal 
pendulum 


A new design approach involves moderniz¬ 
ing the principle of the centrifugal pendulum 
for application on the torsion-damped 
clutch disk (Figure 6). Simulation exercises 
and vehicle trial testing reveal that vibrations 
can be isolated across a wide range of en¬ 
gine speeds when such a setup is used. 
Instead of generating heat by dampening 
friction levels, the centrifugal pendulum 
uses in-phase inertia forces to reduce fluc¬ 
tuating engine speeds more effectively and 
efficiently. 

Further details about this innovation are 
explained in [4]. 


Shifting comfort 

The ease with which gears are shifted is a 
telltale sign of the quality of modern manu¬ 
al transmissions. Achieving this effect fre¬ 
quently poses an inherent conflict to de¬ 
signers, however, who must balance the 
integration of additional components such 
as the centrifugal pendulum and judder 
damper, which improve NVH comfort lev¬ 
els but also increase the mass moment of 
inertia to be synchronized. This increased 
inertia not only leads to higher transmis¬ 
sion synchronizing loads, but also requires 
more effort and time on the part of the 
driver to shift gears. The answer therefore 
lies in systematically optimizing all compo¬ 
nents of the clutch disk with the end goal of 
minimizing the mass moment of inertia as 
far as possible. 

Optimization measures at the design 
level serve as the perfect starting point. 
Up to now, the sheet metal parts of 
clutch disks have primarily been designed 
with functional performance aspects in 
mind. As such, areas that make poor use 
of material and thus offer the potential to 
reduce mass can be found with relative 
ease. By leveraging FE analytical tech¬ 
niques, these components can be opti¬ 
mized to such an extent from a bionic per- 




Clutch Systems 


7 


119 



► 


Figure 7 Clutch disk with reduced mass 
moment of inertia 

spective that areas that do not contribute 
to operative functionality or that are sub¬ 
jected to only minimal loads are removed 
(Figure 7). 

Further reducing the mass moment of 
inertia allows a cushion deflection system 
to be constructed out of single segments of 
thin spring steel. The resulting thinner lining 
structures can then be further optimized 
with respect to the wear reserves or strength 
required in the target application. Design 
measures can also be implemented for the 
centrifugal pendulum or judder damper 
themselves. 

The combined effect of these mea¬ 
sures in turn make it possible to maintain 
the mass moment of inertia of a clutch 
disk with centrifugal pendulum or judder 
damper at the level of current clutch 
disks. Without these additional damping 
elements in place, it would even be 
conceivable to undershoot this level 
(Figure 8). 




Comfort at high stress 

When a vehicle is driven along mountain 
passes, it is much more likely for the clutch 
assembly to overheat, especially under ext¬ 
reme circumstances such as repeated hill 
starts while towing a trailer or due to clutch 
misuse. In some cases, the toll this takes can 
even be smelled! From a technical stand¬ 
point, the thermal deformations that occur 
on the flywheel and the pressure plate at this 
time reduce the effective friction radius and 
induce localized temperature peaks. 

It goes without saying that the clutch 
should offer sufficient performance in ex¬ 
treme situations such as those mentioned a 
certain number of times before friction lev¬ 
els drop so far (fading) that the friction lining 
starts to slip and deteriorate. Better thermal 
resistance can be achieved with systems 
that maintain the target friction radius and 
friction coefficient constant for as long as 
possible under a wide variety of operating 
conditions. This is why cushion deflections 
systems that have a high compensatory 
capacity were developed. 


Figure 8 


Current level 
Shiftability 

Potential for reducing the mass of 
clutch disks 



Centrifugal 
pendulum 
clutch disk 


Optimized 
for inertia 


Conventional 
clutch disk 
















120 


Standard wave form 



Optimized wave form 1 



Optimized wave form 2 




— High 
progressivity 

— Reduced 
progressivity 

-- Initial gradient 


Distribution of pressure 
between potted plates 



High Reduced 

progressivity progressivity 


Figure 9 Cushion springs with high compen¬ 
satory capacity 


The increasing sensitivity of vehicles when it 
comes to dealing with fluctuations in torque 
resulting from the slipping clutch (judder ef¬ 
fect) necessitates a cushion deflection char¬ 
acteristic that has a small initial gradient. For 
this purpose, spring elements made from thin 
steel are typically used. Already when sub¬ 
jected to forces below the maximum clamp¬ 
load the elements are pressed completely flat 
and show a high level of progressivity in this 
range with almost zero spring travel. The 
problem with this design is that these ele¬ 
ments are relatively incapable of counteract¬ 
ing thermal deformation of the flywheel and 
pressure plate. Pressure distribution mea¬ 
surements taken under high-load conditions 
with a deformed pressure plate confirm this. 

Developing specific wave forms for the 
thin cushion spring elements resolves the 
conflict of realizing the small initial gradient 


required while providing for high compensa¬ 
tory capacity. The wave forms are designed 
in such a way that when a defined spring 
travel position is reached, additional waves 
that summon much more energy are activat¬ 
ed, for a combined effect. This, in turn, leads 
to a performance curve with substantially less 
progressivity and a higher compensatory ca¬ 
pacity as maximum clamp-load is reached. 
Pressure distribution measurements taken 
under a load in the presence of a deformed 
pressure plate attest to this improved design 
response, since the friction radius is held 
largely consistent. The results of hill-start tests 
conducted in real-world conditions under¬ 
score the potential of this concept. 

Without requiring any additional space or 
increasing the mass moment of inertia, the 
high-capacity cushion deflection elements en¬ 
hance the thermal durability and power trans¬ 
fer capabilities of the clutch (Figure 9). 

To improve load capabilities and launch 
comfort in the aforementioned situations, 
Schaeffler is also currently developing new 
organically-bound friction materials for 
strip-wound linings. The target objective for 
these constant-p linings is not so much to 
achieve as high a friction coefficient as pos¬ 
sible, but to realize one that is largely con¬ 
sistent (Figure 10). 

The thinking behind this strategy is that by 
minimizing changes in the friction coefficient of 
the lining across a wide range of operating con- 



— Today 
—* Tomorrow 

Figure 10 Constant-|j lining 



















Clutch Systems 


7 


121 


ditions and parameters while sustaining an un¬ 
wavering average performance value, a higher 
minimum friction coefficient can be attained. 
This not only improves power transfer reliability, 
but also facilitates lower clamp-loads, which in 
turn lead to lower release forces for a given 
clutch with specific rated dimensions and iden¬ 
tical power transfer capabilities. An alternative 
approach is to fit a smaller clutch assembly, 
whose reduced maximum friction coefficient 
limits the amount of torque that can be trans¬ 
ferred and, in so doing, softens peak loads in 
the powertrain under dynamic load conditions. 
Automated clutch systems also profit from the 
design, as a constant friction coefficient makes 
it easier to actively regulate the build-up of 
torque along the engagement and release trav¬ 
el respectively. 


Environmental compatibility 

Automakers are presently looking for any 
and all ways to reduce the C0 2 emissions of 
the models they produce. An optimized 
clutch can help in this regard, since reduc¬ 
ing the mass and mass moment of inertia of 
the assembly further improves the efficiency 
of the overall vehicle. 

To this end, applications could be con¬ 
ceived that involve reducing the mass of the 
pressure plate. The limiting factor here is the 
cast materials that are currently in use, how¬ 
ever. In order to safeguard compliance with 
defined criteria such as burst strength, 
thermal durability, and feasibility from a 
manufacturing perspective, the mass of the 
pressure plate frequently cannot be re¬ 
duced to the theoretical minimum. 

Addressing the issue can take the form of 
higher-grade cast materials to allow these 
performance limitations to be marginally shift¬ 
ed. Manufacturing pressure plates from rolled 
steel offers greater potential, however, since a 
steel plate design gives rise to new design 
configurations that leverage closer tolerances, 
thinner cross sections, and increased durabil- 



Pressure plate Pressure plate in 

in cast iron formed sheet metal 



Mass Mass inertia Bursting strength 


■ With pressure plate in cast iron 

■ With pressure plate in sheet metal 

Figure 11 Pressure plate made from rolled sheet 
steel in comparison to a cast variant 

ity to make better use of available installation 
space while reducing mass and the mass 
moment of inertia (Figure 11). 


Comfort at engine startup 

With the advent of an ever larger number of 
new vehicles equipped with start-stop sys¬ 
tems comes the requirement to find solu¬ 
tions that allow the engine to restart with little 
to no delay. In response to this development, 
the last Schaeffler Symposium was used as 
a venue to present a new sprag clutch de¬ 
sign for a permanently engaged starter as¬ 
sembly, or PES [5]. The benefit of this con¬ 
cept is that the starter drive pinion no longer 
has to be engaged. As a result, combustion 
engines can be started and stopped faster, 
quieter, and with less wear from a standstill 
as well as when coasting to a stop. The con¬ 
cept-bound lifting motion of the sprags after 

















122 


startup, which is controlled using centrifugal 
force, is completely void of friction through¬ 
out the entire operating range, thus allowing 
the potential of a start-stop system to reduce 
C0 2 exhaust emissions to be maximized. 

Since the Symposium, the design effect has 
been investigated using vehicle demonstrators 
and the system further enhanced. By improv¬ 
ing the operating direction of the spring used to 
generate the lift movement, it was possible to 
reduce the contact force present throughout 
the sliding process during freewheel overrun. 
The usable wear volume of the sprags was 
also increased and the wear properties of the 
friction partners optimized. The combined ef¬ 
fect of these measures is good for around one 
million starts, a performance benchmark that 
was verified on an actual combustion engine 
(2.0-liter diesel). The positive impact of the con¬ 
cept on the wear exhibited by the starter ring 
gear was confirmed as well. If manufacturers 
experience a heightened need for this configu¬ 
ration, the PES could be used not only in the 
start-stop systems of combustion engines, but 
also in the repeat-start systems designed for 
hybrid applications (Figure 12). 


Electrification 

As the powertrains in modern automobiles 
become increasingly electrified, Schaeffler is 
currently in the process of developing an 
electrically operated clutch. One of the de¬ 
sign objectives of this project is to keep the 
actuation energy as low as possible. The un¬ 
derlying operation of the electrical integrated 
actuator clutch (elAC) is based on the boost¬ 
er principle [1] and encompasses a pre-con¬ 
trol and a main clutch unit (Figure 13). 

Booster clutches generate contact pres¬ 
sure by producing a minimal pre-control 
torque that is converted into an axial force by 
a ball ramp system. With this design, the pre¬ 
control element can be realized by a small 
conventional clutch or an electrically operat¬ 
ed variant. Options here include a magnetic 
or solenoid clutch and an eddy current 
brake. The energy required to close the 
clutch assembly can be taken from the 
powertrain itself. 

Future applications for the elAC involve 
hybridized platforms whereby the clutch, which 
is fitted inside a ring-shaped electric motor, is 




New sprag 
concept 


Permanently 
engaged starter 

Dry sprag clutch 

Ring gear mounted 
to housing 


on crankshaft 


Sprag and 
ring gear after 
10 6 Starts 



Figure 12 Freewheel for permanently meshed starter assembly (PMSA) 










Clutch Systems 


7 


123 


Ball ramps 

Planetary 
gear set 
Eddy current 
brake (ECB) 



— Eddy current brake torque 
-- Starting torque 


Figure 13 Drive clutch with electrical actuation 

called on to mechanically link the combustion 
engine with the powertrain as required. 

When the vehicle is operated in electric 
mode only, the elAC is actuated to disen¬ 
gage the engine from the rest of the pow¬ 
ertrain as efficiently as possible. To this end, 
the system is designed with a “normally 
open” configuration. 

As the combustion engine is started via 
the electric motor, the elAC can be actively 
closed very quickly using an eddy current 
brake. Since this brake is wear-free by de¬ 
sign, the torque transferred can be regulat¬ 
ed with exacting precision across the entire 
service life of the clutch. 

To facilitate a smooth transfer of torque 
to the powertrain while the engine is run¬ 
ning, a freewheel is used as a pre-control 
element. Part of the torque generated by 
the engine is siphoned off over the one-way 
clutch to close the clutch. 

One of the benefits of the electrical inte¬ 
grated actuator clutch is the accurate control of 
overrun torque with minimum response time as 
afforded by the eddy current brake. This per¬ 
formance can be maintained throughout the 


entire service life of the unit, since the brake is a 
wear-free assembly. In addition, no energy is 
required to actuate the clutch when the vehicle 
is driven in electric mode or together with the 
combustion engine, thereby realizing the op¬ 
erative conditions of a “normally stay” clutch. 

When suitable pre-control elements are 
chosen, the elAC can also be used in other 
applications to: 

- Activate an alternative drive system 

- Couple an additional driven axle 

- Distribute drive force, or driving power 
(torque vectoring) 

- Connect/disconnect other assemblies 


Motorcycle clutches 

Almost four million motorcycles are regis¬ 
tered in Germany alone, with low six-digit 
registration numbers of new models each 
year testifying to the ongoing attraction of 
this form of personal transportation. This 
also applies to many other regional markets, 
although there are pronounced differences 
in what people expect of such machines. 

In Germany, for example, customers 
want a motorcycle that provides a level of 
comfort similar to that of a passenger car. 
Trends in technology are also very apparent 
in motorcycle applications as is the case 
with automobiles. Continually increased 
power densities, the never-ending pursuit to 
minimize mass, and efforts to reduce the 
somewhat excessively high actuation forces 
of certain clutch assemblies are just a few 
examples of improvements being sought 
out in this field. The situation in the south¬ 
east Asian markets could not be more dif¬ 
ferent. There, a motorcycle is simply viewed 
as another form of transportation that 
should offer high everyday practicality more 
than anything else. In this context, the de¬ 
velopment activities that surround motorcy¬ 
cle clutches are almost as multifaceted as 
those observed in passenger car applica¬ 
tions. When appropriate solutions are de- 









124 




■ Competitors 

■ LuK clutch 


Figure 14 Motorcycle clutch for improving 
actuation comfort 


vised, however, it is possible to transition to 
an entirely new level of technology. 

For example, the actuation forces re¬ 
quired to operate a motorcycle clutch can 
be significantly reduced by applying the de¬ 
sign principles of the electrically actuated 
drive clutch to a multi-disk clutch assembly. 
By realizing a modular construction in the 
sense of an interconnected system of build¬ 
ing blocks, engineers can quickly adapt the 
mechanicals as required for different engine 
variants (Figure 14). 

The modular concept of the clutch as¬ 
sembly also lends itself to integrating a func¬ 
tion that limits the engine braking torque 
generated in overrun mode as it is trans¬ 
ferred to the rear wheel. This “anti-hopping” 
function considerably improves driving safe¬ 
ty, since it prevents the motorcycle’s rear 
wheel from losing some or all of its traction. A 
critical aspect in this regard is that the brak¬ 
ing effect generated by the engine, which 
can cause wheel blockage when the vehicle 
experiences a dynamic shift in weight toward 
the front wheel as the rear wheel becomes 


severely unweighted during periods of heavy 
braking combined with quick downshifts, 
must be limited to maintain safe handling 
characteristics. 

Another development angle is to sim¬ 
plify the amplification function of a multi¬ 
disk clutch assembly to greatly reduce the 
forces required to actuate it. Although the 
market currently offers clutches that realize 
this type of amplification using slide ramps, 
the problem with their design construction 
is that the changes in the coefficient of fric¬ 
tion (static, dynamic friction) can lead to 
fluctuations in torque delivery when com¬ 
bined with these ramps. 

An innovative new development from 
Schaeffler circumvents these friction- 
based effects by allowing the torque yield¬ 
ed by the contact pressure in the inner 
cage to be transferred via leaf springs to 
the inner hub. Since these springs have a 
tilting angle, a force amplification or reduc¬ 
tion function is realized with practically no 
friction, similar to an articulated lever, de¬ 
pending on the angle of attack. The leaf 


Primary gearing 
Leaf spring pack 
Release bearing 



Release travel 


■ Competitors 

■ LuK clutch 

Figure 15 Reducing actuation forces with an 
amplification function 









Clutch Systems 


7 


125 


Diapraghm spring 
Support spring 

Release bearing 

Primary gearing 




■ Competitors 

■ LuK clutch 


Figure 16 Concept of a motorcycle clutch for 
the Asia-Pacific region 


springs also center the inner cage and 
apply the contact pressure. 

This concept, which was purposely 
devised with simplicity in mind, requires 
comparably little installation space, and can 
be quickly adapted for different engine vari¬ 
ants thanks to its modular construction. An 
“anti-hopping” function can likewise be inte¬ 
grated if needed (Figure 15). 

The development activities being pur¬ 
sued for a motorcycle clutch targeted for 
the Asia-Pacific region take a completely 
different direction, whereby the key objec¬ 
tives are to optimize operative functionality 
while reducing costs by leveraging Schaeffler’s 
manufacturing expertise in the areas of 
stamping, punching, and metal forming. To 
this end, a diaphragm spring is integrated in 
place of compression springs as an energy- 
storage mechanism to lower release and 
holding load when the assembly is new. At 
the same time, this setup also increases 
stability with respect to centrifugal force. 

An additional compulsory disengagement 
facility is fitted between the clutch disks as a 


further design measure and ensures that the 
disks are ventilated in a uniform, consistent 
manner to minimize drag torque. The modular 
construction of this component also makes 
the clutch a universally compatible assembly. 
Adding to this are the benefits of low weight 
and compact dimensions (Figure 16). 


Outlook 


Although the clutch has over 100 years of 
development behind it, it still offers consider¬ 
able potential to be optimized further. The 
broad and diversified portfolio Schaeffler has 
assembled for clutch-based technologies 
can be leveraged to realize solutions for 
many different applications in the automotive 
and motorcycle industries as future innova¬ 
tions target new design criteria established 
to achieve higher levels of comfort and effi¬ 
ciency while reducing C0 2 emissions. 


Literature 


[1] Freitag, J.; Gerhardt, F.; Flausner, M.; 

Wittmann, C.: The clutch system of the future. 
9 th Schaeffler Symposium, 2010 

[2] Welter, R.; Wittmann, Ch.; Hausner, M.; Kern, A.; 
Ortmann, S.: Deckelfester Zentralausrucker fur 
Kupplungen. VDI report, 2013, no. 2206, 

pp. 67-79 

[3] Hausner, M.; HaBler, M.: Kupplungsscheibe mit 
Frequenztilger gegen Rupfschwingungen. 

ATZ 114, 2012, no. 1, pp. 64 - 69 

[4] Kooy, A.: Isolation in the drive train. 

10 th Schaeffler Symposium, 2014 

[5] Zink, M.; Hausner, M.: LuK clutch systems and 
torsional dampers. 9 th Schaeffler Symposium, 
2010 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 






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128 


Introduction 


Currently, the value added chain of manu¬ 
al transmissions is characterized by the 
fact that major automobile manufacturers 
in the triad (EU, USA, Japan) buy their in¬ 
dividual synchronization system compo¬ 
nents from different suppliers (Figure 1). 
However in new markets, manufacturers 
have for some time preferred to work with 
suppliers who design and develop the en¬ 
tire synchronization system and deliver it 
ready to install. It is becoming apparent 
that the value chain will be reorganized 
along these lines in Western industrialized 
countries too. A key driver of this develop¬ 
ment is the need for lightweight designs, 
which are now increasingly finding their 
way into the powertrain. If transmissions 
are to become lighter and more compact, 
then the subsystems such as the syn¬ 
chronization must become more efficient. 
The solution to this lies in the need for 
components to require less installation 
space and material and be even better 
matched to each other at the same time. 

Schaeffler is prepared for this new situa¬ 
tion. The final module required at the com¬ 
ponent level is the development of efficient 


friction linings for synchronization systems 
and this has already been completed. Typi¬ 
cally, synchronizer manufacturers need to 
limit the size of their systems to the space 
available between the gears to be shifted. 
Schaeffler has additional expertise in the 
design of the connecting components - 
such as the bearings supporting the shafts 
and speed gears - as well as gear teeth in 
general. In addition, there is comprehensive 
power transmission expertise available 
throughout the Group. Thus, from the clutch 
to the transmission output, the power trans¬ 
mission system can be tuned so that from 
a systemic point of view an optimum is 
reached in terms of cost, space, weight and 
gearshift comfort. 

Development and 
manufacture from a single 
source 


Practice has shown that system expertise 
will lead to the best solutions if it is accom¬ 
panied by corresponding expertise at the 
component level. Schaeffler therefore de- 


Current Situation: 


New Situation: 



Figure 1 New requirements in the value added chain of manual transmissions 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_8, © The Author(s) 2014 


























Synchronisation Systems 


8 


129 



Figure 2 Schaeffler develops and manufactures the complete synchronization system in-house 


velops and manufactures all synchroniza¬ 
tion system components exclusively in- 
house. These include (Figure 2): 

- Selector sleeve 

- Presynchronization detents 

- Selector hub 

- Clutch body 

- Ring package incl. 

- Carbon-based friction linings 

With the development of friction linings, 
Schaeffler has completed its product port¬ 
folio and is now able to offer a complete 
synchronization system from a single 
source. This vertical integration is unique 
worldwide. When it comes to friction linings, 
it ranges from the selection of raw materials 
to the development and manufacture of the 
friction material and its attachment to the 
carrier. 

However, the aggregation of optimized 
components often does not result in an 
optimum overall system. The development 
history of the selector hub is one such ex¬ 
ample: By reducing the size of the detent, 
the component can in principle become 
more robust and in a further step narrower. 
Converting the manufacturing process 
from sintering to metal forming allows an¬ 
other reduction in size and weight. How¬ 
ever, if it weren’t for the development of the 


strut-in-sleeve design described in greater 
detail below, this step would lead down a 
dead end. A number of lightweight effects 
would remain out of reach since the cor¬ 
rect function could no longer be guaran¬ 
teed. 

Other potential benefits can be 
tapped if the system expertise extends 
beyond the synchronization unit. The ad¬ 
vantages of such an approach have al¬ 
ready been demonstrated in specific 
customer orders. 

Opportunities resulting from the 
systems approach 

A practical example 

Economic considerations prompted the 
customer to use a clutch disk with a cen¬ 
trifugal pendulum-type absorber from 
LuK. However, this would have led to a 
level of gearshift comfort that would have 
been classified as not in line with market 
expectations (Figure 3). The reason for 
this was a 45 % increase in mass inertia at 
the transmission input, which would have 
led to increased gearshift forces if not 
countered. The default parameters of 
driver behavior, gearshift system, trans- 


130 


mission temperature level and the type of 
transmission oil to be used had to be tak¬ 
en into account. 

Therefore, the development work focused 
on two main approaches: 

- New tooth geometry of the selector 
sleeve and clutching teeth 

- Verification of the friction lining and 
modification of the cone geometry and 
friction system 

The measures taken regarding the first 
field of work were so effective that the ex¬ 
isting friction lining was retained and a 
satisfactory overall result achieved. The 
testing and evaluation of the modified 
transmission took place both at Schaeffler 
and at the customer’s premises. In par¬ 
ticular, the operating life and the gearshift 
behavior were investigated and evaluated 
in detail. The simulation had already indi¬ 
cated that the new gearshift curve would 
be much more harmonious and this result 
was confirmed during the test stand trial 
(Figure 4). The comparison of the opti¬ 




Figure 3 Optimization of gearshift comfort, 
matched to a clutch disk with 
centrifugal pendulum-type absorber 

mized transmission with the standard 
transmission in the vehicle finally corrobo¬ 
rated that the optimization makes itself felt 
not only on the data sheets: The perceived 
gearshift comfort achieved a better value 
on the ATZ rating scale than the target 
specified by the customer. 


Mass production (5.6 g m 2 ) 
Mass prod. w. CPA (8.1 g m 2 ) 


Optimized with CPA (8.1 g m 2 ) 


Customer 

target 

7.25 

6.5 

7.75 


Perceived gearshift comfort (ATZ scale) 



Ic 

CO 



Optimized, with CPA 


2. Gear 


1. Gear 


Shift travel 


Figure 4 Measurement of the gearshift gearshift curve before (above) and after. The perceived 
gearshift comfort was improved despite higher mass inertia. 






























Synchronisation Systems 


8 


131 


Carbon-based friction 
linings developed and 
manufactured by Schaeffler 


Requirements to minimize the transmission 
design envelope and weight, and to offer 
increased gearshift comfort and higher 
power density require comprehensive, opti¬ 
mized synchronization systems. 

In double clutch transmissions, the 
high performance requirement results 
from the skipping of gears: The synchro¬ 
nization system must compensate for a 
speed difference that normally does not 
occur in manual transmissions - and all 
this within a very short time. Drivers 
expect much quicker gearshifts with 
automized transmissions than they them¬ 
selves could manage. Such operating 
conditions require high-quality synchroni¬ 
zation systems which usually feature car¬ 
bon-based friction linings. Schaeffler has 
already presented such a material with its 
“Friction Pad System”. After further devel¬ 
opments, the new STC 300 friction lining 
is now available. The acronym STC stands 
for “Schaeffler Technologies Carbon” and 
refers to the base material. 

The second new friction lining, STC 600, 
is based on this material too, however, 
it is a completely new development de¬ 
signed to meet the most exacting re¬ 
quirements. 

STC 300 - Carbon-based composite 
friction material 

STC 300 is manufactured according to the 
method of molded friction material - a man¬ 
ufacturing technology in which Schaeffler 
Friction Products has been proficient for 
many years now. The friction lining is made 
of a composite of carbon and other materi- 



Figure 5 STC 300: Friction lining made of 
carbon-based composite material 

als, which are bound by resin (Figure 5). 
Schaeffler has developed and industrialized 
the production process. STC 300 offers sig¬ 
nificantly enhanced friction coefficient sta¬ 
bility and wear characteristics compared 
with brass and bronze-sintered products, 
whilst having a similar cost level. 

STC 600 - Carbon fiber friction 
material 

STC 600 is a carbon-based friction lining of 
the highest performance class. The lining is 
manufactured using a process derived from 
paper production (Figure 6). This manufac¬ 
turing process, which was also developed 
by Schaeffler, offers significant cost benefits 
compared with woven material and pro¬ 
vides equal and in some cases even better 



Figure 6 STC 600: Premium class carbon- 
based friction lining 




132 


STC 600 Schaeffler 

A high-performance 
™ carbon-based friction lining 

A 

rat-hnn STC 300 Schaeffler molded 

/~\ carbon-based 
friction lining 


Molybdenum 

“ o 

Sinter bronze 


o 

Brass 


higher Cost rating 


lower 


Figure 7 Performance of different types of 
friction lining in relation to costs 


results compared with products of the same 
performance class (Figure 7). 

Performance 

STC 600 friction lining achieves excellent 
results in all relevant fields, such as consis¬ 
tent friction coefficient characteristics through¬ 
out the period of use, friction coefficient 
gradient and friction level within one gear¬ 
shift operation as well as wear resistance. 
STC 600 is highly robust and can perma¬ 
nently sustain a high level of friction energy. 
This is demonstrated by the measurement 
and test results in absolute terms and in 
comparison with products that are current¬ 
ly leading in the market. These are present¬ 
ed in more detail below. 


Dynamic friction coefficient 

The speed at which the friction lining in con¬ 
tact with the synchro ring builds up the fric¬ 
tion coefficient, as well as the friction coeffi¬ 
cient curve during the gearshift operation 
both have a substantial effect on how the 
gearshift comfort is perceived by the driver. 
Ideally, the friction coefficient rises sharply to 
its maximum level and remains constant until 
the transmission shaft and the gear are syn¬ 
chronous. In practice, the maximum friction 



— Friction coefficient 

— Rotational speed 



Shifting time in s 

— Friction coefficient 

— Rotational speed 


Figure 8 Comparison of unfavorable (below) and 
good friction coefficient curves (above) 

coefficient is reached only gradually, and af¬ 
ter the displacement of the oil in the contact 
gap. The lower graph in Figure 8 shows an 
undesirable example: Such a friction coeffi¬ 
cient curve either can no longer ensure the 
proper function or it negatively affects the 
gearshift comfort. The measurements show 
that the actual friction coefficient curve of 
STC 600 differs only slightly from the ideal 
curve (Figure 8 above). Among other factors, 
this result is due to the excellent drainage ca¬ 
pacity of STC 600 (Figure 9). 

Friction coefficient level 

The coefficient of friction is decisive for the 
maximum achievable friction performance. 
The increased friction coefficient leads to 
higher friction performance, which means 

















Synchronisation Systems 


8 


133 



Figure 9 Surface structure of STC 600 friction 
lining 

that the synchronization can occur within a 
shorter time. Values of 0.11 and above indi¬ 
cate that the STC 600 friction lining is a 
high-end product (Figure 10). Thus the fric¬ 
tion material contributes to an increase in 
power density. Thanks to its high load ca¬ 
pacity - both in absolute terms and relative 
to the benchmark - it is possible to reduce 
the necessary contact area and thus short¬ 
en the entire synchronization system. The 
gain in design envelope is added across all 
gear combinations and leads to a more 
compact and simpler transmission design. 

Another indicator for quality is friction 
coefficient stability. This has practical rele- 



- STC 600 

— Benchmark 



Shifting cycle 
— STC 600 
— Benchmark 

Figure 11 Low spread of coefficient of friction 
during the operating life 

vance in that the gearshift feel remains the 
same over time. With STC 600, the start 
level remains practically unchanged over 
the entire operating life (Figure 11). The 
curve can be interpreted as a successful 
development outcome because a high fric¬ 
tion coefficient and high friction stability are 
achieved at the same time. Technologi¬ 
cally comparable materials statistically 
show a friction coefficient level of low unifor¬ 
mity over their operating life. In comparison 
with the materials commonly used on the 
market it is clear that the STC 600 friction 
lining is superior to those in particular in 
terms of friction coefficient gradients and 
curves (Figure 12). 



— Benchmark 1 — Benchmark 3 

— Benchmark 2 — STC 600 


Figure 10 Convincing coefficient of friction in 
comparison with the benchmark 


Figure 12 Friction coefficient gradient of 
various materials in comparison 
















134 


Oil sensitivity and wear 

Depending on the oil, a friction material 
shows different behavior, in particular with 
respect to friction coefficient level and 
wear. A friction material is ideal from a 
customer’s perspective, if it is equally ef¬ 
ficient in all criteria in conjunction with any 
oil. In practice, this has not yet been 
achieved. When selecting the transmis¬ 
sion oil, the primary focus is not usually on 
optimizing the gearshift comfort, but on 
protecting the gear teeth against wear 
and minimizing drag losses. 

STC 600 friction lining shows relatively 
low sensitivity to the oils tested to date 
(Figure 13). The next development stage 
involves the extension of potential appli¬ 
cations, for example, to a preferred type 
of transmission oil in a specific applica¬ 
tion. Schaeffler prefers to take this step 
hand in hand with the customer to ensure 
the best possible result. 

When it comes to wear, STC 600 fric¬ 
tion lining performs significantly better 
than the benchmark: Under the given ex¬ 
perimental conditions and depending on 
the oil used, the necessary wear reserve 
for STC 600 needs to be only half as large, 
so that less installation space is required 
(Figure 14). 


^ 0.16 

c 

0 ) 

| 0.12 
o 

8 0.08 

c 

o 

% 0.04 - 


ip** 


0.1 0.15 0.2 0.5 0.3 0.35 


Shifting time in s 
— Esso Gear Oil 
- Mobil Oil FE75W 
Burmah MTF95 

Figure 13 Low oil sensitivity of STC 600 


"I 

0.4 



Shifting cycle 

- STC 600 
Benchmark 


Figure 14 Low wear over the operating life 


Results in overview 

STC 300 and STC 600 have been de¬ 
signed for two different product catego¬ 
ries. Both were developed by Schaeffler, 
starting with the selection of raw materi¬ 
als through to the finished product in¬ 
cluding the manufacturing processes, 
and they are manufactured using the 
company’s own machines exclusively. 
The linings are positioned in different 
performance classes, but they all share 
the same carbon-based friction material. 
STC 300 offers higher performance in re¬ 
lation to friction coefficient stability and 
wear performance compared to brass 
and bronze-sintered products, but it 
comes at similar cost. 

STC 600 achieves better results with 
regard to the essential criteria of dynam¬ 
ic friction coefficient, friction coefficient 
level and stability than the best products 
currently available on the market. In the 
combined analysis of friction coefficient 
stability over the operating life and the 
margin with which the different friction 
coefficients deviate from each other dur¬ 
ing the individual gearshifts, STC 600 
is close to the optimum (Figure 15). Its 
sensitivity to the transmission oils tested 
so far is low. STC 600 also compares 








Synchronisation Systems 


8 


135 


High power application 



t t 

Friction level during life-cycle time 


Figure 15 Overview of performance character¬ 
istics of STC 300 and STC 600 

favorably to benchmark in terms of wear 
resistance. 

Innovative components 
with system impact 


reduced by about 90 g. With a six-speed 
transmission this comes to about 350 g 
- secondary effects at transmission level 
not included. 


Operating principle 

In the conventional design, the rib of the 
selector hub houses the pressure springs 
of the presynchronization detent. This in¬ 
stallation space requires a specific me¬ 
chanical strength which is achieved by 
appropriate material thickness. The fur¬ 
ther development of this basic design is a 
version using flat struts. These reduce the 
required depth of recess in the rib of the 
selector hub, which also reduces the 
stress peaks in the critical cross-section. 
In this way, higher torques can be trans¬ 
mitted with unchanged geometry. And 
vice versa: For an equally high transmis¬ 
sion torque a narrower rib will suffice. 

In principle, this approach would allow 
a narrower design for the entire selector 
hub and thus for the selector sleeve too. 
But since the shift path is a given, this op- 


Smaller design envelope and 
lower weight 

The goal of designing systems and compo¬ 
nents that are as light as possible while 
maintaining the cost targets also applies to 
the transmission and its subsystems. In ad¬ 
dition, the minimization of the required de¬ 
sign envelope is gaining more and more 
importance. The “strut-in-sleeve” concept is 
a big step forward in this regard. The name 
refers to the consistent further development 
of the selector sleeve, detent and selector 
hub and the optimized harmonization of 
these components. If all options are used, 
then the mass of each synchronizer can be 



Figure 16 With conventionally assembled 

detents, the scope for producing a 
narrower selector hub design is 
limited 










136 


tion is limited. This is because when the 
gear is engaged, there is a risk of the de¬ 
tent balls getting stuck since the selector 
sleeve no longer covers them completely 
(Figure 16). 

The strut-in-sleeve concept path paves 
the way towards a narrower selector hub. 
In this case, the strut is not mounted in 
or on the selector hub, but in a recess in 
the internal teeth of the selector sleeve. 
So during the gearshift operation, it 
is now guided by the selector sleeve 
(Figure 17). 

With this innovation, the selector hub 
is no longer impaired in any way so that 
there are no longer any critical cross- 
sections. Now the width of the selector 
sleeve can be chosen freely and the op¬ 
portunity of choosing a narrower rib can 
be fully exploited. This results in a far- 
reaching benefit: Every single synchroni¬ 
zation system is about 2 mm shorter. 
Consequently the gears move closer to 
each other. This, in turn, allows the use of 
shorter shafts and ultimately a shorter 
transmission housing. 



Figure 17 With a strut-in-sleeve selector 
sleeve, the strut is guided in the 
direction of the gearshift 



Figure 18 Selector sleeve with integrated 
detent (strut-in-sleeve) 

System requirements 

The strut-in-sleeve concept can be imple¬ 
mented at no additional cost. The basis for 
this is the selector sleeve manufactured by 
Schaeffler in volume production using form¬ 
ing methods. In contrast to components 
that are manufactured in a metal-cutting 
process from forged blanks, the integration 
of the usual three recesses for locating the 
struts does not require an additional opera¬ 
tion. The recesses are designed so that the 
struts only have to be pushed in (Figure 18). 

Selector hubs for car transmissions are 
now manufactured almost exclusively from- 
sintered metal. Since the selector sleeve 
introduces the torque into the selector hub 
off-center, it is subject to high torsional and 
bending loads. Therefore the selector hub 
has a solid design and weighs several 
hundred grams. New product concepts 
based on sheet steel designs focus on two 
courses of development: One on weight 
optimization and the other on strength 
optimization. 

At the current state of development, the 
strength of the weight-optimized design (Fig¬ 
ure 17) is still in the same range as that of a 
powder metal sintered component. In the case 
of a six-speed manual transmission in the 
350 Nm torque class, the weight advantage 
is about 350 g, which is equivalent to 25 % 
with regard to the synchronization units. 



Synchronisation Systems 


8 


137 



Figure 19 Selector hub of steel (right) 

compared to one made of sintered 
metal (left) 

The strength potential of sheet steel de¬ 
signs can be used to reduce the design 
envelope. As a result, the rib width as well 
as the width of the internal teeth of the 
sheet steel selector hub can be reduced so 
that the slightly higher density of steel is 
over-compensated (Figure 19). 


Transmission savings 


Schaeffler has evaluated the possible sec¬ 
ondary effects of a transmission optimized 
with strut-in-sleeve and improved synchro 
ring packages (Figure 20). For front trans¬ 
verse installations, the transmission hous¬ 
ing is about 8 mm shorter due to shorter 
shafts. Depending on the conditions in the 
vehicle, this gain in design envelope can 
make a difference in compensating for the 
necessary enlargement of other compo¬ 
nents in the engine compartment. The 
weight savings from secondary effects 
alone add up to approximately 450 g. Pri¬ 
mary and secondary effects reduce the 
weight by about 800 g. 

In a longitudinally mounted transmission, 
the overall length is reduced by about 12 mm. 
The secondary effect in terms of weight and 
material cost is roughly equivalent to that of a 
front transverse transmission. 


Weight reduction 
of components: 

• Gear box 
housing 

• Main shaft 

• Secondary shaft 

slij 

• Output shaft 

• Gear wheels 

1 st - 6 th gear 

• Synchronization 

!IS!|* 

Weight reduction 

Front-transverse gearbox 

Secondary 

approx. 450 g 

Primary 

approx. 350 g 

Total 

approx. 800 g 

Design space potential 

-8 mm 



Figure 20 Saving potential regarding weight and design envelope 

























































138 


Summary and outlook 


Customers in growth markets and increas¬ 
ingly in industrialized countries are looking 
for suppliers who offer not only individual 
components but rather complete synchro¬ 
nization systems for their transmissions. 
Schaeffler has thus decided to become a 
system supplier and has complemented its 
product portfolio with the development of 
carbon-based friction linings. It now con¬ 
sists of a sheet steel selector hub, selector 
sleeve, presynchronization detent, ring 
package and gear cone body. Schaeffler 
has developed all these components and 
manufacturing methods and manufactures 
them in-house worldwide. 

In the course of this product range ex¬ 
tension, the company has continuously 
developed the expertise necessary to de¬ 
sign synchronization systems using an 
integrated approach. This involves not 
only the validation of the specific compo¬ 
nent characteristics, but also the func¬ 
tional optimization of power transmis¬ 
sions from the clutch to the speed gear 
including vibration isolation, in vehicle 
tests if necessary. Schaeffler has already 
demonstrated its expertise in this area in 
related projects. 

In system optimization Schaeffler can 
draw on its long-standing expertise in com¬ 
ponent development. Thanks to the high 
degree of vertical integration, the compo¬ 
nents can be precisely matched to each 
other so that an optimum result is achieved 
at the system level. Concepts such as strut- 
in-sleeve combined with a selector hub 



Figure 21 Further potential for reducing the 
design envelope can be tapped by 
including the freewheel and bearing 
support in the optimization of the 
synchronization system. 

made from sheet steel illustrate the poten¬ 
tial: The secondary effects of a shorter syn¬ 
chronization system result in a more com¬ 
pact and lightweight transmission. 

The aim of a smaller design envelope 
will be given even more attention in the fu¬ 
ture. A development approach that extends 
beyond system boundaries opens up the 
opportunity of achieving even better results 
than the ones described above - for exam¬ 
ple, if the speed gear is included in the opti¬ 
mization of the synchronization system (Fig¬ 
ure 21). However, designs that extend far 
beyond those currently encountered on the 
market require new solutions for the bear¬ 
ings and gears. Therefore, the company’s 
combined expertise in the area of tooth sys¬ 
tems and bearings is gaining increasing im¬ 
portance in the further development of syn¬ 
chronization systems. 


Open Access. This chapter is distributed under the terms of the Creative Commons 
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and reproduction in any medium, provided the original author(s) and source are credited. 



140 


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Dr. Christoph Brands 


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142 


tomotive and supplier industry. To relieve some 
of the pressure exerted by these processes, 
virtual product development has become an 
essential component of design and engineering 
work. When incorporated in time, technical cal¬ 
culations can facilitate quick response in 
many different areas and thereby effectively 
shorten development cycles and times. 

Very few technologies have made as great 
of an impact on product development pro¬ 
cesses as the move toward digitalizing pro¬ 
cess flows. In the process, the only aspects 
that have changed are the tools used and the 
procedures followed. The core development 
tasks for engineers remain the same. Figure 1 
shows the main tasks associated with making 
technical calculations, which include: 

- Analyzing and modeling the system 

- Carrying out the steps involved in the anal¬ 
ysis (i.e. the actual calculation exercises) 

- Deriving design and concept proposals 
from the results obtained 


Development Product and manufacturing 

order technology ready for volume production 



Figure 1 Tasks associated with carrying out technical calculations during the product development phase 

Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 

DOI 10.1007/978-3-658-06430-3_9, © The Author(s) 2014 


Introduction 


Ever shorter development times, an increased 
range as well as continually more complexity 
and diversity of engine systems, are the mega¬ 
trends driving the value-added chains in the au- 





















































Simulation Engine Systems 


9 


143 


System Analysis and 
Modeling 


Modeling always revolves around simplify¬ 
ing the properties and characteristics of the 
real, physical system. A simulation only pro¬ 
vides results that relate to the model at hand 
and its performance limitations. Therefore, 
in order to ensure that the work carried out 
at this time can be transferred as required, 
steps must be taken to verify that the model 
(depending on the level of detail it incorpo¬ 
rates) exhibits the right tendencies and 
leads to the correct quantitative results 
when used under defined system limits as 
performance parameters are varied in real 
time. This validation is typically realized by 
drawing comparisons with trial testing data. 
Different models with varying levels of detail 
can be created depending on the knowl¬ 
edge acquired about the real system. If the 
physical correlations of the real system are 
not known, a mathematical model can at 
least be made that provides corresponding 
results for changes to input parameters. In 
the most basic scenario, this takes the form 
of a regression of existing trial testing data 
(data-based modeling). 

If the physical correlations are known, 
however, and sufficient, reliable input data is 
available, analytical approaches or complex 
physical models can be devised and solved 
numerically. 

Unlike the realm of natural science, 
engineering involves simplifying matters 
and concepts on a daily basis. This ab¬ 
straction or simplification is a key tool 
used to systematically approach complex 
systems. A good example of this is the 
pendulum as considered in the context of 
a point mass with respect to a thread that 
has no mass. When small angles are ob¬ 
served, sin (<D(t)) = 0(t) is then used by way 
of the Taylor approximation. Modelling 


also involves accepting the risk of incom¬ 
pleteness. To this end, when system anal¬ 
yses are carried out, all required effects 
must be identified and factored into the 
model so that technical and design-relat¬ 
ed questions can be answered. In order to 
safeguard reasonable calculation times 
with respect to numerical computability 
(stability), the model must also not be 
overloaded with performance data. After 
all, the right model must be processed us¬ 
ing the right tool, depending on the ques¬ 
tion to explore. 

One of the main tasks involved in techni¬ 
cal calculation exercises is therefore to con¬ 
duct an initial system analysis and create a 
model to establish a baseline. The next step 
is to check the plausibility of the external 
and internal input data available and to pro¬ 
vide this data in a suitable format for the 
simulation as required. This includes geo¬ 
metric data from CAD systems as well as 
functional data such as plotted rigidity 
curves. Checking and verifying the input 
data is critical, since every result obtained 
directly correlates with the quality of the 
data itself. All process steps must be ac¬ 
companied by a defined change manage¬ 
ment policy such that when geometric or 
other relevant data is changed, this is com¬ 
municated appropriately. 


External Influential Factors 

As an integral part of the product develop¬ 
ment process, the technical calculation de¬ 
partment must deal with and account for 
external influential factors as is required of 
all other participating departments. These 
factors can encompass the needs and pref¬ 
erences of specific markets and economic 
requirements as well as key technical as¬ 
pects (Figure 2). The automotive industry is 
currently being pressed to design and build 
vehicles that offer ever better levels of effi¬ 
ciency. 



144 



Figure 2 Trends in the automotive industry 

Expressed in specific terms, this translates 
to such developments as: 

- Optimizing engine output by making 
the air path variable in design (VT, VCT, 
ECP) 

- Reducing mechanical loss by enhanc¬ 
ing tribological systems 

- Integrating lightweight materials 

- Electrifying the powertrain 

In addition to satisfying customers by offering 
more fuel-efficient passenger cars and meet¬ 
ing self-imposed obligations, complying with 
ever stricter C0 2 emission regulations further 
motivates the entire industry to design and 


engineer engines and vehicles that consume 
less fuel than their predecessor models. 

It goes without saying technical calcula¬ 
tion experts need to address these con¬ 
straints by amassing new knowledge and 
devising methods that cater to this trend. 
Correctly evaluating internal engine mea¬ 
sures requires a great deal of knowledge 
about thermodynamics, for example, and 
the increasing complexity of modern sys¬ 
tems designed to enhance variable re¬ 
sponse characteristics need to be thor¬ 
oughly understood to accurately integrate 
them in a simulation model. The same ap- 


Efficiency chain “well to wheel” 



2 % Convection 
111 % Refinery and transport 


Figure 3 Energy lost from mining crude oil to operating a vehicle (“well to wheel”) 














Simulation Engine Systems 


9 


145 


plies to components and systems used to 
electrify or hybridize powertrains. In addi¬ 
tion, methods must be devised to reliably 
predict the outcome of friction-reducing de¬ 
sign measures. Figure 3 provides a starting 
point for achieving higher levels of efficien¬ 
cy. As various sources indicate that by 
2020, up to 1.5 billion vehicles will be in use 
around the world, of which well over 90 per¬ 
cent will have an internal combustion en¬ 
gine, it pays to further optimize the internal 
combustion engine. 

Not only have these trends in technolo¬ 
gy made a significant impact on the devel¬ 
opment work and technical calculations 
carried out by the automotive industry, but 
also the recent move toward globalization. 
In the process, basic engines (world en¬ 
gines) are now being assembled in large 
numbers and subsequently adapted to dif¬ 
ferent vehicle classes by varying the levels 
of performance and equipment accordingly. 
This, in turn, necessitates highly robust 
methods when it comes to technical calcu¬ 
lation, since any inherent design flaw has 
the potential to affect that many more units. 
The models used must also accurately rep¬ 
resent each individual variant. 

Globalization has likewise led to a change 
in production locations, which are now spread 
across multiple geographical regions that are 
served by a separate group of suppliers offer¬ 
ing different material mixes. This brings with it 
the consequence that the development teams 
themselves are also distributed around the 
globe and must collaborate to resolve the in- 
tercultural, regional, and method-based prob¬ 
lems that arise in this context. 

If the full potential that technical calcula¬ 
tion has to offer is to be leveraged, the prac¬ 
tices that it entails must be integrated in the 
overall design process as early as possible, 
and all departments need to collaborate ef¬ 
fectively on a daily basis. This applies to new 
developments and products in particular. Es¬ 
tablished components and systems require 
less commitment, since specifications and 


standard performance criteria are already in 
place and used around the world. 

When new, highly sophisticated sys¬ 
tems and hardware are designed, rapidly 
constructing simulation models around de¬ 
fined performance criteria is not an option, 
since this approach does not guarantee reli¬ 
able, accurate results confirming that the 
function required operates within the target 
parameters assigned to it. Complex sys¬ 
tems can sometimes take years to establish 
the right development environment includ¬ 
ing models and processes. The benefit, 
however, is that validated models and pro¬ 
cedural approaches are created that are 
robust and can provide qualified answers to 
a wide range of questions in minimal time, 
including to ones that are asked on short 
notice. This, in turn, reduces outlay and un¬ 
derscores the true value of technical calcu¬ 
lation. 


Internal Influential Factors 

The individual phases of the product devel¬ 
opment process (PDP) correlate with differ¬ 
ent technical questions and issues that per¬ 
tain to aspects of manufacturing and 
product development and also have a no¬ 
ticeable effect on modeling. This effect be¬ 
comes apparent as soon as a project is 
started, when reliable input data is frequent¬ 
ly not available. At the same time, the manu¬ 
facturer and suppliers are busy making a 
great deal of changes such that the initial 
priority is to limit efforts to investigating the 
primary effects that will point to the best 
possible concept to be adopted (design 
definition and finalization). When familiar 
components or systems are integrated, a 
lack of data can be temporarily substituted 
with values from existing databases. The re¬ 
sults provided by the simulation must then 
be taken into account with this constraint in 
mind and replaced with qualified, realistic 
values later on. In addition to this time- 


146 


System model 


Lookup table for control valve 
Movement and flow cross sections 



Boundary conditions 


Degree of detail 


Figure 4 Modeling and system analysis 


based component in the product develop¬ 
ment process, technical calculation work is 
also characterized by the experience that 
has already been gained with the system 
being developed. 

For existing products, all recurring pro¬ 
cesses have usually been automated or at 
least defined in a specification (Figure 4). 
This is absolutely essential, especially in 
the case of global projects. After tools have 
been automated in line with technical cal¬ 
culation data, they can be handed over 
to the project engineers, who then make 
smaller calculations on their own and profit 
from expedited response times. When a 
finger follower is designed, for example, 
the question of rigidity becomes relevant. 
Schaeffler has fully automated this calcula¬ 
tion and integrated it in its CAD system. 
Ninety-nine percent of the time, the system 
e-mails an automatically generated report 
to the project engineer at the click of a 
mouse just a few minutes after the start of 
the calculation. 


The situation is different when new applica¬ 
tions are developed, however, which are 
characterized by different levels of modeling 
detail as a result of the individual phases of 
the product development process and vari¬ 
ous questions fielded by specialists. Pro¬ 
duction planning personnel, for example, do 
not ask the same questions as the software 
department tasked with programming the 
functions for the engine control unit. Con¬ 
structing a complete model that can answer 
all of these questions is usually too com¬ 
plex, requires too much calculation time, 
and is sometimes not even possible. Mod¬ 
els are therefore constructed to target a 
specific question pertaining to a certain 
technical aspect and do not map an overall 
scenario. 

One example of a scenario in which 
many different questions are asked through¬ 
out the product development process in¬ 
volves the multiphysics simulation model for 
the quick-acting valve used in the fully vari¬ 
able UniAir valve train (Figure 5). 











Simulation Engine Systems 


9 


147 



Figure 5 UniAir system 


Yoke ring 
cap 


Yoke ring 


Housing 

Sealing 

Sealing 

spring 



Magnetic 

core 

Valve 

body 


Valve plate 
Figure 6 Quick-acting switching valve 


Valve body 
spring 


The UniAir system comprises a camshaft- 
controlled actuator with an integrated, 
quick-acting hydraulic valve and corre¬ 
sponding valve timing software [1]. The 
switching valve is a de-energized, open 2/2- 
way switching valve by design that displac¬ 
es a valve body relative to the valve seat to 


open and close the connection linking the 
high and intermediate pressure chambers. 

To bring the quick-acting valve into the 
development environment (Figure 6), a mul¬ 
tiphysics simulation model was created that 
maps and correlates all mechanical, hy¬ 
draulic, magnetic, and electrical design as- 



Figure 7 Physical interactions with the quick-acting switching valve 





































148 


pects. Figure 7 shows the coil current. After 
an initial current is introduced, the maxi¬ 
mum rated voltage is applied to produce a 
magnetic flux while generating the coil cur¬ 
rent and magnetic force required to close 
the valve. When the maximum current is 
reached, the closing time is characterized 
by a bend in the current signal (“V shape”) 
as a result of an actuation triggered in the 
presence of a constant pulse width modula¬ 
tion (PWM) of between 0 and 12 V. During 
the hold phase, an electrical current lower 
than the one observed in the peak phase 
ensures that the closed position of the valve 
is reliably held. Although the energy con¬ 
sumed at this time continues to be high, it is 
in line with operating requirements [2]. 

The valve is opened when the coil is re¬ 
verse connected to the Z diode, which causes 
the magnetic force to quickly deplete and trig¬ 
gers a fast opening movement. The opening 
time is detected by short-circuiting the mag¬ 
netic coil so that the remaining magnetism 
produces a current coincidental with the mo¬ 
tion pattern of the current signal via the mag¬ 
netic-mechanical coupling. Raising and over¬ 
shooting the anchor as a magnetically active 
component, however, means that the exact 
opening time can only be determined using 
higher outlay than that for the closing time. 


Coil Current 



Simulation 
— Measurement 


Figure 8 Comparison of measurement and 
simulation results 


When all required correlations are taken into 
account, a high level of conformity between 
the results of the measurement and the 
simulation is achieved (Figure 8). This, in 
turn, makes it possible to use the model to 
quantify the influence of a leakage gap vari¬ 
ation on the operative function, for example. 
Questions of this nature are typically fielded 
by production planning experts, since the 
size of the gap can lead to varying costs. 

The model does not lend itself to an¬ 
swering questions fielded by the software 
developers responsible for realizing the 
functions for the control unit. Instead, a 
model with real-time capability is required 
whose simulation time corresponds with 
the time spent in the real world, much as is 
the case in a flight simulator. 


Analysis 

After system analysis and modeling have 
been carried out, the actual analysis work 
takes place. In the most basic of scenarios, 
calculations are run using an appropriate soft¬ 
ware application. This step can also involve 
model verification or a sensitivity analysis, 
however, which can retroactively affect the ini¬ 
tial modeling. The objective of this verification 
of unknown or new models is to identify sensi¬ 
tive parameters to keep the number of param¬ 
eters targeted for investigation as low as pos¬ 
sible, thereby minimizing calculation times. 
When a finite element calculation is made, for 
example, the influence of temperature on the 
steel components in the cylinder head is not 
varied, since the elasticity module that relates 
to the temperatures prevailing in this area ex¬ 
hibits almost no change as the dominating 
influential parameter. 

Model verification answers the question 
of which and how many parameters should 
be varied. Once this list has been defined, 
optimization algorithms such as DoE (Design 
of Experiments) allow the input parameters 
to be varied during the analysis until the de- 







Simulation Engine Systems 


9 


149 


sired characteristic statements are quanti¬ 
fied. These methods can also be applied to 
leverage the calculation work so that recom¬ 
mendations for reference samples can be 
made to the testing department. Numerous 
additional methods are likewise available for 
optimizing earlier development stages. 


Support for Design Drafting 


Deriving draft or concept-based proposals is 
included among the core tasks assigned to the 
technical calculation department. The following 
example shows a holistic simulation for opti¬ 
mizing a timing drive to minimize friction. 

In dynamic systems, friction provides for 
the necessary level of damping while at the 
same time exerting a negative effect on op¬ 
erating efficiency. To answer the question of 
the extent to which reducing the friction ex¬ 
perienced in the timing drive can reduce 
fuel consumption, calibrated engine models 
must be created using corresponding data 
about the vehicle. In so doing, the same 
methods and models that were constructed 
to assess the potential improvements af¬ 
forded by complex valve train strategies can 
be applied here as well. 


The first step is to realistically map the ef¬ 
fects on the internal combustion process in 
the model. Due to the many combinations 
of input data possible, the pronounced ef¬ 
ficiency for projecting the rates at which 
heat is released makes quasi-dimensional 
internal combustion models the ideal 
choice in this regard. Altered operating 
conditions such as engine speed, load, re¬ 
sidual gas content, air/fuel ratio, and 
changes in charge movement can then be 
evaluated. To analyze changes in knocking 
tendency and the resulting main combus¬ 
tion point, Schaeffler adopts an Arrhenius 
approach, while a physical-based method 
according to Fischer is used to account for 
mechanical losses. The parameters for re¬ 
alizing the best fuel consumption are de¬ 
termined at stationary mapping points that 
result from the frequency distribution for 
the combined engine and vehicle investi¬ 
gated in the respective driving cycle. In or¬ 
der to improve the design draft parameters 
for a large number of possible variants in 
the relevant section of the data map, sto¬ 
chastic optimization methods are lever¬ 
aged. Final evaluation of the varying design 
draft strategies for the combined engine 
and vehicle is made in different driving cy¬ 
cles in conjunction with the overall vehicle 
simulation (Figure 9). 



Figure 9 GT-Power modeling 






150 


Heat release 

Quasi-dimensional internal combustion model 



Knocking model 
Arrhenius approach 



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Physically-/empirically- 
based approach 


Heat flow / hot end 
Chemically-/physically- 
based models 


* # % / 



Optimization tool 
IAV engineering 
toolbox 


Figure 10 Calculating fuel consumption with GT-Power 


The optimization models and tools shown in 
Figure 10 are leveraged to project combined 
fuel consumption [3]. 

To enter the corresponding data in the 
GT-Power model, the reduction in friction in 
the timing drive that is responsible for 0.5 to 
2 percent of the overall loss in efficiency [4] 
must be examined in greater detail. 

Friction can never be eliminated alto¬ 
gether. At the same time, however, losses 
must be minimized and the influential fac¬ 
tors and interactions within the systems 
must be understood. By utilizing friction in 
a targeted manner to optimize the chain 
drive, the damping properties it affords 
can make a significant impact on limiting 
peak points in dynamic force. The majority 
of tribological systems in an internal com¬ 
bustion engine as it is operated or being 
started encounter the different types of 
friction (static, boundary, and hydrody¬ 
namic friction) at different frequencies. 


After specialists have identified the friction 
phenomena that occur at specific times and 
in specific areas while taking the interac¬ 
tions in the relevant systems into account, 
suitable measures can be selected and 
combined to optimize efficiency as far as 
possible. 

Figure 11 shows the friction types pres¬ 
ent in the chain-driven timing drive. This fric¬ 
tion encompasses the mesh points of the 
chain (A), the friction in the chain link joint 
(B), and the friction between the chain and 
guides (C). Adding to this is the friction ob¬ 
served on the bearings of the crankshaft 
and camshaft as well as any auxiliary drives 
(D), and the losses within the tensioning ele¬ 
ments (E). 

Two problem areas arise when mod¬ 
eling friction in multi-body systems. The 
first involves correctly describing the 
configurations associated with static and 
sliding friction by making differential 











































Simulation Engine Systems 


9 


151 


equations and numerically resolving 
them for a transient simulation. Frequent¬ 
ly, the possibilities for physically describ¬ 
ing static and sliding friction are defined 
by the solution algorithms available. With 
respect to the dissipation of energy and 
the pronounced dynamic characteristics 
of the timing drive and chain drive sys¬ 
tem, the variability of the coefficient of 
friction under static conditions is of mini¬ 
mal significance such that a breakaway 
torque can be specified. The second 
problem is determining the coefficient of 
friction under sliding conditions, or char¬ 
acterizing the kinematic variables and 

A 



physical parameters that influence this 
friction. 

The transition from static to sliding 
friction is described by a new model that 
is being used in a simulation program at 
Schaeffler for the very first time. To this 
end, the static potential of a contact that 
is momentarily stationary is balanced 
against a virtual displacement. This po¬ 
tential is calculated using the parameters 
that describe the contact and include the 
coefficient of static friction, speed, and 
normal force. When the static potential of 
a contact is exceeded, sliding friction 
occurs. 


Static/boundary friction 





Additives 



Viscosity 



Load 


■ 

Shape of contact 



Geometry of gap 


i 

Roughness 


■ 

Coating and heat treatment 



Figure 11 Friction points in the timing drive 


























152 


Data map solution 

Coefficient of friction, joint/bush roller chain 



Velocity in m/s 


Figure 12 Data maps for the coefficient of friction 

As soon as the system starts to move, the 
coefficient of sliding friction must be deter¬ 
mined to quantify the friction at the contact. 

A single variable factor that changes de¬ 
pending on the type and state of the tribo¬ 
logical system enters the equation at this 
point, which is why describing the coefficient 
of friction during the sliding phase is perti¬ 
nent to observing the system from an energy 
perspective. When the system experiences 



Velocity in m/s 


and mixed-friction state 

sliding friction, the coefficient of friction is de¬ 
scribed by data maps using state variables 
such as speed and load. 

The left side of Figure 12 shows, by ex¬ 
ample, the coefficient of friction determined 
for the contact point of the chain link joint in 
a bush roller chain during model testing 
with respect to the data map. This map was 
plotted in relation to the sliding speed and 
normal force. The coefficient of static fric- 



Specification for the highly dynamic chain test stand 

- Hydrostatically supported shafts - Constant brake torques possible 

- Drive motor with rotational irregularities - Direct chain guide friction measurement 

similar to a crankshaft - Conditioning of oil temperature and oil quantity 

- Brake torques similar to a camshaft 


Figure 13 Highly dynamic chain test stand 




















Simulation Engine Systems 


9 


153 


tion is approximately 0.25. When local slid¬ 
ing speeds increase, the coefficient drops 
to around 0.01. Higher normal forces cause 
the friction level to rise as a result of the in¬ 
creased proportion of solid content. The 
diagram on the right side (Figure 12) shows 
the friction value map for the contact point 
between the chain and guide in identical 
fashion. These data maps are determined 
for every frictional contact point (A, B, and 
C) in the timing drive of each chain type tak¬ 
ing into account all additional, relevant pa¬ 
rameters such as oil quantity and quality, 
the material mix, and roughness. The maps 
are then made available to mark the bound¬ 
ary conditions for the simulation and allow 
friction losses to be quantified. 

To validate the model and define param¬ 
eters for friction modeling, Schaeffler and IFT 
designed and constructed a highly dynamic 
chain test stand that comprises a separate 
electric motor to produce driving and braking 
forces. This makes it possible to simulate not 
only the dynamic rotational imbalance near 
the crankshaft, but also the braking torque of 
a crankshaft assembly to reproduce realistic 
performance constraints. Figure 13 provides 
a schematic representation of the design 
configuration. 

The supply unit for lubricating the chain 
is realized by an external oil assembly that is 
positioned in the immediate vicinity of the 
test stand. Heating and cooling systems al¬ 
low the oil quantity and quality to be condi¬ 
tioned, and the friction observed between 
the chain and guide can be determined. 
Figure 14 shows the high level of conformity 
of the measurement data and simulation re¬ 
sults. The method can therefore be used as 
a predictive tool for product development. 

A high-resolution elastohydrodynamic 
simulation technique (EHD) is also em¬ 
ployed to determine the dependencies sur¬ 
rounding the different frictional states. This 
technique accounts for the elastic charac¬ 
teristics of the contacting partners in con¬ 
junction with geometry, contact curvature, 


Comparison Measurement/Calculation 


Bush roller chain with guide 



Tooth chain with guide 



■ Measured tensioning force 0.5 kN 

■ Calculated tensioning force 0.5 kN 

■ Measured tensioning force 1 kN 

■ Calculated tensioning force 1 kN 

■ Measured tensioning force 2 kN 

■ Calculated tensioning force 2 kN 

Figure 14 Comparison of measurement data 
and simulation results 

roughness, oil properties, and the variables 
of load and speed that change with respect 
to time and location. Figure 15 shows the 
models used to determine the tribological 
system attributes of the contact point be¬ 
tween the chain back and the guide. 




154 


90 mm guide radius 
9.61 m/s chain velocity 
675 N chain force 
55 N normal force 



Guide deformation 



Chain velocity in m/s 



Speed in rpm 

— Total 

— Proportion of solid bodies 

— Hydrodynamics 


250 mm guide radius 
9.61 m/s chain velocity 
675 N chain force 
22 N normal force 



Guide deformation 




Chain velocity in m/s 



Speed in rpm 

— Total 

— Proportion of solid bodies 

— Hydrodynamics 


Figure 15 Influence of the guide radius on the friction between the chain and guide 


The EHD simulation technique is capable 
of determining the proportional relation¬ 
ship of hydrodynamic response and solid 
body contact, which it turn makes it pos¬ 
sible to systematically optimize the con¬ 
tact point between the chain and guide. 
Figure 15, for example, shows that with a 
guide radius of 90 mm, the proportion of 
the solid-body friction encountered domi¬ 
nates. The mixed-friction range is also 
constantly present, even when the system 
is operating at high speeds. When larger 
guide radii are introduced, however, the 
hydrodynamic proportion increases much 
more quickly as operating speed builds, 
while the overall friction value is much less 
pronounced. 


Outlook 


Digital tools have greatly accelerated the 
planning and development processes car¬ 
ried out at Schaeffler, and the transition 
from the pilot stage to commonly used 
practices has largely been finalized. Now is 
the time to firmly establish the tools in the 
organizational structures and further opti¬ 
mize the ratio of outlay to usable gain by le¬ 
veraging the variety of digital methods avail¬ 
able. Starting points include standardized 
processes, methods, and IT solutions as 
well as improved integration of production 
data in the product development process. 










Simulation Engine Systems 


9 


155 


Literature 


[1] Haas, M.: Just Air? UniAir - The first fully- 
variable, electro-hydraulic valve control system. 
9 th Schaeffler Symposium 2010, pp. 250-263 

[2] Mayer, A. et al.: Multiphysikalische Simulation 
eines schnellen Schaltventils fur einen elektro- 
hydraulisch vollvariablen Ventiltrieb. 13 th MTZ 
symposium, “Virtual Powertrain Creation”, 2011 

[3] Kirsten, K.; Brands, C.; Kratzsch, M.; Gunther, 
M.: Selektive Umschaltung des Ventilhubs 
beim Ottomotor. MTZ 11/2012, pp. 834 - 839 

[4] Schlerege, E; et al.: Optimierung von Steuer- 
triebketten durch erweiterte Reibmodellierung. 
5 th VDI symposium on cylinder heads and 
friction, 2011 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



156 


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Smart Phasing 

Needs-based concepts for 
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158 


Introduction 


Increasing numbers of gasoline engines 
have a camshaft phasing system - either 
on the intake side only or on the intake 
and exhaust side. A volume-produced 
diesel engine with a phasing system on 
the intake camshaft recently went into 
production for the first time. Systems with 
hydraulically-actuated swivel motors have 
become established [1]. The trend to¬ 
wards downsizing and downspeeding will 
increase the rate with which these sys¬ 
tems are fitted because power and torque 
can be increased and raw emissions re¬ 
duced by changing the relative angle be¬ 
tween the camshaft and the crankshaft. 
Electric phasing units would be the opti¬ 
mum solution from a technical perspec¬ 


tive. They are superior to hydraulically- 
activated variants but are associated with 
higher costs. This is why it is advisable to 
further optimize the systems currently 
used. Further development of these sys¬ 
tems must focus on meeting increasing 
requirements at comparatively low oil 
pressures. 


Requirements 


The most important requirements for cam¬ 
shaft phasing systems are illustrated by the 
load-speed data map of the internal com¬ 
bustion engine; the engine oil temperature 
is also a decisive factor for optimum timing. 
The maximum torque can be increased by 



Figure 1 Functional requirements for camshaft phasing systems 

Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_10, © The Author(s) 2014 





























Camshaft Phasing Systems 


10 


159 


changing the timing on the intake side 
through appropriate design of the cam 
contour. This is in line with the trend to¬ 
wards downspeeding. The phasing system 
must be able to change the timing as 
quickly as possible during transient opera¬ 
tion, i.e. the timing must be changed as 
quickly as possible during the transition to 
another operating point in the data map 
(Figure 1). 

Current adjustment speed require¬ 
ments are up to 500 degrees of crankshaft 
angle per second. If the adjustment speed 
is insufficient, this can be compensated by 
the ignition and injection system; this usu¬ 
ally results in disadvantages with regard to 
fuel consumption. 

Opportunities open up for controlling 
the internal exhaust gas recirculation if 
the exhaust valves are actuated by their 
own adjustable camshaft. This allows to 
reduce the raw emissions. A prerequisite 
is that the phasing unit can adjust the tim¬ 
ing of both camshafts, represented by the 
adjustment angle a, as precisely as pos¬ 
sible at a constant operating point. The 
aim is a deviation of 1° crankshaft angle 
from the set point stored in the data map. 
Future combustion methods, such as ho¬ 
mogeneous charge compression ignition 
(HCCI), place even higher requirements in 
this respect than current gasoline en¬ 
gines with direct injection. 

Another important parameter is to 
what extent the timing can be freely se¬ 
lected when starting the engine. The tim¬ 
ing set during continuous operation is 
sometimes not suitable for starting the 
engine. The phasing unit is locked after 
switching off the engine because the 
build-up of oil pressure during starting is 
too low to actuate it. This is why only pre¬ 
defined timing is currently available dur¬ 
ing starting. In the future, variable timing 
could be desirable for different starting 
conditions (for example, hot or cold 
start). 


Systems and function 


Electric system 

The electric phasing system comprises an 
electric motor and a three-shaft adjustment 
gearbox, which is mounted on the camshaft 
in the same way as a hydraulic phasing unit 
(Figure 2). The output shaft is permanently 
connected to the camshaft. The adjustment 
shaft of the three-shaft gearbox is connect¬ 
ed with the electric motor, which adjusts the 
phase angle between the crankshaft and 
camshaft. The third shaft of the adjustment 
gearbox forms the gearbox housing, which 
is coupled with the belt pulley or sprocket of 
the timing drive. 

If the phase angle is to be changed, 
the speed difference between the output 
shaft of the electric motor and the gear¬ 
box housing is increased. The shaft ro¬ 
tates faster to make an adjustment in the 
direction “advanced” and more slowly to 
make an adjustment in the direction “re¬ 
tarded”. The adjustment angle is held 



Figure 2 Electric camshaft phasing unit 




160 



A Shifting velocity 
■ Camshaft timing accuracy 
# Flexibility for engine start camshaft timing 


HCP with cartridge OCV 

I®/ 

HCP with central OCV 



Figure 3 Comparison of the cost and performance of different phasing systems 


constant when the output shaft of the 
electric motor rotates at the same speed 
as the camshaft or gearbox housing. Typ¬ 
ical gearbox ratios are in the range of 40:1 
to 100:1. 

This electric system allows the great¬ 
est degree of freedom when selecting 
the timing for starting. It offers higher ri¬ 
gidity if torque is applied to the camshaft 
via the crankshaft and therefore achieves 
the highest adjustment accuracy. The 
adjustment speed is also higher com¬ 
pared with the best hydraulic systems 
(Figure 3). 

The electric system is also the only 
system to offer the option of free selection 
of the timing when the engine is started 
[2]. This high performance is also associ¬ 
ated with a higher technical effort. Such a 
system will go into volume production for 
the first time at Schaeffler in 2015. It is de¬ 
signed so that no modifications to the cyl¬ 
inder head are required. 


Rotor Sprocket Stator 



Figure 4 Design principle of a hydraulic 
camshaft phasing unit 









Camshaft Phasing Systems 


10 


161 


Hydraulic system 

Design and function 

The internal part of the camshaft phasing 
unit comprises a vane-type rotor, which is 
firmly attached to the camshaft. The exter¬ 
nal part (stator) is driven by the crankshaft 
via a chain or belt (Figure 4). 

The range of motion of the rotor in the 
stator defines the maximum adjustment 
angle; currently, a crankshaft angle of ap¬ 
proximately 30° in the directions “advanced” 
and “retarded” is standard on the intake 
side. In the neutral position, the rotor vanes 
are in the advanced or retarded position 
and are locked in this position when the en¬ 
gine is switched off. The chambers are filled 
with oil, which means the stator’s torque is 
transmitted to the rotor. The angular posi¬ 
tion of the camshaft relative to the crank¬ 
shaft is changed depending on the change 
of oil pressure on both sides of the rotor. A 
4/3 proportional valve connected to the oil 
circuit controls the relevant oil inlet and out¬ 
let. This valve is controlled by the engine 
control unit and operated magnetically (Fig¬ 
ure 5). Optimum timing data for every load 
and speed case is stored in the engine con¬ 
trol unit. The engine control system detects 
any deviations between the angular position 



Figure 6 Effect of alternating torque on the 
camshaft during valve actuation 


Camshaft phaser A B 



Figure 5 Function of the proportional valve 


of the camshaft and the nominal value from 
the signals sent by the camshaft and crank¬ 
shaft sensors and carries out continuous 
readjustment. 

For the sake of simplicity, the adjustment 
of the timing is usually characterized as a lin¬ 
ear process but adjustment is actually an it¬ 
erative process. The motion of the cam act¬ 
ing on the valve actuation system slows this 
process during the adjustment from “retard¬ 
ed” to “advanced”. In contrast, this alternat¬ 
ing torque accelerates the phasing operation 
when the timing is adjusted from “advanced” 
to “retarded” (Figure 6). The frequency with 
which an impulse occurs in a process de¬ 
pends on the adjustment distance and the 



























162 


engine speed. The magnitude of the alternat¬ 
ing torque depends on the engine speed and 
the valve train. 

The accuracy, with which the timing can 
be adjusted is essentially determined by the 
compressibility of the oil and the leakage 
system. Systems equipped with a solenoid 
located centrally in the phasing unit therefore 
have an advantage compared to units fitted 
with a decentralized arrangement because 
the leakage-prone transfer of oil between the 
camshaft and the cylinder head via control 
ducts is eliminated. The speed, with which 
adjustment can be carried out, depends on 
the available power and thus the oil pressure 
and the alternating camshaft torque. 

The camshaft phasing unit is locked in 
the “advanced” or “retarded” position after 
switching off the engine because the oil 
pressure during engine starts is insufficient 
to set the timing. The solenoid valve is not 
supplied with current. The phasing unit can 
be moved to the “advanced” base position 
using the assistance of a spring designed 
specifically for the application. Different tim¬ 
ing settings are only possible if the oil pump 
supplies the full oil pressure. 

Pressure accumulator 

Schaeffler makes a distinction between ac¬ 
tive and passive pressure accumulators. 
The latter increases the adjustment speed 


of the hydraulic camshaft phasing system 
so that it can be classified between a phas¬ 
ing unit without a pressure accumulator and 
an electric phasing system. In simple terms, 
this pressure accumulator can be described 
as a spring mass system, which is pressur¬ 
ized with oil. The system is in equilibrium if 
the force of the oil pressure is equal to the 
spring force. The compression spring forc¬ 
es are characterized by the preload force in 
the base position and the spring rate that 
defines the increase in force via the travel of 
the piston up to the end position. If the ac¬ 
cumulator is pressurized, the piston con¬ 
verts the oil pressure provided by the oil 
pump into potential energy that is stored in 
the compression spring. The spring un¬ 
winds during the next phasing operation 
and provides additional assistance to the oil 
pressure during movement of the vanes. 
The pressure accumulator is arranged in 
front of the hydraulic solenoid and connect¬ 
ed with the oil supply. It comprises a cup¬ 
shaped piston, compression spring, guid¬ 
ance element and a thin-walled housing 
with a closing plug mounted on the end face 
(Figure 7). 

The piston is guided inside the housing 
and its movement is limited by two stops. In 
the released base position, the piston in 
contact with the inside of the closing plug 
and in the end position, it contacts the guid- 


Central Camshaft 
solenoid phaser 



valve 


Figure 7 Passive oil pressure accumulator 





Camshaft Phasing Systems 


10 


163 



r 3.0 


CD s- 
■ t 10 

> m -o 

— <D 

O -C 

; w "5, 
'i o 


1.0 


— Shifting angle with pressure accumulator 

— Oil pressure cyl. head with pressure acc. 
Shifting angle w/o pressure accumulator 

— Oil pressure cyl. head w/o pressure acc. 


Figure 8 Test results for a passive accumula¬ 
tor during idling, at 90 °C and under 
zero load 

ance element. A check valve located be¬ 
tween the accumulator and the solenoid 
prevents a return flow of engine oil from the 
phasing system to the engine or oil sump. 
This means the phasing system remains ad¬ 
justable at all the operating points. A com¬ 
parison of a system with and without a pres¬ 
sure accumulator shows (Figure 8) that the 
system with a passive pressure accumula¬ 
tor (black curve) reaches the end stop in the 
stator more quickly than the system without 
a pressure accumulator (green curve). 

In the case of the system with the pres¬ 
sure accumulator, the oil pressure decreas¬ 
es more slowly during adjustment than in 
the system without a pressure accumulator. 
This is due to the fact that the majority of the 
required oil volume is provided by the pres¬ 
sure accumulator and therefore more ener¬ 
gy is made available to the phasing system 
for the phasing operation. The reduction in 
oil that occurs here is primarily determined 
by the design of the compression spring. 
The greater the oil volume that can be 
forced out of the accumulator during a dif¬ 
ference in pressure, the lower the decrease 
in oil pressure in the oil circuit. This advan¬ 
tage in terms of the adjustment speed does 


not depend on whether adjustment is car¬ 
ried out away from the base position or to¬ 
wards the base position. The frictional 
torque on the camshaft alone causes the 
adjustment to be unsymmetrical in both di¬ 
rections. If the engine is switched off, the oil 
immediately flows back into the oil sump via 
the leakage points. A system with a passive 
pressure accumulator is therefore unable to 
change or set the timing during starting and 
must also be locked. 

The active pressure accumulator can 
store the oil reservoir for a limited amount of 
time. This is sufficient to supplement a start- 
stop system so that the optimum timing can 
be set when restarting the engine. If the en¬ 
gine is switched off, unpressurized engine 
oil remains in the reservoir for some minutes 
and is not immediately forced out after the 
engine is switched off. If the engine is start¬ 
ed, the accumulator spring is activated and 
pressurized oil is supplied to the phasing 
system so that the oil pressure in the cam¬ 
shaft phasing unit immediately increases. 
This is why adjustment from the base posi¬ 
tion starts earlier than it would without a 
pressure accumulator. The pressure reser¬ 
voir only empties after long stationary peri¬ 
ods, for example, if the vehicle is parked 
over night. This is due to leakage via the 
circumferential groove and radial bores in 
the first camshaft bearing. 

The active pressure accumulator can be 
characterized as a switchable coupling 
mechanism that creates a detachable lock 
for the piston when the reservoir is full. The 
relevant actuator is located on the rear end 
of the accumulator (Figure 9). 

The locked condition of the piston is con¬ 
sidered for the description of functions. If the 
piston is to be unlocked, an electromagnetic 
actuator located on the cylinder head push¬ 
es a rod against a switching pin with a cir¬ 
cumferential groove. As soon as the balls 
can move in the groove, they are pushed in¬ 
wards by means of the compression spring 
force. This releases the piston. If the accu- 






164 


Central Camshaft 
solenoid phaser 



valve 


Figure 9 Design of the active pressure accumulator 


mulator is full, the piston automatically en¬ 
gages in the coupling mechanism. During 
this process, the piston locking unit pushes 
the sliding plate back against the sliding plate 
spring until the base of the piston mates to 
the coupling mechanism. In this position, the 
switching pin is moved in an axial direction 
via the return spring and the balls are pushed 
outwards from the groove in a radial direc¬ 
tion, i.e. the piston is secured. During this 
process, the rod and the actuator are moved 
back to their original position. The piston can 
be unlocked again by feeding the actuator 
with current. The relevant signal comes from 
the engine control unit if it initiates an engine 
start. The discharge process when the en¬ 
gine is started is decisive for the dimension¬ 
ing of the working pressure. The required 
working pressure level is higher than the op¬ 
timum pressure level of the passive pressure 
accumulator that would be necessary to im¬ 
prove the adjustment speed during hot idling. 

Challenge posed by oil 
pressure 


One of the most important boundary 
conditions for hydraulically actuated 


camshaft phasing units is the pressure in 
the oil circuit. Only mechanically driven 
oil pumps were used in the past. They 
are designed for the worst case, i.e. a 
high oil temperature, low speeds and 
long service life. However, the oil con¬ 
sumption of the engine at increasing 
speeds does not increase as rapidly as 
the delivery rate of the oil pump, which 
increases in approximate terms propor¬ 
tionally to the speed in unregulated de¬ 
signs [3]. This is why part of the delivery 
is fed directly to the intake side of the 
pump again at medium and high speeds. 
The pump therefore has a low efficiency 
in this operating range. 



Figure 10 Development of engine oil pressure 
2004 to 2012 


























































Camshaft Phasing Systems 


10 


165 


Regulated oil pumps are increasingly re¬ 
placing unregulated oil pumps as part of 
measures to increase the efficiency of en¬ 
gines. The designs are becoming smaller 
at the same time. The aim is to reduce the 
amount of ineffective work to an absolute 
minimum. A look back at the last eight 
years shows that the maximum power of 
regulated pumps now only reaches a val¬ 
ue that is less than the base power of un¬ 
regulated pumps at the start of the com¬ 
parative period (Figure 10). 

There has been a very significant re¬ 
duction in the overall pressure level be¬ 
cause low-friction bearings are now used 
in the entire engine and leakage has been 
greatly reduced. It can be assumed that 
this trend has now reached its lower limit. 
Flowever, low oil pressure is a challenging 
boundary condition for new and further 
developments of camshaft phasing sys¬ 
tems. The lower the oil pressure, the lower 
the amount of energy available for phas¬ 
ing of the camshaft. 


Pressure-free oil volume 
accumulator 


Design and function 

The adjustment speed of hydraulic cam¬ 
shaft phasing units is mainly determined by 
the performance of the oil circuit. Until now, 
only a system with a passive pressure ac¬ 
cumulator could achieve a higher adjust¬ 
ment speed than a conventional system 
with a central valve. This also results in in¬ 
creased system costs. Schaeffler has there¬ 
fore developed another option: This option 
is based on an oil volume accumulator lo¬ 
cated within the phaser itself. The concept 
is positioned between the above mentioned 
systems both with regard to costs and per¬ 
formance. The oil volume accumulator is ar¬ 
ranged in additional bores in the rotor of the 
camshaft phasing unit - directly next to the 
oil chambers. The adjustment process is 
triggered when these chambers are filled 
(Figure 11). This oil volume accumulator is 


Without oil volume accumulator 


With oil volume accumulator 


Sprocket 


Stator 


Rotor 


Oil volume 




Figure 11 Mounting position of the oil volume accumulator 











166 



Figure 12 Connection of the oil volume 

accumulator (cross-section of a 
phasing system) 


The reservoir is immediately available 
again for the phasing system via only a 
short bore so that it is mainly used. The 
solenoid valve remains connected to the 
oil circuit, which means a second oil feed 
is always available. 

The effect on the oil pressure supply 
to the adjustment chambers is compara¬ 
ble with the aspiration of a syringe: The 
faster the oil can be replenished, the 
faster the piston can be withdrawn (Fig¬ 
ure 13). The oil volume accumulator re¬ 
sults in a number of advantages, as ex¬ 
plained in the test and simulation results 
presented below. 


not pressurized, but improves the adjust¬ 
ment speed by accelerating the flow of oil 
into and out of the adjustment chambers. 

The oil volume accumulator is fed from 
the oil that is forced out of the chamber, in 
whose direction the adjustment is carried 
out (Figure 12). Oil is not discharged into 
the oil sump until the accumulator is filled. 


Simulation and test results 

Simulations and tests are being carried out 
to investigate the influence of the oil volume 
accumulator upon the adjustment speed and 
the required oil flow from the oil circuit while 
taking different cam contours into consider- 



F Push= const 


Environment 


PEnvironment 


Vacuum 


Oil flow 

n*r 4 


8*7*1 

^ Syringe — ^ 


Bowl 


+ Q 


Tank 


Figure 13 The possible suction volume determines the speed with which the piston can be withdrawn 



















Camshaft Phasing Systems 


10 


167 



— Without oil volume accumulator 

Figure 15 Simulation of a phasing operation with a phasing system fitted 
with an oil volume accumulator 


ation (Figure 14). 

The simulation of an 
adjustment at 0.5 bar 
oil pressure leads 
to the conclusion 
that the adjustment 
speed increases 
and the oil flow de¬ 
creases significantly 
with all the cam con¬ 
tours considered 
here (Figure 15). 

It can be ex¬ 
pected that the oil 
requirement is at 
least halved above 
an alternating torque 
on the camshaft of 
10 Nm. At the same 
time, the adjust¬ 
ment speed in¬ 
creases, for exam¬ 
ple, from 175° to 280° crankshaft angle per 
second at 20 Nm alternating torque. Mea¬ 
surements carried out on a test engine 
confirm this simulation: Across the entire 
speed range, the system carries out ad¬ 
justments faster in both directions with the 
oil volume accumulator than without the oil 
volume accumulator (Figure 16). 



Figure 14 Simulation of a phasing operation 
with a conventional phasing unit 


The test results prove that the oil volume ac¬ 
cumulator also has a positive effect on the 
critical variable oil pressure. The measuring 
duration investigated in detail here comprises 
the time from when the cam starts to move 
the valve actuation system until when the 
process is completed during the phasing 


Oil temperature 90 °C 



Engine speed in rpm 
— With oil volume accumulator 
— Without oil volume accumulator 

Figure 16 Comparison of the adjustment speed 
across the speed range with and 
without an oil volume accumulator 















168 


Cylinder Head / Camshaft 



Without oil volume accumulator 


i — i 

<D 



Figure 17 Effect of an alternating torque 
impulse on the camshaft and 
development of the oil pressure 

operation at a crankshaft angle of 31° from 
the neutral position in the direction “ad¬ 
vanced”. The engine rotates at 1,200 rpm 
and the oil temperature is 90 °C. The mea¬ 
surement is carried out in the pressure line 
and in the oil chambers A and B of the cam¬ 
shaft phasing unit (Figures 17-20). 



Without oil volume accumulator 



Figure 18 Oil pressure curve in chamber A 
during alternating torque 

The comparison of a system with an oil vol¬ 
ume accumulator (Figure 20 right) and with¬ 
out an oil volume accumulator (Figure 20 
left) shows that the oil pressure of chamber 
A decreases significantly less during filling in 
the system, if an oil volume accumulator is 
used. This means that the formation of a 
temporary vacuum is largely prevented. In 


























Camshaft Phasing Systems 


10 


169 



Without oil volume accumulator 




Figure 19 Oil pressure curve in chamber B 
during alternating torque 

contrast, a temporary vacuum of up to 1 bar 
is formed in the system without an oil vol¬ 
ume accumulator. At the same time, a 
slightly lower pressure builds up in the 
chamber to be emptied B so that the oil 
flows away more slowly there. 

It is important to prevent the formation 
of a vacuum in the controlled chamber for a 


number of reasons. First and foremost, it 
impairs the adjustment speed: The system 
with an oil volume accumulator is capable of 
carrying out an adjustment with 20° more 
crankshaft angle than a system without an 
oil volume accumulator in the same time. 
The vacuum also causes the entire system 
to oscillate and operate less precisely. 

The oil volume accumulator is an effec¬ 
tive approach for increasing the adjustment 
speed of hydraulic camshaft phasing units. 
It does not achieve the same degree of im¬ 
provement as a passive pressure accumu¬ 
lator, but also does not have its level of 
complexity. The pressure-free oil volume 
accumulator also reduces the conflict of ob¬ 
jectives between the advantage of a higher 
adjustment speed through the use of a pas¬ 
sive pressure accumulator and the improve¬ 
ment that would arise from the use of an 
active pressure accumulator in combination 
with a start-stop system. The oil volume ac¬ 
cumulator can also be combined with an 
active accumulator. 


Summary 


With increasingly stringent emission regula¬ 
tions, it is now becoming an essential re¬ 
quirement to also use camshaft phasing 
systems in diesel engines. Hydraulic sys¬ 
tems have now largely replaced phasing 
units with helical gear teeth and axial pis¬ 
tons. They do not achieve the same perfor¬ 
mance as electric phasing units but are 
more attractive when cost aspects are tak¬ 
en into consideration. The oil pressure is a 
boundary condition critical for the success 
of hydraulic systems. Comparisons show 
that both the base and the peak power of oil 
pumps has been significantly reduced in re¬ 
cent years. Accordingly, less power is avail¬ 
able for the phasing system. One approach 













170 


Without oil volume accumulator 


With oil volume accumulator 


0 

O) 

c 

(0 

k. 

0 c 
c/> E 
ro co 

e = 
(0 


E 

CO 

O 


0 

3 - 

Q. 



Figure 20 Comparison of the vacuum in chamber A and the adjustment distance 


is to use a passive pressure accumulator. 
This significantly increases the adjustment 
speed. The costs are moderate but an even 
more cost-effective system may be required 
in cost-conscious markets. Schaeffler has 
developed a pressure-free oil volume accu¬ 
mulator for such application. The additional 
oil volume reduces the requirements placed 
on the oil circuit and prevents the risk of a 
vacuum forming during rapid and extensive 
phasing operations. The adjustment speed 
is also increased across the entire speed 
range with this concept. The oil volume ac¬ 
cumulator can be combined with an active 
pressure accumulator. The latter stores oil 
under pressure for several minutes, for ex¬ 
ample, if the start-stop system has switched 
off the engine. Sufficient energy is available 
during a restart to change the timing during 
starting. 


Literature 


[1] van Basshuysen, R.; Schafer, F. (Hrsg.): 
Handbuch Verbrennungsmotor: Grundlagen, 
Komponenten, Systeme, Perspektiven. 

5. Auflage. Vieweg+Teubner, Wiesbaden, 2010, 
p. 483 

[2] Schafer, J; Balko, J.: High Performance Electric 
Phasing System, SAE paper 2007-01-1294 

[3] StrauB, A.; Schaefer J.; Dietz J.; Busse M.; 
Boeggershausen M.: Quo vadis hydraulic 
variable camshaft phasing unit? 9 th Schaeffler 
Symposium, 2010 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 
























172 



Z GWR 
L Q N V 


173 


Cylinder Deactivation 

A technology with a future 
or a niche application? 


Arndt Ihlemann 
Norbert Nitz 


D Vv ^ 

JZMHa. 
A G Q S W I 
F I M B C H 


E ft . 

S E H E b 



CECBS 

T 

P 

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174 


Introduction 


One of the ways manufacturers can mini¬ 
mize fuel consumption is to downsize the 
engines they offer. A cylinder’s volume can 
only be restricted to a certain extent, how¬ 
ever, if the thermodynamically ideal volu¬ 
metric capacity of 400 to 500 cm 3 per cylin¬ 
der is to be retained. In practice, downsizing 
therefore frequently leads to a reduction in 
the number of cylinders. 

“Temporary downsizing” in the form of 
cylinder deactivation offers an attractive 
compromise, since this allows an engine to 
shift its operating mode to achieve the spe¬ 
cific consumption figures it is rated for, espe¬ 
cially when low loads and operating speeds 
are encountered. At the same time, the driver 
still has a sufficiently powerful engine at his or 
her disposal that ensures the same level of 


driving pleasure and comfort with regard to 
acoustics and vibration characteristics. 

An additional key success factor that 
can help this technology to be deployed in a 
more mainstream fashion is that it can be 
integrated into existing engine concepts at 
acceptable costs. 


Designs 


The most consistent form of cylinder deac¬ 
tivation is to not only to cut injection and ig¬ 
nition for the respective cylinders, but also 
to stop all moving parts (including the pis¬ 
tons). This, in turn, utilizes the entire thermo¬ 
dynamic potential available and consider¬ 
ably reduces the friction that occurs inside 
the engine. It goes without saying that com- 


Manufacturer 

Type of engine 

Valve concept 

Status 

GM 

6.0-liter V8-6-4 engine 

Pushrod actuation, 
switchable rocker arm pivot point 

SOP/EOP 1980 


3.9-liter V6 engine 

Switchable roller tappet 

EOP 2008 


5.3-liter V8 engine 

Switchable roller tappet 

Volume production 


4.3-liter V6 engine 

Switchable roller tappet 

Volume production 


6.0-liter V8 engine 

Switchable roller tappet 

Volume production 

Daimler 

5.0-liter V8 engine 

Switchable rocker arm; MB 

EOP 2005 


5.8-liter VI2 engine 

Switchable rocker arm; MB 

EOP 2002 

Chrysler 

5.7-liter V8 engine 

Switchable roller tappet 

Volume production 


6.4-liter V8 engine 

Switchable roller tappet 

Volume production 

Honda 

3.5-liter V6 engine 

Switchable rocker arm 

Volume production 

AMG 

5.5-liter V8 engine 

Switchable pivot element 

Volume production 

VW 

Group 

1.4-liter inline 
4-cylinder engine 

Cam shifting system, VW/Audi 

Volume production 

4.0-liter V8 engine 

Cam shifting system, Audi 

Volume production 


6 3/4-liter V8 engine 

Switchable roller tappet 

Volume production 


6 3/4-liter V8 engine 

Switchable roller tappet 

Volume production 


6.5-liter V12 engine 

Only the fuel injection supply is cut 

Volume production 


Figure 1 Examples of engine concepts featuring cylinder deactivation 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_ll, © The Author(s) 2014 












Valvetrain Systems 


11 


175 


Complete engine operation 



Engine speed 


Cylinder deactivation mode 



Figure 2 Operating data map and driving resistance curve: The operating ranges associated with the 
lowest specific fuel consumption are approached in cylinder deactivation mode (graphic on 
the right) and not when all cylinders are operating 


promises must be made when it comes to 
the ignition sequence and dynamic balanc¬ 
ing. What is much more significant, howev¬ 
er, is the outlay required to separate the en¬ 
gine into an area that continues to run while 
the other area is activated and deactivated 
as required. Even the coupling mechanisms 
on the crankshaft and camshaft cannot be 
justified by a cost-benefit analysis, which is 
why implementation of the system looks 
somewhat bleak at present. 

Almost all cylinder deactivation systems 
currently used interrupt the injection and ig¬ 
nition as well as valve actuation sequences 
for the cylinders to be deactivated (Figure 1). 
Today’s applications range from engines 
with 4 to 12 cylinders. Analyses conducted 
by Schaeffler, however, reveal that tempo¬ 
rarily deactivating one of the cylinders in a 
three-cylinder engine can also further re¬ 
duce consumption. 

To ensure that the engine continues to 
run smoothly enough, only certain cylinders 
are deactivated in accordance with the igni¬ 
tion sequence. 


Effect and potential 


When there is a specific performance re¬ 
quirement, the cylinders that are still being 
operated following cylinder deactivation 
must generate a higher mean pressure. This 
load-point shifting leads to a reduction in 
the throttle losses of the engine and ulti¬ 
mately helps to conserve fuel (Figure 2). De¬ 
activating the valves also reduces friction 
loss in the cylinder head, which further min¬ 
imizes consumption. 

The potential for reducing consumption 
when an engine is operated on two as op¬ 
posed to four cylinders can be illustrated in 
a simulation exercise carried out on a 1.4-li- 
ter four-cylinder engine. Line “a” plots the 
mean pressures at which the engine oper¬ 
ating in two-cylinder mode can achieve its 
optimum combustion point (8 crankshaft 
degrees after TDC) (Figure 3). 

When higher mean pressures are intro¬ 
duced in two-cylinder mode, the ignition se- 





























176 



Engine speed in rpm 


Figure 3 Reduction in fuel consumption as a 
result of cylinder deactivation 
(simulation result) 

quence must be retarded to avoid knocking. 
The resulting effect is that combustion no lon¬ 
ger achieves its peak efficiency, and addition¬ 
al fuel is consumed. Opening the throttle valve 
further counteracts this and has a positive 
impact on consumption in cylinders running 
higher mean pressures. Line “b” represents 
the theoretical switchover or transition line, as 


Exhaust gas is trapped 



+ Gas spring 
+ Slow cool down 

- Increased compression -> 
Highly irregular engine running 

- No torque neutrality 


operating the engine above these plotted 
points in two-cylinder mode leads to addition¬ 
al fuel consumption. This line can also drop 
considerably below line “b”, depending on the 
application and customer requirements. 


Technical implementation 


Deactivation mode 

When an engine switches to cylinder deacti¬ 
vation mode, there are two basic strategies 
that can be implemented for introducing a 
charge in the cylinders (refer to Figure 4): 

- Confine the exhaust gas in the com¬ 
bustion chamber after the combustion 
process has been completed 

- Introduce fresh air 

Both variants allow the gas confined to act 
as a pressure or thrust spring. 


Fresh air is trapped 



+ Gas spring 

+ Normal compression -> 
Smooth engine running 
+ High torque neutrality 


Source: MTZ "The New AMG 5.5-liter V8 Naturally Aspirated Engine with Cylinder Shut off“ 


Direct injection allows the realization „Fresh air trapped" 


Figure 4 Possible options for introducing a cylinder charge and their effects in cylinder deactivation mode 























Valvetrain Systems 


11 


177 


The heat generated by the confined ex¬ 
haust gas not only makes the cylinder cool 
down more slowly; the larger quantity of 
gas also produces very different pressures 
inside the cylinder and thus to greater ir¬ 
regularities on the crankshaft. The gas 
pressures that form during initial compres¬ 
sion when the exhaust valves are closed 
can even be higher than those experienced 
during combustion. The support forces not 
only place substantial loads on the piston 
and cylinder, but also lead to considerable 
frictional losses. The deactivation phase 
must then be maintained for a longer peri¬ 
od of time to ensure that a positive overall 
effect is achieved. 

As Figure 4 shows, peak pressures drop 
when the residual gas cools down as well 
as when gas diffuses from the combustion 
chamber into the crank assembly (blow by). 
Simulation calculations reveal that after an 
engine has gone through approximately ten 
revolutions, the pressure in the cylinder 
reaches the level that was present when 
fresh air was confined. 

The latter is only possible with a direct- 
injection engine. The differences in com¬ 
pression between the cylinders are less 
pronounced in this application, and the 
switchover phase can be better coordinat¬ 
ed as a result. This variant also requires 
compromises to be made, however, since 
the air in the combustion chamber loses all 
tumble or swirling motion produced at the 
intake point after just a few cycles. De¬ 
pending on the geometry of the combus¬ 
tion chamber, it may still be possible to re¬ 
fire the engine in this operating state. The 
ignition timing will have to be adjusted, 
though, whereby the efficiency of the com¬ 
bustion process suffers by a correspond¬ 
ing amount. Care must also be taken to 
ensure that no suction or vacuum effect is 
produced in the combustion chamber, 
since this would lead to engine oil being 
drawn in. 


Alternating cylinder deactivation 

Current technology dictates that specific 
cylinders in an engine be targeted for deac¬ 
tivation. Schaeffler is currently researching a 
concept for four-cycle engines that will al¬ 
low all cylinders to be deactivated after ev¬ 
ery ignition cycle and reactivated during the 
next. Cylinder deactivation thus alternates 
within a single deactivation phase and not 
each time a new deactivation mode is intro¬ 
duced (Figure 5). The benefit is a more well- 
balanced temperature level inside the com¬ 
bustion chambers and consistent firing 
intervals for three-cylinder engines operat¬ 
ing in deactivation mode. 

Especially when such a design setup is 
used, the losses encountered when transi¬ 
tioning from operating mode to deactivation 
mode must be kept as low as possible. This 
is why residual gas is not confined, as the 
above illustration depicts. Filling the cylin¬ 
ders with fresh air also brings with it draw¬ 
backs due to the lower level of charge 
movement. 

One variant appears to be particularly fa¬ 
vorable in this context because it allows a 
small, precisely measured quantity of residu¬ 
al gas to be confined in the combustion 
chamber. The suction or induction effect that 
results from the expansion does not last long 
enough to lead to a noticeable loss in engine 
oil. The inherent benefit is that when the 
working cycle starts again, the required 
quantity of fresh air can be introduced with¬ 
out any restrictions in flow. The first and fol¬ 
lowing combustion strokes then take place 


“Rolling” Cylinder deactivation 3-cylinders 
Cylinder number 

13 2 13 2 



0 240 480 720 960 1,200 1,400 


Figure 5 Pattern of alternating cylinder 

deactivation (the red phase designates 
the active operating mode) 




178 


without a decrease in efficiency. To ensure 
that the quantity of residual gas and the vac¬ 
uum pressure assume optimum levels, the 
exhaust valves must be controlled very pre¬ 
cisely as is the case with the fully variable 
UniAir system developed by Schaeffler. This 
system realizes any required stroke in the 
cycle and can completely close the valves 
when needed. At least one two-stage switch 
must be fitted to deactivate the valves on the 
intake side. Simulations carried out on a 
three-cylinder engine point to lower overall 
fuel consumption figures being achieved 
when such a refined alternating cylinder de¬ 
activation concept is used in place of a con¬ 
ventional setup (Figure 6). 

Alternating cylinder deactivation could 
also prove interesting when it comes to 
counteracting engine-induced vibration, es¬ 
pecially in the case of three-cylinder engines. 

All deactivation systems introduced in 
the section following the next are considered 
for a basic cylinder deactivation concept. 


Switchover mode 

One of the logical requirements of this mode is 
that the driver should not be made aware of it 
when the switch is made. In other words, the 
switchover must take place in a torque-bal¬ 
anced manner. The transition between both 
modes must also occur very quickly so that the 
engine can provide good response at all times. 

When the switch is made from operation 
on all cylinders to operation on half of the 
cylinders, the position of the throttle valve 
(cylinder charge), ignition timing, and fuel 
supply are adapted accordingly to prevent a 
drop in torque (refer to Figure 7). To this end, 
the charge is first increased and the ignition 
timing is delayed. When the target charge is 
reached, the valve train is switched over and 
the ignition timing for the cylinders activated 
is realigned with the optimum performance 
setting. As soon as the injection and timing 
sequence for the cylinders to be shut down 
is deactivated, the switchover is complete. 


12 -| 


0 

c 

0 

n 

c 

o 


3 

0) 

c 

o 

O 


8 - 

4 - 

0 - 

-4 - 

-8 - 


0.00 


-12 - 


n = 2,000 min -1 , p me = 2 bar 



-12.35 


11.66 



Base 

Cylinder 

Cylinder 

Cylinder 

Cylinder 

complete 

deactivation 

deactivation 

deactivation 

deactivation 

engine operation 

Cylinder 2 

rolling 

rolling 

rolling 


deactivated 

Vacuum 

Fresh air 

Exhaust 


Camshaft 

optimized 

is trapped 

is trapped 

is trapped 


Figure 6 Different configurations for alternating cylinder deactivation in a fuel consumption comparison 









Valvetrain Systems 


11 


179 


Control of 



Engine torque 

Cylinder filling 


Ignition angle- 
efficiency 


Figure 7 Active regulation at the switchover point 

Since retarding the ignition timing momen¬ 
tarily consumes more fuel, the deactivation 
mode must remain engaged long enough 
for an overall fuel economy benefit to be 
achieved. It goes without saying that the 
longer the engine stays in this mode, the 
more fuel is saved. Such is the case when 
traveling at constant speeds on the high¬ 
way. 

The following requirements are placed 
on the switchover mechanisms: 

- The switchover process for all cylinders 
must take place in exactly the cycle 
that the control unit stipulates. 

- The aforementioned design measures 
for compensating torque must be opti¬ 
mally coordinated and harmonized. 

- The switchover point must occur dur¬ 
ing the ignition sequence. 

- Both operating states must be stable 
and reliable so that no inadvertent swi¬ 
tchovers are made. 

- Since faulty switchovers and missed 
switchovers are relevant from an ex¬ 
haust-gas perspective, a monitoring 
function must be implemented. 


Requirements in a system 
environment context 


Even if the switch from one operating state 
to another is made successfully, in a torque- 
balanced fashion, the vibration characteris¬ 
tics of the engine and acoustic output still 
change. This may, in turn, necessitate mod¬ 
ifications to the following components (refer 
to Figure 8): 

- Phasing unit 

- Timing drive 

- Auxiliary drive assembly 

- Clutch and dual-mass flywheel 

- Exhaust system (sound engineering) 

- Engine mounts 

Depending on the application scenario 
and the requirements it entails, it is typi¬ 
cally a good idea to integrate an active 
noise compensation facility for the pas¬ 
senger compartment. Nonetheless, it is 
generally necessary to operate the engine 
on all cylinders until the engine speed 



















180 


ECU 


Must be 
changed 


- Active requirements 

- Cycle control 

- undetectable switching 

- OBD-Function 

- Driver command prediction 


Valve train 
Accessory drive 


Modifications 
may be required 


Cam phaser / 
Timing drive 



NVH measures 


Active 

noise 

cancellation 


Exhaust system 


Torque- 
converter 
lockup clutch 
(automatic 
transmission) 


Dual mass 
flywheel 


Engine mounts 


Figure 8 Overview of the measures accompanying cylinder deactivation 


reaches approximately 1,500 rpm, de¬ 
pending on the engine concept, as this 
will ensure the desired level of comfort for 
passengers. In addition, cylinder deacti¬ 
vation cannot be engaged if the engine oil 
has not reached operating temperature, 
or engaging the mode would cause the 
catalytic converter to drop below its light- 
off temperature. 


Valve stroke deactivation 


As already mentioned, it is not practical to 
also disengage the moving parts of the 
crank drive during cylinder deactivation. De¬ 
activating the valve stroke sequence, on the 
other hand, can be realized with compara¬ 


bly moderate outlay. The following options 
are available for this purpose: 

- Switchable bucket tappets 

- Switchable finger followers 

- Switchable pivot elements 

- Cam shifting systems 

- Fully variable mechanical valve train 
systems based on detent cam gears 

- Fully variable electrohydraulic valve 
train systems such as the UniAir sys¬ 
tem from Schaeffler 

Most of the switchable elements are actu¬ 
ated using oil pressure, which is con¬ 
trolled and regulated by an upstream 
switching valve. The concept requires an 
additional switching or shifting oil circuit 
to be implemented, whereby special at¬ 
tention must be paid to ensuring the cor¬ 
rect positional arrangement and geometry 
of the oil channels in order to create a 





































Valvetrain Systems 


11 


181 


hydraulically robust 
system and avoid 
air pockets as well 
as throttling or 
restriction points. 
Figure 9 shows a 
basic sketch of a 
system that has 
one switching valve 
per cylinder. 




Cyl 1 ” 

Cyl 2 

Cyl iL 

Cyl 4” 

\(( 

Exhaust 

£ 

P 

13> % 

p -> t 

'P T P 

r 

D if ^ 

£ 

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l 

¥> £ 

'P T P 

l 

w 

5I> 

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l 

Figure 9 

Basic illustration of a valve stroke deactivation system featuring 


one switching valve per cylinder 


Design configuration of the shifting 
oil circuit 

Several solutions are conceivable for con¬ 
trolling hydraulically actuated, two-stage 
valve train components and arranging the 
switching valves in the cylinder head. The 
positional arrangement of the switching 
valves and the design configuration of the 
oil channels produce different switch time 
intervals and system-related constraints. 
The following depicts two different options 
for deactivating cylinders 2 and 3 in a four- 
cylinder engine with an ignition timing se¬ 
quence of 1-3-4-2 and describes the inher¬ 
ent benefits and drawbacks in detail. 

Figure 10 shows the variant with one 
switching valve per cylinder, which means 
that one switching valve at each cylinder 
controls the respective intake and exhaust 
valves. 


The benefit of this design lies in the short oil 
channels and small oil volume. Any oil foam¬ 
ing that could occur would therefore be min¬ 
imal, which is why the system is highly insus¬ 
ceptible to fluctuations in the shifting or 
switching times. This concept enables a 
switching time interval of approximately 
250 camshaft degrees, which equates to 
a theoretical switching time of 28 ms at 
3,000 rpm. On engines with camshaft phas¬ 
ing units, the influence of the adjustment 
range must also be factored into determining 
the interval. By design, this variant can be 
enhanced or extended in such a way that all 
cylinders can be actively switch-controlled, 
which in turn means that in a four-cylinder 
engine application, the engine management 
system can deactivate one, two, or three of 
the four cylinders. One drawback, however, 
is the comparably expensive design configu¬ 
ration associated with the oil channel be¬ 
tween the intake 
and exhaust sides. 

An alternative ar¬ 
rangement is also 
possible by control¬ 
ling the oil circuit 
using one switching 
valve on the intake 
and exhaust sides 
(Figure 11). The in¬ 
take and exhaust 
valves are then ac¬ 
tuated by two sep- 


Intake 







Cyl 1 


Cyl 2 




Cyl 4 




Exhaust 


P 


Figure 10 Oil circuit with one switching valve per cylinder 



































































182 


Intake - -- ^ 

- p 

if 

/6~c\ 

/( ) ( MO) COA /CO) COA 

Cyl 1 Cyl 2 Cyl 3 Cyl 4 

vu c mo coy \o coy vu c y 

Exhaust 

A 

P 

-1 

u 


Figure 11 Oil circuit with one switching valve per side 


arate switching 
valves. The benefit of 
this arrangement is 
that the switching 
time interval can be 
governed indepen¬ 
dently of the adjust¬ 
ment range of the 
camshaft phasing 
unit. In addition, the 
oil channels can be 
designed in a more 
simplistic manner, 
and the switching 
valves can be inte¬ 
grated more easily. This design facilitates 
a switching time interval of approximately 
180 camshaft degrees, which corresponds 
to a theoretical switching time of 20 ms at 
3,000 rpm. The longer oil channels do pose 
a limitation, however, as they require a higher 
oil volume, which in turn makes the system 
more susceptible to fluctuations in the shift¬ 
ing or switching times as a result of the great¬ 
er potential for oil foaming to occur. 

The shifting oil circuit and switching valve 
linkage can also be implemented in ways 
other than the ones described here. Critical 
design aspects that apply in this context are 
the ignition timing sequence and configura- 



Figure 12 Switchable finger follower 


tion of the cylinder head and oil channels of 
the target engine, whereby the main focus of 
the design work should be on maximizing 
the switching time interval as far as possible 
using justifiable levels of outlay. 


Deactivation via switchable elements 
Finger followers 

Since the design configurations for the 
switchable finger follower can also be ap¬ 
plied to the switchable bucket tappet, we 
will not explore this topic any further. 

The solutions that are based on finger 
followers or hinged-lever designs that can 
be coupled with one another and have a 
locking mechanism at the pivot point are 
numerous. All systems that rely on oil pres¬ 
sure require spring-actuated elements to 
return the deactivated components to their 
starting position after cam elevation (Figure 12). 
The shift mechanism must be designed in 
such a way that the entire valve stroke is 
traveled when no oil pressure is present 
(zero-pressure lock), since this safeguards 
operation in limp-home mode and is re¬ 
quired for cold-starting the engine. 

Although cylinder deactivation brings 
with it many benefits, the concept also has 
several drawbacks. The additional contact 
points and increased number of compo- 






























Valvetrain Systems 


11 


183 


nents, for example, reduce rigidity as com¬ 
pared to a standard finger follower and 
negatively affect the vibration of the valve 
train. The added components also increase 
the mass moment of inertia of the follower, 
which in turn means that stronger valve 
springs need to be fitted, and the valve train 
assembly encounters higher levels of fric¬ 
tion as a result. Potential space restrictions 
necessitate narrower rollers, a design that 
inherently increases the surface contact 
pressures between the roller and camshaft. 

Switchable finger followers that brace 
themselves against a zero-stroke cam in 
deactivation mode create a more stable 
system than the variant that does not pro¬ 
vide for this effect. The only drawback is 
that the camshaft then requires two differ¬ 
ent profiles per valve. If a zero-stroke cam is 
not provided, the acting forces must be pre¬ 
cisely coordinated with each other; in the 
decoupled state, the lost-motion spring 
needs to be strong enough to prevent “infla¬ 
tion” or “pump-up” (undesired elongation) of 
the support element. On the other hand, the 
spring must not be so rigid that the motor 


valve inadvertently opens in the direction 
opposing the valve spring pressure. 

Support element 

The switchable pivot element also lends it¬ 
self to being deactivated. Similar to the 
switchable roller tappet, the inner part of the 
element can be telescopically extended into 
the outer part (Figure 13). Here too, a spring 
or spring assembly is required to return the 
moving part to its starting position. The oil 
pressure, which is controlled by an up¬ 
stream switching valve, is also used to actu¬ 
ate the coupling mechanism. The distance 
traveled by the oil to this mechanism is 
shorter, however. The same restrictions that 
apply to the switchable finger follower with 
regard to the oil pressure also hold true for 
this application. 

The rigidity of the valve train is only re¬ 
duced by the structural integrity of the cou¬ 
pling point in the switchable pivot element. 
The geometry (with the exception of the 
valve contact surface) and mass moment of 
inertia of the finger follower are unaffected. 
As a result, the valve spring pretension force 



Figure 13 Switchable pivot element 










184 


does not have to be changed in comparison 
to that of the conventional valve train as¬ 
sembly, which means that the surface con¬ 
tact pressures between the roller and cam¬ 
shaft remain at low levels. 


Deactivation via a cam sliding system 

The cam shifting system allows the valve 
stroke to be switched in up to three stages. 
The switchover process occurs when a 
cam piece positioned in an axially movable 
arrangement on a splined shaft is displaced. 
This sliding cam piece comprises several 
cams sectioned into two groups, which are 
arranged in relation to the two valves on 
each side of a cylinder (Figure 14). 

A control groove is integrated into the 
sliding cam piece. When the cam lift is to be 
adjusted, an electromagnetically actuated 
pin extends into this groove to force the en¬ 
tire unit to change its respective groove 



Figure 14 Two-stage cam shifting system 



Figure 15 Three-stage cam shifting system 


contour position. A second cam profile (or a 
third one in the case of three-stage sys¬ 
tems) thus acts on the finger follower to 
transfer the new cam lift (which can also be 
a zero-stroke) to the valve so that each valve 
pair can be actuated individually. The inher¬ 
ent benefits of this system are that the cylin¬ 
ders and camshaft can be switched selec¬ 
tively and the sequence of the elements to 
be switched is variable. 

After the actuation sequence has taken 
place, a relay signal generated by the actua¬ 
tor pin as a result of a voltage shift in the 
electric coil is sent to the actuator. Although 
this signal provides clear indication of a shift 
occurring and the direction that was taken, 
it is not sufficient for determining positional 
arrangements as operation continues (OBD 
requirement). The cam shifting system of¬ 
fers a benefit here that initially appears to be 
the exact opposite: Both valves are forced 
to switch at the same time. This, in turn, 
makes it considerably easier to detect cor¬ 
rect position during active operation by way 
of sensors (pressure or oxygen sensors) on 
the intake and exhaust side or by evaluating 
torque imbalance than when systems with 
individual switch logic are used. 

When viewed from the perspective of a 
cost-benefit analysis, it is important to note 
that the cam shifting system requires more 




Valvetrain Systems 


11 


185 


outlay than switchable elements, since in 
four-cylinder engine applications, both 
camshafts must be equipped with a deacti¬ 
vation function - a design aspect that also 
affects positional elements that are not 
switchable. Consequently, the cam shifting 
system is a commercially viable option for 
cylinder deactivation if an existing two- 
stage system for varying the valve stroke is 
enhanced to include a third stage dedicated 
to the cylinder deactivation process (refer to 
Figure 15). 

Theory-based investigations conducted 
by Schaeffler indicate that a three-stage 
system can offer further significant potential 
compared to a two-stage solution in con¬ 
sumption testing cycles carried out under 
higher load conditions. When the cam shift¬ 
ing system is designed so that all intake and 
exhaust valves can be deactivated, it is pos¬ 
sible to deactivate any desired number of 


cylinders. This setup also facilitates the inte¬ 
gration of an alternating cylinder deactiva¬ 
tion pattern [3]. 

Cylinder deactivation via UniAir 

UniAir not only controls and regulates 
valve stroke travel in a fully variable fashion, 
but can also completely deactivate any 
cylinder (Figure 16). This deactivation is 
achieved by actuating the system's integrat¬ 
ed switching valves as required. In its cur¬ 
rent version, UniAir actuates both valves in a 
uniform manner. As a result, both intake 
valves are always closed in deactivation 
mode. The operating state of the valve train 
can thus be easily determined with the Uni¬ 
Air system as well. When UniAir is only used 
on the intake side, switchable support ele¬ 
ments can be fitted in the relevant positions 
on the exhaust side (as is the case with the 
fully-variable me¬ 
chanical system). 

Schaeffler is cur¬ 
rently working on ad¬ 
ditional valve stroke 
configurations that 
approach the poten¬ 
tial afforded by cyl¬ 
inder deactivation 
while making it pos¬ 
sible to forego valve 
deactivation on the 
exhaust side. The 
genuine appeal of 
this type of configu¬ 
ration is that it allows 
any number of cylin¬ 
ders to be deacti¬ 
vated without having 
to implement further 
design measures. 
Detailed informa¬ 
tion is provided in 
an additional article 
[4] in this book. 



Figure 16 Electrohydraulic, fully-variable UniAir valve train system 


186 


Summary and outlook 


Temporarily deactivating cylinders offers an 
attractive compromise between downsizing 
an engine to reduce fuel consumption and 
retaining high levels of comfort and driving 
pleasure. Even three-cylinder engines can 
profit from the economical benefits of cylin¬ 
der deactivation. Simulations point to the 
potential that an alternating cylinder deacti¬ 
vation system has for maintaining a bal¬ 
anced temperature level in the engine and 
reducing vibrations, particularly in three- 
cylinder engine applications. 

Several options are available for tempo¬ 
rarily deactivating valves, especially in the 
context of finger follower regulation sys¬ 
tems. When cylinder deactivation is the only 
variable aspect required, switchable pivot 
elements offer a very cost-effective solution 
without noticeably compromising the basic 
functions of the valve train assembly. In the 
case of multi-stage systems or entire engine 
families, cam shifting systems are more fa¬ 
vorable because they can be easily adapt¬ 
ed. Fully-variable valve train systems go 
hand in hand with cylinder deactivation in 
the presence of discretely switchable ele¬ 
ments as a minimum expenditure item. 
Depending on the size of the engine and the 


expectations customers have regarding 
comfort levels, additional design measures 
may be required for the engine and overall 
vehicle that conflict with the potential for re¬ 
ducing fuel consumption, which can be es¬ 
pecially prominent in lightweight vehicles 
equipped with powerful engines. 

In the future, it is highly probable that 
cylinder deactivation will play an ever in¬ 
creasing role in optimizing powertrains that 
use engines with three or more cylinders. 


Literature 


[1] Middendorf, H.; Theobald, J.; Lang, L; Hartel, K.: 
Der 1,4-l-TSI-Ottomotor mit Zylinderabschal- 
tung. MTZ, 3/2012, pp. 186-193 

[2] Kirsten, K.; Brands, C.; Kratzsch, M.; Gunther, M.: 
Selektive Umschaltung des Ventilhubs beim 
Ottomotor. MTZ, 11/2012, pp. 834-839 

[3] Faust, H.: Powertrain Systems of the Future: 
Engine, transmission and damper systems for 
downspeeding, downsizing, and cylinder deac¬ 
tivation. 10 th Schaeffler Symposium, 2014 

[4] Plaas, M.; Piecyk, T.: Get ready for the combustion 
Strategies of tomorrow. 10 th Schaeffler Sympo¬ 
sium, 2014 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



188 


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189 


Get Ready for the 

Combustion Strategies of Tomorrow 


Michael Haas 
Thomas Piecyk 



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190 


The variable valve train - 
a tool for greater 
efficiency 


In the future, internal combustion engines 
will have to fulfill increasingly stringent re¬ 
quirements with regard to carbon dioxide 
emissions and exhaust pollutants, and this 
means a decisive role for the valve train: On 
the one hand, it should be designed in 
such a way that the losses occurring dur¬ 
ing the charge cycle are low, and on the 
other hand it creates the prerequisite for 
the best possible mixture preparation in 
the cylinder and thus a combustion pro¬ 
cess that provides optimum efficiency and 
low emissions. In addition, the valve timing 
directly influences the combustion process 
by way of the compression, which is ad¬ 
justable within limits, and the residual gas 
in the cylinder. 

The variability of valve trains has there¬ 
fore increased dramatically in the last few 
years. Two basic approaches for a higher 
degree of variability must be observed in 
this context: 

1. The temporal shifting of the valve lift 
curve using camshaft phasing units 

2. The variation of the valve lift curve in 
terms of the lift height and the opening 
and closing point, and thus of the result¬ 
ing opening period. 


Camshaft phase adjustment 

Ever-increasing numbers of gasoline en¬ 
gines have a camshaft phasing system - ei¬ 
ther on the intake side only or on the intake 
and exhaust side - and a volume-produced 
diesel engine recently went into production 
with a phasing system on one camshaft for 
the first time. Systems with hydraulically- 
actuated swivel motors have become es¬ 


tablished. The trend towards downsizing 
and downspeeding will increase the rate 
with which these systems are fitted because 
power and torque can be increased and 
raw emissions reduced by changing the 
relative angle between the camshaft and 
the crankshaft. 

Electric phasing units are the optimum 
solution from a technical perspective. An 
electric system allows the greatest degree 
of freedom when selecting the timing for 
starting. It offers higher rigidity when 
torque is applied to the camshaft via the 
crankshaft and therefore achieves the 
highest adjustment accuracy. The adjust¬ 
ment speed is also higher compared with 
the best hydraulic systems. The electric 
system is also the only system to offer the 
option of free selection of the timing when 
the engine is started. Such a system will go 
into volume production for the first time at 
Schaeffler in 2015. It is designed so that no 
modifications to the cylinder head are re¬ 
quired. 

The high performance of electric phas¬ 
ing units is also associated with a higher 
outlay, however. This is why it is advisable to 
further optimize the systems currently used. 
Further development of these systems must 
be focused on meeting increasing require¬ 
ments at comparatively low oil pressures. 
How this can be done is described in a sep¬ 
arate article [1]. 


Variable valve lift curves 

In gasoline engines, incrementally variable 
valve trains on the intake and exhaust side 
have been known for many years, and a 
diesel engine with cam profile shifting im¬ 
plemented on the exhaust side went into 
volume production a few months ago. 

When an incrementally variable valve 
train system is designed to act on the cam, 
this is referred to as a “shifting cam”. In this 
type of system, an electromagnetically op- 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3 12, © The Author(s) 2014 



Variable Valvetrain 


12 


191 


erated actuator axially shifts a cam assem¬ 
bly that is mounted on the camshaft. An 
advantage here is the independence from 
the engine’s hydraulic circuit with its depen¬ 
dency on temperature and viscosity. 

There are many known variants of valve 
lift curve shifting on the cam contact part¬ 
ner or the hydraulic pivot element, in which 
the task of switching between the valve lift 
curves is performed by a hydraulically-actu¬ 
ated locking piston. 

Partially-variable valve trains allow both 
valve lift curve shifting and valve/cylinder 
deactivation to be implemented. These sys¬ 
tems are attractive in that they offer signifi¬ 
cant benefits in terms of fuel consumption 
with only a moderate cost outlay. 

Engine designers have been looking for 
a way to regulate the lift of both intake and 
exhaust valves for a long time. The ideal 
situation would be to have a valve lift curve 
that is adjusted to suit the engine’s current 
operating point and condition and can be 
defined as desired, which would make it 
possible to set the timing in such a way that 
it is not a compromise between the diverse 
requirements of individual sub-targets. 

In 2009, Schaeffler and Fiat collabora¬ 
ted to put the UniAir electrohydraulic valve 
train system, the design and function of 
which has been described multiple times 
[2-4] into volume production. The scope of 
delivery from Schaeffler comprises the fol¬ 
lowing: 

- The electrohydraulic actuator module 

- The software required for controlling 
the valve control system, which is inte¬ 
grated into the customer's engine con¬ 
trol system 

- A calibration data set for the relevant 
application 

This system has since been adapted for a 
range of volume-production engines with 
capacities ranging from 0.9 to 2.4 liters and 
more than 500,000 units have been deliv¬ 
ered to customers in Europe and in North 
and South America. 


Development level of UniAir 


The basic function of the UniAir valve train 
has not changed since its market launch. 
Figure 1 shows a typical assembly. The 
camshaft acts on a finger follower, which 
drives the pump (4) that fills a high-pressure 
chamber (6). Depending on the position of 
the solenoid valve (5), the oil pressure acts 
on the engine valve via a piston or is re¬ 
duced via outflow into the medium-pressure 
chamber (3) and pressure accumulator (1). 
By adjusting the solenoid valve - temporar¬ 
ily decoupled from the camshaft position - 
various pressure levels in the high pressure 
chamber and thus variable engine valve lifts 
can be achieved. In today’s applications, 
the maximum pressure in the high-pressure 
chamber is approx. 150 bar in continuous 



Figure 1 Typical setup of a UniAir actuator 
with the components: 

1 - Pressure accumulator 

2 - Oil supply 

3 - Medium-pressure chamber 

4 - Oil pump 

5 - Solenoid valve 

6 - High-pressure chamber 

7 - Valve brake 





192 


operation. The peak pressures that are ac¬ 
ceptable for short periods are as high as 
200 bar. For energy reasons, a portion of 
the oil flows from the medium-pressure 
chamber into a pressure accumulator (1). 
The oil supply (2) is provided by the engine 
oil circuit. 

After the pressure in the high-pressure 
chamber drops, the engine valve is closed 
by the valve spring. The guide for the piston 
that is responsible for opening the valve has 
small bores that allow the oil to flow out in a 
controlled manner and thereby act as a 
brake (7). 

In addition, a temperature sensor is also 
required in order to compensate for the hy¬ 
draulic effects produced by the tempera¬ 
ture-dependent viscosity of the oil. All other 
parameters required for controlling the Uni- 
Air system - such as the camshaft speed - 
are provided by sensors that are already 
employed. 

The UniAir system is not restricted to 
application in engines with one intake valve 


per cylinder, however. Two intake valves of 
the same cylinder can be operated using 
either individual activation or with a hydrau¬ 
lic or mechanical bridge (Figure 2). For cost 
reasons, only the variant with a hydraulic 
bridge is currently in volume production. In¬ 
dividual activation would, however, provide 
an even higher degree of flexibility. 

The systems that are already installed in 
volume-production applications today allow 
a significant degree of variability to be 
achieved in both the valve lift and the open¬ 
ing times (Figure 3). The maximum lift and 
the earliest opening point are specified by 
the envelope curve of the cam that is used 
to drive the system. The same applies for 
the latest possible closing point. Within 
these limits, regulation is carried out exclu¬ 
sively via the current controlling the solenoid 
valve. 

The overall result is that the valve open¬ 
ing and closing times and the valve lifts can 
be optimally adjusted for all engine operat¬ 
ing points (Figure 4). 


Single actuation 
Camshaft 


Simultaneous actuation 
(hydraulic bridge) 


Simultaneous actuation 
(mechanical bridge) 



Engine 

valve 



Figure 2 Alternative solutions for operating two intake valves 

































































































Variable Valvetrain 


12 


193 


Full lift 



Late valve closing 




Time 


Early valve closing 



Time 


No lift 




CO — 

55 8 2E£1 


- 2 
C O 
Q> C 

t 0) 

o 8 


Time 


Figure 3 Variability of valve lift and opening/closing point with the current UniAir system 


During the closing time, the current curve 
required to close the solenoid valve - i.e. 
to open the engine valve - displays the 
typical V-shape that is caused by the 
valve reaching its end stop position. The 
position of the turning point in the V-pro- 
file indicates the extent to which the de¬ 
sired lift was achieved by the solenoid 
valve. The reproducibility can be traced 
by making a comparison of several con¬ 
secutive events. If deviations occur here 
that are not within the tolerance limits, 
e.g. due to aging components, these can 
be compensated by changing the current 
curve. 


Experience with volume-production en¬ 
gines to date has shown that UniAir displays 
very good values, both in terms of repro¬ 
ducibility - i.e. deviations from cycle to cycle 
in one cylinder - and of the system’s preci¬ 
sion - i.e. the spread across several en¬ 
gines. UniAir thus achieves a repeat accu¬ 
racy of 0.4 crankshaft degrees at 3,000 rpm 
and a system temperature of 120 °C for the 
“early intake closure” function. The opening 
angle during “late intake valve opening” - 
which is decisive for cylinder balancing - 
also achieves a precision of 0.4 crankshaft 
degrees. Under the conditions described, 
the deviations between various volume-pro- 










































194 



Figure 4 Valve lift curves within the engine map 

duction engines are at a slightly higher level 
for “early intake valve closure” than for indi¬ 
vidual cylinder actuation. A compensation 
function that is integrated into the UniAir 
software also ensures that the cylinders of 
the respective engine are correctly bal¬ 
anced here, however. 

The market launch of UniAir has 
proved that the system can be applied in 
such a way that no changes to the design 
envelope are required. The prerequisite 
for this is that the UniAir system and the 
remaining standard valve train are both 
driven by a common camshaft that is still 
installed. In the case of a direct injection 
gasoline engine with centrally-positioned 
injectors, the injector and spark plug must 
be arranged perpendicular to the cam¬ 
shaft assembly. If this is carried out, the 
intake camshaft (for example) can be 
omitted, which means that the additional 
costs of the UniAir system can be partially 
compensated. In the future, however, 
there will also be applications in which 
both camshafts are retained. 


Expansion of the scope 
of functions 


Applications for gasoline engines 

The trend for low-consumption gasoline en¬ 
gines with direct-injection and increasingly 
small engine capacities is continuing un¬ 
checked. The introduction of new standard 
cycles for measuring fuel consumption, par¬ 
ticularly the WLTP, mean that operating 
points with higher loads are being achieved 
at the same time, in which the benefits pro¬ 
vided by the combination of downsizing and 
turbocharging cannot be exploited to the 
same degree. When combined with turbo¬ 
charging, a fully variable valve train can 
therefore contribute towards reducing fuel 
consumption even further. 

In turbocharged engines, the flushing 
of the cylinder with fresh air (“scavenging”) 
provides significantly faster response times 



















































Variable Valvetrain 


12 


195 


at low speeds and high loads. The valve 
overlap that this requires is conventionally 
achieved through the use of camshaft 
phasing units. This type of camshaft phas¬ 
ing unit, which is characterized by its high 
adjustment speed, is available from 
Schaeffler [4]. When designed correctly, 
UniAir can partially replace camshaft phas¬ 
ing units of this type. Although it is not pos¬ 
sible for this system to influence the point 
at which the valve’s envelope curve begins, 
the variation of the opening point together 
with a special lift curve (see Figure 8) can 
be used to only activate valve overlap when 
it is required. 

Dethrottling at low speeds is known to 
have a very positive effect on the fuel con¬ 
sumption. It is important that the smaller 
valve lifts are not designed in such a way 
that a vacuum that could cause load cycle 
losses occurs in the cylinder during the 
intake stroke. This is an argument in favor 
of systems in which the lift height and lift 
duration can be varied to the same extent 
(Figure 5). Compared with an engine op¬ 
erated with a standard valve train and 
camshaft phasing units, an 8.4 % reduc¬ 
tion in the specific fuel consumption that 
has been measured and verified is achieved 
at the operating point of 2,000 rpm and 
2 bar. 



— Exhaust — UniAir 

— Intake — 2-step system 


Figure 5 Lift curves for dethrottling at low 
speeds: UniAir compared with a 
two-stage cam profile 



° Crankshaft 


— Exhaust — Valve 1 
— Intake Valve 2 

Figure 6 Lift curves of two intake valves of a 
cylinder when operated individually 

An increase in the charge motion in the cyl¬ 
inder (“swirl”) can contribute to improving 
the formation of the mixture and thus mak¬ 
ing the combustion process more efficient, 
especially with low loads of the kind typi¬ 
cally found in city traffic. If both of a cylin¬ 
der’s intake valves are individually operated 
by a UniAir system, individual lift profiles for 
the valves can be illustrated (Figure 6). 

Valve lift curves for city traffic/operation 
under low load conditions are also being 
developed (Figure 7) that can only be illus¬ 
trated with UniAir and not with mechanical 
solutions for a fully variable valve train. The 
“hybrid lift” function combines late opening 
of the intake valve with early closure, a pro- 



° Crankshaft 

— Exhaust 
Intake 


Figure 7 Special valve lift curves: 1) Hybrid lift 
and 2) Multilift 










196 


cess in which the ramps are not symmetri¬ 
cal with one another. 

The “multilift” function opens and closes 
the intake valve twice within the intake stroke, 
which produces an optimum combustion 
process and low pumping losses at low 
speeds and under low to medium load con¬ 
ditions, and is therefore particularly suitable 
for optimizing fuel consumption in city traffic. 

Under medium load conditions, opera¬ 
tion using the Miller cycle - i.e. rapid and pre¬ 
mature closing of the intake valve - is a good 
option. The improved expansion ratio im¬ 
proves the engine’s degree of efficiency. The 
Miller cycle can easily be implemented with 
the UniAir system. The same applies for op¬ 
eration with the Atkinson cycle under high 
load conditions, during which the intake 
valve is opened for longer. This provides the 
desired reduction in compression without 
the charge motion in the cylinder being de¬ 
stroyed. The lower degree of compression in 
the high load range reduces the tendency 
towards knocking, which is of relevance for 
modern, supercharged gasoline engines. 

Schaeffler and Continental have tested 
the application suitability of a combination 
of Miller and Atkinson cycles controlled ex¬ 
clusively by UniAir in a joint advance devel¬ 
opment project. For this purpose, a 1.4-liter 
volume-production engine that was already 
equipped with UniAir was fitted with a differ¬ 
ent engine control system in order to corre¬ 
spondingly optimize the process control. 
The results show significant potential with 
regard to fuel consumption: 

- The use of the Miller cycle reduces com¬ 
pression at the operating point of 2,000 rpm 
and 12 bar, which produces lower final 
compression temperatures and thus 
a reduction in the tendency towards 
knocking. The specific fuel consump¬ 
tion is improved by 4.4 % compared to 
the volume-production engine. 

- If the mean pressure is increased to 
15 bar at the same speed, a significant¬ 
ly lower tendency towards knocking is 


observed due to the closure of the in¬ 
take valve after the beginning of the 
compression stroke (Atkinson cycle). 
The specific fuel consumption is re¬ 
duced by 4.6 %. 

- At a typical full-load point (3,000 rpm 
and 18 bar), the use of the Atkinson 
cycle makes it possible to reduce the 
degree of enrichment that would other¬ 
wise be required in order to lower the 
temperature in the combustion cham¬ 
ber. This leads to a 4.6 % reduction in 
the specific fuel consumption. 

The use of UniAir allows not only the fuel 
consumption but also the exhaust emis¬ 
sions to be reduced. This particularly ap¬ 
plies to nitrogen oxide (NO x ) emissions, as 
was verified for diesel engines as early as 
2008 [6]. Related designs combined inter¬ 
nal and external (low-pressure) exhaust gas 
recirculation in the engine. 

With the first-generation UniAir system, 
it was only possible to adjust the valve lift 
within the “conventional” envelope curve as 
specified by the cam. This makes it difficult 
or even impossible to achieve the kind of 
large valve overlaps required for high resid¬ 
ual gas content. This obstacle has now 
been overcome thanks to the introduction 
of a correspondingly designed “two-stage” 
cam profile for the UniAir system. This is re- 



— Exhaust 
Intake 


Figure 8 Valve overlap for internal exhaust 
gas recirculation with a two-stage 
cam as the UniAir drive 




Variable Valvetrain 


12 


197 


ferred to as a “boot cam” due to its boot-like 
shape. It is now possible for the first time to 
achieve a large valve overlap without a sig¬ 
nificant reduction in the maximum filling 
level during the intake cycle (Figure 8). 

For operating points that do not require 
internal exhaust gas recirculation (e.g. under 
full load conditions), no current reaches the 
solenoid valve that is responsible for the 
UniAir’s switching until the first stage of the 
cam has already passed over the contact 
surface with the high-pressure pump. 


UniAir for diesel engines 

Over the coming years, the further develop¬ 
ment of the diesel engine will continue to fo¬ 
cus on the reduction of NO x and soot ex¬ 
haust emissions. In order to minimize the 
outlay for costly exhaust gas aftertreatment, 
especially selective catalytic reduction (SCR), 
a clean combustion process is of the utmost 
importance for every engine designer. In ad¬ 
dition to other measures such as increasing 
the injection pressure, combustion with a 
high rate of exhaust gas recirculation is an 
important prerequisite for keeping these raw 
engine emissions to a minimum. 

The two-stage cam is not a suitable solu¬ 
tion for the diesel engine, because a signifi¬ 
cant valve overlap is not possible due to the 



Figure 9 Internal exhaust gas recirculation for 
diesel engines through early exhaust 
valve closure and late intake valve 
opening 


UniAir on intake side: 
EGR storage 
in intake manifold 



° Crankshaft 


UniAir on exhaust side: 
EGR rebreathing 



Figure 10 Internal exhaust gas recirculation for 
diesel engines with double cams for 
the UniAir drive 


lack of free-running clearance between the 
piston crown and the engine valve at the top 
dead center. An alternative concept would be 
to control the quantity of residual gas remain¬ 
ing in the combustion chamber within wide 
limits through early closure of the exhaust 
valve and late opening of the intake valve (Fig¬ 
ure 9). The first-generation UniAir system can 
already provide the lift curve variability that this 
requires - however, UniAir must also be used 
on the exhaust side in this case. 

Double cams to actuate the UniAir sys¬ 
tem on the intake or exhaust side provide a 
further solution for achieving even higher 








198 


internal exhaust gas recirculation rates if re¬ 
quired. It is ensured here that the valve lift 
curve of the smaller cam can also be con¬ 
trolled (Figure 10) and that it is thus possible 
to control the exhaust gas recirculation rate. 
All of the UniAir system’s modes still remain 
available for the primary/main cam. Opera¬ 
tion without exhaust gas recirculation is 
thus possible at all times, which is important 
at extremely low temperatures (below -10 °C), 
for example. 

For the sake of completeness, it 
should be mentioned that this is exactly 
how the thermodynamically positive ef¬ 
fects of an effective compression ratio 
that has been reduced using the Miller/ 
Atkinson cycles with UniAir can be 
achieved for diesel engines [5]. However, 
the objective here is to reduce the final 
compression temperature and thus the 
maximum final combustion temperature 
for emission reasons. 


Supporting future 
combustion processes 


Cylinder deactivation 

Despite the unmistakable trend towards 
highly charged low-displacement en¬ 
gines, engines with four cylinders and up¬ 
wards will continue to be used in large 
vehicles and those designed with high 
performance in mind. Engines of this type 
have to achieve better specific fuel con¬ 
sumption figures, especially under low 
load conditions and at low to medium 
speeds, so that the vehicle’s overall emis¬ 
sions are reduced in the common stan¬ 
dard cycles. A solution for this that has 
already been put into volume production 
by manufacturers including Audi, Chrysler, 



UniAir 


Figure 11 Cylinder deactivated by shutting 
down the intake valve 

GM, Honda, Mercedes, and Volkswagen 
is cylinder deactivation, which increases 
the load of the operating point of the cylin¬ 
ders that are not deactivated. The me¬ 
chanical systems that have been intro¬ 
duced for this purpose require a high level 
of outlay. 

It is already possible to perform simple 
cylinder deactivation using the current 
UniAir system (Figure 11). However, this 




Figure 12 Cylinder deactivation with additional 
intake valve opening 















Variable Valvetrain 


12 


199 


100 - 


96 “ 


a 

E 

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o 

o 


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CD 


Cylinder deactivation with UniAir 



<D 

C 

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c Q 
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(0 

00 + 


Figure 13 Simulated fuel consumption benefits for different cylinder deactivation strategies 


type of system does not utilize the full po¬ 
tential for reducing C0 2 emissions that is 
available with the latest mechanical de¬ 
activation systems, which is estimated at 
up to 4 % in addition to the use of the fully 
variable valve train. This is because of the 
high level of charge cycle work caused by 
the exhaust valves that are still being ac¬ 
tuated. 

If the use of additional measures on the 
exhaust side, e.g. the use of switchable piv¬ 
ot elements, has to be avoided, there is a 
further variant that is also based on the use 
of a double cam. In this case, the cam is 
used on the intake side to briefly open the 
intake valve of an unfired cylinder during the 
exhaust stroke (Figure 12), allowing the ex¬ 
haust gas to flow into the intake system. The 
intake valve in turn is then opened and then 
rapidly re-closed during the intake stroke, 
so that almost only the “stored” exhaust gas 
is allowed to flow into the combustion 
chamber. This alternating effect means that 
the overall charge cycle work is significantly 
reduced, and a large proportion of the sav¬ 
ings potential available from the complete 


deactivation of all cylinders can be achieved 
(Figure 13). 

Homogeneous charge compression 
ignition 

Homogeneous charge compression igni¬ 
tion (or HCCI for short) has been undergo¬ 
ing development for volume production for 
some time . The section of the engine data 
map in which the thermodynamic benefits 
of self-ignition can be utilized has been con¬ 
tinuously expanded, but still only covers 
part of the engine data map (up to a maxi¬ 
mum of mid-range loads and speeds). This 
is because charge stratification with pre¬ 
cisely defined composition and a high 
quantity of residual gas is the decisive factor 
for a stable HCCI combustion process. In 
addition to injection, the precise guidance of 
the charge motion combined with precise 
metering of the exhaust gas recirculation 
rate and adjusted compression can also 
have a significant positive effect on the sta¬ 
bility of the combustion process [6]. 








200 




Figure 14 Valve lift curves for an HCCI 
combustion process: Greater 
variability through the combination 
of a phasing unit (left) and UniAir 
system (right) 

In order to ensure the correct charge mo¬ 
tion at high speeds and under high load 
conditions, Schaeffler relies on a combi¬ 
nation comprising a camshaft phasing 
unit (electromechanical or hydraulic) and 
a UniAir system with a double cam drive 
(Figure 14). This fast actuator system 
makes it possible to set the correct com¬ 
pression and mixture ratio for every op¬ 
erating point. Switching between sec¬ 
tions of the data map with compression 
ignition and external ignition can also be 
achieved in a significantly faster and 
more reliable way. 


Outlook 


The UniAir system has firmly established itself 
on the market. The significantly higher degree 
of valve lift curve flexibility displayed by an 
electrohydraulic system compared to me¬ 
chanical systems, which are also in volume 
production or development, makes far more 
dynamic process control possible even today. 
The second-generation UniAir system also 
makes additional functions available. The ad¬ 
justment of the UniAir drive using two-stage or 
double cams that is required for this purpose 
has achieved a high level of maturity. 

From 2015/16 onwards, UniAir will be used 
in a range of further passenger car applica¬ 
tions, including engines equipped with different 
numbers of cylinders from those in today’s 
volume-production applications, and it is also 
set to be put into volume production by more 
automobile manufacturers. Intensive prepro¬ 
duction testing is currently being carried out on 
a four-cylinder diesel engine application. 

The first motorcycles to be equipped with 
the UniAir system will also be seen in the 
near future. In parallel to this, Schaeffler is 
also collaborating with ABB Turbo Systems 
on a project to market the UniAir system for 
use in stationary engines. An area of particu¬ 
lar interest here is the use of gas-operated 
stationary engines for energy generation. 
Even these engines will have to be controlled 
with a significantly higher degree of flexibility 
in the future without their high degree of effi¬ 
ciency being sacrificed. In applications of 
this kind, the cost savings that are achieved 
through the targeted improvements in fuel 
economy are significantly higher than the 
cost of the variable valve train system. 







Variable Valvetrain 


12 


201 


Literature 


[1] Dietz, J.; Busse, M.; Racklebe, S.: Needs- 
based Concepts for Camshaft Phasing Sys¬ 
tems. 10 th Schaeffler Symposium, 2014 

[2] Haas, M.: Just air? UniAir - The first fully- 
variable, electro-hydraulic valve control system. 
9 th Schaeffler Symposium, 2010 

[3] Bernard, L.; et al: Elektrohydraulische Ventil- 
steuerung mit dem “MultiAir”-Verfahren. 

MTZ 70, 2009, No. 12, pp. 892-899 

[4] Haas, M.; Rauch, M.: Elektrohydraulischer 
Vollvariabler Ventiltrieb. MTZ 71, 2010, No. 3, 
pp. 160-165 

[5] Ruggiero, A.; et al.: Combustion Technologies 
to meet EURO 6 Emission Standard on Diesel 
D-segment Passenger Car. SIA International 
Conference „Diesel Engine: the low C0 2 and 
emission Reduction Challenge”, Rouen, 2008 

[6] Schutting, E.; et al.: Miller- und Atkinson-Zyklus 
am aufgeladenen Dieselmotor. MTZ 68, 2007, 
No. 6, pp. 480-485 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



202 


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203 


Rolling Bearings in Turbochargers 


A real bargain with regard 
to C0 2 emissions 


Chris Mitchell 
Christian Schaefer 
Oliver Graf-Goller 
Peter Solfrank 
Martin Scheldt 


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204 


Introduction 


Although the internal combustion engine 
is likely to still dominate the automotive 
landscape for the next decade or so, the 
increasing mismatch between energy 
consumption and available resources, to¬ 
gether with tighter legal restrictions on en¬ 
gine C0 2 emissions, is creating an in¬ 
creased demand for improvements to 
existing automotive technologies and the 
development of reduced friction, more en¬ 
ergy efficient, ‘greener’ alternatives. At the 
same time, an increased awareness of 
air pollution has resulted in more and 
more stringent regulations on automotive 
engine emissions that drive technology 
developments. 

Gasoline and diesel fuel internal com¬ 
bustion engines are positioned com¬ 
pletely differently with regard to the 
conflicting aims of fuel consumption and 
emissions (Figure 1). 

The gasoline engine is clearly in the low 
emission category due to its very efficient 
after treatment of exhaust gases. 


Emissions target 


1 

Gasoline 




It 




1 


Dies< 

3 l C0 2 target 

Target 

range 

1 

rj 




Emissions 

(HC, NO x , particulates) 


■ Transmission technology 

■ Hybridization 


However, the spark ignition engine has in¬ 
herently lower thermodynamic efficiency 
and hence has high fuel consumption. 

The diesel engine, on the other hand, 
has a place in the low fuel consumption cat¬ 
egory due to its favorable, thermodynamic 
efficiency and advantageous low end torque 
characteristics. This supports the trend to¬ 
wards downspeeding for a further reduction 
in fuel consumption. 

However, the compression ignition en¬ 
gine suffers with high exhaust emissions, 
HC, NO x and particulates. 


Forced Induction 


In order to support the growing demand 
for more energy efficient, low carbon emis¬ 
sion vehicles, manufacturers of forced in¬ 
duction systems, particularly for passenger 
cars and commercial vehicles, are being 
asked to provide more compact, higher ef¬ 
ficiency systems that are both durable and 
affordable. 

Forced induction, achieved by both turbo¬ 
charging and supercharging, allows an en¬ 
gine to burn more fuel and air mixture by 
packing more oxygen molecules into the 
existing cylinders. Thus, the engine is able 
to deliver more power output per combus¬ 
tion stroke. 

Forced induction is a key strategic tech¬ 
nology for engine downsizing, permitting a 
small displacement engine to deliver a pow¬ 
er output similar to larger naturally aspirated 
engines, as well as downspeeding, permit¬ 
ting the same power output with lower en¬ 
gine speed. Friction reduction and further 
improvement of thermodynamic efficiency 
at high specific loads are the drivers for this 
development. 


Figure 1 Fuel consumption/emissions 

Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle, 
DOI 10.1007/978-3-658-06430-3_13, © The Author(s) 2014 







Rolling Bearings for Turbochargers 


13 


205 


Turbocharger 


A turbocharger is a device that uses the 
energy of exhaust gases emitted from an 
engine to compress the air going into the 
engine. 

The core of a turbocharger is a rotating 
shaft coupling two wheels, a turbine wheel 
and a compressor wheel. The rotation of 
the shaft is supported by a bearing system. 

The turbine wheel is positioned in the 
exhaust stream of the internal combustion 
engine. The exhaust gas from the cylinders 
passes through the turbine blades, causing 
the turbine to spin. The more exhaust gas 
that goes through the blades, the faster they 
spin. 

The compressor wheel is positioned 
before the air intake to the cylinders and 
through rotation of the shaft supplies 
compressed air to the combustion cycle 
by increasing the number of the oxygen 
molecules. 


Conventional bearing systems associated 
with turbochargers are oil film based. The 
shaft and wheel assembly is supported by a 
controlled oil film thickness to facilitate both 
rotation and ensure dynamic stability at very 
high speeds. 

Advantages of ball rolling element 
bearing turbochargers over the conven¬ 
tional oil film turbocharger bearing systems 
originate from the fundamental change in 
the friction mechanism present in the sys¬ 
tem. Multiple rolling elements replace a thin 
oil film under high-shear conditions, signifi¬ 
cantly reducing system friction. This results 
in a significant improvement in system fric¬ 
tion at operating temperature (typically up 
to 50 %) and even greater improvements 
during the first minute of an engine cold 
start (Figure 2). 

With the more conventional oil film 
turbocharger bearing systems, the oil is 
very viscous in cold conditions. At this 
time, the viscous drag of the bearing 
system prevents the effective rotation of 
the shaft and hence does not supply 




— Plain bearing 

— Ball bearing 

Source: Honeywell Turbo Technologies 


— Plain bearing at 20 °C 

— Ball bearing at 20 °C 

— Plain bearing at -20 °C 
Ball bearing at -20 °C 


Figure 2 Friction loss benefits for ball bearing turbochargers 









206 



Figure 3 Schaeffler ball bearing cartridge 


sufficient boost air to the combustion pro¬ 
cess. This means reduced power output 
and increased emissions. However, with 
ball bearing turbochargers the fundamen¬ 
tal frictional change means that the flow of 
the exhaust gas, even at cold start, is suf¬ 
ficient to provide rotation to the shaft so 
that the compressor wheel can provide 
the necessary boost air to the engine sys¬ 
tem immediately. This results in a more 
energy efficient system with reduced 


emissions and also means that the driver 
experiences increased engine torque from 
the very beginning of the drive. 

Turbocharger studies have shown that 
the ‘ball bearing effect’ is most pronounced 
at low engine speeds, just where a down¬ 
speeding or downsizing concept needs 
the most help from the turbocharger sys¬ 
tem. For engine operation the reduced 
bearing friction results in higher turbo¬ 
charger speeds for the lower engine speed 
conditions mentioned above. Specifically, 
in the event of a sudden engine load re¬ 
quest during idle or low load conditions, 
the increased turbocharger speed results 
in a significant improvement of engine re¬ 
sponse due to the turbocharger’s instanta¬ 
neous ability to supply compressed air. It is 
not only the ability to avoid the “turbo lag” 
that is striking but also the improvement of 
raw emission quality due to improved fresh 
air supply. 



Figure 4 Friction power loss at various oil flow regimes 



















Rolling Bearings for Turbochargers 


13 


207 



Oil jets into bearing 



1-D CFD simulation of oil path (left); cross section of bearing with oil duct features (right) 


Figure 5 

Schaeffler ball bearings 
for Turbocharger 


Schaeffler ball bearings for turbochargers 
(Figure 3) are of the angular contact type. 
Typically, these bearings utilize ceramic 
balls, cages, anti-rotation devices, an outer 
ring, a compressor side inner ring, a tur¬ 
bine side inner ring and a series of oil 
duct features for lubrication, cooling and 
for supplying the squeeze film damper 
areas. 

Ball bearings for turbochargers rotate at 
very high rotational speeds. If the common 
characteristic speed value of bearings is 
considered, taking diameter and rotational 
speed (n • dj into account, turbocharger 
bearings run six times faster than any other 
bearing in a vehicle. By speed value, they 
compete with the peak of jet engines and 
textile machines. For these high speed con¬ 
ditions, requirements for lubrication are 
delicate: Sufficient lubricant must be pro¬ 
vided at all times, but an excess of it might 


rapidly result in significant churning losses 
(Figure 4). Hence for the Schaeffler ball bear¬ 
ing cage, consideration was given to the 
design of the internal surfaces and compo¬ 
nent geometries. 

As the ball bearing cartridge represents 
a single component of a larger system, we 
must take a closer look at the entire system 
and mutual effects. 

The oil flow to the turbocharger (Fig¬ 
ure 5), for both the oil film and the ball 
bearings, must also provide squeeze film 
damping. A squeeze film is a viscous flu¬ 
id zone which provides structural isola¬ 
tion between elements, reduces the am¬ 
plitude of rotor response to imbalance 
and also suppresses rotor-dynamic in¬ 
stability. 

The entire fluid path, and necessary 
uses, can be simulated and fine tuned 
according to the ultimate performance 
requirements for the full range of 
temperatures. 

A ball bearing can be understood as a 
series of springs of known, predictable 
stiffness. However, once coupled to the 
shaft, the wheels and the housing of the 



































































208 




Figure 6 Bearinx model, multi-body structure, 
modal shapes 


turbocharger and then revved at very high 
speed, the bearing must work well in the 
system. 

The influence of the dynamic system 
must be considered. The radial load ap¬ 
plied to the bearing typically comes from 
the residual imbalance. Here the interac¬ 
tion of the various system components 
comes into play as the structural stiffness 
of the shaft wheel assembly, the bearing 
stiffness as well as the properties of 
the squeeze film surrounding the bearing 
influence the bearing loads. At very high 
rotational speeds that are known to have 
a critical effect on shaft deflection this can 
cause damage to the internal kinematics 
of the bearing design and ultimately affect 
the life of the system (Figure 6). The axial 
load component is generated from “gas” 
pressure loading forces of the compressor 
or turbine wheel. 

Once we consider the rotational speeds 
and modal shapes, operational tempera¬ 
tures and loads, we are better able to un- 



Figure 7 CABA 3D bearing simulation 


































































Rolling Bearings for Turbochargers 


13 


209 



Figure 8 Combined FEM-CFD thermal analysis 

derstand the internal kinematics of the ac¬ 
tual bearing system. 

The motion of the balls, their interac¬ 
tions with the raceways and the cages 
have a very complex relationship. Using 
special purpose simulation tools such as 
Schaeffler’s CABA 3D, tailored to the re¬ 
quirements of bearing analysis, this mo¬ 
tion can be computed and detailed re¬ 
sults on the existence, location, extension 
and load in each individual contact be¬ 
tween the components of the bearing 
can be obtained (Figure 7). This variety of 
results can be combined to achieve a 
detailed understanding of the bearing. 
Hence, the bearing system can be de¬ 
signed for optimal performance, life 
expectancy, reduction of friction and 
materials sensitivity. 

Turbochargers must operate in ex¬ 
treme temperatures. We have already 
considered cold starts with regard to fric¬ 
tion reduction and should now explore 
the hot running requirements. The turbine 
wheel is driven by the exhaust gases of 
the internal combustion engine and is 
therefore exposed to exhaust gas tem- 


Average temperature of bearing parts 



Position in axial direction 



Figure 9 Temperature distribution in the 
bearing parts 


Temperature in 
























210 


peratures. In normal operating cycles this 
high temperature ultimately flows by 
means of thermal conduction through the 
multibody system to the ball bearing 
where it is led away by oil flow. More criti¬ 
cal are the thermal shut down conditions 
where the oil flow is stopped. 

In normal operation, the bearing can 
reach temperatures of around 300 °C on 
the turbine side, whereas in thermal shut¬ 
down conditions, these bearings can even 
reach temperatures of up to 400 °C. 

Applying both CFD and FEM to the 
problem (Figure 8), we can obtain thermal 
characteristics (Figure 9) with regard to ra¬ 
dial and axial growth parameters in addition 
to the composition or thermal conditioning 
of the material growth. 

There is a significant temperature differ¬ 
ence across the components, and the con¬ 
sequences must be taken into account in 
the internal bearing design as well as the 
material characteristics. The bearing sys¬ 
tem must withstand extreme temperatures 
and extreme running speeds and be dur¬ 
able for the long term operation of the sys¬ 


tem. It is therefore necessary to select the 
materials very carefully for the relevant ap¬ 
plication environment. 


Outlook 


The year 2014 signifies a great achievement 
for turbocharger ball bearings supplied by 
Schaeffler and its group of companies. 

For 10 years, Schaeffler has been leading 
the way in the development and supply of 
low-friction double row angular contact ball 
bearings for turbocharger technologies. Dur¬ 
ing this time, we have perfected our applica¬ 
tion analysis, design tools and manufactur¬ 
ing methods. These precision ball bearings 
have helped set new turbocharger perfor¬ 
mance benchmarks for the future, particu¬ 
larly in the passenger car, light duty and 
heavy duty truck markets, and this year we 
will deliver our 1 millionth ball bearing cage 
for turbocharger applications in this sector. 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



212 


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213 


Who’s Afraid of 48 V? 

Not the Mini Hybrid with Electric Axle! 

Modular electric axle drive in a 48-volt on-board electric system 


Thomas Smetana 



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214 


Axle drives for hybrid 
vehicles 


Development level of the eDifferential 
in a high-voltage design 

In the entire automobile industry, there is a 
discernible trend towards hybrid vehicles 
in order to meet future C0 2 requirements. 
The test cycles used for determining C0 2 
emissions favor vehicles with a long range 
of electric operation. Plug-in hybrid vehi¬ 
cles are increasingly appearing on the 
market, whose batteries can be charged 
using public or private power supply sys¬ 
tems. The driving performance required 
from these vehicles requires relatively high 
levels of electric power with low space re¬ 
quirements. 

At the Schaeffler Symposium 2010, 
Schaeffler presented a technical solution 
for these vehicles with the first genera¬ 


tion of the so-called “active electric dif¬ 
ferential” [1]. This electric axle enables 
both an optimum use of space as an axle 
drive and also active torque distribution 
to the wheels so that very good values 
for driving dynamics are achieved as 
well. 

Schaeffler has been consistently devel¬ 
oping the electric axle drive ever since. The 
third generation currently being tested is 
matched to the topology of a plug-in hybrid 
vehicle with a front mounted engine and 
front-wheel drive. The drive unit (Figure 1) is 
still designed to be fitted coaxially in the rear 
axle and is characterized by the following 
features: 

- Water-cooled electric motors in hybrid 
design (permanently excited synchro¬ 
nous motors with a high proportion of 
reluctance) are used. These meet auto- 
motive-specific requirements in con¬ 
trast to the industrial motors used in the 
first generation. 

- The transmission is still in planetary de¬ 
sign and now has two ratio stages. 



Figure 1 Section through the electric axle in a high-voltage design 

Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_14, © The Author(s) 2014 








48-volt Electric Axle 


14 


215 




Gen 1 (VEP) 

Gen 2 

Gen 3 

Application 

EV, E-AWD 

HEV, E-RWD 

HEV, E-RWD 

Dimensions 

300 x 562 

230 x 550 

230 x 525 

Weight w/o power electronics 

120 kg 

89 kg 

79 kg 

DC Battery Voltage 
min / nom / max 

350 / 400 / 450 V 

270 / 320 / 360 V 

264 / 323 / 361 V 

Max. power 

105 kW (< 10 s) 

55 kW (< 30 s) 

65 kW (< 60 s) 

Continuous output 

- 

38 kW 

45 kW 

Max. torque 

1,200 Nm 

1,800 Nm 

2,000 Nm (60 s) 

Continuous torque 

- 

- 

1,200 Nm 

Max. speed (electr. motor) 

8,900 rpm 

13,000 rpm 

14,000 rpm 

Ratio 

i = 7 

i 1 = 12.3; i 2 = 4.2 

i 1 = 12.3 ; i 2 = 4.2 

V max vehicle (electric) 

<150 km/h 

< 250 km/h 

>= 262 km/h 

Max. torque vectoring torque 

1,500 Nm 

1,150 Nm 

1,200 Nm (10 s) 

Max. differential lock 

- 

- 

1,600 Nm (5 s) 


Technical data for three generations of high-voltage axles 


Figure 2 

- The drive unit has increased power den¬ 
sity and a modular design so that trac¬ 
tion and active torque distribution can 
be offered as separate functions. 

The progress achieved in development is 
apparent if one considers the main key 
figures of the third generation (Figure 2). 
For example, the diameter was reduced 
by 70 mm to 230 mm and the weight of 
the unit was reduced by 41 kg to 79 kg. 
The peak power was reduced to 65 kW 
due to the voltage range of 270 to 360 V 
of batteries used in plug-in hybrids. Peak 
power is now available for up to 60 sec¬ 
onds. The maximum torque is 2,000 Nm 
due to the high ratio of the two-speed 
transmission. The continuous torque of 
1,200 Nm is sufficient for all convention¬ 
al driving situations. Torque vectoring 
with torque differences of up to 1,200 Nm 
can also be implemented at very high 
speeds. 


Schaeffler’s high-voltage electric axle can 
achieve the high levels of electric power, 
which are typical for hybrid and plug-in hy¬ 
brid vehicles as well as range-extenders 
and electric vehicles. The system is cur¬ 
rently undergoing field tests with automobile 
manufacturers. 

The mini hybrid with 48-volt on-board 
electric system 

After it became clear that mini hybrid vehi¬ 
cles with 48-volt on-board electric subsys¬ 
tems would be introduced in increasing 
numbers in the coming years [2], the ques¬ 
tion arose at Schaeffler as to whether the 
electric axle drive could also be used for 
these vehicles. The objective of using a 
48-volt hybrid must be considered: A sig¬ 
nificant C0 2 reduction must be achieved at 
acceptable costs. The key to achieving this 


















216 


Front-Wheel-Drive 


Rear-Wheel-Drive 


eAWD 


All-Wheel-Drive 





Traction 

Support 



Figure 3 Basic topologies of a 48-volt hybrid powertrain 


objective is not only the battery, which is still 
the largest cost block, but also the lower 
overall safety requirements for drive sys¬ 
tems with a peak voltage of less than 60 V. 
A low-voltage system is the subject of sig¬ 
nificantly lower requirements in all steps of 
the value added chain, from assembly 
through to maintenance. 

The maximum C0 2 saving is also de¬ 
pendent on the electric power of a 48-volt 
hybrid system. The decisive factor is not 
only the acceleration to be achieved by 
the vehicle, but above all the maximum 
braking energy to be recuperated. The 
maximum power achievable with current 
technology is approximately 12 kW. This 
electric power not only enables recupera¬ 
tion in generator mode, but also a dis¬ 
placement of the operating point of the 
internal combustion engine in the data 
map and electric driving in a low speed 
range, for example, during maneuvering 
or in traffic jams. 

The integration of a corresponding 
low-voltage electric axle into the pow¬ 
ertrain can be carried out in different con¬ 


figurations (Figure 3). The driven axle can 
be provided with motor assistance in both 
front-wheel and rear-wheel drive vehicles. 
An electric rear axle drive can also be im¬ 
plemented in a front-wheel drive vehicle, a 
configuration, which is occasionally de¬ 
scribed as an “electric all-wheel drive”. 
Lastly, the electric drive force can also be 
distributed between the front and rear 
axle, although this means that two electric 
motors and two power electronics units 
are required. 

With regard to the following consider¬ 
ations, Schaeffler assumes that vehicles 
with an electric axle based on 48-volt sys¬ 
tem will always have a belt-driven starter 
generator with a nominal voltage of 12 or 
48 V because it is not possible to start the 
internal combustion engine with the elec¬ 
tric axle motor. In addition, the starter 
generator is already part of the modular 
system from the vehicle manufacturer’s 
point of view. This has the advantage that 
safety-critical functions such as electro¬ 
mechanical torque vectoring are always 
available irrespective of the battery’s state 





















































































48-volt Electric Axle 


14 


217 


of charge because the battery can be re¬ 
charged by the starter generator at any 
time. The use of the electric axle as a 
“electric four-wheel drive” is also depen¬ 
dent on the state of charge. Four-wheel 
drive functionality is available without limi¬ 
tations when moving off. 


Design of the 48-volt axle drive 

After taking the fundamental decision to 
derive an electric rear axle from the high- 
voltage system based on a 48-volt system, 
Schaeffler began development of a rele¬ 
vant system, which would fit in an actual 
current volume-produced vehicle with 
rear-wheel drive. It was apparent that the 
electric motor could be fitted around the 
propshaft without having any affect on the 
space requirement (Figure 4). An existing 
asynchronous motor from a belt-driven 
starter generator is used. The 48-volt axle 
drive is also equipped with a two-speed 
transmission as in the latest generation of 
the high-voltage variant. 

The drive was designed so that it can 
be offered as an additional variant with a 
single-speed or two-speed transmission in 
a volume-produced vehicle without neces¬ 



Figure 4 Prototype of an electric rear axle 
drive for a rear-wheel drive vehicle 


sitating changes to the vehicle body or 
chassis. This required a very compact de¬ 
sign. In this particular case, the diameter of 
180 mm was less than the diameter of the 
propshaft tunnel. The entire drive unit is lo¬ 
cated coaxially relative to the propshaft di¬ 
rectly in front of the axle drive (Figure 5). 
Water cooling was also not required. 

In first gear, the force flows from the 
electric motor via the sun wheel of the first 
planetary gear set (Figure 6). The planet car¬ 
rier is connected with the sun wheel of the 
second gear set and the force is transmitted 
to an intermediate shaft via the planet car¬ 
rier. The force flows between the planet car¬ 
rier and the intermediate shaft via a switch- 
able selector sleeve. In second gear, 
however, the planet 
carrier of the first 
gear set is connect¬ 
ed with the intermedi¬ 
ate shaft so that 
only the ratio of this 
planet carrier is effec¬ 
tive. The second 
gear set rotates free 
of load. The trans¬ 
mission is station¬ 
ary when disen¬ 
gaged and the 
vehicle behaves like 
a conventional ve¬ 
hicle. 



Figure 5 Section through the electric rear axle drive integrated around 
the propshaft 










































218 



Figure 6 Flow of force in the electric rear axle drive 


Avery high transmission ratio has been se¬ 
lected for first gear in order to achieve a 
sufficient starting torque of at least 1,000 Nm 
despite the relatively small electric motor. 
In the prototype, a ratio i of 19.6 was se¬ 
lected for first gear taking into account the 
ratio of the hypoid stage of the rear differ¬ 
ential. In second gear the ratio i is 4.4. The 
relatively high ratio steps were selected be¬ 
cause the asynchronous motor reaches its 
maximum speed in first gear at slightly 
above 20 km/h. 

Functions of a 48-volt 
mini hybrid 


C0 2 optimization 

Without a doubt, the reduction of C0 2 
emissions is the primary motivation for in¬ 
troducing a mini hybrid drive system. The 
decisive reduction for homologation 
should also be reflected in the lowest pos¬ 
sible actual fuel consumption for end cus¬ 
tomers. Schaeffler has therefore carried 
out simulations of several driving cycles 
for a vehicle with an electric rear differen¬ 
tial (Figure 7). The simulations were based 
on a very heavy luxury class vehicle with a 


weight of more than 2 tons and a V-8 gas¬ 
oline engine. 

The simulations show that consump¬ 
tion can be reduced by up to 9 % in the 
NEDC, compared with a vehicle equipped 
with a start-stop function. In the ARTEMIS 
cycle, which is aimed at simulating the 
actual fuel consumption of a theoretical 
average customer, there is a reduction in 
consumption of around 6 %. These simu¬ 
lations were created using models, which 
take the overall efficiency chain into con¬ 
sideration. For example, the actual reduc¬ 
tion in power of the electric motor with 
increasing temperatures was also consid¬ 
ered. 

It was however assumed in the simula¬ 
tions shown that the internal combustion 
engine was switched off during sailing. 
This is not always the case in all foresee¬ 
able applications for the future so that the 
fuel consumption of the internal combus¬ 
tion engine during idling must also be add¬ 
ed if required. 


Electric driving functions 

Schaeffler’s electric axle differential has 
sufficient torque to enable driving using 
electric power only in a low-speed range 
of 0 to 20 km/h. We prefer to use the term 
“moving off using only electric power” as 
a synonym to ensure that any reference to 














































48-volt Electric Axle 


14 


219 


+ Start-Stop 


+ Smart Alternator 


Total C0 2 Reduction 12 V 
up to ca. 8 %** 



+ eSailing (ICE off) 



Reference: E-Segment V8 gasoline engine, benefits in driving cycles, without start-stop 


* Drivability of eSailing (with ICE=off) not yet considered 
: Depends on strategy & vehicle 


Figure 7 Potential reduction in C0 2 emissions of a mini hybrid drive system in different driving cycles 


“electric driving” does not lead to unreal¬ 
istic customer expectations. There are 
major advantages in terms of comfort for 
the customer, particularly in stop-and-go 
traffic and during maneuvering. Control of 
longitudinal dynamics can be carried with 
the brake pedal alone, as in a vehicle with 
an automatic transmission - and with the 
internal combustion engine switched off. 
The torque that can be currently achieved 
on the axle is sufficient to accelerate a ve¬ 
hicle from a standstill on a gradient of up 
to 10 %. The potential for reducing C0 2 by 
moving off solely under electric power is 
less than 3 % in the premium segment 
sedan considered above. The possible 
range of electric operation is also limited 
to several hundred meters or just a few ki¬ 


lometers depending on the size of the cur¬ 
rently available low-voltage batteries. The 
described advantages in terms of comfort 
and the experience of electric driving, in 
combination with the minimal additional 
complexity for end customers, are a thor¬ 
oughly convincing argument for deciding 
to buy a hybrid vehicle. 


Active torque distribution 

If the unit is fitted coaxially relative to the ve¬ 
hicle’s axle, the electric differential in 48-volt 
design can also be used in order to operate 
active torque distribution in a transverse di¬ 
rection (so-called torque vectoring). This 
form of variable drive torque distribution be- 















220 


Electric Driving 



Torque Vectoring 

8 48-volt axle drive with an electric motor and a two-speed transmission 


tween the wheels has two basic advantag¬ 
es: 

- Increased traction if the friction coef¬ 
ficients of both wheels are unequal, for 
example, when driving on snow-cov¬ 
ered or icy roads. 

- Improved lateral dynamics due to tar¬ 
geted adjustment of the torque, which 
counteracts understeer or oversteer of 
the vehicle during cornering. 

Active torque distribution is increasingly 
regarded as a comfort function. For ex¬ 
ample, it would be possible to completely 
compensate for the influence of strong 
side winds on the direction of travel in an 
energy efficient manner by using torque 
vectoring. The input variable for such 
functions is the yawing moment about the 
vertical axis of the vehicle, which is al¬ 
ready continuously recorded by the ESP 
sensors. The introduction of such func¬ 
tions is the subject of detailed discussions 
about the personal responsibility of the 
driver. 


The design of a 48-volt mini hybrid with 
an electric rear axle is based on the idea 
of torque distribution so that a single 
electric motor can be used - in contrast 
to the high-voltage module shown in Fig¬ 
ure 1. In addition, the architecture of the 
two-speed transmission should be used 
for both the drive and torque distribution 
(Figure 8). The two-speed transmission 
with a torque vectoring function can be 
combined with a planetary differential 
but also with a standard bevel gear dif¬ 
ferential. 

Shifting between the three planetary 
gear sets is carried out sequentially with a 
single actuator, which reduces the com¬ 
plexity and costs of the gearshift system. 
This type of actuation concept with one ac¬ 
tuator offers additional advantages with re¬ 
gard to functional safety because the risk of 
faulty gearshift operation (double gearshift 
operations) can be reduced. The ratios are 
designed so that the vehicle can be driven 
at approximately 20 km/h using electric 

















































48-volt Electric Axle 


14 


221 


power only. Subsequently, the system shifts 
from first to second gear. Boosting, recu¬ 
peration and load point shifting of the inter¬ 
nal combustion engine are possible within a 
speed range of approximately 20 to 80 km/h. 
Planet gears 1 and 3 are used for the traction 
mode. 

Active torque distribution is possible 
from second gear after passing through 
another neutral position. The force now 
also flows via the center planet gear, which 
is connected with both the differential cage 
and the side shafts. The side shafts are 
“rotated” in relation to each other due to 
the torque applied by the electric motor, 
resulting in a difference in speed. Torques 
of up to 1,200 Nm (peak) and 800 Nm 
(continuous torque) can be achieved with 
this type of system, which is comparable 
with the hydraulic systems already estab¬ 
lished on the market. It must be empha¬ 
sized that the torque vectoring position is 
independent of the actual vehicle speed, 
i.e. it can also be selected when the vehicle 
is stationary. 

Torque vectoring or electric drive can 
be selected automatically by means of 
suitable sensors and prioritization depend¬ 
ing on the vehicle speed and other input 
variables. An additional option is the tar¬ 
geted activation of functions by the driver 
using a “sport button”, “economy button” 
or “city mode button”. 

This has the following advantages for 
the electric axle based on a 48-volt system 
with integrated electromechanical torque 
vectoring: 

- Moving off using electric power only 
and active torque distribution are pos¬ 
sible in contrast to a standard rear dif¬ 
ferential. 

- A significant reduction in fuel consump¬ 
tion is possible compared to a hydrau¬ 
lic system for active torque vectoring. 
An electromechanical system has max¬ 
imum actuating speeds of 60 ms, virtu¬ 
ally independent of the temperature. 


- The 48-volt system is significantly less 
complex and therefore more cost ef¬ 
fective compared with the high-voltage 
system according to Figure 1, which is 
the “non plus ultra” in technical terms. 
With its recently presented system, Schaef¬ 
fler is pursuing a strategy of maximizing the 
integration of functions by means of innova¬ 
tive drive technology and minimum product 
complexity. Schaeffler has succeeded in 
integrating three functions into the rear dif¬ 
ferential using an electric motor, an actuator, 
and transmission architecture: Moving off 
using only electric power, a significant po¬ 
tential for reducing C0 2 in hybrid mode and 
an increase in vehicle agility and comfort by 
means of torque vectoring. 

This type of “three-in-one” modular 
concept combines the demands for effi¬ 
cient mobility with the maximum require¬ 
ments for vehicle dynamics and emotion¬ 
ality of future vehicles and acceptable 
purchase costs. The resulting added val¬ 
ue for end customers can be a decision¬ 
making criterion for the acceptance of 
low-voltage hybridization and accelerate 
the hybridization of vehicle drives world¬ 
wide. 

Schaeffler is currently equipping a 
sporty coupe in the compact vehicle class 
with an electric axle and integrated torque 
vectoring based on a 48-volt system in or¬ 
der to test these advantages, which are di¬ 
rectly noticed by end customers. 


Outlook 


The C0 2 reductions that can be achieved 
with a mini hybrid drive are of course signifi¬ 
cantly less than the values, which can be 
achieved with a high-voltage electric drive. 
However, the ratio of costs and benefits ac¬ 
cording to the first simulations is so positive 



222 


that Schaeffler is continuing intensive fur¬ 
ther development. The potential identified in 
the simulations will be checked by design¬ 
ing a demonstration vehicle and carrying 
out practical tests. 

There is a strong correlation between 
the C0 2 reduction and the electric power of 
the system as described above. This is clear 
if the speeds driven in the NEDC are plotted 
over the corresponding axle torque and 
compared with the data map of the electric 
motor (Figure 9). 

Consequently, a significantly higher pro¬ 
portion of operating points could be cov¬ 
ered with a performance-enhanced electric 
motor of 12 to 18 kW. This also applies for 
the braking performance and thus the 
quantity of recuperated energy. Schaeffler 
is therefore also working on the further de¬ 
velopment of an electric drive with higher 
power in addition to a prototype equipped 
with a 12-kW motor. 

An increase in the available continu¬ 
ous output would also be possible by 


changing the method of cooling used in 
the electric motors in the prototypes from 
cooling via the air gap to oil cooling and 
this is therefore also part of further devel¬ 
opment work. 

Schaeffler can also envisage that radi¬ 
cal optimization of the rolling resistance of 
the tire in combination with active electro¬ 
mechanical torque distribution will become 
a further field of research. This work is 
based on the idea of compensating the re¬ 
duced cornering forces of particularly nar¬ 
row tires with a low rolling resistance by 
means of torque vectoring. Initial estimates 
indicate a potential reduction in rolling re¬ 
sistance of up to 30 % - without any risk to 
the active safety. The implementation of 
this idea still raises many questions. For 
example: How can it be ensured that these 
types of tires are only fitted on vehicles 
with active torque distribution? Is perma¬ 
nent roll stabilization of a vehicle by means 
of the active intervention of an electrome¬ 
chanical system permitted? 



Figure 9 Operating points in the NEDC and torque output of the electric motor 


















48-volt Electric Axle 


14 


223 


At Schaeffler, we regard innovation as a 
continuous search for new concepts and 
we pursue radical ideas, whose one hun¬ 
dred percent feasibility must still be proven 
as part of research and development 
projects. We are always open to further 
inspiration and ideas from our customers, 
suppliers, and development partners! 


Literature 


[1] Smetana, T.; et al.: Schaeffler active eDifferen- 
tial: The active differential for future drive trains. 
9 th Schaeffler Symposium, 2010 

[2] Gutzmer, P.: Individuality, efficiency and com¬ 
fort: Paradigms for future mobility. 

10 th Schaeffler Symposium, 2014 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



224 


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225 


Double Clutch Systems 

Modular and highly efficient 
for the powertrain of tomorrow 


Matthias Zink 
Uwe Wagner 
Clement Feltz 


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226 


Introduction 


Alongside the established stepped auto¬ 
matic transmission and CVT, the double 
clutch transmission in particular has 
achieved considerable market penetration 
in the last few years (Figure 1). 

Significant growth has been seen in the 
European and Chinese markets in particular, 
and current forecasts indicate that in ten 
years’ time, every fifth automatic transmis¬ 
sion will be a double clutch transmission. 

The following basic requirements apply 
to automatic transmissions in accordance 
with current definitions: 

- Maximum comfort achieved through 
powershift capabilities combined with 
a dynamic driving experience 

- Ideal spreading and the highest poss¬ 
ible level of efficiency across all op- 
er-ation modes 

- Actuating mechanism operated with 
minimal losses and, where possible, 


without the need for additional effort while 
the combustion engine is turned off 
- Hybrid function presented in the simplest 
and most flexible manner possible 
A double clutch transmission, which fulfils 
all of these requirements, including an inte¬ 
grated hybrid function, entered into series 
production in Japan at the end of 2013 under 
the name “i-DCD”. Figure 2 shows the dry 
double clutch transmission and the wet ver¬ 
sion for the application “SH-AWD”, as well 
as the parts supplied by Schaeffler, which 
are both modular and highly efficient! 

In addition to using highly efficient, hydro¬ 
statically operated double clutches (equipped 
with a concentric slave cylinder, or “CSC”), a 
newly developed actuator with an integrated 
control unit is also used in these applications 
(a hydrostatic clutch actuator, or “HCA”), and 
a gear actuator featuring an “active interlock” 
concept. Thanks to the “power-on-demand” 
actuator elements present in this design, it is 
possible to reduce the NEDC power con¬ 
sumption levels for transmission and clutch 
operation to values lower than 20 W. 



Figure 1 Production volumes for different automatic transmissions (selected regions) 

Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 

DOI 10.1007/978-3-658-06430-3_15, © The Author(s) 2014 
















Double Clutch Systems 


15 


227 


180 Nm “dry” 


370 Nm “wet” 



■ Parts supplied by Schaeffler 


Dry double clutch transmission “i-DCD”, wet double clutch transmission for “SH-AWD” 


Figure 2 

The double clutch: 
“dry or wet” or 
“dry and wet”? 


As was previously the case, comfort when 
driving off and shifting gear remains a key fo¬ 
cus for the double clutch system — with re¬ 
gard to driving off in particular, there exists a 
strong, established and extremely vigilant 
competitor when it comes to the NVH, com¬ 
fort and dynamics of the torque converter. In 
this respect, the new, highly innovative de¬ 
sign of the iTC torque converter [1] in particu¬ 
lar will see double clutch systems face a new 
challenge across all disciplines. 

Back in 2006, a chapter was dedicated to 
the question of “dry or wet?” as part of the LuK 
Symposium, and following on from a large 
number of series production projects (Figure 3) 
that have since been carried out, it has become 
apparent that the clutch technology used for 
the corresponding applications is determined 
on the basis of the technical rationale for the 
specific clutch load described at the time. 


In terms of efficiency, the dry double clutch 
continues to be the first choice, wherever 
the torque capacity permits this. Combined 
with electromechanical actuating mecha¬ 
nisms, the dry double clutch represents a 
system solution that sets a benchmark for 
efficiency. In addition, this version of the 
double clutch places few requirements on 
the peripheral equipment of the transmis¬ 
sion, as no additional oil cooling is required 
for the clutch. As a result, implementation is 
achieved with greater ease and with a ten¬ 
dency towards increased cost-effectiveness. 

However, this clutch variant also poses 
specific challenges. Cooling is achieved by 
means of air convection, meaning frictional 
heat needs to be stored temporarily in the 
pressure plates, which increases weight 
and inertia levels. Furthermore, the ability to 
control dry clutches tends to be more criti¬ 
cal than for wet systems. Clutches with 
wear adjustment mechanisms in particular 
are more difficult to control due to the more 
complex mechanics of the internal struc¬ 
ture, but also due to the greater variance of 
the friction coefficients of the dry linings in 
general. A dry friction system must be fully 




228 



2006 2007 2008 2009 2010 2011 2012 2013 2014 2015 

Year 


DCT dry 
DCT wet 


Source: IHS + Schaeffler Knowledae/Dec. 2013 


Figure 3 Production volumes for different double clutch transmissions 


functional throughout the entire life of the 
clutch; oil changes - and therefore the addi¬ 
tion of fresh additives, such as for a wet 
clutch - is not necessary or possible for a 
dry clutch. 

The primary development objectives are 
to reduce inertia levels and, in particular, to 
enhance controllability and thus optimize 
the comfort characteristics of the dry dou¬ 
ble clutch when driving off and shifting gear. 
Reducing internal friction and compensat¬ 
ing the geometric torsional vibration exci¬ 
tation using a new design featuring direct 
actuation (DCC), the optimization of the ac¬ 
curacy of the individual parts, new friction 
linings with significantly improved damping 
and actuators with special control algo¬ 
rithms to control juddering (anti-judder con¬ 
trol system) will also considerably improve 
the comfort characteristics. These mea¬ 
sures are explained in detail in [2]. 

For applications with higher specific 
clutch loads, wet double clutch systems are 
generally used, as this oil-cooled version 
has the advantage of a higher cooling ca¬ 
pacity in comparison with the dry version. 


On current applications, the transition from 
dry to wet occurs at driving torques of be¬ 
tween 250 and 350 Nm. In addition, the wet 
double clutch is also smaller and lighter in 
terms of its transmission capacity. To date, 
this version has also featured a more simple 
mechanical arrangement, as it does not re¬ 
quire a fixture for wear adjustment. 

However, today’s wet clutch systems 
cannot fully utilise the benefits of reduced 
weight and inertia, as they require additional 
masses in the damper system to achieve 
the necessary level of torsional vibration 
isolation. In addition, the oil cooling system, 
which has a positive effect on performance, 
together with the peripheral equipment re¬ 
quired for oil cooling, also represents a con¬ 
siderable additional effort with respect to 
design and energy, which has a negative 
effect on weight, cost and efficiency. 

At the very least, optimised wet double 
clutch systems should therefore be capable 
of utilising the lower inertia levels owing to 
an appropriately powerful damping system. 
Furthermore, attention should be paid 
to improving efficiency by reducing drag 





Double Clutch Systems 


15 


229 


losses and using a highly efficient actuating 
mechanism. 

In comparison with dry double clutches, 
future wet clutch systems should also in¬ 
clude design elements to reduce geometric 
excitation, as well as optimised linings and 
dynamic actuator systems with the option of 
an “anti-judder control system” [3] in re¬ 
sponse to the increased challenges facing 
modern powertrains in relation to NVH. 


The actuating mechanism 


From a functional viewpoint, the key re¬ 
quirement placed on the actuating mech¬ 
anism of double clutch transmissions is 
most certainly the need to combine ade¬ 
quate dynamics and performance with the 
highest levels of efficiency. Actuators there¬ 
fore not only require a minimum amount of 
energy to operate the clutch and transmis¬ 
sion, but should, for example, also be capa¬ 
ble of supporting the aforementioned an¬ 
ti-judder control system. 

A modern actuator system should there¬ 
fore only use power when required (“power 
on demand”). Furthermore, it must also be 
possible to operate the actuator system 
when the combustion engine is not running, 
in order to support start-stop and hybrid 
functions. For these operating states, spe¬ 
cial attention must be paid to noise ge¬ 
neration, as the masking noises of the 
combustion engine are not present in this 
instance. In terms of design, the actuating 
mechanism should take up as little space as 
possible, and a modular design may be 
beneficial in order to reduce the costs for 
different applications. 

Forward-thinking ideas for expanding 
the functions of such modular actuator sys¬ 
tems, while reducing their level of complexi¬ 
ty at the same time, are detailed in [4]. 


Outlook/transition 


Double clutch transmissions have all the 
prerequisites and real potential for becom¬ 
ing the basic architecture for the powertrain 
of tomorrow. The i-DCD transmission, pro¬ 
duced in series since 2013, is a trend¬ 
setting example of a modular, highly effi¬ 
cient and even integrated hybridised double 
clutch transmission. 

The consistent use and inclusion of 
hybrid elements in the design - for example, 
for driving off using electric power - will fur¬ 
ther strengthen the position of the double 
clutch transmission in relation to alternative 
transmission concepts from competitors. 


Literature 


[1] Lindemann, P.; Steinberger, M.; Krause, T., 
iTC — innovative torque converter solutions 
paving the way to the future. 10 th Schaeffler 
Symposium, 2014 

[2] Kimmig, K.-L.: The highest levels of comfort: 
The dry double clutch rises to the challenge. 
10 th Schaeffler Symposium, 2014 

[3] Englisch, A.; Goetz, A.; Baumgartner, A.; Endler, T.; 
Lauinger, C.; Steinmetz, S.: The wet double 
clutch: Thinking in systems. 

10 th Schaeffler Symposium, 2014 

[4] Mueller, B.; Rathke, G.; Grethel, M.; Man, L.: 
Gearshift actuation: Less complexity, more 
functionality. 10 th Schaeffler Symposium, 2014 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 





230 


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231 


Highest Level of Comfort 


The dry double clutch 
faces the challenge 


Karl-Ludwig Kimmig 
Dr. Peter Buehrle 
Dr. Ralph Kolling 
Rene Daikeler 
Michael Baumann 


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232 


Introduction 


The competitive environment facing the 
dry dual clutch has already been intro¬ 
duced in [1]. A large number of series-pro¬ 
duction applications highlights the fact that 
dry double clutches have proven success¬ 
ful in the market (Figure 1). 

The dry double clutch system also 
provides an ideal alternative for future, 
automated powertrains used in compact 
and mid-sized vehicles on account of its 
very high level of overall efficiency and 
the fact that oil cooling is no longer re¬ 
quired for the clutch system. One chal¬ 
lenge is that the NVH and comfort de¬ 
mands placed on the powertrains will 


continue to increase and the mass mo¬ 
ment of inertia of the double clutch sys¬ 
tem should be kept to a minimum as re¬ 
gards the driving dynamics. In order to 
face this challenge, further development 
of dry double clutch systems and the 
associated vibration damping concepts 
is required. New friction linings tailored 
specifically to requirements and ad¬ 
vanced software functions contribute to 
a large optimisation step for the overall 
system. For instance, by operating the 
clutches via appropriate software control 
strategies, vibrations on the powertrain 
can be eliminated (anti-judder control 
system). Additional potentials for improv¬ 
ing comfort are also brought about by 
electrical launches when used with hybri¬ 
dised powertrains. 



Figure 1 Dry double clutch applications in series production 

Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_16, © The Author(s) 2014 








Dry Double Clutch 


16 


233 


Dry double clutch systems 
with minimal vibration 
excitation 


The NVH behaviour of modern powertrains 
depends on a range of system-specific fac¬ 
tors, such as damping, transmission behav¬ 
iour of the powertrain, vibration couplings 
and, of course, the excitation between the 
engine, clutch and transmission. The fac¬ 
tors especially relevant from the point of 
view of the clutch are isolating the engine's 
torsional vibrations via the dual mass fly¬ 
wheel and additional damping measures in 
the clutch, including a controlled isolation 
system and disturbance excitation of the 
slipping clutch (known as judder excitation). 
Disturbance excitation of a slipping clutch, 
in particular, is very much the focus when it 
comes to double clutch systems. The 
causes of why disturbance excitation poses 
a significantly greater challenge in double 
clutch systems than in manual vehicles, are 
illustrated in Figure 2. 



Time in s 

— Motor 

— Transmission 


One key difference in the launch and shift¬ 
ing processes with automated systems 
compared to manual driving behaviour is 
that the state in which the clutch is oper¬ 
ated with slippage is maintained for sig¬ 
nificantly longer (creep function) in order 
to enhance driving comfort. This is be¬ 
cause, especially in the lower gears 1 st , 2 nd 
and reverse, the natural powertrain fre¬ 
quency runs virtually always non-stop 
during slip and the smallest torsional vi¬ 
brations at the clutch output can result in 
noticeable vibrations or associated noises 
on the vehicle. These vibrations and nois¬ 
es are amplified by the fact that during the 
launches of double clutch transmissions, 
only one sub-transmission is pre-loaded 
by the torque and the inactive sub-trans- 
mission can vibrate freely. This action 
causes additional noises to occur. In the 
launch simulation (Figure 2), identical 
clutch parameters (including geometry, 
friction lining properties and starting 
torque) were used as the basis, and iden¬ 
tical powertrain damping was also used 
as the starting point. It is evident that by 
extending launch the judder vibrations 



Figure 2 Comparison of speed curves when launching with a manual transmission and a double 
clutch transmission with creep function (simulation) 








234 


vibrate against each other much more vio¬ 
lently and for significantly longer. There¬ 
fore, in terms of subjective feeling, auto¬ 
mated, scattered launch feels significantly 
worse. 

Transmissions with a torque converter 
represent the benchmark for comfortable 
and low-vibration launches with creep 
function. To ensure that this level of com¬ 
fort can also be reached by double clutch 
systems, clutch disturbance excitation 
must be significantly reduced. This is par¬ 
ticularly the case because modern pow¬ 
ertrains have as low-friction designs as 
possible, for efficiency reasons, and 
therefore do not feature vibration damp¬ 
ing. The sources of disturbance excita¬ 
tion on clutches are already known; 
however, these sources have become in¬ 
creasingly important due to the facts out¬ 
lined above [8, 9]. 

The proven analogous clutch model 
for the belt sander in an enlarged form 
can be used to explain the physical prin¬ 
ciples and to demonstrate the optimisa¬ 
tion potential (Figure 3). 

Coupling a vibrating powertrain via a 
friction system in slippage can produce 
additional excitations or damping, de¬ 
pending on the friction characteristics as 
a function of the slip speed [4, 8]. Mod¬ 



ern, dry friction linings tend only to excite 
frictional vibrations to a small extent. In 
the majority of operating states, the fric¬ 
tion system supports the powertrain 
damping characteristics during the slip¬ 
page phases by means of a positive fric¬ 
tional coefficient gradient. However, it 
was demonstrated for the first double 
clutch applications that new and previ¬ 
ously unknown causes of damage can 
occur as a result of specific driving con¬ 
ditions for automated clutch systems. As 
a result, the tribological system is 
changed by the formation of special sur¬ 
face layers that decrease the damping 
characteristics. It is normal practice on 
dry clutches that wear on the clutch con¬ 
stantly renews the surface of the friction 
system, so that there is no drop in damp¬ 
ing over the life in real driving conditions. 
However, this renewal process can be 
slowed down by particularly light-duty 
loads. Therefore, dry double clutches 
definitely benefit from occasional higher 
thermal loads. Based on these findings, 
it is possible to achieve further increases 
in damping characteristics and therefore 
greater comfort benefits with new friction 
linings and friction mating surfaces tai¬ 
lored specifically to the loads of double 
clutch systems. 



Figure 3 Basic model of “slipping clutch system”; left-hand image of tribological system without 
superimposed geometry errors, right-hand image with geometry errors 

















Dry Double Clutch 


16 


235 


In addition to these friction-induced exci¬ 
tations, there is a second source for dis¬ 
turbance excitation in the slipping clutch. 
This is the result of geometric errors, 
which must always be present in a mini¬ 
mum of pairs and with interaction [8, 9]. 
This is illustrated in the analogous clutch 
model (right, Figure 3). Both surfaces in 
the friction contact exhibit warping that, in 
the case of relative movement, can pro¬ 
duce contact force modulation with the 
rigidity of the clutch (cushion deflection) 
and the rigidity of the operating system. 
In order to minimise the geometric errors 
of double clutch components and sub¬ 
systems, a host of ideas for solutions have 
already been developed, some of which 
are already being implemented or may be 
implemented over the coming months in 
high-volume production; these ideas in¬ 
clude pairs of components for reducing 
geometric deviations [9]. 

However, development work has not 
just been limited to reducing the geomet¬ 
ric errors. The focus has also been on 
coming up with solutions with a stable 
and lasting impact on compensating the 
effects of geometric deviations on com¬ 
ponents, such as by means of a cardanic 
support. The current line of thinking is that 
this compensation is ideally achieved with 
directly actuated clutches, carried out 
by means of a concentric slave cylinder 
(CSC). The first dry double clutch system 
of this design went into series production 
in mid-2013. This double clutch and the 
new, derivative clutch series is explained 
in greater detail in the second part of this 
article. 


Damping powertrain 
vibration using the 
anti-judder control system 


In addition to the causes of vibration exci¬ 
tation, the analogous clutch model (Figure 3) 
can also be used to outline the idea of 
damping vibrations by using an anti-jud¬ 
der control system. In essence, the idea is 
that inversely phased, active contact force 
modulation is used, which is initiated and 
monitored via a software control circuit. 
The result is that additional damping of 
the powertrain is indirectly achieved, with¬ 
out the disadvantage of increasing con¬ 
sumption. The challenges posed by this 
system are processing of the available 
speed signals and having as accurate as 
possible a picture of the overall system 
characteristics, as determined by the ve¬ 
hicle, powertrain, clutch and its actuating 
mechanism. Today, an anti-judder control 
system can be achieved for 1 st gear and 
reverse with an improvement of 1-2 ATZ 
scores. Initial vehicles featuring this kind 
of software solution have been in series 
production since the beginning of 2013. 
The excellent effect that an anti-judder 
control system has in the vehicle during 
creep launch is shown in Figure 4. The 
judder vibrations have almost been com¬ 
pletely eliminated. 

Additional potentials can be tapped 
into in conjunction with hybrid pow¬ 
ertrains. Therefore, when combined with 
electric motors, launch can be performed 
by completely electrical means. Pro¬ 
longed clutch slippage (Figure 2, right) 
is therefore largely avoided. In addition, 
small inversely phased torsional vibrations 
can be generated in the powertrain by 
regulating the speed of the electric motor; 
this also enables judder vibrations to be 
completely eliminated. 



236 



o 

C N 

S = 

CJ c 

<D ■” 




— Motor 

— Transmission 




Figure 4 The effect of the anti-judder control system in the vehicle: left without, right with anti-judder 
control system 


Optimising the tribological 
system for dry double 
clutch applications 


In order to prevent powertrain vibrations, 
clutch linings combined with cast or steel 
mating friction surfaces should only dis¬ 
play damping-supporting properties over 
a large application range during the slip¬ 
page phases. This requires a slightly in¬ 
creasing frictional coefficient over the 
slip speed. As the loads on the clutch 
friction system differ between manual 
clutches and double clutches, new fric¬ 
tion linings had to be developed for opti¬ 


misation purposes and a full range of 
testing had to be performed. Extensive 
systematic tests have shown that the in¬ 
organic filler and friction material in the 
lining compound in particular are respon¬ 
sible for changes to the lining damping 
during usage in the double clutch-specif¬ 
ic slippage phases. The mode of action 
in the friction contact can be described 
using a two-phase model. 

Stage 1 - enrichment of inorganic 
substances in the friction layer: 

In many similar clutch slipping phases with 
low friction energy, but with average, spe¬ 
cific frictional power, organic components 


Position acceleration in m/s 2 





























Dry Double Clutch 


16 


237 


of the lining compound are partially, ther¬ 
mally broken down on the lining surface. 
The associated lining wear is not high 
enough to renew the friction surface suffi¬ 
ciently. As a result, an increasing numbers 
of inorganic components build up in the fric¬ 
tion layer close to the surface. 

Phase 2 - enrichment of casting wear 
particles in the friction layer: 

The increased proportion of inorganic 
components in the friction layer leads to 
increased wear of the contact material, 
comprised of cast iron or steel. As the 
surface of the friction lining is not renewed 
due to the comparatively low thermal 
stress, the metallic wear particles are en¬ 
riched in the friction layer. The result of the 
layer being enriched leads to a negative 
change in the frictional coefficient gradi¬ 
ent. 

New fillers and fiber combinations 
have been developed for linings for opti¬ 
misation purposes. The positive effect of 
these aspects has since been proven in a 


variety of component and system dura¬ 
bility tests. Today, Schaeffler can recom¬ 
mend B 8040 and RCFlo as two friction 
materials ideally suited for double clutch 
applications. In terms of taking a final de¬ 
cision on the respective friction lining in a 
specific vehicle application, it is not only 
the frictional coefficient gradient that is 
important; other parameters such as the 
wear behaviour and absolute frictional 
coefficient value are also decisive in the 
various operating states. Development 
work carried out over the past few 
months has shown that it is highly prob¬ 
able that further improvements with re¬ 
gard to the lining damping characteris¬ 
tics are possible with advanced friction 
linings. 

In addition, the way in which the contact 
friction mating surface is designed also en¬ 
ables a slowdown and reduction in cause of 
the damage. 

Possible measures include specific 
surface roughness and also special radial 
grooves on the mating friction surface. 



Figure 5 Damping characteristics of the tribological system for double clutch applications 




238 


Reducing geometric 
torque excitation 


The first generation of dry double clutch¬ 
es was designed featuring the extremely 
compact three-plate arrangement. In 
these designs, the double clutch is 
mounted on the transmission’s hollow in¬ 
put shaft by a support bearing located in 
its central casting plate. The actuating 
lever springs and the components of the 
wear adjustment device are arranged on 
a common clutch cover on the side fac¬ 
ing the transmission. This arrangement is 
extremely compact; however the sheet 
steel and casted individual parts used in 
this arrangement must meet stringent 
requirements for flatness and parallelism. 
They must meet these on account of 
the considerable disturbance excitation 
stresses. A wealth of experience in tool 
design and also in adjusting production 
dies is required in order to achieve this. 
However, when this arrangement fea¬ 
tures several nested sheet metal parts, it 
also offers the option of pairing these 
parts in series production so that a mini¬ 
mum of parallelism errors occurs [9]. 

Although these optimisation mea¬ 
sures have already proved extremely 
successful (in some cases leading to 
geometric disturbance excitations being 
reduced by half), further improvements 
can be achieved by using cardanic actu¬ 
ation as well as by compensating any fi¬ 
nal geometric inaccuracies. And it was 
for this reason that the new, directly ac¬ 
tuated double clutch system with hydro¬ 
static control by means of CSC was de¬ 
veloped. 


The new, directly actuated 
double clutch system with 
hydrostatic control 


There are essentially three main ways of re¬ 
ducing geometric disturbance excitation. 

1. Minimising geometric errors. 

2. Linear clutch mapping characteristics 
(torque over engagement travel) with 
low gradient. Where geometric errors 
are present, this results in low contact 
force or torque modulation. 

3. Reduction in tilting rigidity of the contact 
force transfer elements, with the aim of 
using the cardanic compensating effect 
in the clutch system. 

This last point in particular results in a new 
double clutch arrangement (Figure 6), the 
directly actuated double clutch system 
with hydrostatic control by means of 
CSC. With this system, geometric devia¬ 
tions in the friction contact can occur as 
a result of a virtually hysteresis-free car¬ 
danic angular alignment of the pressure 
plate via the load transfer plate up to the 
CSC piston, without unequal contact 
forces being created. This compensating 
function is only possible if the system has 
a low tilting rigidity, i.e. it behaves in a 
cardanic manner. As a result, when geo¬ 
metric errors are present, only minimal 
torsional vibrations occur in the clutch. 
Overall, this will result in a simpler overall 
structure with clutch control using CSC, 
which also can be used in wet double 
clutch systems [3, 10]. 



Dry Double Clutch 


16 


239 




Figure 6 Structure of the directly actuated double clutch with double-CSC actuation with cardanic 
tilt compensation via the CSC piston 


Evolution of new directly 
actuated dry double 
clutches 


This directly actuated dry double clutch first 
went into series production in mid-2013. For 
this system, the four pressure plates re¬ 
quired for this concept were again designed 
in a cast material commonly used for clutch¬ 
es. The double clutch was also secured via 
a support bearing on the clutch cover by 
means of a flex plate attached to the trans¬ 
mission housing; securing the clutch in this 
way also has the advantage of providing ex¬ 


tremely good vibration isolation. Only a 
maximum engine torque of 180 Nm can be 
covered with the direct actuation (contact 
force is transferred directly from the CSC 
piston via the engagement bearing to the 
pressure plates). A modular design based 
on the previous system was developed 
(Figure 7) in order to now also be able to 
meet torque requirements of up to approx. 
250 Nm for enhanced system specifications 
and to meet the requirement for a low mass 
moment of inertia. 

A new feature is that, through the use of 
a new bearing concept (power flow via the 
transmission shaft closed), it has been pos¬ 
sible to simplify the system further (Figure 8). 
The new bearing concept prevents virtually 





























































240 




Series production 




concepts 

One-disc concept 

Two-disc concept 


dry 

wet 



Torque 
in Nm 

250 

280 

150 

250 

Torque 
in Nm 

12.5-16.6 

8.4-9.3 

10 

12 

Sec. inertia of 
masses in kgm 2 

0.09-0.15 

0.055- 

0.065 

0.08 

0.09 


Figure 7 Modular design of dry, directly actuated double clutches 


all vibration feedback from the engine and 
transmission onto the clutch system, thus 
improving the overall NVH behaviour of the 
powertrain. With this bearing concept, axial 
vibrations of the crankshaft or even the 
transmission input shafts caused by the 
forces of the helical gearing system do not 
generate any disturbing clutch torque fluc¬ 
tuations in any operating state. 


Furthermore, the intention is for steel mat¬ 
ing friction surfaces to be used for this 
concept. These surfaces offer a range of 
new design possibilities, such as reducing 
the thickness, integrating functions (e.g. a 
tone wheel directly integrated into the 
pressure plate on the engine side), new 
friction surface design (e.g. embossed 
grooves to protect against damage to the 



Figure 8 Bearing concept, directly actuated double clutches, series concept on left, new concept 
with closed power flow through the transmission input shaft on right 




































































Dry Double Clutch 


16 


241 


tribological system) etc. From a design 
and project point of view, it is beneficial if 
a radially smaller single-disc clutch is 
used for applications up to 150 Nm and a 
two-disc double clutch with a smaller di¬ 
ameter is used for higher torques and/or 
higher specific loads. A considerable ben¬ 
efit in terms of the mass moment of inertia 
of the clutch system can therefore also be 
achieved for a wide range of applications 
(20 - 30 % reduction in the mass moment 
of inertia), without this greatly increasing 
the overall costs due to the wide diversity 
of options or significantly reducing the 
thermal mass. With the two-disc concept, 
the increase in torque is achieved by in¬ 
creasing the number of friction surfaces 
from 2 to 4 for each partial clutch. The in¬ 
termediate pressure plate is secured in 
the same way as the pressure plate via 
leaf spring packages. In the axial direc¬ 
tion, the intermediate pressure plate al¬ 
ways extends about half as far as the 
pressure plate. Figure 9 shows the new, 
simplified directly actuated two-disc dou¬ 
ble clutch. 



Figure 9 New directly actuated two-disc 
double clutch 


Summary and outlook 


Dry double clutch systems offer a wide 
range of options for optimising the system 
characteristics with regard to NVH, comfort, 
complexity and mass moment of inertia. 
Specifically with the most recent develop¬ 
ments, it has been possible to make real 
progress with issues such as NVH and 
comfort, and this progress will further boost 
the success of the system, as other benefits 
such as excellent fuel consumption and low 
overall costs continue to be valid. 

The most important measures for sig¬ 
nificantly increasing NVH and comfort are: 

- New clutch linings with improved damp¬ 
ing properties (B8040 and RCFlo, as 
well as B9000 in the future) 

- Geometric optimisations for reducing dis¬ 
turbance excitation caused by geometric 
factors, such as the pairing of clutch com¬ 
ponents or the use of these by more ac¬ 
curate clutch components through the 
application of optimised and, in some 
cases, new manufacturing processes 

- Further development of the double 
clutch bearing concept in conjunction 
with the new, directly actuated double 
clutch system, thereby eliminating the 
negative effects of axial vibrations 

- Compensation of geometric errors by a 
“cardanic function” of the clutch and of 
the engagement system in the directly 
actuated double clutch concept 

- Active vibration damping through minor 
software-controlled force modulation 
of the clutch, the anti-judder control 
system 

- For hybrid vehicles, supporting the 
anti-judder control system by means 
of counter excitation via the electric 
motor, as well as the avoidance of vi¬ 
bration excitation and the reduction of 
thermal clutch loads through electrical 
launches 

























242 


Reduction of mass moment of inertia of dry 
double clutch systems can also be achieved 
thanks to the new modular and directly 
actuated concept with a reduced outside 
diameter. Directly actuated two-disc 
double clutches are used for applications 
with engine torques greater than approx. 
150 Nm for each partial clutch. The particu¬ 
larly special feature of this system is its low 
complexity. 

Using the options outlined, the dry dou¬ 
ble clutch system for the lower to mid-range 
vehicle segment will set new standards for 
efficiency and comfort. 


Literature 


[1] Wagner, U.; Zink, M.; Feltz, C.: Double Clutch 
Systems - Modular and Highly Efficient for the 
Powertrain of Tomorrow. 10 th Schaeffler 
Symposium, 2014 

[2] Wagner, U.; Buehrle, R; Mueller, B.; Kimmig, K.-L.; 
Kneissler, M.: Dry double clutch systems - In¬ 
novative components for highly-efficient vehicle 
transmissions, ATZ 11/2009, pp. 826-833 


[3] Mueller, B.; Kneissler, M.; Gramann, M.; Esly, 

N.; Daikeler R.; Agner I.: Smaller, More Flexible, 
More Intelligent - Developed Components for 
Double Clutch Transmissions. 9 th Schaeffler 
Symposium, 2010 

[4] Albers, A.; Herbst, D.: Grabbing - Causes and 
Solutions. 6 th LuK Symposium, 1998 

[5] Kimmig, K.-L.; Buehrle, P.; Henneberger, K.; 
Ehrlich, M.; Rathke, G.; Martin, J.: Efficiency 
and Comfort Lead to Success - The Dry 
Double Clutch Established in the Automatic 
Market, 9 th Schaeffler Symposium, 2010 

[6] Rudolph, F.; Schaefer, M.; Damm, A.; Metzner, F.-T.; 
Steinberg, I.: The Innovative Seven Speed Dual 
Clutch Gearbox for Volkswagen’s Compact Cars, 
28 th International Vienna Motor Symposium, 2007 

[7] Wagner, U.: What makes a transmission operate 
- Tailored actuation systems for double clutch 
transmissions; CTI Symposium, 2011 

[8] Steiger, S.; Treder, M.; Neuberth, U.; Reuschel, M.: 
Innovative Weiterentwicklungen bei trockenen 
Doppelkupplungssystemen; VDI-Berichte 

Nr. 2206, 2013 

[9] Kimmig, K.-L.; Rathke, G.; Reuschel, M.: The 
Next Generation of Efficient Dry Double Clutch 
Systems, VDI Congress, 2013 

[10] Mueller, B.; Ubben, H.; Gantner, W.; Rathke, G,: 
Efficient Components for Efficient Transmis¬ 
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Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



244 


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245 


Wet Double Clutch: 
Thinking in Systems 


Andreas Englisch 
Andreas Goetz 
Andreas Baumgartner 
Thomas Endler 
Christian Lauinger 
Stefan Steinmetz 


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246 


Introduction 


Status 

In ten years, the number of wet and dry 
double clutch transmissions (DCT) will make 
up approximately 20 % of the total automat¬ 
ic transmission market. 

Against this backdrop, customers are 
faced with the question of which double 
clutch system is right for their application. 
Both dry and wet double clutches have prov¬ 
en themselves in volume production. There 
are various designs and various forms of ac¬ 
tuation within the two systems, with key differ¬ 
ences being in the torque capacity, space re¬ 
quirements, weight and inertia of masses [1]. 


A whole range of wet double clutch sys¬ 
tems have since been developed by LuK 
in order to be able to serve a vast array of 
applications. The first wet double clutch 
went into volume production in 2013. 

In addition to the actual double 
clutches, other components such as 
dampers, centrifugal pendulum-type ab¬ 
sorbers and actuators are also available. 
The focus is on a perfectly matched 
overall system that meets the target pa¬ 
rameters of comfort, consumption and 
costs in the best possible manner. To do 
this, components and assemblies need 
to be standardized to pool volumes and 
thus be able to continue to offer appeal¬ 
ing solutions in the future. 

Throughout the development phase, 
various different concepts were analyzed 
and compared on a broad basis. The clutch 


Type 


Arrangement 


CPA 
wet/dry 

Torque/energy 

capacity 

Available 

space 

Dynamic 
requirements j 

Hybrid 
yes/no 


Hydraulic 

Hydrostatic 

Actuation 


V s Piston « 

^ Diaphragm 
r spring 

7 CSC 

^Compression 

Pump 

actuator 

/ 

\ springs 
’ Sealing 

Hydraulic 

Low 

/pressure 
k High 


Hydrostatic 

\ pressure 


Pump 

\Compen- 
i sation 


actuator 

yes/no 



♦ 



Consumption 

Comfort 

Costs 


Friction 

system 



Oil quantity 
^Design IPC 
Input 

connection 

Losses \ Output 
connection 


Figure 1 Wet double clutch system 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_17, © The Author(s) 2014 










Wet Double Clutch 


17 


247 


components were examined in detail and 
developed accordingly. 

The tribological system (comprising 
a friction plate, steel plate and oil), in 
particular, plays a key role in the design 
and comfort characteristics of the 
clutch. In addition to examining different 
friction linings and friction lining tech¬ 
nologies, geometry, grooving, as well as 
the distribution of cooling oil and pres¬ 
sure are all important. Furthermore, the 
gathered findings and experience will 
be used to develop our own linings for 
wet clutches. 

The customer’s 
perspective 


When choosing a system, customers are 
faced with a whole range of difficult deci¬ 
sions and questions that will significantly 
impact the architecture of the clutch sys¬ 
tem. 

a) Which engines are to be used in the future? 

- Are 3- and 2-cylinder applications to 
be taken into account? 

- Does cylinder deactivation need to 
be taken into account? 

- What are the minimum drive speeds 
that should be achieved? 

b) What measures are to be included to 
further reduce fuel consumption? 

- Is hybridization of the powertrain 
expected? 

- What are the maximum torques to 
be taken into consideration? 

- How is the clutch to be actuated? 

c) How do these measures impact on 
fuel consumption and driving perfor¬ 
mance? 

d) What kind of oil and oil flow rate is re¬ 
quired for cooling and, if necessary, for 
actuating the double clutch? 


All of these functions have a crucial impact 
on the consumption, comfort and, of course, 
the cost of the system. 


Design 


Clutch 

According to the specified space require¬ 
ments, wet double clutches can be de¬ 
signed in radial and axial forms, with the 
different designs offering various advantag¬ 
es depending on the application. 

The axial design allows moments of 
inertia to be kept to a minimum in relation 
to the transmission input shaft. It also 
provides the option of cooling both 
clutches independently of each other, 
thereby continuing to reduce the drag 
losses of the open clutch. The radial con¬ 
struction continues to represent the pre¬ 
ferred solution for front transverse and 
rear longitudinal applications. The trend 
towards minimizing the inertia of masses 
could also make axial solutions more in¬ 
teresting for transverse applications. 

There is the option of combining both 
systems with a centrifugal pendulum- 
type absorber in the wet area; this can 
help to further reduce fuel consumption 
and improve comfort. Figure 2 shows 
clutches for 180 Nm with extremely 
compact dimensions and a centrifugal 
pendulum-type absorber integrated into 
the clutch. In some cases, using a cen¬ 
trifugal pendulum-type absorber even re¬ 
duces the total space required, as an ad¬ 
ditional secondary mass on the DMF is 
not required. 




248 


Max. Input torque 180 Nm 



Figure 2 A wet double clutch for 180 Nm in axial and radial design with a centrifugal pendulum-type 
absorber [2] 


Centrifugal pendulum-type 
absorbers in wet area 


Using the centrifugal pendulum-type absorber 
(CPA) in the double clutch enables a sub¬ 
stantial reduction of the drive speeds, there¬ 


by also reducing fuel consumption. Figure 3 
shows the quality of decoupling by the cen¬ 
trifugal pendulum-type absorber, on the ba¬ 
sis of excitation by a 3-cylinder engine. The 
engine can be operated at very low engine 
speeds, without compromising on comfort. 

Due to the minimal additional space re¬ 
quired, these kinds of solutions in the wet 



Required space for 
additional secondary mass 



Speed in rpm 

— Engine Speed irregularities 

excitation — without additional secondary mass 

— with additional secondary mass 

- with CPA 


Figure 3 Quality of vibration decoupling with 3-cylinder engine [2] 














































Wet Double Clutch 


17 


249 




+ CPA 

+ J_sec 

Total secondary moment of inertia in kgm 2 

0.040 

0.105 

Weight difference in kg 

0 

+ 7.0 

Additional axial length required in mm 

0 

10 

Distance difference after 4 s in m 

0 

- 1.9 

Time to accelerate from 0 to 100 km/h in % 

0 

+ 4.1 


Figure 4 Double clutch with and without centrifugal pendulum-type absorber: differences in the 

mass moment of inertia, in weight, in the axial length as well as in the driving performances 


area can be easily implemented in almost 
any transmission. Using the pendulum- 
type absorber enables the total mass of 
the clutch and DMF, and therefore also of 
the inertia of masses, to be reduced, in 
addition to reducing the minimum drive 
speeds. For smaller engines in particular, 
the result is improved dynamics while si¬ 
multaneously lowering fuel consumption. 
Comparing driving performances of a ve¬ 
hicle fitted with the appropriate equipment 
also documented the positive effects 
(Figure 4). 


Actuators 


The clutch can be actuated hydraulically, 
hydrostatically or by using pump actuators. 
Hydrostatic systems (HCA) offer the advan¬ 


tage that power-on-demand systems, and 
thus significant benefits to efficiency, can 
be realized. This is reflected in Figure 5: 
Clutch and gear actuators only contribute 
4.6 % to overall transmission losses in the 
NEDC (New European Driving Cycle). How¬ 
ever, individual operating points need be to 
examined in much greater detail, such as if 
the clutch is to be used at low temperatures 
(from -20 °C to -30 °C). 

The hydraulic systems offer the advan¬ 
tage of high power density, but a perma¬ 
nent power input is also required. 

A combination of clutch, CSC (con¬ 
centric slave cylinder), hydrostatic clutch 
actuator (HCA) and gear actuator repre¬ 
sents the best version in terms of energy. 
In order to prove this, consumption simu¬ 
lations will be performed using a detailed 
model of the powertrain, so that the effi¬ 
ciency of individual components can 
be evaluated. Furthermore, different con¬ 
cepts can be evaluated, such as concepts 








































250 


12.1 % 

Mechanical pump 


4.6 % 
Clutch and 
gear actuation 


19.2 % 

Cl & C2 

losses It -- 

(starting, L 

shifting) H 

0.6 % v 

Cl drag losses / \ \ 

0.9 % 4.9 % 


S 


4.2 % 



53.7 % 
Gear teeth 
of bearings 


C2 drag losses C2 bearings Cl bearings 


Figure 5 Distribution of transmission losses in the NEDC for 180 Nm DCT with HCA and gear 
actuators 


for the low-pressure pump used and the 
operating strategy for clutch and bearing 
lubrication. 

The result of the NEDC simulation (Fig¬ 
ure 5) shows that DCT-specific losses from 
actuators and clutch drag losses can be 
reduced to a fraction of the mechanical 
losses. 

Figure 5 also shows that the low- 
pressure pump driven on the primary 
side that is used as a basis for the calcu¬ 
lation indicates a share of 11.5 % of total 
losses. It is possible to significantly re¬ 
duce this share of total losses to around 
3 % if the pump driven by the engine 
speed is replaced with an electrically 
driven one (Figure 6). The reason for this 
is the small time portion of approximately 
9 % when starting or shifting in which a 
higher cooling oil flow is required during 
driving mode for the slipping clutch. In 
contrast, the pump can be operated at a 
lower speed and therefore lower drive 
power with considerably higher time por¬ 
tions of approx. 70 % (start-stop system 


taken into account here) in order to pro¬ 
vide the minimum oil quantity for the 
non-actuated clutch and bearings. 

One aspect already mentioned has a 
positive impact on both pump concepts: 
The optimized groove geometry of the 
lining plates results in an improved oil dis¬ 
tribution requiring a significantly smaller 
quantity of cooling oil. The pump can 
therefore be designed for a smaller flow 
volume. 


Dynamics 


The previous sections looked at the ben¬ 
efits of hydrostatic control with HCA, in 
particular at the small share of clutch ac¬ 
tuators in the overall transmission losses 
in the NEDC. In this section, the dynamic 
behaviour of the line, comprising a hy¬ 
drostatic clutch actuator, the CSC and 




Wet Double Clutch 


17 


251 



— Cl actuator position 

— Cl actuator pressure 
Cl torque 


Figure 6 Measuring the dynamics of clutch Cl: 

the standardized parameters are 
shown - actuator position, actuator 
pressure as well as the torque 
transmitted by clutch Cl. 

the clutch, is explained using measure¬ 
ments. 

Figure 6 shows the measured profile 
for operation of clutch Cl. The parame¬ 
ters shown are standardized to the re¬ 
spective maximum values to enable uni¬ 
form presentation. Starting from the initial 
value of approximately 21 % the actuator 
position starts to change after 50 ms. On 
account of the clutch characteristic curve, 
the clearance must first be overcome un¬ 
til the touch point (TP) is reached. From 
this point onwards, the actuator pressure 
increases significantly. With a small time 
delay the torque transmitted by clutch Cl 
also increases. The actuator position 
reaches the target value (100 %) after ap¬ 
prox. 200 ms, which in this case is equiv¬ 
alent to the maximum driving torque ac¬ 
cording to the design. It takes approx. 
100 ms until the maximum pressure is 
reached. The maximum torque is trans¬ 
mitted as early as 120 ms after pressure 
has started to build up. The measurement 
data relates to a 550 Nm DCT. For smaller 
clutches and therefore lower actuation 


forces, the dynamics can be increased 
further. 

In order to keep the time difference be¬ 
tween the actuator position starting to 
change and reaching the required torque as 
small as possible in a real driving situation, 
the actuator position is not moved to 0 mm 
in a waiting position (WP) of the non-actuat- 
ed clutch; instead, it is moved in the clear¬ 
ance range just underneath the TP. The fric¬ 
tion system is described in greater detail in 
one of the following chapters, which also 
lists the measures designed to minimize 
drag losses in the clutches. Using these 
measures, it is possible to keep the spacing 
of the waiting position from the TP and the 
drag torque through the non-actuated 
clutch as small as possible. Doing so 
achieves short actuation times for adjusting 
a required torque. 


Axial or radial design 


Double clutches in axial and radial designs 
are compared in detail in this section. This is 
again based on the NEDC simulation, and 
the share of the individual components in 
the overall transmission losses are dis¬ 
cussed. According to Figure 2, the axial de¬ 
sign comprises two wet clutch release 
bearings with CSC. Rotary connections 
with sliding ring seals are taken into con¬ 
sideration for the radial design concept. 
However, in principle, a radial double clutch 
can also be actuated via a CSC. 

Due to the geometric ratios, the drag 
torques of Cl and C2 and the resulting 
shares in the overall transmission losses 
(Figure 8) are approximately 1 percentage 
point smaller for the axial arrangement than 
for the radial concept. The drag torques are 
calculated based on the measurements on 
the NEDC operating points. The measure- 







252 


merits regarding drag losses of various 
plate geometries are described in greater 
detail in a subsequent section. 

Diaphragm springs or 
compression springs 


Using diaphragm springs for opening the 
clutch pack represents a space-saving al¬ 
ternative to spiral springs. The question that 
therefore needs to be asked is how do these 
elements influence of the clutch hysteresis. 
As part of this question, a solution using 
diaphragm springs does not automatically 
lead to higher hysteresis values. The exam¬ 
ple shown indicates that, with the right de¬ 
sign of the diaphragm spring, contact 
surface and overall system, it is possible to 
achieve hysteresis values comparable to 
those of compression spring solutions. 


Rotary connection or CSC 


Rotary connections are used in many of to¬ 
day’s double clutch transmissions and pro¬ 
vide a robust solution in conjunction with 
hydraulic systems. Wet clutch release bear¬ 
ings with CSC can be used as an alterna¬ 
tive; the losses of these bearings are much 
lower in comparison with the rotary connec¬ 
tions. Applications up to 700 Nm are cur¬ 
rently in the development phase. 

Furthermore, the bearing concept of the 
axial arrangement with a deep groove ball 
bearing, four axial needle roller bearings 
and two clutch release bearings for CSC in¬ 
dicate a benefit of more than 3 % points in 
comparison with sliding ring seals. Again, 
measured values are converted into NEDC 
operating points. Furthermore, the CSC 
offers the option of minimizing geometrical 
deviations in the clutch to be similar to the 
dry system. 


<D 

3 

<T 



0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 


Pressure in MPa 


— LuK DC with compression springs 

— LuK DC with diaphragm spring 


Figure 7 Comparing the clutch hysteresis 














Wet Double Clutch 


17 


253 


Drag losses Cl and C2 


Electric pump/ 


Clutch release bearings 



^ sliding ring seals 




Electrically Mechanically 
driven driven 


HCA Rotary 
with CSC connection 


Figure 8 


Comparison of axial and radial designs: share of C1/C2 drag torques, as well as of the bearings 
(CSC) and the sliding ring seals for rotary connections in the overall losses in the NEDC 


The friction system 


Design 

Waves and grooves play a key role in how 
the wet clutch works. The manufacturing 
method also significantly influences the 
friction value structure and the resulting 


drag losses. In addition, the lining’s geo¬ 
metric characteristics play an essential role 
in the uniformity of torque transmission. 
Depending on the volume flow, there is a 
considerable increase in the clutch drag 
torques. 

The aim of the development phase is 
therefore to minimize the cooling oil vol¬ 
ume flow, optimize the groove geometry 
and ensure the correct wave of the plate. 


E 

z 


Different designs c 

it 1 l/min 



T-1-1- 

-1-1 


Design 3 at different cooling oil volume flows 

2 n 



-1,500 -1,000 -500 0 500 1,000 

Speed difference in rpm 

— Design 1 — Design 3 — Design 5 

— Design 2 Design 4 — Design 6 


t-1 

-1,500 -1,000 -500 0 500 1,000 

Speed difference in rpm 


- 100 % 
- 50 % 


25% 

10 % 


-5% 

2 % 


Figure 9 Impact of the plate geometry on drag losses 

















254 


Simulation 


Test result 



Figure 10 CFD analysis and comparison with test results 


Cooling 

The cooling effect is heavily influenced by 
the connection of the friction plates (primary 
or secondary side), the layout of the bores in 
the inner plate carrier and the lining groove 
geometry. 

By using CFD analyses, these effects 
can now also be demonstrated with simula¬ 
tions. Figure 10 shows the influence of the 


bore pattern in the inner plate carrier and 
the influence of the groove geometry. 

Discoloration of plates clearly shows 
the poorly cooled areas. Good consistency 
can be seen between the simulation and 
the thermal load of the plate. 

Furthermore, the thickness of the plates 
must be optimized depending on the spe¬ 
cific application. 


Stability of friction value in endurance test 


Friction plate after 20,000 hill starts under full load 



0.05-1-1-1— 

0 1,000 2,000 

Speed difference in rpm 


— New — 15,000 launches 

— 5,000 launches - 20,000 launches 

10,000 launches 



Figure 11 Endurance test with LuK friction lining: 20,000 hill starts under full load 






Wet Double Clutch 


17 


255 


LuK lining 


The friction plate really comes into its own 
within the tribological system. The plate has 
a significant impact on the friction value, the 
friction value gradient, the wear behaviour 
and the thermal capacity of the clutch, 
which is why LuK pushed the development 
of its own double clutch linings. The suit¬ 
ability of LuK lining in functional and endur¬ 
ance tests has since been demonstrated 
(Figure 11). The characteristic curves clearly 
show that the friction value gradient is fol¬ 
lowing a very strong and stable, positive 
course. It should therefore be possible to 
reliably rule out clutch judder due to the fric¬ 
tion value. 

Running in parallel with the develop¬ 
ment of wet double clutch linings is the ad¬ 
vancement of their industrialization. 


Wet double clutch 


The modular design 


tently optimizing damping components, the 
clutch and the clutch and gear actuators. 

LuK friction linings for wet double 
clutches have demonstrated their suitability 
for DCTs. 

In combination with the software, it is 
possible to find special solutions that im¬ 
prove the driving performance, signifi¬ 
cantly increase driving pleasure and further 
reduce fuel consumption for every customer. 

In conjunction with modern hybrid sys¬ 
tems, future powertrains can therefore be 
realized with a high level of comfort and 
minimum fuel consumption. 


Literature 


[1] Zink, M.; Wagner, U.; Feltz, C.: Double Clutch 
Systems - Modular und Highly Efficient for 
the Powertrain of Tomorrow, 10 th Schaeffler 
Sympoisum, 2014 

[2] Baumgartner, A.; Lauinger, C.; Lorenz, E.; 
Fischer, N.; Goetz, A.; Krause, T.: Reducing 
C0 2 emissions with wet double clutch trans¬ 
missions with a simultaneous improvement in 
comfort and driving dynamics. VDI Getriebe in 
Fahrzeugen, 2013 


Wet double clutches 
are now available in 
in radial and axial 
designs for torques 
between 100 Nm 
and 3500 Nm (Fig¬ 
ure 12). The double 
clutches can be com¬ 
bined with centrifugal 
pendulum-type ab¬ 
sorbers as an option. 

The efficiency of 
the overall trans¬ 
mission can be en¬ 
sured by consis- 



200 400 700 3,500 

Enqine torque in Nm 


Figure 12 Modular design 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 















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257 


Turning New Directions: Surprising 
Potential in Planetary Transmissions 

Part 1: Planetary gear set 


Rainer Schuebel 
Martin Gegner 
Frank Beeck 


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258 


Introduction 


The automotive industry and suppliers 
have implemented numerous innovations 
with the objective of reducing the C0 2 
emissions of individual transport. Exam¬ 
ples are general lightweight designs and 
optimizations to the exhaust gas system 
as well as numerous detailed solutions for 
engine technology. For many years, trans¬ 
mission technology has also been con¬ 
tributing to the continuous reduction of 
fuel consumption and emissions. This has 
usually been accompanied by an increase 
in the number of gears. This increased 
number of gears and the smaller trans¬ 
mission ratio spread result in smooth, 
more comfortable and hardly perceivable 
gearshift operations. At the same time, 
each additional gear has enabled reduc¬ 
tions in fuel consumption and emissions 
by several percentage points by approxi¬ 
mating the optimal tractive force hyper¬ 
bola (Figure 1). 

With the increased number of gears, 
the number of planetary gear sets in auto¬ 
matic transmissions also tended to be in¬ 
creased. This trend was not linear in rela¬ 
tion to the number of gears due to the 



4-speed 5-speed 6-speed 6-speed 8-speed 
Gen I Gen II 


Figure 1 Reductions in fuel consumption 

based on transmission development 


intelligent control of the flows of force. The 
design envelope of the transmission, 
however, remained the same. The indi¬ 
vidual transmission components therefore 
had to become smaller and more com¬ 
pact. This requirement often created spe¬ 
cial challenges for the design and dimen¬ 
sioning of components. At the same time, 
the requirements for the materials and 
manufacturing technologies used have in¬ 
creased. 

Schaeffler has been able to make sig¬ 
nificant contributions to reducing emis¬ 
sions and fuel consumption by continu¬ 
ously optimizing planet gear bearings and 
axial needle roller bearings. Recent analy¬ 
ses have shown that even inconspicuous 
new developments can offer great poten¬ 
tial. The most recent example is the new 
axial needle roller bearing support for 
planet gears. This development is consid¬ 
ered a first in rolling bearing technology 
and can contribute to reducing C0 2 emis¬ 
sions by up to 1 g/km with low additional 
costs. 


Trends and challenges 


The further development of transmission 
technology has increased the subsequent 
requirements for modern planetary gear 
sets by more than 50 % during the last 
few years. This is because fuel consump¬ 
tion can only be reduced by means of 
smaller jumps in speed, which requires a 
wider transmission ratio spread and 
causes additional outlay for the design of 
the transmission, for example due to an 
additional planetary gear set required. 
The center distances in the planetary gear 
set are also becoming larger in order to 
achieve the required forces and moments 
through the ratios. The center distance 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_18, © The Author(s) 2014 



















Planetary Transmissions 1 


18 


259 


Wider transmission ratio spread 

4 

10 

More planetary gear sets 

3 

4 

Larger center distance in % 

+ 20 

Higher planet carrier speed in rpm 

4,000 

10,000 

Higher planet gear speed in rpm 

10,000 

20,000 

Higher needle roller speed in rpm 

50,000 

100,000 

Increased acceleration value in g 

1,000 

6,700 


Figure 2 Challenges for the transmission and effects on the components 


and the speed of the planet carrier cause 
high centrifugal force loads at individual 
operating points of the transmission. Of¬ 
ten, the centrifugal forces even increase 
considerably. This results in higher loads 
acting on the planet carrier, the planet 
gear, and the planet gear bearing. The 
maximum carrier speed and the maxi¬ 
mum centrifugal acceleration are now 
slightly above the currently valid limits. 
However, these values can be controlled 
with adequate development outlay. 

The number of gears in automatic trans¬ 
missions can be increased by using addi¬ 
tional shift elements or more planet stages. 
Looking back at the six-speed automatic 
transmission, the number of gears could be 
increased further by using a fourth plane¬ 
tary gear set. This additional planet stage 
has proved ideal with regard to constant or 
reduced drag losses due to open shift ele¬ 
ments. [1] 

Certain cases, however, require an au¬ 
tomatic transmission with a particularly 
space-saving design in order to use it in 
front transverse applications, for example. 
The most suitable method for saving 


space is the nesting of planetary gear 
sets. As a result, the outer nested gear set 
has a comparably large center distance 
(Figure 2) [2]. 

Previous requirements for planetary 
gear sets have been implemented by using 
state-of-the-art bearing supports for the 
planet gears. This means that bearing loads 
of up to 3,500 times the acceleration due to 
gravity (g = 9.81 m/s 2 ) can be managed us¬ 
ing current manufacturing methods. How¬ 
ever, the latest generation of automatic 
transmissions must meet bearing load 
specifications of up to 6,700 g, and future 
transmissions must even be designed for 
up to 8,000 g. The basis for this is the de¬ 
pendency of the centrifugal acceleration on 
the speed of the load-bearing component 
(Figure 3). 

The following comparisons can give a 
better idea of the occurring forces. For ex¬ 
ample, forces of around 4 g act on the hu¬ 
man body in a carousel. A fighter pilot is 
subjected to up to 25 g when the ejection 
seat is fired. The forces of up to 6,700 g 
acting on a planet gear bearing can also 
be explained using this example: If a nee- 









260 


6,5001 
o> 


§ 5,000- 

(0 
> 

g 

'■g 3,500 H 
o 
<D 

o 

ra 2,000- 

>< 

CD 

2 


500- 

4,000 


High planet carrier speeds 












8,000 10,000 


6,000 


Planet gear carrier speed in rpm 


• 4 AT • 6 AT • 8 AT 

• 5 AT • 7 AT • 9 AT 


Figure 3 Changes in planet carrier speed 


die roller in a planet gear bearing weighs 
one gram, an acceleration of 6,700 g re¬ 
sults in a mass of around 6.7 kg. Such val¬ 
ues push the limits of design and materi¬ 
als technology and require innovative 
development solutions. 


Power losses 


Even though transmissions generally have 
relatively high levels of efficiency, losses 
cannot be completely prevented in indi¬ 
vidual assemblies. They are mainly caused 
by the lockup clutch and its actuation sys¬ 
tem in the converter. The one-way clutch 
supporting the stator is considered as 
relatively negligible in this regard. The ad¬ 
jacent oil pump is increasingly actuated 
by means of a separate electromechani¬ 
cal system. This means that energy must 
be used only when a specific oil quantity 
or oil pressure is required. The focus of 
planetary gear sets is on the influence of 


centrifugal force, the rolling friction of the 
gear teeth, the bearing friction, and the 
axial sliding friction of the planet gears. 
The main focus of bearings for automatic 
transmission components is on axial 
bearings, ball bearings, or radial plain 
guidance systems. Additional losses are 
due to the viscosity of the oil, the throt¬ 
tling/pump effects of various rotating 
components, and churning losses. 


Reducing frictional 
power 


Planet gear bearings in general 

The main purpose of planet gear bearings 
is to position the planet gears [3]. In addi¬ 
tion, they must support forces and mo¬ 
ments and ensure rotation of the planet 
gears with minimal friction. Bearings with 
needle roller and cage assemblies are 
mainly used due to the speed and cen¬ 
trifugal force requirements. The cage, 
which is usually made of steel, guides and 
positions the rolling elements. The cages 
are generally manufactured using form¬ 
ing, punching and welding methods in or¬ 
der to ensure efficient production of the 
high quantities required for automotive 
applications. 

The raceways for the rolling bearing 
are located directly on the gear set com¬ 
ponents. The outer raceway is defined by 
the bore of the planet gear. The stud that 
is firmly located in the planet carrier pro¬ 
vides the inner raceway. Special thrust 
washers facilitate the axial contact be¬ 
tween the planets and the bearing cages. 
This design represents a cost-effective 
and highly functional bearing arrange¬ 
ment. 






Planetary Transmissions 1 


18 


261 


The bearing itself comprises a specific 
number of load-bearing rolling elements 
(needle rollers) and a so-called cage, 
which guides the needle rollers both axi¬ 
ally (cage rib) and in circumferential direc¬ 
tion (cage crosspieces). The cage cross¬ 
pieces have outer and inner retentions 
that retain the needle rollers radially and 
prevent them from falling out. The position 
of these retentions has been optimized so 
that cage stresses can be kept as low as 
possible. The pocket corner radii have 
also been optimized in order to reduce the 
component stresses. 


Cage design 

Another important point in the design of 
planet gear bearings is the friction behav¬ 
ior. If the diameter, length or number of 
rolling elements are modified, this can re¬ 
sult in significant changes in friction. The 
following comparison of the calculated 
friction components of two bearing de¬ 
sign alternatives subjected solely to cen¬ 
trifugal force loads demonstrates this as 
an example. Compared to the standard 
design with rolling elements with a diam¬ 
eter of 2.5 mm, the alternative has thinner 
rolling elements with a diameter of 2.0 mm 
and therefore a slightly larger inner race¬ 
way diameter. With the same static load 
ratings, the friction of the rolling elements 
with the smaller diameter is reduced by 
around 11 %. 

Cage friction depends on the selected 
rolling element diameter and therefore has 
a significant influence on the friction be¬ 
havior of the bearings. This presents var¬ 
ied challenges for the design of planet 
gear bearings as it must combine suffi¬ 
cient load ratings and cage strength with 
minimal friction. 


Coating 

Adapted coatings for bearing cages also 
play an important role in increasing the ef¬ 
ficiency of planet gear bearings. The bear¬ 
ing cage is subjected to sliding contact 
due to the occurring loads. For example, 
the rolling elements are pressed against 
the outer retentions of the crosspieces 
when subjected to centrifugal force, and 
the outside surface of the cage is pressed 
into the planet bore. 

Simulations have shown that optimizing 
the contact surfaces of the cage is advan¬ 
tageous as they account for up to 70 % of 
the friction. The rolling surfaces of the nee¬ 
dle rollers and the raceways are less 
suitable for low-friction coatings as they 
already meet the highest surface require¬ 
ments in order to function as rolling race¬ 
ways. 

Coatings based on zinc phosphate 
(Durotect Z) or manganese phosphate (Du- 
rotect M) improve the sliding and wear be¬ 
havior of metal contact surfaces that slide 
against each other and also contribute to 
corrosion protection. Due to their capabili¬ 
ty of storing oil, the layers are also used 
specifically for sliding contact surfaces. 
The previously used Durotect Z layer has 
been replaced by the Durotect M layer for 
planet bearing applications. Practical ex¬ 
perience has shown that this layer achieves 
significant advantages in terms of friction 
and wear. 

If additional increases in efficiency 
are required, a specially developed coat¬ 
ing based on nickel and phosphorus (Du¬ 
rotect NP) is used on the bearing cages 
to reduce friction. This layer offers very 
good adhesive wear resistance, excellent 
dry running characteristics, and temper¬ 
ature resistance in combination with out¬ 
standing sliding and anti-adhesive char¬ 
acteristics. 

The measured frictional power using 
Durotect M is approximately 10 % lower 


262 


compared to the previously used Durotect Z 
layer. A significant reduction in wear has 
also been achieved. The Durotect NP coat¬ 
ing additionally reduces friction by a further 
13 %. 


Planet gear design 

In order to investigate possible optimiza¬ 
tions for planet gear design, Schaeffler 
modified the parameters of the bearing 
support of a planet gear while maintaining 
the same gear teeth. The objective was to 
achieve the maximum possible planet 
gear inside diameter, which at the same 
time requires the minimum possible wall 
thickness and the maximum possible out¬ 
er raceway of the bearing. Varying rolling 
element diameters and the correspond¬ 
ingly required number of rolling elements 
in the pitch circle diameter were used to 
determine the radial bearing support ge¬ 
ometry with the longest rating life. This 
specified geometry also determines the 
raceway inside diameter and the size of 
the planet gear stud. 

The smaller wall thickness reduces the 
mass of the planet gear and the resulting 
centrifugal forces. 


The investigations led to the following 
conclusions: The strategy of using a 
smaller rolling element diameter and sub¬ 
sequently increasing the number of rolling 
elements results in a bearing support with 
reduced friction and a higher load carry¬ 
ing capacity. Calculations have shown an 
increase in rating life by 55 % for the opti¬ 
mized bearing in conjunction with a re¬ 
duction in the planet gear mass by 30 %. 
This leads to a reduction in friction by 50 % 
in the radial bearing and a reduction in 
the centrifugal force by 60 %. Figure 4 
provides an overview of the successes 
achieved in development. 

Latest findings from 
axial bearing supports for 
planet gears 


In a planetary gear set, the annulus, the 
planet gear and the sun wheel mesh with 
each other. The planet gear plays a spe¬ 
cial role as it meshes with both the an¬ 
nulus and the sun wheel. The gear teeth 


Optimizations compared 
with the previous product 

Friction behavior 
compared with the 
standard as basis 
(100%) 

Responsibility 
for development 

Optimized planet gear bearing 
cage design 

-11 % 

Schaeffler 

Optimized planet gear bearing Durotect -NP 

cage coating 

-23 % 

Schaeffler 

Weight-optimized 
planet gear 

-50 % 

Currently responsibility 
of the customer 


Figure 4 Overview of improvements 













Planetary Transmissions 1 


18 


263 



■=>Nominal 


axial force — Nominal 

■=>Circumferential force helix angle 

>=> Resulting —Modified 

axial force helix angle 

Figure 5 Force conditions on the planet gear 

are helical. Meshing forces are generat¬ 
ed when the planet gear meshes with 
the annulus and the sun wheel. They 
cause a force that acts on the planet 
gear stud in circumferential direction 
(Figure 5). 

Furthermore, centrifugal forces occur 
when the planet carrier rotates. The rota¬ 
tion of the planet gear generates frictional 
forces due to the contact with the planet 
gear bearing. Forces are also generated 
by the angular acceleration of the planet 
gear, and additional frictional forces result 
from the rolling contact itself. All three 
forces act against one of the two meshing 
forces depending on the direction of the 
power flow. This causes irregularities in 
the sun/planet and planet/annulus sys¬ 
tems, resulting in an axial force that acts 
on the end faces of the planet gears (red 
arrows). 


As part of the development work in this 
field, Schaeffler modified the last position 
without rolling bearing supports in the plan¬ 
etary gear set. This position has a sliding 
contact surface. It comprises either a non- 
ferrous metal washer or a steel washer, and 
sometimes also a combination of materials. 
The non-ferrous metal washer is the friction 
partner for the planet carrier made of un¬ 
hardened steel. The steel washer is used as 
a thrust washer for the planet gear and its 
bearing. 

The plain washers are very small. They 
can have various characteristics despite 
their dimensions, for example an inside 
diameter of 17 mm, an outside diameter of 
30 mm, and a thickness of 1 mm. These 
include: 

- special oil feed grooves or oil ways, 

- retentions for simplified final assembly, 

- anti-rotation locking devices (to prevent 

abrasive wear), and 

- various coatings. 

The small design envelope represents a 
major challenge for the development of an 
adequate rolling bearing. A reliable stan¬ 
dard axial needle roller bearing has a prod¬ 
uct width of 2.13 mm. The needle roller di¬ 
ameter is 1.5 mm, and the washer thickness 
is 0.63 mm. 

The objective was to develop a new ax¬ 
ial needle roller bearing with the same 
smaller dimensions. Schaeffler has suc¬ 
cessfully achieved this with its latest axial 
needle roller bearing with an axial washer. It 
has an inside diameter of 17 mm, an outside 
diameter of 29.9 mm, and a thickness of 
1.2 mm. Schaeffler was able to reduce the 
needle roller diameter to 1.0 mm and the 
washer thickness to 0.2 mm (Figure 6). 

This design places very high demands 
on the quality of the material and its sur¬ 
face and heat treatment in order to fulfill 
the requirements for rolling bearings and 
withstand the occurring loads. The film¬ 
like washer thickness also represents a 
special challenge for the production pro- 


264 


L - lLi - J 



r\ 


► 



◄ 



Dimensions 

Thrust washer 

New axial needle 
roller bearing 

Standard axial needle 
roller bearing 

Inside diameter 

17 mm 

17 mm 

50.8 mm 

Outside diameter 

30 mm 

29.9 mm 

67.5 mm 

Width 

1.0 mm 

1.2 mm 

4.0 mm 

Needle roller dimensions 

- 

1.0x2.25 mm 

3 x 4.3 mm 

Washer thickness 

1.0 mm 

0.2 mm 

1.0 mm 

Needle roller speed 

- 

500,000 rpm 

120,000 rpm 


Modifications to the axial planet gear bearing support 


Figure 6 

cess. The needle roller with a diameter of 
1 mm and a length of 2.25 mm is the 
smallest rolling element ever used in 
transmission applications. The axial nee¬ 
dle roller cage must have a very filigree 
design in order to securely guide and re¬ 
tain the needle rollers. 

The behavior of the axial bearing sup¬ 
port in the planetary gear set for an entire 
transmission and its effects on fuel con¬ 
sumption were investigated for four differ¬ 


ent planetary gear sets using simulation 
tools. The simulation was based on the 
NEDC with a reduced number of load 
points of 1,400. The engine data map 
based on the NEDC and the mass inertia 
values correspond to those of a premium 
vehicle. The friction parameters were de¬ 
termined on the basis of test stand runs 
(Figure 7) [4, 5]. 

The results in Figure 8 show a compari¬ 
son of the power loss of the individual plan- 



Friction parameters p R from the heat balance 

Scenario 

Thrust load 

Churning 

Wear 

P_AXK (basis) 

Mr_AS 

1 

3.5 % 

yes 

yes 

74.40 W 

0.018 


2 7.0% yes - 14.88 W 0.070 

AS: Axial washer AXK: Axial needle roller and P_AXK: Power loss of axial needle 
cage assembly roller and cage assembly 


Figure 7 Simulation 




















































Planetary Transmissions 1 


18 


265 



Gears 

■ Frictional power with thrust washer ■ Frictional power with axial needle roller bearing 
Figure 8 Simulation results 


etary gear sets under various load condi¬ 
tions using plain bearings or rolling bearings. 
The values can be accumulated for the rel¬ 
evant gears only. Under the specified load 
conditions, the axial bearing achieves rela¬ 
tively high values in two planetary gear sets, 
which represent a highly effective reduction 
in friction if a suitable rolling bearing support 
is used. 

In the third gear, for example, the maxi¬ 
mum frictional power is 470 W if a thrust 
washer is used, but only 50 W if an axial 
needle roller bearing support is used. For 
the entire transmission with four planetary 
gear sets, this means a reduction in fric¬ 
tional power by 420 W or 90 % in the third 
gear. Based on the simulation, a reduction 
in fuel consumption by around 0.5 % in the 
NEDC can be expected if the plain wash¬ 
ers are replaced with axial needle roller 
bearings. 

Comparative tests with axial bearings 
and plain washers on an axial bearing high¬ 
speed test stand also confirmed reduced 
frictional torque and additional temperature 
differences. To determine the speed limits 


of the axial bearing depending on the axial 
load, the oil temperature and the oil flow 
rate, the axial load is introduced into the test 
stand using a hydraulic system. The hydro¬ 
static system enables measurements of the 
frictional torques at high speeds. 

Speeds of 6,000 rpm and 20,000 rpm 
were specified as test conditions. The axial 
load was 500 N. A reduction in frictional 
torque from 0.23 Nm to 0.024 Nm was 
achieved at a speed of 6,000 rpm. At a 
speed of 20,000 rpm, the frictional torque 
was reduced from 0.13 Nm to 0.03 Nm. This 
means that the frictional torque can be re¬ 
duced by around 90 % with the new axial 
needle roller bearing. At the same time, the 
temperature on the bearing position de¬ 
creases by 5 to 10 °C (Figure 9). 

As an alternative to complex and costly 
tests with planetary gear sets in entire trans¬ 
missions, Schaeffler has a component test 
stand that provides the option of investigat¬ 
ing the function and operating life of entire 
planetary gear sets subjected to centrifugal 
force and specified loads. The moment is 
variably introduced using two coupled plan- 



























266 


with an axial load of 500 N 


with an axial load of 500 I 


0.25 


= 0.15 



0.05 


6,000 20,000 
Speed in rpm 

p R axial washer 

Mr axial needle roller bearing 


6,000 20,000 
Speed in rpm 

I Temp, of axial washer 
Temp, of axial needle 
roller bearing 


Functional tests on the axial bearing high-speed test stand 

Axial load, bearing and oil temperature, oil flow rate, frictional torque, 
speed and operating time can generally be measured. 


Figure 9 Results of the friction and temperature measurements on the 
axial bearing high-speed test stand 


etary gear sets. Ad¬ 
ditional influencing 
parameters of the 
test setup are the 
supplied quantity of 
oil and its tempera¬ 
ture for lubricating 
and cooling the 
gear set. The mea¬ 
sured bearing tem¬ 
perature has proven 
to be a reliable in¬ 
spection criterion 
for monitoring. 

Temperature sen¬ 
sors measure the 
temperature directly 
in the gear set on 
each planet gear 
bearing support. 

This enables con¬ 
clusions to be 
drawn about the 
functional capability and the behavior of the 
system during the test. If a sudden increase 
in temperature is measured, this is a reliable 
indication of damage to the planet gear 
bearing support. In most cases, this means 
that the bearing cage is defective. 

This test stand can also be used for 
comparative tests with planetary gear sets 
using axial bearings and planetary gear 
sets using plain washers. For this test set¬ 
up, the annular gears were preloaded 
against each other by up to 1,000 Nm and 
located. The planet carriers were subject¬ 
ed to a drive speed of up to 6,000 rpm. 
This resulted in a maximum planet gear 
speed of 20,000 rpm. 

The measurement results in Figure 10 
show the mean and maximum values of the 
axial bearing and plain washer tempera¬ 
tures and the corresponding temperature 
differences. A comparison of the results for 
an input moment of 100 Nm and at a speed 
of 6,000 rpm, and for an input moment of 
1,000 Nm and at a speed of 2,500 rpm 


shows significant differences between the 
two design variants. The mean temperature 
difference is 1.5 °C in the first case and 3 °C 
in the second case. The maximum values 
even show a temperature difference of 4 °C 
and 5 °C, respectively. 

The reduced temperatures and fric¬ 
tional torques determined in the simula¬ 
tion of the entire transmission and in 
actual component tests of individual bear¬ 
ings (Figure 11) have a significant effect on 
the system’s emission characteristics. For 
example, the calculated reduction in C0 2 
emissions of the entire transmission in 
the NEDC is around 1 %. Replacing the 
planet thrust washers in an automatic 
transmission comprising four planetary 
gear sets with axial bearings results in a 
reduction in C0 2 emissions of 1 % with 
additional costs amounting to less than 
ten euros. 










Planetary Transmissions 1 


18 


267 


Absolute temperature 
Stress torque 


Temperature difference 
Stress torque 


| 1,000 Nm 100 Nm 6-i 1,000 Nm 100 Nm 

i _11.1 


1,800 


2,500 


4,000 


6,000 


Planet carrier speed in rpm 
Thrust washer ■ Axial needle roller bearing 


1,800 2,500 4,000 6,000 

Planet carrier speed 
in rpm 


The planetary gear set system with the axial bearing was tested and compared with the plain 
washers on this test stand. 

For this test setup, the annular gears were preloaded against each other by up to 1,000 Nm and 
located (speed 0 rpm). The planet carriers reached a drive speed of up to 6,000 rpm. 

The temperature is measured on the planet gear stud. 


Figure 10 Functional tests on the planetary gear set test stand and results 


Investigations 

Thrust washer 

Axial needle 
roller bearing 

Reduction 
in friction 

Reduction 
in friction 

Reduction in 
C0 2 emissions 

Simulation (3 rd gear) 

470 watts 

50 watts 

420 watts 

90% 

1 % 

Components test 
(6,000 rpm) 

0.23 Nm 

62 °C 

0.02 Nm 

57 °C 

0.21 Nm 

5 °C 

90% 

1 % 

Planetary gear set test 
(6,000 rpm) 

150 °C 

146 °C 

4 °C 

- 

- 



The axial needle roller 
bearings achieve a 
reduction in C0 2 
emissions of 1 % 
with additional costs 
of only 10 euros 


Figure 11 Effects on emission characteristics 

































268 


Modular design in planetary gear 
sets 

Today, Schaeffler offers and supplies a 
large number of individual components 
and additional parts for planetary gear 
sets. Examples are the planet gear bear¬ 
ings that radially position the planet gear, 
support bearing forces and moments, and 
ensure rotation with minimal friction. The 
planet gear stud represents the inner race¬ 
way of the planet gear bearing. Schaeffler 
offers planet gear studs in all oil feed and 
geometry variants. The axial contact be¬ 
tween the planet gear and the planet gear 
bearing is supported by the plain washers 
described above or by the newly devel¬ 
oped axial needle roller bearings, which 
are also Schaeffler components. The plas¬ 
tic oil collector and the axial needle roller 
bearings running on the planetary gear set 
complement Schaeffler’s product range 
(Figure 12). 

The carrier usually comprises formed 
parts that are drawn and punched and 
have gear teeth manufactured by forming 
methods. Schaeffler uses its core exper¬ 


tise in these manufacturing technologies 
for a wide range of products, such as 
planet gear carriers or multi-disk clutch 
carriers. 

Welding is the preferred joining method 
for manufacturing planet gear carriers. 
Schaeffler also has extensive experience in 
riveting technology, which is used for dual 
mass flywheels, torque converters or an¬ 
nulus carriers, for example. This expertise 
can also be used for the assembly of plan¬ 
et carriers. The advantages of riveting 
compared to thermal joining are that no 
thermal distortion occurs and no welding 
spatter must be removed due to welding. 
Schaeffler also develops the gears for 
planetary gear sets in-house in order to be 
able to offer comprehensive assemblies. 
During this development work, Schaeffler 
has gained a great deal of experience in 
high-performance planetary gear sets. 
Schaeffler has therefore been able to posi¬ 
tion itself as a development partner and 
supplier for mechanical assemblies or 
comprehensive solutions for planetary 
gear sets for manufacturers of entire trans¬ 
missions. 


Radial rolling bearing support 
for planet gears (KZK) 


Axial plain bearing support 
for planet gears (AS) 


Axial rolling bearing support 
for planet gears (SAX) 


Oil distributor and metering 
elements, plastic parts 



Planet gear stud with 
optimized oil feed 


Riveted planet 
carrier joints 


Integrated axial needle roller 
bearing supports for main shafts 


Additional modules, 
e.g. clutch carrier 


Figure 12 Schaeffler components with modular design for planetary gear sets 



Planetary Transmissions 1 


18 


269 



Planet gear carrier with three 
planet gears 



Planet gear carrier with 
different axial bearing 
supports 



Planet gear carrier with three 
planet gears and integrated 
spline hub 



Planet gear carrier with four 
planet gears 


Double planet gear carrier 
arrangement with three 
four planetary 
gear sets 


Planet gear carrier with 
integrated clutch 



carrier 


Figure 13 Variations in planetary gear set modules 

Schaeffler offers modular concepts for 
planetary gear sets that can easily be in¬ 
tegrated into existing transmission de¬ 
signs. The gear sets are characterized by 
gear teeth (spline teeth, engaging teeth) 
that are manufactured using forming 
methods. The planet carrier can be de¬ 
signed and manufactured using welding 
or riveting technology. Due to lower pur¬ 
chasing costs and shorter cycle times, 
riveting is particularly suitable if new ma¬ 
chinery must be purchased. The entire 
gear set including additional parts is 
matched to the specific application in or¬ 
der to achieve the best possible oil lubri¬ 
cation and the lowest possible friction. 
The planet carriers can be designed with 
adjacent components such as multi-disk 
clutch carriers and annulus carriers, or in¬ 
tegrated into load stages and differential 
stages (Figure 13). 

Collaboration between Schaeffler’s 
specialists and the transmission manu¬ 
facturer early on in the concept phase is 
useful if the modular strategy is to develop 
its full potential. 


Literature 


[1] Scherer, H.; Wagner, G.; Naunheimer, H.; 

Dick, A.: Das automatische Getriebe 8HP70 
von ZF. Getriebesystem, konstruktiver Aufbau 
und mechanische Bauteile. VDI reports 2029, 
Duesseldorf: VDI Verlag 2008 

[2] Greiner, J.; Scherer, H.; Girres, G.; Dick, A.: 
Transmission Kit for Front-Wheel-Drive Appli¬ 
cations from ZF. VDI reports 2130, Duessel¬ 
dorf: VDI Verlag 2011 

[3] Pabst, A.; Beeck, F.: Increasing the perfor¬ 
mance of planetary bearings for modern 
automatic transmissions. VDI reports 2158, 
Duesseldorf: VDI Verlag 2012 

[4] Koch, O., Weber, J., Zintl, G., Gronau, B.; En- 
ergieeffiziente Auslegung von Walzlagerungen, 
VDI conference on plain bearing and rolling 
bearing supports, 2011 

[5] Koch, O., Gao, G.: Energy-efficient design of roll¬ 
ing bearings: An important contribution towards 
reducing C0 2 emissions and increasing power 
density in automotive transmissions, CTI Sympo¬ 
sium Innovative Automotive Transmissions and 
Hybrid & Electric Drives, 2012 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 







270 


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271 


Turning New Directions: Surprising 
Potentials in Planetary Transmissions 


Part 2: Shifting clutches 


Jeff Hemphill 
Philip George 
Vural Ari 
Chris Luipold 
Patrick Lindemann 
Greg Copeland 


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272 


Drivers for change Controllability and drag 


The wet friction shifting clutch was devel¬ 
oped in the 1930’s and has been very 
successful [1]. Multi-plate wet clutches 
are used in large volumes not only for 
shifting automatic transmissions and 
CVT’s, but also as launch devices in some 
CVT and DCT transmissions. The early 
development of this technology was so 
successful that it has met the needs of in¬ 
dustry with relatively few changes in basic 
construction for many years [2]. 

Current trends in the market however, 
are placing new demands on shifting 
clutches. The number of speeds in auto¬ 
matic transmissions is increasing dra¬ 
matically, as shown in Figure 1. The de¬ 
mand for improved shift comfort is 
likewise stronger than ever. The push for 
sustainable mobility continues to in¬ 
crease and includes the environmental 
effects of manufacturing processes. Fi¬ 
nally, fuel economy standards are rising 
steeply around the world, making drag 
torque and mass reduction ever bigger 
problems. 



Figure 1 Increasing number of speeds in 

planetary automatic transmissions 
over time 


Multi-plate wet friction clutches are sub¬ 
ject to a paradox: Low lift-off gaps im¬ 
prove controllability while large lift-off 
gaps help reduce drag torque. The con¬ 
trollability is influenced by the two-stage 
nature of the pack characteristic. That is, 
the piston must first close the lift-off gap 
against little significant resistance and 
then clamp the pack to provide torque ca¬ 
pacity. A typical clutch pack schematic 
can be seen in Figure 2. 



Figure 2 Schematic representation of a clutch 
pack 

Since the axial movement of the piston to 
close the gap is 1-3 mm while the com¬ 
pression of the pack is 0.1-0.3 mm, the 
bulk of the oil volume used to actuate the 
piston in dedicated to closing the lift-off 
gap. However, the piston area has to be 
sufficient to allow it to generate enough 
clamping force to provide the needed 
torque. This means that it is normally a 
slow process to close the gap. Further¬ 
more, the transmission controller has no 
way of knowing when the gap is closed. 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_19, © The Author(s) 2014 





































Planetary Transmissions 2 


19 


273 



0 5 10 15 20 25 30 

Volume in cc 


Figure 3 Pressure vs. volume for a typical shifting clutch 


This can lead to torque errors if the con¬ 
troller doesn’t guess accurately. A pres¬ 
sure vs. volume characteristic is shown in 
Figure 3. 

The usual solution to this problem is a 
pre-fill strategy. The controller keeps a 
look-up table of time to touch point for 
given temperatures and furnishes a very 
high flow to close the gap. When it’s time 
estimate is reached, the flow is reduced 
and torque control 
takes over. If there 
is a torque error, 
the controller can 
adapt the time in 
the table. This 
method works but 
gives rise to sever¬ 
al errors. 

The drag is im¬ 
proved by a larger 
lift-off gap, since it 
is largely viscous 
drag and the shear 
forces in the oil are 
smaller in larger 
gaps. A measure¬ 
ment of a clutch 
pack with various 
lift-off gaps is 


shown in figure 4. It is also important to 
keep in mind that each clutch has a dif¬ 
ferent tolerance situation. Therefore, the 
nominal drag may be acceptable but the 
maximum drag can be significantly higher. 

Another factor which can raise clutch 
drag is plates sticking together, even 
though there is a lift-off gap. This can 
happen because the oil between the fric¬ 
tion material and the steel plate forms a 



Figure 4 Multi-plate clutch drag for various lift-off gaps 








274 



Figure 5 Two-stage clutch apply 
mechanism 

seal which allows atmospheric air pres¬ 
sure to hold the plates together. 

Given these physics a new approach 
is needed. Since we recognize that the 


friction pack offers a two-stage character¬ 
istic, a two stage mechanism could be 
used to break the paradox. A schematic 
of such a mechanism is shown in Figure 5. 
In this principle sketch, a ramp and crank 
mechanism are introduced between the 
piston and the clutch pack. This mecha¬ 
nism is actuated by the main piston and, 
due to the high ratio of the crank, fills the 
gap with very little piston travel. This 
means that a larger lift-off gap can be 
closed quickly and with little oil demand, 
avoiding a pre-fill strategy. Furthermore, 
when the clutch reaches the touch point, 
very little torque is exerted due to the high 
ratio of the mechanism. This minimizes 
torque errors. The pressure over volume 
curve for a clutch with this mechanism is 
shown in Figure 6. 

The design for a mechanism which can 
meet these functional requirements is 
shown in Figure 7. 

This mechanism functions as follows: 
High pressure oil enters the area behind 
the main piston. The oil enters the rotary 
actuator through several holes in the main 
piston. The oil pressure rotates the actua¬ 
tor, which, in turn, rotates the ramp ring. 
As the ring moves down the ramps, it 
closes the gap to the friction plates. Once 



Figure 6 Pressure vs volume characteristic for a clutch with a two-stage mechanism 











































Planetary Transmissions 2 


19 


275 


Vane 

Ramp ring actuator 

/ / Return 



Figure 7 Cross-section of two-stage piston 

the ramp ring hits the friction plates, it 
stops. As oil pressure continues to in¬ 
crease, the main piston now begins to ad¬ 
vance. Since the only travel that the main 
piston needs to make is to compress the 


clutch pack, it is designed as a membrane 
piston, this allows the piston to accom¬ 
plish the roughly 0.3 mm displacement 
and also allows it to act as its own return 
spring. The clamping action of the main 
piston is applied through the rotary actua¬ 
tor. This clamps the actuator closed and 
provides two additional benefits: The ac¬ 
tuator is sealed and the ramp ring is 
clamped in place, preventing any unwant¬ 
ed adjustment. Figure 8 shows a simula¬ 
tion of this type of clutch versus a normal 
clutch. Here we can see a faster engage¬ 
ment time, smaller torque error, and a 
larger lift-off gap. 

This leads to the following concrete ad¬ 
vantages: 

- Allows lift-off gaps of 3 mm or more 
without shift time penalty. 

- Reduces drag torque by allowing such 
lift-off. 

- Reduces oil flow required for clutch ac¬ 
tuation, potentially allowing a smaller 
transmission pump. 



— Baseline pressure — Baseline displacement 

— Gap filler pressure — Gap filler displacement 


Figure 8 Simulation results of two-stage mechanism vs normal mechanism 









276 


- Improves controllability by reducing 
torque error when reaching the touch 
point. 

- Eliminates the need for tolerance cor¬ 
rection in clutch pack assembly. 

Friction plate with a 
built-in separator feature 


Now that we have shown a mechanism 
which can open a larger lift-off gap with¬ 
out penalty, we need to assure that we 
separate the friction disks in this gap. 
Over the years, several things have been 
tried to accomplish this including separa¬ 
tor springs, hydrodynamic forces, etc. 
These concepts usually suffer from toler¬ 
ance problems and often can space either 
the friction plates or steel plates but not 




— without plate separators 

— with plate separators 


Figure 10 Drag torque measurement with and 
without plate separators (SAE#2 
plates, 0.71pm, 60C) 

the one from the other, which is the im¬ 
portant interface. 

Figure 9 shows a friction plate with a 
built-in separator feature. In this design, 
tabs have been formed in the plate itself 
and slightly twisted so that the edges of 
the tabs protrude above the friction mate¬ 
rial by the amount of the desired lift-off 
gap. Since these tabs only have line con¬ 
tact with the separator plate, the surface 
area for viscous drag is dramatically re¬ 
duced. This cannot be accomplished with 
other methods, such as “waving” the fric¬ 
tion plates. The thin arms on the tab act 
as torsion springs, allowing the tab to be 
compressed back in line with the friction 
material during engagement. 

A measurement of a friction pack with 
and without the plate separators is shown 
in Figure 10. Here a reduction in drag 
torque of more than 60 % can be seen. It 
should be noted that this concept also 
has a significant tolerance stack-up. How¬ 
ever, when using it with the gap filling pis¬ 
ton, the additional tolerances do not pres¬ 
ent a penalty. 


Figure 9 Friction plates with separator feature 






Planetary Transmissions 2 


19 


277 


Friction Disk Production 


The concepts reviewed so far have en¬ 
abled better performance of a clutch pack 
in operation. Now we turn our attention to 
improving the manufacturing process. 
Friction plates today are made by stamp¬ 
ing a steel ring, acid etching the ring, ap¬ 
plying adhesive, placing friction paper on 
the adhesive, clamping between hot, par¬ 
allel plates, and cutting oil flow grooves if 
required. This process has several disad¬ 
vantages: 

- Environmentally harmful, and therefore, 
difficult to dispose of chemicals are 
used for cleaning the parts and as ad¬ 
hesive. 

- The process can be difficult to control 
especially since friction performance 
is influenced by the amount the paper 
is cured during bonding. This means 
the adhesive and paper must both be 
cured to the right level in one pro¬ 
cess. 

- Multiple process steps are required to 
reach the end result. 

An improvement can be made by eliminat¬ 
ing the adhesive and using a mechanical 
connection between the paper and the 
steel. An example of such a construction is 
shown in Figure 11. 

In this design, two thin steel plates are 
pressed into a paper ring from either side. 
The steel plates meet at the teeth around 
the inside diameter and at a series of 
holes in the middle of the paper ring. At 
each hole, the steel plates are joined by a 
coining operation similar to riveting. The 
resulting pressed grooves provide a me¬ 
chanical means for transmitting torque 
from the spline teeth to the paper. They 
also provide a means for allowing cooling 
oil to flow. 

This design eliminates most of the is¬ 
sues with current production methods. 



Figure 11 Composite friction Disk 
construction 

There is no need for adhesive and, there¬ 
fore, no need for acid to prepare the steel 
for it. It is much easier to control since only 
paper curing is involved. It can be a one 
step process wherein the steel is com¬ 
pressed into the paper, the coining is ac¬ 
complished, and the paper gets a final cure 
and flattening. This process can also be 
much faster than a bonding process since 
the time required for the adhesive to flow 
and cure is eliminated. 

Further advantages include reduction 
in mass and inertia of the friction plate. 
This can result in a savings of 0.5 kg in a 
typical automatic transmission. The meth¬ 
od can be used with various friction pa¬ 
pers, allowing the same range of friction 
performance as with bonded plates. In 
some cases, even better performance 
can be achieved since the paper is rough¬ 
ly 3 times thicker than in a bonded design. 
This allows a softer stiffness which is 
more forgiving to the additives in the oil. 



278 




Figure 12 Comparison of friction behavior between composite facing (top graph) and bonded facing 
(HOC, 2,700/3,500rpm, SAE J2490 test profile) 


Figure 12 shows a comparison of friction 
performance between a composite facing 
and a bonded facing with the same paper 
and the same total paper thickness. As 
expected, there is virtually no change in 
friction behavior. Tests are underway to 
quantify the advantage of the composite 
facing compared to typical thickness 
bonded friction plates. Various groove 
patterns are possible. In fact, the resulting 
groove geometry is similar to a pad de¬ 
sign, without the intensive processing 
which is normally needed for that con¬ 


struction. Finally, the design also lends it¬ 
self to the plate spacers described in the 
previous section. 


Conclusion 


The demands on shifting clutches are in¬ 
creasing with the new generation of auto¬ 
matic transmissions and increasing con- 














Planetary Transmissions 2 


19 


279 


sumer demands. These new requirements 
can be met by breaking some of the old 
paradigms of clutch design: 

- Creating a two-stage apply character¬ 
istic allows better controllability with 
larger lift-off gaps. 

- Plate separators maximize the advan¬ 
tage of this larger gap. 

- Eliminating adhesive provides an envi¬ 
ronmentally friendly production method 
while decreasing mass and inertia. 

Together, these lead to some notable im¬ 
provements on the vehicle level, such as: 

- fuel economy improvement of 1-2 % 

- mass reduction of 0.5 kg 

- reduced space requirement (see Figure 13) 

- shift time improvement of up to 100 ms 
with reduction in shift shock complaints 

These advantages are useful not only in 
planetary automatic transmissions, but 
also in CVT and DCT transmissions, which 
use multiple-plate wet clutches. Even in an 
80 year old technology, new ideas can pro¬ 
vide new functions and continue to meet 
the growing demands for fuel efficient mo¬ 
bility. 



Figure 13 Improved shifting clutch piston 

assembly compared to conventional 
piston assembly (red outline) 

Literature 


[1] Gott, P.: Changing Gears - The Development 
of the Automatic Transmission, SAE Interna¬ 
tional, 1991 

[2] SAE Transmission/Axle/Driveline Forum Com¬ 
mittee: Design Practices: Automatic Transmis¬ 
sions, 1994 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 








280 


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281 


sMJ - Innovative Solutions for Torque 
Converters Pave the Way into the Future 


Patrick Lindemann 
Markus Steinberger 
Thorsten Krause 


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282 


Preface 


The torque converter has been a stable 
choice as a launch device for automatic 
transmissions for several decades. Glob¬ 
al vehicle production in 2013 is estimated 
to be 83 million, with 43 % of the vehicles 
are being equipped with a torque con¬ 
verter [1]. In particular, the North Ameri¬ 
can and the Asian market show a high 
ratio of torque converters in new vehicles 
[2]. Additionally, the European market is 
experiencing a trend away from manual 
transmissions as some vehicles - espe¬ 
cially in the luxury and higher torque seg¬ 
ment - are offered only with planetary 
automatic transmissions and a torque 
converter. 

The choice of transmission type is 
largely driven by its impact on the pow¬ 
ertrain efficiency and comfort. With strict¬ 
er regulation on C0 2 emissions and the 
prospect of further tightening of emission 
regulation, the automotive industry has 
made designing for fuel efficiency a core 
goal, resulting in drag reduction, more de¬ 
fined combustion processes and in¬ 
creased electrification. Despite electrifica¬ 
tion, internal combustion engines are a 
core element of powertrain strategies and 
their optimization will drive improvements 
in the drivetrain. 

Supercharging of downsized engines is 
a primary path to achieving the required ef¬ 
ficiency improvements [8]. This technology 
has been used in motorsport applications 
for some time. However, the more wide¬ 
spread application of turbochargers in gas¬ 
oline engines required additional develop¬ 
ments such as direct injection, availability of 
durable turbochargers from TDI engines, 
and increased development pressure 
through the reduction of C0 2 targets. 

The increased specific power and 
torque compared to a naturally aspirated 


W 120 n 

c 



> 2013 2015 2020 


Year 

■ Manual transmissions 

■ Automatic transmissions 
without torque converter 

■ Automatic transmissions 
with torque converter 

Figure 1 Global vehicle production [1] 

engine allows the most often used driving 
conditions to be shifted to lower engine 
speeds. The reduced rotational speed mini¬ 
mizes losses caused by friction and im¬ 
proves the combustion efficiency. 

With the improvements in engine tech¬ 
nology, the driver does not have to accept a 
loss in performance for an increase in fuel 
efficiency. On the powertrain side however, 
the engine improvements change the 
boundary conditions for durability and com¬ 
fort. The reduced number of cylinders, to¬ 
gether with downspeeding and increased 
torque per cylinder leads to higher torsional 
vibrations. As a result, measures have to be 
taken to increase the durability of the drive- 
train. The impact on comfort in the form of 
seat vibration, boom and rattle noise can be 
even greater and has to be met with highly 
capable damper technology. Finally, the use 
of a turbocharger can introduce a degrada¬ 
tion of launch performance as a result of the 
turbo lag. In particular, small gasoline en¬ 
gines with 3 or 4 cylinders do not reach the 
peak torque until mid-operating speed. 

In this environment, the drivetrain re¬ 
quires an element that is able to reduce 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_20, © The Author(s) 2014 




Torque Converter 


20 


283 


torsional vibrations, provide the desired 
launch performance and achieve this 
with a minimum of added inertia and ax¬ 
ial space. For automatic transmissions, 
the torque converter is the launch device 
of choice (Figure 1). Despite develop¬ 
ment in areas such as double clutch or 
automated manual transmissions, the 
majority of automatic and continuously 
variable transmissions are equipped with 
torque converters. This success of the 
torque converter raises the question of 
its origin and development potential for 
the future. 

History of the 
torque converter 


Torque converters were not the initial 
choice for a launch device. Early trans¬ 
missions used friction clutches that were 
shifted manually or by means of centrifu¬ 
gal acceleration. With the advent of auto¬ 
matic transmissions and more refined 
passenger vehicles, the comfort and con¬ 
trollability aspect received more weight. 


This lead to the first mass production use 
of a fluid coupling in 1940 by GM. 

Torque converters and fluid couplings 
in ships 

The history of torque converters and fluid 
couplings did not start in the automotive 
industry where they would later reach a 
production volume of millions of pieces 
per year. Instead, it started in the maritime 
industry. Hermann Fottinger designed a 
converter and a fluid coupling in 1905 
(Figure 2) - which both have their specific 
advantages and disadvantages. The fluid 
coupling has a higher efficiency when the 
turbine speed is close to the impeller 
speed and the converter is able to provide 
a torque ratio to increase the output 
torque, which can be considered as an 
additional gear. After the initial patent, 
Fottinger also created several variations of 
his design that allowed him to change the 
torque transmission characteristic manu¬ 
ally [4]. 

At this point, neither the converter nor 
the fluid coupling were envisaged as a 
launch device but changed the ship’s pro¬ 
peller torque or decoupled the propeller 



Figure 2 Hermann Fottinger’s design for a fluid coupling (left) and a converter (right) [5], [6] 























284 


from the drive shaft to prevent the propa¬ 
gation of a torque spike. 

The first mention of a fluid coupling 
for passenger vehicles was made by Her¬ 
mann Rieseler in 1925 [3]. The device 
consisted of a multitude of turbines and 
impellers. It remained an idea and did not 
reach the production phase because of 
its complexity. 


Improved torque converter controls 

A simplification of the torque converter 
design was achieved in 1928 when H. Kluge, 
K. von Sanden and W. Spannhake (TRILOK 
Group) combined Fottinger’s designs into 
a subassembly. It was the first time that 
the converter’s stator was mounted on a 
one way clutch. This allowed the resulting 
design to provide the torque ratio of a 
converter and the high speed ratio effi¬ 
ciency of a fluid coupling. There was no 
need for additional controls to switch from 



Figure 3 Daimler Company’s fluid flywheel 
from 1928 [9] 


torque converter to coupling. The fluid’s 
angle of attack relative to the stator blades 
provides the signal for the stator to spin 
freely and turn the hydrodynamic circuit 
into a coupling. 

Even after the breakthrough of the 
TRILOK design, torque converters were 
not widely used in passenger vehicles. 
The first attempt to use a hydrodynamic 
clutch in a car was made in 1933 by the 
British Daimler Company Limited, which 
used a fluid flywheel in conjunction with a 
synchronized gearbox to avoid shift 
shocks. 


Mass production torque converter 

Although the use of a torque converter in 
automobiles was suggested in 1928, 
mass production did not start until 1940. 
The designs combined a torque convert¬ 
er with the planetary automatic transmis¬ 
sion and the fluid coupling was seen as 
an integral part of the transmission. The 
first mass production torque converter 
was introduced in the Oldsmobile Hydra- 
Matic as a safety, comfort and perfor¬ 
mance device. Since there was no shift 
lever and no clutch pedal, the driver had 
less interfaces with the car and could pay 
more attention to steering and braking. 
Comfort and performance are addressed 
by superior launch performance, re¬ 
duced vibrations and improved shift 
quality. 

Oldsmobile sold 10 million Hydra- 
Matic units [1], establishing planetary au¬ 
tomatic transmissions with a fluid cou¬ 
pling in the automotive industry. 

Following the stepwise introduction 
of features to the torque converter, the 
Packard Ultramatic introduced a lockup 
clutch in 1949. The so-called ‘Direct Drive’ 
hard locked the torque converter at high 
speeds and gave this transmission the 
fuel efficiency of a manual transmission. 



Torque Converter 


20 


285 



Oldsmobile Hydra-Matic with first large-scale production fluid coupling [2] 


Figure 4 

The introduction of the lockup clutch did 
not lead to immediate widespread adop¬ 
tion. Until the emphasis of efficiency that 
followed the oil crises in the 70’s, the loss¬ 
es in an unlocked torque converter did not 
warrant the additional components and 
controls of a lockup clutch. 


Torque converter dampers 

A torque converter advantage that has not 
been mentioned so far is that the hydrody¬ 
namic circuit does not transmit the engine 
vibrations to the transmission. This allows 
the engine to run at speeds which would 
otherwise lead to excessive drivetrain vi¬ 
brations. The undoubtedly beneficial intro¬ 
duction of a lockup clutch exposes the 
transmission to the previously avoided tor¬ 
sional vibrations. Therefore, the wide¬ 
spread adoption of lockup clutches in the 


1980s also required devices to control vi¬ 
brations in the drivetrain. This led to the 
advent of torque converter dampers. 

The first torque converter damper was 
built by LuK in 1983 for use in the Ford 
AOD torque converter. With the damper, 
the engine torsional vibrations were atten¬ 
uated to increase the driving conditions in 
which the lockup clutch can be fully locked. 
Initially, the lockup clutch was only en¬ 
gaged during cruising but with increasing 
demand for fuel efficiency, the duty cycle of 
the lockup clutch increased. This required 
increasingly complex dampers. The tur¬ 
bine damper that was introduced by LuK in 
1994 did not only use springs to prevent 
the propagation of torsional vibrations 
along the drivetrain. This damper locked 
the turbine mass that used to be on the in¬ 
put shaft after the damper to the engine 
side side, thus eliminating a vibration 
mode. 




























































































286 



Figure 5 LuK torque converter damper with 
centrifugal pendulum absorber 

Reaching a limit for spring volume and 
available inertia, torque converter damp¬ 
ers had to be based on a different prin¬ 
ciple to reduce the torsional vibrations of 
modern engines. As described above, 
improvements in engine efficiency direct¬ 
ly lead to the demand for improved 
dampers. 

Serving this demand, LuK introduced 
a torque converter damper in 2010 that 
used the centrifugal pendulum absorber 
principle. This allowed the lockup clutch 
to be engaged at engine speeds down to 
1,000 rpm, covering the majority of typical 
driving, further improving the powertrain’s 
efficiency. 


Drivers for torque 
converter development 


Following their long history, torque con¬ 
verters are an indispensable component 
of modern automatic transmissions. Their 
evolution leads to permanent adaptations 
and the current generation can be best 
understood with the guiding principle of 
value enhanced design. Current LuK 
torque converters are developed with the 
focus on the areas of performance in¬ 
crease, space reduction, cost reduction 
and higher efficiency. 

Using the value enhanced design phi¬ 
losophy, components and subassem¬ 
blies are strictly designed to meet the 
performance, space and cost targets. 
Supporting this goal, a modular design 
approach is used. This allows the torque 
converter to be customized to meet dif¬ 
ferent objectives with a maximum focus 
on cost. 

Performance improvements target the 
drivetrain efficiency and are achieved 
through reduction of the torque convert¬ 
er’s weight and inertia, improved damper 
performance and improved efficiency of 
the hydrodynamic circuit. 

Measures can be introduced gradu¬ 
ally, such as stress optimization to re¬ 
duce sheet metal thickness or directional 
design changes such as component re¬ 
placements. A comparison between a 
torque converter design from 2005 and 
its successor from 2013 shows that the 
weight was reduced by 2.1 kg while the 
maximum torque converter efficiency re¬ 
mained stable at 90 % and the damper 
windup increased by 31 %. 

The axial length of the torque con¬ 
verter more so than its radial size is a key 
element of the power train’s size. Crash 
test and aerodynamic requirements limit 



Torque Converter 


20 


287 




Figure 6 Torque converter designs from 2005 (left) and 2013 (right) 


the available space and are therefore di¬ 
rectly opposed to requirements for in¬ 
creased damper performance. In the ex¬ 
ample above, the axial distance between 
the stud plane and the torus was reduced 
by 2.9 mm while the damper windup was 
increased. These improvements were 
mainly achieved by reducing the torus 
width and improving the piston attach¬ 
ment method. 

Typical piston attachments for clutch¬ 
es with 2 friction surfaces require a rivet 


connection outside the piston area. In 
the 2005 design, this attachment is be¬ 
tween the cover and the piston drive 
plate. For the value enhanced design of 
2013 a Schaeffler riveting connection 
was developed. It establishes a direct 
leaf spring connection from the cover to 
the piston without the need for a piston 
drive plate. 

For the Schaeffler riveting process, the 
leaf springs are attached to the cover as¬ 
sembly. The domed rivets for the leaf 



Figure 7 Schaeffler riveting at piston connection 














288 



Figure 8 Modularity of the value enhanced torque converter designs from 2013 


spring-piston connection are already in 
place - in a pattern that matches the pis¬ 
ton holes. During the piston assembly, the 
piston is placed on the rivets and the rivet¬ 
ing tool pushes it towards the cover until 
the domed rivet heads make contact with 
the cover. At this point the rivet head can 
be formed, establishing a permanent con¬ 
nection between the leaf springs and the 
piston. The connection was developed so 
as not to overstrain the piston and to avoid 
contact between the domed rivet head 
and the cover in the application. 

By eliminating the inner drive plate, the 
piston could move closer towards the cover, 
creating more space for the damper. 

Reduction of the torque converter cost 
is a perpetual goal. With the value enhanced 
philosophy, methods for minimizing the cost 
were modularization and a reduction in the 
number of components. 

In the 2013 design, the number of 
components were reduced using the 
Schaeffler Riveting process piston rivet¬ 
ing and redesign of the input shaft inter¬ 
face. The 2005 design uses a riveted hub 
to connect the damper flange with the 
transmission input shaft. For the 2013 
value enhanced design, a flange was de¬ 
veloped that integrates the connection to 
the input shaft. With the spline formed 
from the flange, the space requirements 
were also reduced, leaving more space 
to optimize the remaining components. 


Efforts to focus more on modularity in the 
development resulted in reduced flexibility 
and consequently in the increased impor¬ 
tance of NVH and durability simulations. It 
had to be ensured that torque converters 
would meet customer requirements even 
with the limited possibility of modifica¬ 
tions. The resulting designs are shown in 
Figure 8. They make it possible to choose 
between different clutch capacity and 
gain, damper performance and engine/ 
transmission interfaces while retaining a 
large number of common components. 

Following the goal of improving pow¬ 
ertrain efficiency to reduce C0 2 emis¬ 
sions, torque converters are designed to 
support efficient driving conditions as 
well as increase the efficiency of the 
torque converter hydrodynamic circuit or 
torus. 

Between the value enhanced design 
and its predecessor, the peak torus effi¬ 
ciency was unchanged despite the width 
reduction and the corresponding weight 
reduction. This was achieved by improve¬ 
ments in blade design and production 
methods which allowed the blades to bet¬ 
ter guide the torus flow. 

However, a more significant impact on 
efficiency is achieved with the damper 
technology. Increased damper space 
through the elimination of components in 
the flange and piston connections permits 
the use of larger springs. Furthermore, ad- 








Torque Converter 


20 


289 


vanced damper technology such as cen¬ 
trifugal pendulum absorbers is used to im¬ 
prove the damper function beyond the 
ability of a coil spring damper in the same 
envelope. This allows efficiency improve¬ 
ments on a system level as the lockup 
clutch can be fully locked at lower speeds 
without compromising NVH. 

Centrifugal pendulum 
absorber for 
torque converters 


The main driver for advanced damper 
designs is the challenge of meeting C0 2 
emission requirements that are placed 
on modern combustion engines while 
maintaining or even improving NVH per¬ 
formance. The trend points towards 
smaller supercharged or turbocharged 
engines with fewer cylinders. In order to 
make the low speed range accessible to 
the driver, the torque at low rotational 
speed is increased. From simple physics 
it can be concluded that the reduction of 
both the number of cylinders and the 


drivable rotational speeds result in lower 
excitation frequencies. This leads to a 
considerable increase in the engine’s cy¬ 
clic irregularity and torsional vibrations 
(Figure 9) thus driving the development of 
damper technology. 

The possibilities offered by the new 
generation of engines require suitable au¬ 
tomatic transmissions and drivetrains. 

Transmissions must be adapted in line 
with improvements to engines in order to 
fully utilize the potential for reducing C0 2 
emissions. The shift and lockup speeds are 
reduced to a level that was previously im¬ 
possible because the available torque was 
not sufficient in older engines. The aim is to 
drive at 1,000 rpm and below, not only in the 
fuel consumption cycle at part load, but 
also at full load, while reducing the lockup 
clutch slip as much as possible. 

Torsional vibration damper with 
centrifugal pendulum absorber 

Even with increased installation space for 
conventional torsional dampers, the isola¬ 
tion of vibrations is often insufficient for 
modern downsized turbocharged en¬ 
gines. To achieve further improvements in 
torsional isolation, a speed adaptive ab- 


Fuel Consumption Map 


E 

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£ 



1,000 1,500 2,000 2,500 
Engine speed in rpm 



Fuel Consumption 


Torsional Fluctuations 






Specific fuel 

Fuel 

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consumption 

consumption 

0 

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in g/kWh 

in 1/100 km 

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3.39 ^ 

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— Drive resistance in 5 th gear 




(3) 


Time 


Constant power of 6 kW (@70 km/h) 
Drive resistance in 5 th gear 
with reduced gear ratio (-20 %) 


Figure 9 Development trend for downspeeding: drive at low speed for low fuel consumption 








290 


sorber is added to the secondary side of a 
torsional damper. A speed adaptive ab¬ 
sorber changes its absorbing frequency 
directly proportional to engine speed. The 
centrifugal pendulum absorber (CPA) in¬ 
corporates this function and thus can ab¬ 
sorb the engine’s main firing order opti¬ 
mally. 

In practice, a bifilar CPA with two sus¬ 
pension points is used to guide the pen¬ 
dulum mass. It follows a path that can be 
described as having the pendulum mass 
suspended by hinged parallel connectors. 
Each point of the CPA follows the same 
trajectory and it can be approximated suf¬ 
ficiently as a mathematical pendulum. The 
CPA movement is guided by rollers which 
roll on tracks defined by kidney shaped 


Figure 10 Torque converter with CPA 



— Standard TD 

- DTD with CPA 


Figure 11 Comparison of a standard torsion 
damper with a CPA damper 


cutouts in the 
mass and flange. 
The absorber or¬ 
der is determined 
by the form of the 
raceways and the 
rollers. 

The CPA com¬ 
bined with a 
suitable torsional 
vibration damper 
constitutes a sig¬ 
nificant improve¬ 
ment in the torsion 
damper’s isolation 
efficiency over stan¬ 
dard torsion damp¬ 
ers. Its superiority 
over other damper 
concepts has been 
proven with a cen¬ 
trifugal pendulum 
absorber attached 
to a double damp¬ 
er. This design has 
been in production 
since 2011. The 





Torque Converter 


20 


291 


40 n 


0 

■a 

!i 

30 - 

CO c 
-Q Z 

0 CO 

20 - 

O +3 
&§ 

■O it 

o ■s 

10 - 

1 

CM 



1,000 


— Gen. 1 



i i- 

1,500 2,000 

Speed in rpm 


2,500 



— Gen. 2 


Figure 12 Measurement of vibration amplitude at differential of a 1 st and 2 nd generation CPA 


comparison of this damper design with a 
standard turbine damper shows a signifi¬ 
cant improvement in isolation as shown 
in Figure 11. With this damper it was pos¬ 
sible to reach lockup speeds around 
1,000 rpm. 

2 nd generation centrifugal 
pendulum absorber for torque 
converters 

A further increase in the angular dis¬ 
placement and the weight of the paral¬ 
lel CPA would be necessary to improve 
the isolation capability further and 
achieve additional fuel efficiency im¬ 
provements by lowering the lockup 
speed. Torque converter space and 
mass limitations however set limits to 
the CPA growth. The 2 nd generation 
CPA aims to improve of the pendulum 
efficiency without increasing installa¬ 
tion space. It uses optimized movement 
of the pendulum mass by superimpos¬ 
ing a rotation onto the swinging motion 
of the pendulum mass, similar to a trap¬ 
ezoid (Figure 12). In addition, the 2 nd 


generation allows larger pendulum trav¬ 
el angles to improve the isolation per¬ 
formance. This maximizes efficiency in 
the given radial space, thus reducing 
the required pendulum width. Figure 12 
illustrates the isolation improvements at 
low rotational speeds. 

Vehicle measurements with the same 
installation space and damper configura¬ 
tion show a significant reduction in tor¬ 
sional vibrations of approximately 50 % 
from the 1 st to 2 nd generation CPA. 

3 rd generation centrifugal 
pendulum absorber for torque 
converters: 

track-optimized and spring-loaded 

With ongoing engine and drivetrain optimi¬ 
zations, dampers will have to provide even 
better isolation of torsional vibrations. Vibra¬ 
tion amplitude targets are likely to be re¬ 
duced while the engine’s torsional vibra¬ 
tions increase and the lockup clutch is 
engaged at lower engine speeds. Further¬ 
more, in the development of automatic 
transmission drivetrains the focus is on a 







292 



Figure 13 Design of a spring-loaded CPA 

higher ratio spread as well as the reduction 
of transmission damping and drag to im¬ 
prove the efficiency. In combination with the 
lightweight design of gearboxes and drive- 
trains, the reduced internal damping leads 
to structures that are more vibration sensi¬ 


tive. Further improvements in isolation will 
be necessary for these future transmission 
designs. 

With its superior isolation performance 
the 2 nd generation CPA provides an optimal 
base for additional improvements. Further 


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Cl C 
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1,000 1,500 2,000 2,500 

Speed in rpm 

— Gen. 1 — Gen. 3 

— Gen. 2 


Figure 14 Simulation of isolation with 1 st , 2 nd and 3 rd generation CPA 












Torque Converter 


20 


293 


travel-dependent track optimization can im¬ 
prove the pendulum efficiency. The im¬ 
provements can be focused on situations 
where significant pendulum travel occurs at 
low engine speed. 

The optimized movement of this pendu¬ 
lum masses enables the addition of a coil 
spring in between. 

The spring force stabilizes the pendulum 
motion during rotation at creep speed or 
transition events such as acceleration 
from standstill, further improving the NVH 
of the drivetrain. These additional springs 
also have a speed-dependent effect on 
the CPA order which can be minimized by 
travel-dependent optimization of pendu¬ 
lum tracks. 

Figure 14 compares the vibration ampli¬ 
tude at the differential in a drivetrain with a 
spring-loaded and track-optimized 3 rd gen¬ 
eration CPA with that of the 1 st and 2 nd gen¬ 
eration. 

CPA and 

cylinder deactivation 

A trend with immense implications on the 
damper design is cylinder deactivation. It 
improves the engine’s efficiency under par¬ 
tial load by requiring a higher specific load 
from the active cylinders [11]. The advan¬ 
tage over downsizing and downspeeding 
is that the high torque of the additional cyl¬ 
inders is still available when needed. 

First applications of a CPA with en¬ 
gines capable of cylinder deactivation 
from 8 to 4 cylinders are already avail¬ 
able in the market. Due to the reduction 
in main firing order while maintaining a 
high vibration amplitude, isolation in 
4-cylinder mode is more challenging 
than in 8-cylinder mode. Therefore the 
CPA is tuned to attenuate the 4-cylinder 
vibrations while the double damper is 
designed to isolate vibrations in 8-cylin¬ 
der mode. Besides an increase in sec¬ 


ondary inertia, the CPA does not affect 
the damper performance when the en¬ 
gine runs on all 8 cylinders. 

Engines with deactivation from 6 to 
3 or 4 to 2 cylinders may require the su¬ 
perior CPA isolation even if all cylinders 
are used. In this case, two centrifugal 
pendulum absorbers can be installed 
with one tuned to each order. This 
achieves perfect isolation in all driving 
conditions. 

Torsional vibration dampers with 2 nd 
generation CPA are a key technology 
which allows the lockup clutch to close 
when the engine speed is to idle. Continu¬ 
ous improvements to the CPA, as de¬ 
scribed above, provide additional isolation 
improvement. 



• Operating point 

— 8-cylinder WOT 

— 4-cylinder WOT 
— Driving resistance 

V8 engine 

OftO • Active in cylinder deactivation 
#OOt O Inactive in cylinder deactivation 


Figure 15 Effect of cylinder deactivation shown 
in a engine characteristic map [11] 










294 


CPA for CPA for 

4-cylinder mode 8- and 4- cylinder mode 



■ 4-cylinder CPA masses 

■ 8-cylinder CPA masses 


Figure 16 CPA configurations for an application with cylinder deactivation 
from 8 to 4. 


Torque converter 
innovation 


As discussed in the previous chapter, in¬ 
stallation space for the torque converter 
is becoming increasingly and smaller. 
Nevertheless the requirements for tor¬ 
sional isolation are becoming increas¬ 
ingly demanding. One approach to im¬ 
prove the damper performance is to 
increase the damper space by reducing 
the torus width and creating a squashed 
torus design. 

This conventional approach has been 
used before and its potential to increase 
the damper space is limited. To find even 
more damper space it was necessary to 
overcome the usual restrictions of torque 
converter design by incorporating the 
piston function into the turbine. In a tradi¬ 
tional design the turbine and lockup 
clutch are separate components. Both 


functions will be 
required in torque 
converters for the 
foreseeable future. 
By increasing the 
functions per¬ 
formed by the tur¬ 
bine, one of the 
large components 
of a typical torque 
converter design - 
the piston - can be 
eliminated. 

This requires 
the turbine design 
to change to with¬ 
stand the actuating 
pressure. Howev¬ 
er, in traditional 
torque converters, 
the piston takes up 
more space than 
its thickness. It requires clearance to the 
cover, clearance for piston deflection and 
more clearance to the damper to avoid 
contact during operation. The turbine 
thickness must to be increased with this 
design in order to withstand the lockup 
clutch actuating pressure. Ultimately, the 



Figure 17 Typical FWD torque converter 
design 







Torque Converter 


20 


295 



Figure 18 Initial iTC design 

space gained from the elimination of the 
piston as a separate component out¬ 
weighs the thickness increase of the tur¬ 
bine. 

Turbine and impeller now fulfill the lock¬ 
up function. During a vehicle launch the tur¬ 
bine is active, providing the required torque 
multiplication. At higher vehicle speeds, the 
lockup clutch can engage and create a 
torque path to bypass the hydrodynamic 
circuit. 

By integrating the piston into the tur¬ 
bine, the actuation direction is opposite 
to the typical torque converter design. 
Instead of actuating in the direction of the 
engine, the lockup clutch now actuates 
in the direction of the transmission. This 
means the lockup clutch apply and re¬ 
lease channels have to be controlled in 
reverse to a typical torque converter. 

The lockup clutch engagement con¬ 
trol distinguishes between 2 stages: 
open condition and lockup or slip condi¬ 
tion. In the open condition, oil enters the 
torque converter through the torus, build¬ 
ing a higher pressure on the turbine’s 


transmission side and thus causing the 
clutch to lift off. To avoid cavitation at low 
speed ratios, the torque converter charge 
pressure is elevated even with the lockup 
clutch disengaged. This leads to a cool¬ 
ing flow of 5 to 10 l/min. This is more than 
sufficient to balance the turbine thrust 
and ensure that there is an oil layer be¬ 
tween the iTC friction surfaces. Measure¬ 
ments have shown that this design low¬ 
ers the clutch drag in torque converter 
mode to almost zero. 

To engage the lockup clutch, oil is 
directed through the center of the input 
shaft and generates a pressure differ¬ 
ence on the turbine. A bushing on the 
inside diameter of the turbine seals the 
turbine off from the input shaft and en¬ 
sures that flow has to pass through the 
friction surface and build actuation pres¬ 
sure for the lockup clutch to engage 
towards the impeller. The lockup clutch 
engagement requires a more detailed ex¬ 
amination of the turbine thrust. 



Figure 19 ATF flow in lockup clutch 
disengagement 














































296 



Figure 20 Cause of turbine thrust 

iTC Measurements 

Turbine thrust is a result of different oil veloci¬ 
ties on both sides of the turbine. Inside the 
torus, the oil circles between the impeller and 
the turbine with a velocity that depends on the 
speed difference between both components. 
The velocity is at its highest when the vehicle 
is at a standstill (“staH”).On the engine side of 
the turbine, the oil velocity is considerably 


High air velocity, 
low pressure 



Low air velocity, 
high pressure 


lower. Only shear stress on the cover and im¬ 
peller cause the oil to have a different speed to 
the turbine. Using Bernoulli’s Principle, the 
axial force on the turbine is easily explained. 
High oil velocity causes the oil pressure to 
drop which creates a pressure difference be¬ 
tween both sides of the turbine and results in 
a force towards the higher velocity oil. It is the 
same physical principle that allows airplane 
wings to create lift. 



£ 

E 

z 

E 

o. 


§ 

I - 


o 

CO 


i 

* 


— Torque ratio, blocked iTC — K-factor, blocked iTC 

— Torque ratio, iTC — K-factor, iTC 


Figure 21 Characteristic measurements of iTC and blocked iTC 





















Torque Converter 


20 


297 



Time in s 

— Impeller speed 

— Turbine speed 

— Pressure 


Figure 22 Torque converter characteristic and lockup clutch engagement measurements 


At first glance the turbine thrust might 
seem to be an obstacle. Closer examina¬ 
tion however shows that the turbine 
thrust can be balanced with TC charge 
pressure to disengage the lockup clutch. 
This creates a hydrostatic support at the 
turbine’s friction surface and leads to low 
clutch drag. Engaging the lockup clutch 
creates a force that is oriented towards 
the transmission - the same direction as 
the turbine thrust. For the lockup clutch 
engagement, turbine thrust means that 
the gap between the lockup clutch fric¬ 
tion surfaces has a tendency to be re¬ 
duced every time there is relative speed 
between the turbine and the impeller. 
This is a preliminary stage to the closure 
of the the lockup clutch and allows a 
smooth engagement. 

Measurements of the iTC confirm the 
theoretical considerations about clutch 
drag and engagement. An iTC prototype 
was prepared for the comparison of TC 
characteristics between an iTC design 
and a typical torque converter. The first 
characteristics measurement was taken 
on the iTC prototype, for the second mea¬ 


surement a rolling bearing was placed be¬ 
tween the turbine and the stator to in¬ 
crease the clutch lift-off and prevent it 
from engaging. The measured character¬ 
istic curves are nearly identical and only 
differ in terms of measurement accuracy. 
This shows that the turbine bearing exhib¬ 
its the same resistance the friction sur¬ 
face lifted off by the oil flow, as previously 
described. 

The clutch controllability and engage¬ 
ment quality was checked in a dynamom¬ 
eter test where the turbine and impeller 
spin is at a given speed and the actuation 
pressure is increased gradually. After a 
successful engagement, the actuation 
pressure is reduced to determine when 
the clutch starts slipping again. 

As expected, the engagement starts 
at low actuation pressure values and does 
not lead to a torque spike. The pressure 
value at which the slip ended during the 
engagement phase and the pressure val¬ 
ue at which the slip started during the dis¬ 
engagement phase are also identified as 
very close. This small engagement hyster¬ 
esis verified the iTC slip controllability. 










298 


iTC Advantages 

As described above, the main motivator for 
the iTC design is in the reduction of the 
torque converter width. This goal coincides 
with the trend towards drivetrain weight re¬ 
duction and increasing numbers of gears. 
Besides the space advantage, the iTC pro¬ 
vides the following improvements over a 
typical torque converter. 

The iTC has less components when 
compared to a typical torque converter. 
All of the turbine’s axial force is directed 
through the friction surface and the tur¬ 
bine never makes contact with the stator, 
therefore it does not require a turbine 
bearing. The iTC also reduces the number 
of components by avoiding multiple con¬ 
nections for the damper torque input and 
by removing the piston. This reduces the 


design complexity but its major effect is a 
cost and weight reduction. The turbine 
has to be thickened to withstand the lock¬ 
up pressure but this is more than out¬ 
weighed by removing the components 
mentioned above. 

On the functional side, the iTC pro¬ 
vides very smooth engagements through 
an effect that is comparable to a preload¬ 
ed clutch but without the clutch drag 
caused by a preload mechanism. The 
self-engagement of the iTC is stronger at 
higher relative speeds between turbine 
and impeller because the turbine thrust 
force depends on the speed ratio. This re¬ 
sults in an increased engagement stability 
and reduces the risk of clutch shudder. As 
a result of the lockup clutch proximity to 
the torus, the clutch interface is stiffer 
than in a typical torque converter. The 



Figure 23 Modular damper and torus designs with the iTC layout 


























































































































Torque Converter 


20 


299 


brazed blades add a rib-like effect to the 
turbine and impeller shell and reduce the 
friction surface taper that would other¬ 
wise result from the apply and ballooning 
pressure. With less taper, the power input 
on the friction surface has a more equal 
distribution which avoids localized high 
temperatures and wear. 


iTC modularity 

Modularity was already the goal for the 
value enhanced design described above. 
With the iTC however, modularity receives 
an enabler that hasn’t been available be¬ 
fore. Typical torque converters require the 
damper to have connections to both the 
piston and the turbine. The iTC design has 
only a single input to the damper. This re¬ 
duces the damper complexity which cre¬ 
ates space for higher performing dampers 
and reduces the obstacles to variations of 
the damper. 

The iTC modularity doesn’t only ex¬ 
tend to damper variations as shown in 
Figure 25. Due to the relative simplicity of 
the impeller clutch design, its addition 
becomes modular and the result is more 
predictable. 


One way clutch 

Since 1928, one way clutches have been 
used to switch from converter to coupling 
mode. Typically, roller or sprag designs are 
used in torque converters. Driven by the 
width reduction that is achieved with the iTC 
design, the focus falls on reducing also the 
one way clutch width. 

Two concepts are available to reduce 
the one way clutch width. On a sprag or 
roller one way clutch, the contact surfaces 
are coaxial, therefore the capacity cannot 
be increased without increasing the width. 
In the wedge design, the contact surfaces 
are arranged circumferentially. The wedge 
one way clutch locks if the outer race pulls 
the wedge plate onto the inner race ramps. 
Utilizing the wedge principle, the normal 
force on the wedge plate increases with 
torque and creates a self-energizing friction 
coupling. In the freewheel direction, the 
wedge plate is prevented from spinning by 
the inner race shoulders. The wedge plate 
design is adjusted to only create drag in 
lockup direction to slide up the ramps. In 
the freewheel direction, the plates are de¬ 
signed to reduce the drag. Due to the ar¬ 
rangement of the contact surfaces, this de¬ 
sign has a considerable width reduction 
over typical roller one way clutches. 



Figure 24 Wedge (left), rocker (center) and slim cage one way clutch (right) 




300 


A second design with reduced width aims 
to eliminate non-functional features. Pedes¬ 
tals on the outer race of a typical roller one 
way clutch provide a reaction surface for the 
preload springs but their width is mainly 
driven by the production process. By using 
a plastic cage, the number of rollers can be 
permitting a reduction in their length. 

Yet another one way clutch develop¬ 
ment is the rocker design. This design aims 
to reduce the cost of the torque converter. 
Since the rocker location is fixed with re¬ 
spect to the stator, the outer race can be 
removed and the contact surfaces can be 
integrated into the stator’s aluminum body. 
The contact surface can be increased to 
avoid plastic deformation by the rockers. 
This design creates a challenge since it can 
only engage at discrete positions. It has 
been shown however, that a lash angle of 
2.4° is small enough to avoid any noticeable 
difference between the rocker and a roller 
one way clutch. With the elimination of a 
part as complex as the outer race of a roller 
one way clutch, this design reduces the 
cost of the torque converter. 


Multi Function iTC 

Unlike the typical lockup clutch, the iTC de¬ 
sign allows the implementation of an impel¬ 
ler clutch without large design changes. As 
the addition of a shell to the outside of the 
impeller cannot be avoided, the axial di¬ 
mensions are increased. However, the im¬ 
peller clutch can be integrated into the iTC’s 
turbine clutch and the pressure channels 
can also be adapted with ease. 

The reason for an impeller clutch traces 
back to the increase in the combustion en¬ 
gine’s efficiency. With a trend toward turbo¬ 
charged engines with a smaller number of 
cylinders, the maximum engine power is not 
available below 3,000 rpm. A highly boost¬ 
ed engine can struggle to reach full torque 
in a timely manner under heavy load at low 


speeds, commonly called turbo lag. This 
can cause poor vehicle launch performance 
and feel. To mitigate this effect and allow the 
engine to reach high speeds during a 
launch, the torque converter would typically 
be designed with a high K-factor to make it 
softer. The high K-factor has another advan¬ 
tage in that it results in reduced idle losses. 
As a detriment, high K-factor allows the en¬ 
gine to flare to higher speeds when the 
lockup clutch disengages for torque multi¬ 
plication. In this case, a lower K-factor 
would be desired to make the torque con¬ 
verter stiffen 

Both can be achieved if the impeller is not 
hard connected to the engine. The impeller 
clutch allows the engine to have a higher 
speed than the impeller which allows the en¬ 
gine to provide a higher torque faster, dimin¬ 
ishing turbo lag. This function can then best 
be described as a variable K-factor. The Multi 
Function iTC is designed to the lowest de¬ 
sired K-factor and the impeller clutch slip is 
used to make the system softer. 

The MFiTC controls require the 2 stan¬ 
dard pressure channels and an additional 
channel from the back of the impeller 






























Torque Converter 


20 


301 


clutch into the sump. Pressure between 
the turbine and impeller press the impeller 
against the transmission side cover and 
connect it to the engine. The channel 
through the center of the input shaft has to 
be closed to force the oil flow to exit 
through the impeller hub. The slip that is 
required to modify the torque converter 
stiffness can be controlled through the oil 
flow that enters between the turbine and 
the impeller. For the engine start and in 
idle, the channel through the transmission 
input shaft is opened which eliminates the 
pressure difference on the impeller and 
opens the clutch. The engine can therefore 
start without the drag that is typical of non- 
multi function torque converters. 

Lockup is achieved by providing oil flow 
through the input shaft to the outside of the 
turbine shell. This presses not only the tur¬ 
bine but also the impeller towards the outer 
shell and engages both clutches. Transi¬ 
tions between the lockup and torque con¬ 
verter modes require controlled backpres¬ 
sure on the exit channel to prevent the 
engaged clutch from slipping. 


Summary 


The torque converter as it is used in modern 
transmissions is the result of an evolution 
that spanned more than 70 years. Improve¬ 
ments of the internal combustion engine 
and the automatic transmission caused the 
torque converter design to be adapted but 
torque converter development itself has left 
a mark on the automatic transmission mar¬ 
ket as well. Smooth launch, torque multipli¬ 
cation and the attenuation of torsional vibra¬ 
tions set new directions for both engines 
and transmissions. The iTC is a continuation 
of this evolution and the space that was 
freed by integrating the piston into the tur¬ 


bine increases the degrees of freedom in 
the powertrain design. 

Similarly, torsional vibration dampers 
with 2 nd generation CPA will be a key tech¬ 
nology. It allows the powertrain to be used 
close to idle speed with the lockup clutch 
engaged for fuel efficiency. The CPA is inde¬ 
pendent of the torque converter design. It 
can be paired with a regular TC as well as 
an iTC to meet future demands of downsiz¬ 
ing and downspeeding. 


Literature 


[1] IHS Automotive Production Forecast, Aug 2013 

[2] Naunheimer, Bertsche, Lechner: Fahrzeugget- 
riebe. Berlin: Springer, 2007 

[3] Thompson E. A.: Fluid Coupling Rotor, 

US Patent 2,357,295. 1940 

[4] Rieseler H.: Flussigkeitswechsel- und -wende- 
getriebe, German Patent 435662, 1921 

[5] Fottinger H.: Improvements , GB Patent 
190906861, 1009 

[6] Fottinger H.: Flussigkeitsgetriebe mit einem 
Oder mehreren treibenden und einem Oder 
mehreren getriebenen Turbinenradern zur 
Arbeitsubertragung zwischen benachbarten 
Wellen, German Patent 221422, 1905 

[7] Fottinger H.: Flussigkeitsgetriebe zur Arbeits¬ 
ubertragung zwischen benachbarten Wellen 
mittels treibender und getriebener Rader, 
German Patent 238804, 1905 

[8] Janssen P, Govindswamy K.: Future Automatic 
Transmission Requirements, FEW, 2013 

[9] Autocar Handbook, 13th edition, 1935 

[10] Krause, T.; Kooy, A.; Kremer, E.: Torsional 
Dampers with 2 nd Generation Centrifugal 
Pendulum Absorber for Manual and Automatic 
Transmissions; VDI Congress Getriebe in 
Fahrzeugen, 2011 

[11] Golloch, R.: Downsizing bei Verbren- 
nungsmotoren, Springer, 2005 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 




302 



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303 


Hot & Cold 

Schaeffler’s thermal management 
for a C0 2 reduction of up to 4 % 


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304 


Introduction 


Improved and variable use of the heat 
flows in a vehicle is a requirement for fur¬ 
ther reducing emissions and fuel con¬ 
sumption and increasing the air condition¬ 
ing comfort in passenger cars. The 
integrated turbochargers (ITL) increasing¬ 
ly used in vehicles place increased re¬ 
quirements on cooling systems. ITLs re¬ 
quire a predictive cooling system if 
possible instead of a system, which reacts 
to different operating conditions. This re¬ 
quirement cannot be met with conven¬ 
tional thermostats because thermostats 
have a delayed reaction to energy input 
into the cooling system and also suffer 
from pressure losses. 

Innovative mechatronic components 
are required for making a predictive cal¬ 
culation of the cooling requirements from 
the engine load and speed. Schaeffler’s 
thermal management modules (TMM) are 
able to adjust the coolant flow to zero, for 
example, in order to achieve accelerated 
heating of the engine. At the same time, 


Around 1922 


From the 
engine 

Thermostat (controlled by bellows) 

Figure 1 Early thermostat controlled by 
bellows 


they are able to decouple thermal masses 
and thus dissipate quantities of energy to 
other components such as the engine oil, 
transmission oil, heater or traction battery 
via the residual mass. In contrast to con¬ 
ventional thermostats (Figure 1) TMMs are 
controlled using a load-based calculation 
model. This allows the integration of a 
large number of connected components 
as well as a narrow temperature range of 

_i_/_2 °Q 

The first multifunctional 
thermal management 
module in volume production 


The first volume produced engine to be 
equipped with a multifunctional thermostat 
is the Audi 1.8-liter TFSI engine (four-cylin¬ 
der in-line engine EA888Gen.3). This mod¬ 
ule was developed jointly by Audi and 
Schaeffler (Figure 2). 


Figure 2 Thermal management module in the 
Audi 1.8-liter R4 TFSI engine 




Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_21, © The Author(s) 2014 





Thermal Management 


21 


305 



Engine 

connections 


Fail-safe thermostat 


Water pump 
connection surface 


Engine oil heat 
exchanger connection 


Intermediate 
gear with locking 
function 


DC motor with 
reduction gear 


Radiator feed 


Radiator return line 


Rotary valve 2 for zero flow 


Sensor plate with 
integrated rotation 
angle sensor 


Rotary slide valve 1 


Connection for transmission 
oil heat exchanger and 
interior heating 


Figure 3 TMM design for the Audi 1.8-liter R4 TFSI engine 


In the warm-up phase of the engine, the ther¬ 
mal management module is able to com¬ 
pletely close the coolant inlet in the engine or 
set a minimum flow rate. If the engine is warm 
from operation, the coolant temperature can 
be adjusted quickly and fully variably to dif¬ 
ferent temperature levels depending on load 
requirements and external boundary condi¬ 
tions [1]. The thermal management module 
has two coupled rotary slide valves, which 
are operated by only one drive motor. One of 
these rotary slide valves is on the pressure 
side of the water pump and is designed for 
shutting off the coolant. The second rotary 
slide valve is used for distributing coolant on 
the intake side. The entire cooling circuit also 
has switching valves to enable the flow of 
coolant through the heater and the transmis¬ 
sion oil heat exchanger to be switched on 
and off in a targeted manner. 

Two rotary slide valves, which are coupled 
mechanically, control the flow of coolant in¬ 
side the rotary slide valve module. An electric 
motor drives rotary slide valve 1 via a worm 


gear with a high reduction ratio. Rotary slide 
valve 1 is, in turn, connected with rotary slide 
valve 2 via a lantern pinion. Rotary slide valve 1 
replaces the conventional wax thermostat 
and can very quickly and fully variably adjust 
the coolant temperature between 80 °C and 
110 °C depending on requirements. In addi¬ 
tion, rotary slide valve 1 switches the coolant 
return from the engine oil cooler (Figure 3). 
The coolant water is heated 30 % faster com¬ 
pared to the previous engine with a wax ther¬ 
mostat. The time required to reach the target 
oil temperature is reduced by around 50 %. 

The module essentially comprises high- 
performance plastics. The coolant-carrying 
parts comprise polyphenylene sulfide (PPS) 
with extreme levels of fill. This means the 
material is almost as strong as aluminum, is 
insensitive to media and has thermal stabil¬ 
ity. A search was made for an alternative for 
polytetrafluorethylene (PTFE) during the de¬ 
sign of the seal materials because the plas¬ 
tic known under the trade name Teflon is 
expensive and has a tendency to creep un- 


306 



Figure 4 Rotary slide valve module for full 

electronic control of heat flows in the 
engine and vehicle 

der the influence of temperature. An alterna¬ 
tive material was developed on the basis of 
polyvinylidene fluoride (PVDF). 

The materials used in the gears were 
developed by the Schaeffler Group in- 
house. Particular attention was paid to the 
selection of fiber materials. The gears oper¬ 
ate under dry running conditions because 
lubricants would be ejected over the oper¬ 
ating life and would no longer be effective. 
The seals are not pressure-dependent and 
are able to compensate for angular offsets 
due to the integration of a pretensioning 
spring instead of an O-ring (Figure 4). 

High-precision manufacturing of the ro¬ 
tary slide valves and sealing assemblies al¬ 
lows leak rates of less than 1 liter per hour. 
An auxiliary thermostat ensures protection 
against failure. This means a return spring is 
not required on the drive motor and the en¬ 
ergy consumption of the TMM is minimized. 



Figure 5 Compact module with two to three 
regulated outlets 

Compact to comprehensive 
Schaeffler solutions 


Schaeffler’s thermal management mod¬ 
ules can have different designs depend¬ 
ing on customer requirements and the 


Radiator feed 



(Turbo) 

Figure 6 Multifunctional module with 
integrated split cooling 







Thermal Management 


21 


307 


available space. A particularly compact 
solution, for example, offers up to three 
regulated channels and fits into the design 
envelope of conventional thermostat 
housings (Figure 5). The integration of a 
temperature sensor is also possible. Stan¬ 
dardized actuators also allow efficient de¬ 
velopment. The use of technologies and 
materials validated in volume production 
is an excellent basis for a robust new de¬ 
velopment. 

The development of a multifunctional 
module with separate circuits for the en¬ 
gine block and cylinder head (split cooling) 
is going in another direction. It has up to 
five controlled channels as well as a feed 
and flow control system. A high level of in¬ 
tegration is one of the advantages of the 
multifunctional module. In addition, only 
one interface is required to the control unit 
(Figure 6). 

Maintaining the engine oil 
temperature 


Plate-type heat exchangers of stacked- 
disk design are frequently used for indirect 
cooling with coolant. The plates are pro¬ 
vided with turbulence inserts to improve 
the heat transfer between the media. The 
design of a plate-type heat exchanger 
comprises a number of corrugated plates. 
Chambers are created between the 
plates, in which the heated fluid and the 
fluid to be heated can flow. A chamber 
with heated fluid is followed by a counter¬ 
flow of the fluid to be heated separated by 
a plate (Figure 7). 

The use of an oil/coolant heat ex¬ 
changer has two advantages: The cool¬ 
ant, which heats more rapidly than the 
engine oil during cold starts, can be used 
to ensure the oil reaches its target tem- 



Figure 7 Design of a plate-type heat 
exchanger 

perature more quickly. It also assists heat¬ 
ing of the pistons, which quickly reduces 
the piston clearance. This results in an 
improved level of particle emissions. The 
oil can also reach higher temperatures 
during engine operation. The oil can dis¬ 
sipate this heat to the coolant via the heat 
exchanger. The ability to maintain the oil 
temperature within narrow limits has an 
advantageous effect on the stress placed 
on the lubricant. 


Model verification 

The warm-up behavior of the oil at differ¬ 
ent water temperature levels was verified 
experimentally on an oil cooler at Schaef¬ 
fler. The oil temperature is 20 °C on a spe¬ 
cial test setup (Figure 8) at the start of the 
test. The water inlet temperature is to be 
held constantly at 40, 60, 80 or 100 °C. 
Four measurements with different oil 
pump speeds, oil flows and water flows 
are carried out for every coolant tempera¬ 
ture. The measurement results are shown 
in Figure 9 as an example 

In general, the measurements show 
that a higher level of friction reduction can 
be achieved if the water starts to flow 
through the oil cooler earlier rather than 




308 



Restrictor 


Heat exchanger (test part) 

Ik I 


Coolant feed 


Bypass 


Coolant return f - 


Oil 


pump 


t exchanger for cooling 
I between measurements 


Temperature and 

pressure measurement points: 

1. Heat exchanger inlet (coolant) 

2. Heat exchanger outlet (coolant) 

3. Heat exchanger inlet (oil) 

4. Heat exchanger outlet (oil) 


Figure 8 Test setup for determining the warm-up behavior of the oil 



— Parameter 1 — Parameter 3 

— Parameter 2 — Parameter 4 


Oil pump Oil Coolant 
speed flow rate flow rate 

in rpm in l/min in l/min 


Parameter 1 

640 

8 

Parameter 2 

1,290 

16 

Parameter 3 

1,950 

24 

Parameter 4 

2,470 

32 


4 

8 

12 

16 


Figure 9 Oil outlet temperature over the measur¬ 
ing period at a coolant temperature of 
60 °C for different flow rates 


later. The coolant should be used to heat 
the oil as quickly as possible in order to 
achieve a reduction in C0 2 and fuel con¬ 
sumption. The oil cooler must be taken into 
consideration in the design of the oil circuit 
because the heat exchanger is a restriction 
at low temperatures. 

The NEDC is started with a cold en¬ 
gine. This means that the oil is in a highly 
viscous state and can only flow through 
the heat exchanger with difficulty. If heat¬ 
ed coolant does not flow through the oil/ 
water heat exchanger (OWHE) from the 
start, it is also not advisable to direct the 
oil via the OWHE. The cooler can also be 
bypassed until the oil is within a tempera¬ 
ture range, in which it must be cooled. 
This means the heat in the oil is not dissi¬ 
pated to the surroundings via the cooler 
or to the coolant via the heat exchanger. 
In both cases, this causes the heat to ac¬ 
cumulate in the oil circuit , which, in turn, 
means that the operating temperature can 
be reached more quickly. The installation 













Thermal Management 


21 


309 


of a control valve in the oil circuit would 
also be a possible solution. This would al¬ 
low rapid and requirement-based control 
of the oil. 


NEDC 

When determining the standardized fuel 
consumption it must be taken into consid¬ 
eration that consumption is strongly influ¬ 
enced by the driving style of the driver. 
Today, standardized driving cycles are run 
in order to achieve comparable values. A 
synthetic speed curve, the New European 
Driving Cycle, was defined for Europe. 
Phases of constant acceleration, constant 
speed, constant deceleration and idling 
phases at zero speed are run during this 
cycle. The shifting points for vehicles were 
also defined in the NEDC because engine 
speed also has a large influence on fuel 
consumption. The NEDC is a sequence of 
five cycles, four identical urban cycles 
with a maximum speed of 50 km/h and an 
extra urban cycle with a maximum speed 
of 120 km/h. Figure 10 shows how the 
coolant and oil temperature affect fuel 
consumption. 

Figure 11 also shows the speed curve in 
relation to time. It can be seen that the en- 



Oil and coolant temperature in °C 

- Toil = 60 °C 
— Tcoolant = 110 °C 

Figure 10 Influence of coolant and oil 

temperature on fuel consumption 



Time in s 

Figure 11 Speed curve of the NEDC 

gine is initially subjected to low loads. It is all 
the more important not to lose any energy 
and to quickly bring the motor up to tem¬ 
perature in this early phase. 

Maintaining the temperature 
in the interior 


After cold starting a passenger car, opti¬ 
mum air conditioning should be achieved 
in the passenger compartment as quick¬ 
ly as possible. A defined interior air tem¬ 
perature is recommended for comfort¬ 
able air conditioning of the interior. The 
fed und dissipated heat flows must be 
designed and adjusted to achieve this 
temperature. 

A comfortable mean air temperature in 
the closed rooms of buildings is approxi¬ 
mately 22 °C according to DIN 1946-2. 
The mean interior air temperature in a 
passenger car is calculated from the arith¬ 
metic mean of the mean air temperature in 
the footwell and the mean air temperature 
in the ceiling area. The mean interior air 
temperature required for ensuring comfort 
in the interiors of passenger cars is not 
constant. It is dependent on the physical, 
physiological and intermediate influencing 
factors (Table 1). 








310 


Factors influencing thermal comfort 

Physical 

Physiological 

Intermediate factors 

• Enclosing surfaces 

• Solar radiation 

• Air temperature 

• Air flow 

• Humidity 

• Activity 

• Status 

• Skin moisture level 

• Clothing 

• Number of occupants 


Table 1 Factors influencing thermal comfort 

The interior air temperature perceived as 
comfortable depends strongly on the am¬ 
bient air temperature (Figure 12). If the am¬ 
bient air temperature is 20 °C, the interior 
air temperature perceived as comfortable 
is 22 °C. The interior temperature consid¬ 
ered to be comfortable is higher than 
22°C at lower ambient air temperatures. 
This higher interior temperature is re¬ 
quired, for example, in order to compen¬ 
sate for the thermal radiation dissipated to 
enclosing surfaces. The optimum temper¬ 
ature at high ambient air temperatures 
is also over 22 °C because, for example, 
lighter clothing is worn. 

Influence of different shut-off 
systems on comfort 

As part of the measurements for a master 
thesis supervised by Schaeffler, tests were 
carried out to determine which strategy 
heats the engine and coolant more quickly 
than the standard strategy and what influ¬ 
ence the different strategies have on heat¬ 
ing the passenger compartment. A pas¬ 
senger car was driven on a rolling test 
stand under the specified loads for the 
measurements. 

Measurements were carried out on 
the engine with different strategies for the 
coolant pump. These included: 

- the standard coolant pump, which is 

permanently connected, 



Figure 12 Mean air temperature in a vehicle 
interior depending on the ambient 
air temperature 


- a switchable coolant pump, which is 
controlled by the vehicle control sys¬ 
tem in accordance with the cold-start 
strategy of the automobile manufac¬ 
turer, 

- a coolant pump which is disconnected 
in the warm-up phase and is only con¬ 
nected when a defined coolant tem¬ 
perature is reached and 

- a shut-off element, which prevents the 
thermo-syphon effect 

In order to assess the different coolant 
pump strategies, the engine was initially 
operated with the coolant pump discon¬ 
nected (thermo-syphon effect permitted) 
and subsequently with the coolant circuit 
shut off (thermo-siphon effect prevent¬ 
ed). Table 2 shows details of the test sce¬ 
narios. 








Thermal Management 


21 


311 




Measuring 

time 

Load 

5 kW** 

1 

Standard (cyclical CP) 

15 min 

X 

2 

Standard (SE opens in cycle times as in 1) 

15 min 

X 

3 

Coolant pump switched on after motor start 

15 min 

X 

4 

Coolant pump disconnected (CP switched on 
after coolant temperature reaches 50 °C) 

15 min 

X 

5 

Coolant pump disconnected (CP switched 
on after coolant temperature reaches 80 °C) 

15 min 

X 

6 

Shut-off element closed (SE opens after same time as in 4) 

15 min 

X 

7 

Shut-off element closed (SE opens after same time as in 5) 

15 min 

X 


** 5 kW at 2,000 rpm (crankshaft) 

Table 2 Test scenarios 


The scope of the measurements included: 

- the coolant temperature before and 
after the heater core (HC), 

- the temperature of the coolant after 
the shut-off or after the coolant 
pump, 


- the temperature of the air after the 
heater core and 

- the air temperature in the interior. 

Figure 13 shows a diagram of the test setup. 


> Crankshaft speed 

> Coolant temperature 

> Pedal travel 

> Throttle valve 

> Lambda 

> Temperature of intake air 

> Temperature of engine oil 


Measurement on the roller: 

• Torque 

• Speed 


Temperature on the center 
console ventilation system 




Air temperature before 
and after heater core 
and air speed 


Temperature of 
coolant after shut-off/ 
element and water 
pump 


Coolant temperature 
before and after heater 
core and coolant flow rate 


CP - water pump 
SE - shut-off element 


^Ambient 

temperature 


Water pump speed 


Figure 13 Diagram of the test setup 












































312 



Figure 14 Measurement points in the test setup 


The measurement point for the interior tem¬ 
perature was at the height of the head re¬ 
straint on the passenger side (Figure 14 
left). The measurement of the air speed after 
the heater core is carried out after the 
fan (Figure 14 bottom right). Figure 14 
shows the measurement point before 



Time in s 


— CP switched on after motor start 
CP with SE switched on after 124 s 
CP with SE switched on after 215 s 

— Cyclical CP with SE 


Figure 15 Coolant temperatures for different 
coolant pump strategies 


the coolant pump at 
top right. 

Figure 15 shows 
the coolant temper¬ 
ature curve at the 
measurement point 
before the coolant 
pump depending 
on the switching 
strategy. This curve 
progression is simi¬ 
lar to the coolant 
temperature curve 
after the heater 
core. A temperature 
increase during the 
“stationary coolant” 
phase can be seen. 
For the curves with 
the strategy “cycli¬ 
cal coolant pump” 
and the coolant pump that is connected 
above a coolant temperature of 50 °C, there 
is only a slight effect before connecting the 
coolant pump. For the strategy, in which the 
coolant pump is connected above a coolant 
temperature of 80 °C, a significant increase 
of the coolant temperature is noticeable be¬ 
fore the coolant pump is connected. The 
increase for the measurements with a shut¬ 
off is significantly larger than for the mea¬ 
surements without a shut-off. 

For the measurements at the measure¬ 
ment point before the coolant pump, heat 
transfer is only possible by means of ther¬ 
mal conduction in the coolant if the coolant 
pump is disconnected. The heat transfer 
for the measurement without a shut-off ele¬ 
ment continues in the coolant pipe. The 
coolant in the measurements with a shut¬ 
off element can only be heated as far as the 
shut-off element. The coolant is continu¬ 
ously heated at the measurement point be¬ 
fore the engine inlet without any heat dis¬ 
sipation due to the shut-off in the pipe. 
Therefore, the coolant temperature mea¬ 
sured at this position is higher than the cool- 











Thermal Management 


21 


313 



Time in s 

— CP switched on after motor start 
CP with SE switched on after 124 s 
CP with SE switched on after 215 s 
— Cyclical CP with SE 

Figure 16 Air temperatures with different 

coolant pump strategies after the 
heater core 

ant temperature in the test without a shut¬ 
off element. 

After the pump is connected at 124 sec¬ 
onds and 215 seconds, there is initially a 
short drop in temperature, because cooler 
coolant is fed from the heater core and 
pipes to the measurement point. This is fol¬ 
lowed by a significant increase in tempera¬ 
ture due to the warm coolant, which was 
heated in the motor and now reaches the 
measurement point. 

With the cyclical coolant pump strate¬ 
gy, temperature differences occur with de¬ 
lays after the coolant pump is connected. 
Initially, the warm coolant is moved through 
the circuit by the pump, until it reaches 
the engine inlet. The coolant temperature 
drops only slightly during the periods when 
the coolant pump is disconnected. The 
coolant only loses heat slowly because 
heat continues to reach the measurement 
point from the engine due to heat conduc¬ 
tion in the coolant. The curves with the cy¬ 
cled coolant pump strategy catch up the 
other curves after the pump has been con¬ 
nected four times. The curves for all strate¬ 
gies have the same progression after per¬ 
manent connection of the coolant pump, 


whereby the curve with the cyclical cool¬ 
ant pump is slightly higher. The strategy 
for the cyclical coolant pump has the lon¬ 
gest coolant pump disconnection times. 
This means a very small quantity of heat is 
dissipated from the engine, which is why 
the engine and coolant are heated mini¬ 
mally faster. 

The heating characteristics of the air af¬ 
ter it exits the heater core due to different 
heating strategies can be seen in Figure 16. 
This shows that the air temperature curve 
with a coolant pump that is permanently 
connected cannot be improved by any of 
the air temperature curves of the other 
strategies. Above 550 seconds, the curves 
of all the coolant pump strategies lie on top 
of each other. Heating up the air requires 
different periods of time depending on the 
strategy. The earlier the heat is transferred 
to the HC, the earlier the air will be heated. 
The greater the quantity of heat transferred 
to the HC, the faster the air will be heated. 

The measured interior temperature de¬ 
pending on different heating strategies is 
plotted in Figure 17. These curves follow the 
air temperature curve after the heater core, 



Time in s 

— CP switched on after motor start 
CP with SE switched on after 124 s 
CP with SE switched on after 215 s 
— Cyclical CP with SE 

Figure 17 Comparison of the air temperature in 
the interior. 







314 



Figure 18 Modified naturally aspirated engine with a TMM 


however, they have 
a different gradient. 

The air, which exits 
the heater core, 
mixes with the air 
in the passenger 
compartment after 
it leaves the noz¬ 
zles. A change of 
temperature there¬ 
fore requires a lon¬ 
ger period due to 
the large quantity of 
air in the vehicle in¬ 
terior. The strategy, 
with which the inte¬ 
rior is heated most 
quickly, is the strat¬ 
egy with the coolant 
pump permanently 
connected. The less 
the coolant pump is 
connected in the 
heating phase, the 
more slowly the in¬ 
terior will be heated. 

The measure¬ 
ments carried out 
to determine cool¬ 
ant and air temperatures at different 
measuring points in the warm-up phase 
of the engine show that the strategy with 
a coolant pump that is permanently con¬ 
nected is still the most appropriate for 
heating the interior of a passenger car as 
quickly as possible. Other coolant pump 
strategies with disconnected phases do 
show a faster heating phase after the 
coolant pump is connected, but are not 
an improvement on the curve with the 
coolant pump that is permanently con¬ 
nected. These results show that a switch 
must be made to the strategy with a 
coolant pump that is permanently con¬ 
nected as soon as a passenger operates 
the heater - customer satisfaction is the 
highest priority. 


Cold-start strategies 


Schaeffler modified a conventional naturally 
aspirated engine and replaced the thermo¬ 
stat control system with a thermal manage¬ 
ment module in order to verify the effects of 
a TMM on cold starting (Figure 18). 

The system is able to distribute or shut 
off the coolant due to the combination of a 
coolant pump and two valves instead of a 
coolant pump and a thermostat. The shut¬ 
off function is particularly attractive for the 
cold-start strategy. This has a significant in¬ 
fluence on the fuel consumption figures in 
the NEDC. Schaeffler tested two different 
operating strategies for the TMM with this 





Thermal Management 


21 


315 



— Average value of cylinder head outlet coolant temperature with base engine 

— Average value of cylinder head outlet coolant temperature with TMM 

■ Cylinder head outlet coolant temperature in different TMM control modes 



Figure 19 Load-based temperature control on a modified naturally aspirated engine 


setup: Zero flow for quick heating and load- 
based temperature variations (part load 
110 °C, full load 85 °C) (Figure 19). 

The temperature curve in Figure 19 does 
not correspond with the real values because 
motion of the coolant and a change in coolant 
temperature do not occur until after 100 sec¬ 
onds. The temperature can subsequently be 


20 % Emission Benefit 


15 % - 
10 %- 
5% - 


0 % -| 

1 



o 

o 

NO x 

co 2 'hc+nOx' ch 4 1 


HC 

CO 

NO x 

C0 2 HC+NOx CH 4 

Benefit 8 % 

6% 

18% 

1 % 13 % 8 % 


maintained at a constant level +/- 2 °C using a 
simple calculation model. This system can re¬ 
act immediately to the driver’s load require¬ 
ments and significantly reduce the tempera¬ 
ture. The zero flow strategy alone resulted in a 
reduction in fuel consumption of 1.2 %. In ad¬ 
dition, significant reductions in secondary ex¬ 
haust gases such as HC, NO x or CH 4 were 
achieved by means of the higher exhaust gas 
temperature and operation of the catalytic 
converter at an earlier stage (Figure 20). Even 
though these results are impressive at first 
glance, the full potential can only be realized in 
close collaboration with heat physicists from 
automobile manufacturers. 


Gasoline Technology Car 


Figure 20 Reduction in secondary exhaust 

gases due to operation of the catalytic 
converter at an earlier stage 


Schaeffler has built a concept vehicle called 
the Gasoline Technology Car (GTC) using 
advanced components on the basis of a 














Temperature in 


316 


Original 


Modification with TMM 



Thermostats replaced by a module: ► higher functionality 

► quick response 

► reduced assembly costs 


EGR 

cooler 


Figure 21 Design of the GTC with advanced Schaeffler components 



Mechanical efficiency 
— Cylinder liner center 
Engine oil 

Thermal Efficiency 
— Exhaust valve bridge 
Cylinder liner top 

-- Original 
— Modified 

Faster warm-up offers potential for 
increased efficiency and passenger comfort 

Figure 22 Faster heating for increased 
efficiency and comfort 


Ford Focus with a 1.0 liter Fox engine. The 
original engine has two thermostats. One of 
the thermostats is used for block control, 
the second operates the radiator. These 
two thermostats were replaced in the GTC 
by a TMM, which bundles the functions and 
is also able to switch the oil cooler on and 
off (Figure 21). 

In contrast to the original engine, it is 
possible to realize a zero flow due to the 
integration of the TMM. The required mod¬ 
ule is so compact that it can be fitted in the 
existing design envelope of the main ther¬ 
mostat. The results of the first tests show a 
significant increase in the thermal and me¬ 
chanical efficiency (Figure 22). Also in the 
GTC, the significantly faster increase in the 
temperature of the exhaust gas leads to a 
more rapid response of the catalytic con¬ 
verter and reduced secondary exhaust 
gases. 

The heating of the oil is slower despite 
the steeper heating curve because there is 
no flow through the oil/water heat exchang¬ 
er in the initial phase. The objective is to 
achieve the optimum switching point be- 
























































































Thermal Management 


21 


317 


tween thermal and mechanical efficiency. 
This depends on both the engine architec¬ 
ture and the parameters of the engine oil 
used. The closer the collaboration with the 
automobile manufacturer, the more efficient 
the realization of potential will be. 

Even though the presented results are 
only an approximate model of the first tests, 
these measurements show that the differ¬ 
ence in temperature gradients is significant 
and the system offers an additional degree 
of freedom for engine design. Fine calibra¬ 
tion of the engine control unit at Continental 
will result in a significant smoothing of the 
curves. 

Design of the cooling 
circuit for conventional 
powertrains 


A multi-stage design is recommended for 
future cooling circuits of conventional pow¬ 
ertrains on the basis of the findings present¬ 
ed in this article. A zero flow phase should 
initially ensure that the interior of the engine 
is heated in order to enable a rapid reaction 
of the catalytic converter. A bypass with an 
integrated oil cooler or heater offers the 
required flexibility. The decoupling of the 
OWHE from the bypass with variable inlet 
control allows an additional degree of free¬ 
dom. 

After dealing with the engine control, the 
conditioning of the transmission must be 
taken into consideration. The requirements 
for transmissions will also increase due to 
the increasing number of gear ratios and 
bearing positions. There is still a large 
potential for increasing the efficiency of 
hydraulically actuated transmissions. Initial 
tests have already been carried out on dou¬ 
ble clutch transmissions. 


The radiator’s control system should de¬ 
couple as much thermal mass as possible. 
This means the focus can be placed on ef¬ 
ficiency with normal or warm ambient air 
and on comfort with cold ambient air. The 
use of finely regulated systems instead of 
conventional on/off switches offers signifi¬ 
cant potential. 


Outlook 


Mechatronic systems for coolant control are 
a trend with the potential to optimize the fuel 
consumption and emission characteristics 
of vehicles and at the same time increase 
the air conditioning comfort in vehicle interi¬ 
ors. This results in a wide range of design 
options for specific designs depending on 
the configuration of the powertrain. As a 
partner with a holistic approach in develop¬ 
ment and production, Schaeffler offers con¬ 
cepts with a wide range of options. 


Literature 


[1] Eiser, A.; Doerr, J.; Jung, M.; Adam, S.: Der 
neue 1,8-l-TFSI-Motor von Audi. MTZ 6/2011, 
pp. 466-474 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 





318 


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319 


Hidden Potential 

between the Crankshaft and Valves 

From the optimization of components to the optimum valve train 


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320 


Introduction The entire system 


The design of tinning drives in modem inter¬ 
nal combustion engines is affected by a 
large number of parameters, which influ¬ 
ence each other. Engine development usu¬ 
ally starts from the top down, i.e. with the 
control of the charge cycle. This approach 
carries the risk that target conflicts are 
caused by systems that are taken into con¬ 
sideration at a later date, for example, if ad¬ 
justment of the timing drive to suit the spe¬ 
cial design features of the crankshaft results 
in negative effects for the camshaft phasing 
system. 

Experience shows that the development 
process currently used cannot realize all the 
available optimization potential. The chal¬ 
lenge is to define all subsystems in detail 
at the very beginning of development 
so that the optimum 
is achieved at sys¬ 
tem level. This type 
of demanding de¬ 
velopment work 
can only be man¬ 
aged if all depart¬ 
ments involved - 
both from the 
automobile manu¬ 
facturer and the 
supplier - collabo¬ 
rate even more in¬ 
tensively. The orga¬ 
nization must allow 
component devel¬ 
opment experts to 
use the available in- 
house systems ex¬ 
pertise at any time. 


Definition 

The entire timing drive system includes the 
camshaft drive itself with a chain or belt, the 
camshaft phasing unit and the different de¬ 
signs of valve actuation (Figure 1). This may 
also include a spur gear - if only one of the 
two overhead camshafts are directly driven 
- as well as the connection to the crank¬ 
shaft. 


Higher-level development targets 

The target of designing an optimum valve 
train follows the normal premises of engine 
development: The priority is to safeguard 



Figure 1 The entire timing drive system: Camshaft drive with a chain as in 
this photo or with a belt, camshaft phasing unit and different 
designs of valve actuation 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_22, © The Author(s) 2014 





Timing Drive Systems 


22 


321 


the function of individual systems through¬ 
out the operating life. An additional focus 
of development is placed on minimizing 
friction in the entire system. Almost as 
much importance is now attached to this 
task as achieving the functional targets. Ef¬ 
forts to minimize noise emissions and, in 
particular, the moving mass have also be¬ 
come established development tasks. To¬ 
day’s preferred designs are derived from 
these requirements. The procedures used 
for developing valve train systems for many 
years have led to a great deal of expertise 
in this area. This means it is possible to 
draw up a design proposal for the valve 
train during the concept phase of the en¬ 
gine. And this is precisely where the dilem¬ 
ma starts: What might be the optimum 
valve train design can have grave disad¬ 
vantages for the following systems and 
vice versa. The typical development pro¬ 
cess of a valve train describes this chal¬ 
lenge. 

Initial situation in 
development 


The main approach in engine development 
is to define the charge cycle processes. 
This ensures that the most important re¬ 
quirements in the requirements specifica¬ 
tion such as power, torque and exhaust 
gas values are met. The first individual sys¬ 
tem to be considered is therefore the valve 
train. In this development stage, it is impor¬ 
tant to select the optimum concept for the 
relevant application from numerous possi¬ 
ble variants. During the second stage, a 
decision is made on whether and to what 
extent phase position adjustment of the 
camshafts is required. It is worth mention¬ 
ing that responsibility is frequently trans¬ 
ferred to another department during this 


phase. Different employees design the tim¬ 
ing drive during the third stage and inte¬ 
grate it into the engine while taking the first 
of the above mentioned restrictions into 
consideration. Work is usually carried out 
on a phased basis. With this approach, 
each individual development department 
must adhere strictly to the requirements, 
regardless of the knowledge that the re¬ 
quirements for the adjacent systems can 
force them to use a design for their own 
component, which falls short of the opti¬ 
mum in certain circumstances. Frequently, 
only marginal adjustments of the require¬ 
ments for adjacent functions would be suf¬ 
ficient to open up new options for the sys¬ 
tem under development. 

The assessment of this process 
shows both sides of the coin: On the one 
hand, extensive expertise is created up 
to the subsystem and component levels. 
On the other hand, it enables the realiza¬ 
tion of a global optimum using, in princi¬ 
ple, unnecessary iterations. This dilem¬ 
ma can be solved if the automobile 
manufacturers involve their suppliers in 
development at the beginning of the con¬ 
cept phase in order to combine the ex¬ 
pertise at component and system level of 
all those involved. A comprehensive con¬ 
sultation is most effective during this 
early phase. Some examples based on 
practical experience illustrate the poten¬ 
tial for improvement. 

Approaches for optimizing 
friction 


With mechanical or hydraulic tappets, roll¬ 
er tappets, roller followers or finger follow¬ 
ers, development engineers have a wide 
variety of reliable technologies that have 
been tested worldwide. Well designed 




322 



Figure 2 Example of a “pulled” follower 

roller-type finger follower valve trains in 
combination with hydraulic pivot elements 
usually have significant advantages with 
regard to their friction behavior compared 
to other concepts. This conclusion was 
reached after carrying out extensive tests 
on externally-driven cylinder heads. The 
arrangement of the camshaft in the cylin¬ 
der head in relation to the space between 
the hydraulic pivot element and valve stem 
end can require a mounting position of the 
finger followers, which Schaeffler de¬ 
scribes as a “pulled” follower arrange¬ 
ment (Figure 2). 

Further tests of the finger follower valve 
train have shown that the “pulled” follower 
has advantages over the “pushed” follower 
arrangement. This is because skewing of 
the “pushed” follower occurs under load 
due to its design (Figure 3). The “pulled” fol¬ 
lower prevents this because it is self-align¬ 
ing. 


This is because the load is applied differ¬ 
ently to the the follower by the cam. In a 
“pushed” valve train, the cams can cause 
misalignment in an axial direction at one 
end of the finger follower at low to medium 
speeds. At the other end, the follower is lo¬ 
cated in position on the pivot element by a 
spherical piston head, which acts as a 
guide. “Pushing” in a lateral direction is 
prevented by the geometry at the pivot ele¬ 
ment so that the resultant force acting on 
the follower cannot generate any further 
movement. At the other end, the lateral 
guides of the finger follower are in contact 
with the valve stem so that an equilibrium 
of forces exists between the valve and the 
cam. The forces acting through all the fol¬ 
lowers on the entire shaft are totalized be¬ 
cause the finger followers usually have a 
preferred direction. The simulation of this 
arrangement shows that the entire cam¬ 
shaft is ultimately pressed into its axial 
bearing support. Corresponding tests 
clearly confirm the theoretically derived 
motion for all the tested parameter varia¬ 
tions. 

The increased force in the direction of 
the camshaft axis causes a higher drag 
torque of the cylinder head - which is 
greater if conventional plain bearings are 
used. Tests have confirmed the conclu¬ 
sion that rolling bearings are advanta¬ 
geous here. This effect is reduced with 
increasing speeds because the time avail¬ 
able for the cam lift is not sufficient to 
cause a significant increase in the axial 
force. 

In the case of the pulled valve train de¬ 
sign, the movement of the cam “pulls” the 
finger follower directly away from its fixed 
point - the spherical piston head on the 
pivot element. This process is compara¬ 
ble with pulling a conventional suitcase on 
two rollers: The handle corresponds to 
the spherical piston head as the location, 
and the force is also applied here via the 
rollers, only on the floor instead of via a 















Timing Drive Systems 


22 


323 



Figure 3 Comparison of forces acting on a “pulled” and a “pushed” follower 



— pushed 

— pulled 


Figure 4 Friction of a “pushed” follower compared with a “pulled” follower in relation to the speed 



































324 


cam. The resulting “pulling” force aligns 
the case in a straight line. However, if the 
case is pushed, it will veer to the side after 
a short distance. This pushed arrange¬ 
ment corresponds with the “pushed” fol¬ 
lower. 

The force values recorded in compari¬ 
son measurements correlate consistently 
with the friction measurements and verify 
the theory that with this combination no 
relevant transverse forces act on the cam¬ 
shaft. In a comparison of both directions of 



Figure 5 A positive example: The camshaft 
phasing unit is narrow, the sprocket 
is located on the camshaft 


camshaft rotation, the “pulled” follower 
concept has approximately 40 % less fric¬ 
tion at low speeds (Figure 4). This direction 
of camshaft rotation has around 30 % less 
friction at a speed of 4,000 rpm. 

The camshafts must be suitably posi¬ 
tioned in order to realize this type of finger 
follower arrangement. The decisive refer¬ 
ence point is the position of the finger fol¬ 
lower roller. The boundary conditions for the 
timing drive and particularly for the phasing 
unit change significantly depending on this 
position. 

The distance between the camshafts in 
combination with the maximum section 
height of the engine - this is defined from 
the requirements for the protection of pe¬ 
destrians - are the most important specifi¬ 
cations for subsequent designs. 

Challenges during the 
optimization of the entire 
system 


Camshaft phasing units 

A suitable camshaft phasing system is se¬ 
lected according to the required adjust¬ 
ment speed and adjustment force as well 
as the adjustment angle, which must be 
covered. The oil pressure is also a rele¬ 
vant input variable because hydraulic sys¬ 
tems are normally used. The required per¬ 
formance data determine the effective 
hydraulic surfaces and therefore the mini¬ 
mum size of the system. However, the 
available mounting space is usually limit¬ 
ed. If the ideal position of the camshafts in 
relation to the adjacent construction is not 
possible it will have the following implica¬ 
tions: The maximum possible outside di¬ 
ameter for the phasing unit is automati- 





















Timing Drive Systems 


22 


325 


cally reduced if there is a smaller distance 
between the camshafts. But if the cam¬ 
shafts have to be positioned further apart, 
they may be too close to the lateral limits 
of the cylinder head or valve cover. In this 
case, the only solution is to extend the de¬ 
sign envelope in a longitudinal direction to 
ensure the phasing unit has the required 
hydraulic power. The first conflict of objec¬ 
tives occurs if this solution is not possible. 
However, not all the questions are an¬ 
swered even if a phasing unit with a longer 
design is possible (Figures 5 and 6): The 
latter must be screw mounted, i.e. the 
cover and sprocket are clamped together 
by a number of screws, which pass 
through the phasing unit. The forces re¬ 
quired are relatively high and the design of 
the screw connection is very complex and 
critical. In addition, the sprocket is no lon¬ 
ger directly located on the shaft. In con¬ 
trast to a sprocket fitted directly on the 
shaft, the screw connection leads to toler¬ 
ances, which affect the radial runout of 
the system. This imbalance causes addi¬ 
tional excitations, which can impair the 
adjustment function, or have further dis¬ 
ruptive effects on the smooth running of 
the timing drive. 

The outside diameter of a phasing 
unit with a specified length is not only de¬ 
fined by the power requirements of the 
hydraulic system but also by the speci¬ 
fied ratio of 1:2 between the teeth on the 
crankshaft and camshaft. This requires a 
fixed number of teeth - usually an even 
number - on the camshaft. Not all possi¬ 
ble combinations can be realized in prac¬ 
tice: If the “correct” combination in rela¬ 
tion to the number of teeth required on 
the camshaft results in an outside diam¬ 
eter, which does not allow the specified 
distance between the camshafts, it can¬ 
not be implemented as in the converse 
case. The system is then too small and 
cannot transmit the required power due 
to physical reasons. 


This can ultimately mean that the targets 
must be changed or a completely different 
solution must be developed. One alternative 
is an arrangement where only one camshaft 
is directly driven by the timing drive. This 
means the problem regarding the restricted 
space between the camshafts is rectified 
and the outside diameter is only limited by 
the section height of the engine in the vehi¬ 
cle. This is generally the normal approach 
for timing drives although this measure 




Figure 6 A negative example: The cover and 
sprocket of the phasing unit are 
clamped together by means of 
screws. Any tolerances can affect 
the radial runout. 























326 


does have an impact if the engine is consid¬ 
ered as a whole. 

Firstly, a drive must be provided for the 
second camshaft - either a second chain 
or belt drive or a spur gear stage. This al¬ 
ways creates additional space require¬ 
ments in the longitudinal direction of the 
engine. In certain circumstances, the 
camshafts must also be extended. The 
design of the cylinder head and valve cov¬ 
er is significantly more complex on this 
side. Even an oil supply must be integrat¬ 
ed if a hydraulic tensioner is required for 
an additional chain drive. Firstly, this 
means an additional consumer must be 
taken into consideration in the oil system. 
Secondly, it results in an increased outlay 
when designing the oil ducts. These 
changes inevitably result in a heavier and 
more expensive engine. Additional pro¬ 
cess steps are also required in volume 
production assembly. Because more 
highly-dynamic components are used, the 
system is also more susceptible to vibra¬ 
tion and noise generation, which can 
sometimes only be managed by using 
complex solutions; this applies especially 
to spur gear stages. 

Last but not least, these components 
also generate friction in addition to noise, 
weight and costs. The resultant friction 
losses under unfavorable conditions are 
larger than those, which the change to a 
“pulled” valve train eliminated. In the worst 
case, the net result is worse than a con¬ 
cept where compromises have been 
made in the design of the valve train if it is 
equipped with an optimum camshaft 
phasing system. 

Track position of the chain 
or belt drive 

In addition to the design analysis de¬ 
scribed above, the position of the chain or 
belt track relative to the first camshaft 


bearing has a significant influence on the 
phasing unit concept. If there are many 
unfavorable requirements resulting from 
the design and arrangement of the adja¬ 
cent components, this can lead to solu¬ 
tions as shown in Figure 7. 

There are noticeable restrictions for 
the required phasing unit concept. The 
stator implemented here may be regard¬ 
ed as a sophisticated design but its man¬ 
ufacture is very expensive. Further costs 
are incurred during manufacturing be¬ 
cause additional quality assurance mea¬ 
sures are required. Clever positioning of 
the track for the chain or belt drive would 
allow a phasing unit design, which not 



Figure 7 A negative example: The phasing 
unit gear is connected via a narrow 
web. Manufacture of the stator is 
expensive. 









Technology 

Parameters Challenges competition 


Timing Drive Systems 


22 


327 



Figure 8 A positive example: The chain runs 
centrally in the chain tunnel 


only operates more efficiently but also 
has a positive influence on the timing 
drive due to its reduced mass (Figure 8). 
Lightweight components usually cannot 
fully compensate for the increased weight 
of a design because they soon reach 
their limits with regard to component 
strength. 


Selection of the chain or belt drive 

The decision about whether the require¬ 
ments are better met by using a chain or a 
belt timing drive must be clarified initially 
depending on the complexity and number 
of shafts to be driven. Chain drives and 
dry toothed belt drives have the longest 
history on the market and have reached a 
corresponding level of sophistication (Fig¬ 
ure 9). In contrast, the wet belt or BiO (belt 
in oil) is still a new product. If correctly de¬ 
signed, all three types of drives are com- 


Chain 



► Wear 

► NVH 

► Friction 

► ... 

► Chain width/length 
^ Max. diameter of 

crankshaft sprocket 

► Reasonable kinematics 

► ... 


Belt in Oil 



► Durability 

► Media resistance 

► Axial space 

► ... 

► Belt width/length 
^ Max. diameter of 

crankshaft sprocket 

► Reasonable kinematics 

► ... 



► Contamination 

► Sealing of VCT 

► Axial space 

► ... 

► Belt width/length 
^ Max. diameter of 

crankshaft sprocket 

► Reasonable kinematics 

► ... 


Figure 9 Whether it is a chain, belt in oil or dry belt: The specific requirements determine the 
selection of the chain or belt drive 















328 


parable with regard to their operating life. 
Irrespective of the type of chain or belt 
drive used, it is important to start the de¬ 
tailed design at an early stage in order to 
precisely work out all the potentials and 
risks. Even if it has favorable prerequi¬ 
sites, an effective, low-friction system can 
still fail to meet the targets with regard to 
important parameters. In some cases, 
moving the location of the camshaft by 
just a few tenths of a millimeter is suffi¬ 
cient to turn a good system into an unsat¬ 
isfactory system. 

Belt 

If a decision is made to use a dry toothed 
belt, the engine and camshaft phasing 
units must be sealed against the ingress 
of oil from the engine’s oil circuit. This 
negative effect is compensated by the 
advantage that a high measure of flexibil¬ 
ity is maintained during the implementa¬ 
tion of the design. A belt in oil eliminates 
this disadvantage. This can result in a 
slight advantage with regard to noise 
emissions due to the integration of the 
belt into the engine, although this de¬ 
pends on the application. The belt in oil 
allows the same degree of design free¬ 
dom as the dry belt. In comparison to a 
chain, both designs of belt have the ad¬ 
vantage that timing belts with an odd 
number of teeth can be manufactured. 
This means it is easier to make adjust¬ 
ments to the entire arrangement. In con¬ 
trast, smaller sprocket diameters are 
possible with chains without having a 
detrimental effect on the operating life. 

Chain 

As in the case of a belt, the first step is 
usually to determine the required number 
of sprocket teeth on the shafts. Ideally, 
the number should be as large as possi¬ 
ble in order to minimize the polygon ef¬ 
fect. At the same time, a check is made to 
ensure that the number of chain links re¬ 


quired is not a common multiple of the 
number of teeth on the crankshaft sprock¬ 
et. This results in improved noise and 
wear behavior. The length of chain re¬ 
quired is determined from the number of 
teeth and the distance between the 
shafts. Chain drives can only be manufac¬ 
tured in variants with an even number of 
chain links. This has a significant influence 
on the chain line. It is worthwhile investing 
sufficient time in consideration of the pos¬ 
sible variants with regard to chain pitches, 
the possible number of teeth and design 
of the chain line. An ideally designed chain 
line usually ensures quiet and dynamically 
correct running of the entire system in the 
fired engine. It also forms the basis for a 
low-friction system. 

Interdependence with adjacent 
components 

Individual boundary conditions, for exam¬ 
ple design envelope requirements, are oc¬ 
casionally so restrictive that they lead to 
functional impairments. This can mean that 
chain guides with pronounced curves 
must be inserted in the driving side of the 
chain drive. These inevitably increase the 
friction due to the higher normal forces. 
This also applies for a belt if it requires a 
pulley on the driving side. 

The slack side is assessed differently. 
Excitations on the system resulting from 
dynamic chain loads are damped by the 
chain tensioner via the tensioning rail ar¬ 
ranged on the slack side of the chain 
drive. The form of the rail affects the way 
and intensity, with which these impulses 
are transmitted. However, the first im¬ 
pression does not always correspond 
with the actual result: A clever design - 
even one with small radii - allows lower 
mean chain forces to be achieved than in 
chain drives designed to reduce the nor¬ 
mal force acting on the sliding layers by 
means of an exaggerated “straight” 
guide. 



Timing Drive Systems 


22 


329 


Summary 


The current structures in development de¬ 
partments correspond to the technical 
content of their systems. This is why it is 
difficult to consider all the systematic ef¬ 
fects of valve control components. Experi¬ 
ence at Schaeffler shows that optimization 
at component level does not often lead to 
the best possible result with regard to the 
entire system. In the worse case, for ex¬ 
ample, the net result of an optimized valve 
actuation system is worse than a concept 
which allows compromises in the design of 
the valve train if it is equipped with an opti¬ 
mum camshaft phasing system. This situa¬ 
tion can ultimately mean that the targets 
must be changed or a completely different 
solution must be developed. Current proj¬ 
ects confirm that the automobile manufac¬ 
turers come to the same conclusion in their 
analysis of the development process. Their 
are different approaches for implementing 
a more efficient development process. 
Schaeffler achieves this by bundling all de¬ 
velopment and application engineering de¬ 
partments in one organization. It is also 
helpful if all the components of interactive 
systems are jointly developed at all loca¬ 
tions. This simplifies matching the compo¬ 
nents to one another. It is highly recom¬ 
mended that system suppliers are involved 
in the new and further development of 
valve trains at the earliest possible stage. 
This allows the opening up of potential that 
was previously unused and rapidly leads to 
success. 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



330 



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Friction Tailored to Your Requirements 

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Tim Hosenfeldt 
Edgar Schulz 
Juergen Gierl 
Stefan Steinmetz 


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332 


Introduction Basic principles 


In any discussion about reducing fuel 
consumption, attention quickly turns to 
powertrain hybridization. However, the 
fact that a saving of 15 % can be achieved 
simply by minimizing friction is often 
overlooked. This has been demonstrated 
by Schaeffler in various studies over the 
past few years. The costs required to re¬ 
duce C0 2 emissions by 1 g/km can be 
kept well below those of electrifying the 
drive. 

However, friction is not a parameter 
that must be minimized in every case. 
Without friction, movement is not possi¬ 
ble in our daily lives - and this includes 
driving cars. Both involve requirements 
that can only be met if different friction 
conditions interact in the desired fashion 
- similar to a classic cross-country skier: 
The skier hopes that in his tribological 
system - consisting of the shape of the 
skis, their sliding layer, and the snow - 
friction will remain as low as possible 
when going downhill. By contrast, when 
going uphill, the skier ideally requires ad¬ 
hesive friction to move up the hill quickly 
and without having to use too much en¬ 
ergy. 

Even when seen from a tribological 
standpoint, the optimization of a system 
or of an entire powertrain must always be 
subjected to a cost-benefit analysis and 
ensure that the functional requirements 
for the overall system and its compo¬ 
nents are fulfilled. The service life, for in¬ 
stance, must not be less than that stipu¬ 
lated in the requirements specification. 
That is why the experts working at 
Schaefflers’ Surface Technology Com¬ 
petence Center depend on the systems 
expertise available in the company when 
developing highly specific coating solu¬ 
tions. 


Tribological system 
System details 

According to conventional definitions, a tri¬ 
bological system consists of four elements 
[1]: The part and the counterpart - these 
two move relative to each other - and the 
interfacial medium as well as the ambient 
medium. Now, new types of coating sys¬ 
tems in the micrometer and nanometer 
range are adding another element that has 
a significant influence on the properties of 
tribological systems and that can be used to 
adjust these systems for a specific purpose 
(Figure 1). 




_. 

Counterpart- 

Intermediate— 


Part 






Ambient medium- 


Figure 1 Structure of the tribological 
system [1] 


Friction in tribological systems 

The amount of friction that occurs in a tribo¬ 
logical system is influenced by a range of 
factors. In addition to the load applied and 
the lubricant characteristics, the surface of 
the active areas is particularly significant. 
The most important surface characteristics 
that determine friction and wear include the 
following [2]: 

- The chemical composition of the sur¬ 
faces that can be changed by pretreat¬ 
ment or during operation by reactive 
layers on the component surface. 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3 23, © The Author(s) 2014 













Customized Friction 


23 


333 



delamination) furrowing) tribochemically 

modified barrier layer) 

- fretting corrosion 

- friction oxidation 
- tribooxidation 

Figure 2 Overview of wear mechanisms 


- The hardness and Young’s modulus 
of the materials used. A hard surface 
alone does not protect a component if 
the basic material underneath can eas¬ 
ily undergo plastic deformation. 

- The surface structure not only affects 
lubricant film formation but also the 
force applied to the surface and thus 
surface fatigue. 

- Interaction of the surface and the lubri¬ 
cant. 

Generally, a distinction is made between 
four wear mechanisms (Figure 2). Abrasive 
wear occurs as a result of the mechanical 
impact of a harder active surface on another 
surface or hard particles. Adhesive wear re¬ 
sults from the molecular interaction when 
surface contact occurs in the contact inter¬ 
face. Tribochemical reactions change the 
contact interfaces by oxidation, for instance. 
Surface fatigue occurs if the material micro¬ 
structure changes under mechanical stress. 

There are various types of friction. The 
Stribeck curve is a good way to distinguish 
between the various types, as it can be 


used to plot the relative motion of the active 
surfaces for a lubricated contact in relation 
to the friction torque that occurs (Figure 3). 



Figure 3 Curve of the friction torque in relation 
to the relative speed (Stribeck curve). 
The black curve shows the classic 
profile, the red curve shows the torque 
profile when using nano-structured, 
diamond-like coating systems in the 
valve train. 















334 


Surface technology 

In addition to the lubricant, the material 
and surface quality of a bearing also have 
significant influence on friction and wear 
behavior. The properties of the material 
close to the surface are changed to im¬ 
prove the tribological characteristics of 
engine and transmission bearings. The 
initial approach here is to achieve smooth¬ 
ing by reducing the roughness peaks, 
such as by honing the raceways. Addi¬ 
tional improvement can be achieved by 
changing the surface through heat treat¬ 
ment and coating. Conventional methods 
here are black oxide finishing and carbo- 
nitriding. Chrome-plated surfaces are of¬ 
ten used for engine and transmission 
components that are subjected to high 
stress levels. 


State-of-the-art carbon thin film coating 
systems usually do not consist of a single 
layer but of up to 100 layers in the nano¬ 
meter range that perform certain func¬ 
tions. The exact layer structure is matched 
to the relevant application and require¬ 
ments [3]. 

Analysis, calculation, and 
simulation 


It is not unusual for tribological aggregate 
loading conditions not to be fully known as 
part of component development. That is 
why the services of Schaeffler’s Surface 
Technology Competence Center include a 


Prediction System 



Data mining (e.g. artificial neural networks - ANN) 


Empirical data set 


Mechanical, structural, chemical and tribological parameters 


Lubricant chemistry: 

- Base oil 

- Additives 


Tribological stress collective: 

- Kinematics 

- Pressure, forces 

- Temperatures 


Local and temporal 
overlapping of friction 
and wear mechanisms: 

- Chemical reactions 

- Surface fatigue 

- Adsorption 

- Adhesion 

- Abrasion 


Analytical 
prediction 
not possible! 


Coating, part, 
counter part: 

- Composition 

- Materials 


Figure 4 Method for predicting friction based on empirical data [4] 




























Customized Friction 


23 


335 


Results: 

Parameter optimization [ 
using data mining 




System analysis 

1 Primary body 

2 Counter body 

3 Interfacial medium 

4 Ambient medium 

5 Coating system 
+ stress collective 



Figure 5 Data mining as part of the development process for coatings [4] 


comprehensive analysis of the initial situa¬ 
tion. A standardized procedure ensures, for 
instance, that all relevant parameters are 
entered. 

A good development strategy always 
takes the overall system into consider¬ 
ation. This is the only way to develop a 
rolling bearing that is optimally designed 
for a specific application. Against this 
background, Schaeffler has expanded its 
tried and tested Bearinx® calculation pro¬ 
gram to include an analytical model for 
calculating rolling bearing friction. This 
model takes a wide range of parameters 
into consideration, such as real pressure 
distribution and internal bearing geome¬ 
try. In addition to load distribution and 
service life, it permits the calculation of 
rolling bearing frictional torque and thus 
the power loss of entire shaft systems or 
transmissions. 

Schaeffler has been breaking new 
ground in the development of the tribo¬ 
logical system of components that are oil 


lubricated and coated with customized 
diamond-like carbon coating systems. 
Due to the high complexity and interac¬ 
tions imminent to the system, the possi¬ 
bilities for analytically calculating tribologi¬ 
cal behavior is limited. The development 
and optimization of coatings for the sur¬ 
faces of cams and bucket tappets that 
come into contact with each other has 
thus so far been based on experimental 
investigations and the experience of spe¬ 
cialists. This method can be time-con¬ 
suming and expensive. 

For this reason, Schaeffler has devel¬ 
oped a method that can predict the tribo¬ 
logical behavior of camshaft and bucket 
tappet systems, for example. It is based 
on a combination of data mining and an 
artificial neural network and can be prac¬ 
ticed with available experimental data [4] 
(Figure 4). 

In this process, the artificial neural net¬ 
work learns the phenomenological corre¬ 
lations on which these data are based. 


















336 


Their capability to “learn” non-linear cor¬ 
relations allows artificial neural networks 
to predict an input variable - such as the 
friction value - even in complex tribologi¬ 
cal systems. Influencing factors include 
the type of coating and its hardness, the 
surface quality, lubricant oil additives and 
their concentration, the base oil and its 
viscosity, and the material of the counter¬ 
body and its structure. 

The suitability of an artificial neural net¬ 
work for this kind of application depends 
on its predictive accuracy. This is deter¬ 
mined by entering data from an experiment 
that was not used to practice the model. 
Finally, the prediction is compared to the 
value measured in the experiment. The 
data mining process is thus integrated into 
the development process of coatings for 
tribological systems (Figure 5). 

Although data mining programs al¬ 
ready contain automated optimization al¬ 
gorithms, expertise is required to design 
the optimum network topology of the arti¬ 
ficial neural network. The challenge lies in 
finding the optimum number and arrange¬ 
ment of neurons, the optimum number of 
input variables, the optimum parameters 
of the learning algorithm, and much more 
for specific data with a given number of 
examples. 

Since artificial neural networks only 
approximate functional correlations in 
most cases, the evaluation of a learned 
model is one of the most important steps 
in data mining. This can be achieved 
through various methods. Schaeffler has 
found that a 10x10 cross validation or 
boot strap cross validation supplies the 
best results to predict the tribological be¬ 
havior of a camshaft and bucket tappet 
system. Upon comparison with an exter¬ 
nally driven cylinder head, a deviation of 
only 8 % was found - a very good result, 
especially when considering that the 
measurement error with reference to fric¬ 
tion is at 5 % [4]. 


The use of such methods capable of “learn¬ 
ing” can therefore reduce the time and 
costs spent on experimenting, secure avail¬ 
able knowledge and use it efficiently for 
product development. 

Energy efficiency through 
minimized friction 


Influence of bearing designs 

The friction occurring on the active sur¬ 
faces of bearings is primarily determined 
by the selection of the bearing system 
and the bearing type and then by its de¬ 
sign details. One example here are the 
bearing supports of the main shafts in 
the transmission. Locating non-locating 
bearings are increasingly used as an al¬ 
ternative for conventionally adjusted ta¬ 
pered roller bearings. Schaeffler has an¬ 
alyzed various applications to determine 
the effect a change in the bearing system 
can have on fuel reduction. For a com¬ 
pact car with a double clutch transmis¬ 
sion, consumption was reduced by 3.8 % 
in the NEDC simply by changing the 
bearing system to locating non-locating 
bearings. Tandem angular contact ball 
bearings offer significant benefits in the 
rear axle differential. They replace the ta¬ 
pered roller bearings used in the past 
and develop a smaller contact surface 
and thus a lower friction torque while 
maintaining the same load carrying ca¬ 
pacity (Figure 6). With regard to the ve¬ 
hicle, this results in a potential C0 2 re¬ 
duction of 1.5 % that can be achieved at 
low cost. 

Over the past few years, significant 
progress has been made by using rolling 
bearings instead of plain bearings. For in- 



Customized Friction 


23 


337 




Pinion shaft 


New: 

Bearing supports 
with tandem 
angular contact 
ball bearings 


Rear differential unit 


Differential 


Standard: 

Bearing supports 
with tapered roller 
bearings 


Figure 6 Use of tandem angular contact ball bearings in the rear axle differential (blue) instead of 
tapered roller bearings (red) 


stance, this is true for balance shafts in the 
engine. Changing to rolling bearings while 
also designing the components with an op¬ 
timized weight can reduce C0 2 emissions 
by up to 2 % at a cost of less than ten euros 
per shaft. The cost-benefit ratio is just as 
favorable when switching from plain bear¬ 
ings to rolling bearings in the camshaft 
bearing supports. 


Coatings for specific 
requirements 

The Schaeffler Coatings Center uses all of 
the coating technology methods and has a 
modular system for validated coatings that 
can meet any requirement: Corrotect® 
coatings made from a zinc-iron or zinc- 
nickel alloy provide corrosion protection, 
Durotect® desig¬ 
nates tribological 
coatings that are 
produced chemi¬ 
cally. The coating 
configuration with 
iron oxide com¬ 
pounds is charac¬ 
terized by the fact 
that it has good dry 
running character¬ 
istics in the event of 
insufficient lubrica¬ 
tion. Insutect® - 
which can be used 
as an aluminum ox¬ 
ide coating, for in¬ 
stance - has been 
used primarily in 
energy production 
for current insula¬ 
tion; at Schaeffler, 
it is mainly used 


Corrosion Protection 


Wear 

Protection 



Sensor coating 


Increased life time 
Energy efficiency 
Increased functionality 


Figure 7 Schaeffler modular coating system as a basis for coatings for 
specific requirements 


































338 


for railway bearings, generators, and ship 
engines. With the hybridization of the pow¬ 
ertrain, this coating has become more and 
more important for the automotive industry 
as well (Figure 7). 

Over the past few years, nano-struc- 
tured coating systems based on carbon 
have been used increasingly as an alter¬ 
native for conventional surface technolo¬ 
gy processes, such as those developed 
by Schaeffler under the Triondur® brand 
name. In the power train, this type of coat¬ 
ing system was initially used in bucket 
tappet valve trains because the cost-ben¬ 
efit ratio appeared to be especially inter¬ 
esting: By using a customized Triondur® 
diamond-like carbon (DLC) layer on the 
tappet base, the sliding contact surface 
for the cam, the tribological properties 
have improved so much that friction in the 
valve train has been reduced by half. The 
mechanical bucket tappet thus almost 
reaches the friction values of a roller finger 
follower [5]. In relation to the entire vehi¬ 


cle, this means a reduction in C0 2 emis¬ 
sions of 1 to 2 % (Figure 8). Triondur® 
coating here offers excellent wear protec¬ 
tion and hardly requires any design space 
at all with its layer thickness of only 2 to 
3 microns. 

Schaeffler has standardized both its 
coating processes and its coating facili¬ 
ties. The same machines used in volume 
production are used for new develop¬ 
ments or adjustments from the start. The 
manufacturing process is developed 
along with the product. This ensures that 
the transfer from development in the 
coating process to worldwide volume 
production is stable and free of errors. 
The result is a consistently high level of 
quality irrespective of the manufacturing 
location. 



Mechanical bucket tappet 
polished and coated 

“customized Triondur® coating systems” 


Roller finger follower 
with hydraulic 
valve lash 

adjustment element 


* 



Mechanical bucket tappet with adjusted 
combination of coating, lubricant and 
counter body 


Mechanical bucket tappet carbonitrided, tempered, polished 


50 


100 


Mechanical bucket tappet 
carbonitrided and tempered 


Friction reduction in % 


Figure 8 Triondur® DLC coatings improve friction behavior by up to 50 % and offer a high level of 
wear protection [5] 




Customized Friction 


23 


339 



Friction loss 

Functional friction 

A 

Characteristic 

Mainly undesired effect 
not required for function 

Specific friction value and friction 
behavior required for function 
in friction systems 

Goal 

Reducing friction 

Achieving a specific friction level 





Low friction High friction 

1 


i i r 


0 





Figure 9 Functional targets on the friction axis with product examples 


Functional friction 


Similar to the cross-country skier men¬ 
tioned earlier, friction in an automobile is 
not always a bad thing. In a bearing, the 
aim is to minimize friction, in an engaged 
clutch, a brake, or in a press fit, the aim 
is to maximize it. Here, friction is used for 
a very specific purpose. The latter can 
thus also be called “functional friction” 
(Figure 9). 

Classic application examples of func¬ 
tional friction are clutches with a dry and 
oil-lubricated design, damping systems 
such as torsion dampers for clutch disks, 
and the synchronizing units in the trans¬ 
mission. 

Like all systems in an automobile, the 
tribological system is in line with the 
downsizing trend, resulting in improved 
performance: Specified requirements for 
the friction value must be met with small¬ 
er components. In addition to the param¬ 
eters relevant for tribology, such as the 


sliding speed, temperature, and pres¬ 
sure, the system environment must be 
optimized continuously to supply cus¬ 
tomized friction. 

One of the requirements is under¬ 
standing friction phenomena. Analyses 
were previously limited to specific dimen¬ 
sions or scale levels; friction phenomena 
are often analyzed at the machine dy¬ 
namics level. This means that the friction 
system is tested for a specific applica¬ 
tion, and conclusions are drawn from 
this. However, since friction occurs in the 
friction contact, the scale levels of con¬ 
tact mechanics must also be taken into 
consideration, such as the micro, meso, 
and nano levels. The atomic scale level 
can be left out here as it is used more for 
fundamental scientific research. 

The consistency of methods and tools 
across the various scale levels - from mate¬ 
rials to production - is an essential compo¬ 
nent for the detailed analysis and optimiza¬ 
tion of friction systems. The tools used at 
Schaeffler range from methods for analyz¬ 
ing systems and data to scale-specific test 




















340 


stands and materials analyses. The experts 
of the Schaeffler Group work together in an 
interdisciplinary fashion to use these meth¬ 
ods efficiently. 

Dry running friction linings for 
clutches 

A dry running double clutch system rep¬ 
resents a much greater challenge for the 
friction materials of the clutch than con¬ 
ventional manual transmissions. Inspec¬ 
tions here range from the system level to 
sub-components and partial lining (Fig¬ 
ure 10). 

An essential parameter are the comfort 
properties of the friction material. These 


are assessed by determining the damping 
behavior or excitation behavior of the 
friction material on a judder test stand 
(Figure 12). These inspections often show 
that damping decreases as the mean friction 
value increases; this means that friction 
vibrations increase. This behavior can be 
observed for all friction systems and ap¬ 
pears to be a universally valid principle. 
The consistent application of standardized 
methods and tools has lately achieved 
considerable success. Figure 11 shows an 
example: The damping and excitation val¬ 
ues were taken from a large number of 
tests with components that have different 
load profiles and plotted on a frequency 
scale. The significant improvement in the 
friction materials is clearly visible and re- 



Benefits: - systematic product development 
- profound material knowledge 


Component test stand 

- material inspection 
(screening method) 


Clutch test stand 

- material testing 

- standard tests 


Vehicle 


- validation 


Figure 10 Comprehensive lining development from partial lining investigation to complete clutch 

































Customized Friction 


23 


341 


Judder test, 30,000 cycles, 
running-in at 120 °C, 16 KJ (light load judder test) 



Required damping in Nm/s 


damping excitation 

- Initial state 
Optimization 1 

- Optimization 2 

Figure 11 Optimization of lining damping 

through use of a judder test stand 

fleeted primarily in a much lower disper¬ 
sion. Due to the internal damping of the 
power train, excitation values of more than 
0.05 Nm/s are a disturbance and notice¬ 
able for the driver as judder vibrations. 


Wet linings for twin clutches 

For wet linings, oil is the third tribological 
component in addition to the lining and 
steel or cast iron that function as the fric¬ 
tion contact surfaces. The oil serves to 
dissipate the frictional heat, but it can 
also have negative effects. Too much oil 
between the friction contact surfaces 
leads to hydroplaning, similar to the 
aquaplaning of tires on a road wet with 
rain, and thus results in a low and uncon¬ 
trollable friction value. In addition, drag 
losses occur in open clutches that sig¬ 
nificantly reduce efficiency. If there is not 
enough oil, there is a risk of partial mixed 
friction or even dry friction; this has a 
considerable impact on comfort behav¬ 
ior. For this reason, the macrostructure of 
the lining must be designed in a way that 



— Wet lining with average porosity 

— Wet lining with ideal porosity 

Wet lining with insufficient porosity 


Figure 12 Improvements in friction value 
behavior through lining porosity 

permits the oil to be distributed uniformly 
to the various friction surfaces and to en¬ 
sure that the aquaplaning effect can be 
prevented reliably. This is achieved 
through the use of specific groove geom¬ 
etries. Adequate porosity of the lining 
supports this effect. 

If these and other findings are imple¬ 
mented consistently, the friction value be¬ 
havior of wet clutches can be improved 
significantly. Figure 12 shows an example. 
The correlations described here should 
suffice to show that the development of 
an efficient tribological system is only 
possible if all conceivable boundary con¬ 
ditions have been taken into consider¬ 
ation. 

Challenges for friction linings in 
synchronization 

The friction value structure and consis¬ 
tency are important variables in the devel¬ 
opment of synchronizing materials. The 
transmission oil must be pushed away 
quickly from the friction contact to achieve 
these variables. If the microstructure of 
the friction materials cannot ensure this, 
other solutions must be found. One solu- 










342 



Shifting time in s 


— without grooves 

— spiral groove 
axial grooves 

Figure 13 Effect of macro-geometry on the 
torque curve 

tion is the targeted macro-structuring of 
the friction material by using grooves or 
groove patterns. The diagram in Figure 13 
shows the torque curve of a gear shift 
mechanism with a friction material with a 
groove and without a groove. Without the 
groove, only a very small dynamic initial 
friction value can be observed. It results in 
longer shifting times and, in extreme cas¬ 
es, prevents the transmission from shift¬ 
ing altogether. 

Modifying the macroscopic surfaces - 
in this case by means of spiral or axial 
grooves - can help improve the torque 
curve. As the oil is pushed away rapidly, 
the dynamic friction value is already at a 
much higher level even at the start of the 
shifting operation, guaranteeing a high lev¬ 
el of shifting comfort and short shifting 
times. 


Tribotronics 


The term “tribotronics” is used to describe 
a fairly new field within tribology that inte¬ 
grates mechatronics into a tribological 


system with the aid of an electronic con¬ 
trol system. Mechatronics differs from tri¬ 
botronics in that it only uses information 
from the inputs and functional outputs of 
the mechanical system to control its op¬ 
eration. Such functional outputs supply 
information about speeds, torques, tem¬ 
peratures and loads. 

Tribotronics, on the other hand, not 
only considers additional output param¬ 
eters such as friction, wear, or vibrations, 
but influences them by means of an elec¬ 
tronic control system. The goal is to in¬ 
crease the performance, efficiency, and 
reliability of the tribological system and 
thus of the entire application. In tribo¬ 
tronics, the component becomes a sen¬ 
sor or actuator - or, put another way: The 
sensor or actuator becomes a compo¬ 
nent (Figure 14). This opens up an entire¬ 
ly new range of applications for coating 
technology. 



Figure 14 Sensotect® coating measures the 
force on the rolling bearing. The 
component becomes a sensor. 















Customized Friction 


23 


343 


Its new thin-layer sensor Sensotect® pro¬ 
vides Schaeffler with a basis for imple¬ 
menting tribotronics in automotive engi¬ 
neering and industrial applications. Going 
forward, this will permit output parame¬ 
ters such as the force, torque, and tem¬ 
perature of a component to be measured 
in places where conventional sensors, 
such as glued strain gauges, cannot be 
used because they are susceptible to ma¬ 
terial aging or signal drifting due to poly¬ 
mer glues or transfer foil. 

With Sensotect®, a thin, strain-sensi¬ 
tive PVD coating performs the actual 
measurement function. The coating is 
structured by micromachining. These 
structures are deformed at the same 
rate as the carrier component. Deforma¬ 
tion results in a change in electrical re¬ 
sistance in the sensor layer. This change 
is a measurement, for instance, of the 
contacting torque or the forces impact¬ 
ing on thrust bearings, drive shaft, or 
steering column shafts. Measurements 
are taken during continuous operation 
and with extreme levels of sensitivity 
and precision, or, to be more precise, 
with minimal hysteresis errors and mini¬ 
mal linearity deviations. Schaeffler has 
already been able to show the function 
of this type of sensor system in demon¬ 
stration vehicles - both in passenger 
cars and bicycles. 

One of the greatest challenges of such 
sensory coating systems is manufacturing. 
The use of highly efficient coating sources 
and compliance with very stringent re¬ 
quirements for cleanliness in the manufac¬ 
turing process has helped Schaeffler to 
achieve a quality level even for typical 
three-dimensional rolling bearing compo¬ 
nents that was previously known only for 
planar substrates in the semiconductor in¬ 
dustry. 

In the continued development of tri¬ 
botronics, Schaeffler focuses on pro¬ 
cessing signals from surface sensors in 


an external control unit. Based on tribo¬ 
logical algorithms, these signals are used 
to perform calculations that indicate 
whether the operating temperature of the 
component has to be corrected or 
whether a dimensional change is re¬ 
quired, to name just two examples. 
Actuator coating systems carry out the 
necessary corrections. Completely au¬ 
tonomous and self-regulating systems 
are feasible if additional functions are in¬ 
tegrated into the component surface, 
such as telemetric components or trans¬ 
fer structures for energy supply and en¬ 
ergy production. 


Summary and outlook 


The optimization of tribological systems 
in drives still offers considerable potential 
for reducing fuel consumption. Opportu¬ 
nities can be found in the selection of an 
optimum bearing system as well as a 
coating customized for a specific appli¬ 
cation. Customized, nano-structured di- 
amond-like Triondur® DLC coating sys¬ 
tems can help optimize sliding contacts 
in such a way that their friction losses 
occur in the same range as rolling fric¬ 
tion. The new Triondur® coating systems 
help mechanical bucket tappets achieve 
friction values that are almost identical to 
those of a roller finger follower. At the 
same time, coating offers excellent wear 
protection - without requiring additional 
design space. Since these systems can¬ 
not be calculated analytically, Schaeffler 
has broken new ground in their develop¬ 
ment: Data mining is combined with an 
artificial neural network to generate a 
procedure that can predict the tribologi¬ 
cal behavior of such complex tribological 
systems. 



344 


At the other end of the imaginary “friction 
axis”, one of the development goals is to 
produce the required friction values with 
components that continuously decrease in 
size and weight. The analysis of friction phe¬ 
nomena must be expanded to include all 
dimensions and scale levels in order to 
achieve the best results from a system eval¬ 
uation. 

Tribotronics opens up an entirely new 
range of applications for coating technolo¬ 
gy. In future, the combination of force con¬ 
verters, data transfer, and transfer struc¬ 
tures for energy supply and energy 
production will make autonomous mea¬ 
surement systems a possibility even for ro¬ 
tating parts. 

Schaeffler has combined all of the im¬ 
portant expertise, from fundamental tri¬ 
bological research, coating develop¬ 
ment, and coating facility engineering to 
volume production and quality control 
measures. The most important goal here 
is to derive the best integrated surface 
technology-based solution from custom¬ 
er requirements so that the customer is 
provided with a product that has a sig¬ 
nificant added value - and with the best 
quality. 


Literature 


[1] Musayev, Y.: Verbesserung des tribologischen 
Verhaltens von Stahl/Stahl Gleitpaarungen fur 
Prazisionsbauteile durch Diffusionschromier- 
ung im Vakuum. Friedrich-Alexander-Univer- 
sitat Erlangen-Nurnberg, Dissertation, 2001 

[2] Czichos, H.; Habig, K.: Tribologie-Handbuch. 
Reibung und VerschleiB. 2. Auflage. Wies¬ 
baden: Vieweg, 2003 

[3] Hosenfeldt, T.; Musayev, Y.; Christgen, W.: 
Energieeffizienz durch Reibungsreduzierung 
mittels Oberflachentechnik als Konstruktion- 
selement. 2. ATZ-Fachtagung Reibungsminim- 
ierung im Antriebsstrang, 2011 

[4] Schulz, E.; Musayev, Y.; Tremmel, S.; Hosen- 
feldt, T.; Wartzack, S.; Meerkamm, H.: Reibung 
und VerschleiB vorhersagen - Durch neue 
Methoden der virtuellen Produktentwicklung 
Schmierstoff und Schichtsystem optimal 
aufeinander abstimmen. mo - Magazin fur Ober¬ 
flachentechnik 65, 2011, no. 1-2, pp. 18-21 

[5] Hosenfeldt, T.; Musayev, Y.; Schulz, E.: Cus¬ 
tomized Surface Technology - A Modern De¬ 
sign Element to Increase Energy Efficiency by 
Friction Reduction, VDI Fachtagung Ventiltrieb 
und Zylinderkopf, 2012 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



346 



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The Long Path from Discomfort 
to Customer Acceptance 

Start-stop: Yesterday, today and tomorrow 


Tobias Eckl 

Dr. Eckhard Kirchner 


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348 


Introduction 


Engine start-stop systems mark the entry 
into the electrified powertrain and from a 
cost-benefit point of view they are one of the 
best ways to reduce C0 2 emissions. Sav¬ 
ings measured under the NEDC amount to 
between 4 and 5 %. In heavy urban traffic, 
the reduction in fuel consumption can be 
larger. Applying the current WLTP driving 
cycle it can be expected that the savings 
measured for basic start-stop systems will 
be lower than with the NEDC. This is due to 
the fact that the proportion of time during 
which the vehicle is stationary is estimated 
to be 13 % - significantly lower compared to 
the NEDC which assumes 23 %. By con¬ 
trast, stop-start systems with sailing func¬ 
tion will benefit since according to the new 
test procedure the vehicle is driven more 
dynamically and is accelerated to higher 
speeds. However, considering the reduced 
consumption under real driving conditions, 
basic start-stop systems still remain an af¬ 
fordable option. 

Surveys initiated by Schaeffler - even if 
they are not fully representative - show that 
many motorists would like to permanently 
switch off the start-stop system despite the 
proven benefits in fuel economy. This is due 
to discomfort associated with restarting. 
Here the currently used technologies, for 
example starter pinions, meet their func¬ 
tional limitations. 

A systemic approach to this task opens 
up promising options and potentials, for 
example when additionally taking into ac¬ 
count the belt drives of accessory units as 
well as the second on-board electric sys¬ 
tem with 48V. This allows, for instance, 
comfortable air conditioning even with the 
engine switched off - to culminate in an 
initial mild hybridization without the need 
for a high-voltage on-board system in the 
vehicle. 


Constraints and expecta¬ 
tions of the protagonists 


The OEM view 
Market aspects 

The strongest driver of the anticipated 
further spread of start-stop systems is 
their compelling cost-benefit ratio - es¬ 
pecially as far as the basic version is 
concerned. Such a basic version in¬ 
cludes a reinforced starter and the imple¬ 
mentation of the start-stop strategy in the 
engine control unit, and in some cases 
the capacity or the type of battery is al¬ 
tered. In addition, in Europe in particular, 
the motorists’ desire for fuel-efficient 
cars in urban traffic leads to the same 
technical solution for the reduction of 
C0 2 emissions as that brought about by 
the regulatory requirements for automo¬ 
bile manufacturers. Between 2015 and 
2021, every manufacturer must cut the 
fuel consumption of its fleet of new vehi¬ 
cles offered in Europe by an average of 
27 %. Those who fail to achieve this tar¬ 
get will have to pay high penalties. Under 
the NEDC, turning the engine of a com¬ 
pact car off during the idle fraction of the 
driving cycle will lower the fuel consump¬ 
tion by about 4.5 %, depending on vehi¬ 
cle and engine data. 

Schaeffler expects start-stop systems 
to prevail also in the Chinese and U.S. 
high-volume markets. This is despite the 
fact that start-stop cannot fully bring to 
bear its advantages in these markets due 
to the respective local standard cycles 
and corresponding consumption require¬ 
ments. Moreover, it is foreseeable that a 
waiver of start-stop in the product range 
could increasingly lead to competitive 
disadvantages. 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3 24, © The Author(s) 2014 



Start-Stop 


24 


349 


Technical aspects 

The decision to integrate a start-stop sys¬ 
tem immediately raises the question of 
what technology to use. Schaeffler has 
consolidated external data, its own re¬ 
search results and ongoing projects and 
project requests into a technology outlook 
and the functions that can be derived 
from it. In general it can be stated that Eu¬ 
rope clearly has a pioneering role in this 
area and that the number of models fea¬ 
turing start-stop as standard or optionally 
is increasing year after year. It is to be ex¬ 
pected that in the near future, some auto¬ 
mobile manufacturers will initially imple¬ 
ment a sailing option in the higher speed 
range even in vehicles other than ones 
with full-hybrid drive. Schaeffler defines 
sailing as rolling with the engine switched 
off, either with (active sailing) or without 
(passive sailing) the support of an electric 
machine. By the end of this decade, in 
Europe it should be standard for new ve¬ 
hicles that the air conditioning system 
can operate independently of the internal 
combustion engine. Slighty, all start-stop 
technology variants as well as the various 
possible functions will find their way into 
China and the United States. The active 
sailing function is expected to be avail¬ 
able in low volumes in 2017 for the first 
time. 


By choosing a specific start-stop tech¬ 
nology and the associated operating 
strategy, the expected number of start- 
stop cycles is defined. While a conven¬ 
tional vehicle is designed for only about 
36,000 starts, basic start-stop systems 
are designed to endure between 300,000 
and 500,000 such cycles. The specifica¬ 
tions for a vehicle with sailing function 
are based on an average 1.2 million 
starts. This means: During the 300,000 
kilometers of the vehicle’s expected life¬ 
time, the motor is switched off and on 
again every 250 meters. For comparison: 
Today’s hybrid vehicles arrive at approxi¬ 
mately 600,000 start cycles (Figure 1). 
However, a decrease in the total required 
start cycles is emerging in vehicles with 
sailing function. This is due to the limited 
number of starts due to the operating 
strategy (reduced sailing speed range, 
latency). 

The clear differences stem from the 
fact that a basic start-stop system switch¬ 
es off the engine only when the vehicle is 
stationary or in the last phase of rolling. In 
contrast, a system with sailing function 
stops the engine whenever it is not under 
load, and thus even more often than a hy¬ 
brid drive. The latter deactivates the com¬ 
bustion engine only when a change into 
E mode is possible. 


Basic ESS (Engine Start-Stop) Hybrid vehicle Sailing 

I nil 

Conventional 1 2345678 

Starter Vehicles of different customers 


Figure 1 


Number of start-stop cycles compared by system 




350 


It is not only the components directly involved 
in the start process that must be adapted to 
this requirement. The following aspects also 
have to be taken into consideration: 

- The crank and the valve train are sub¬ 
jected to prolonged periods without oil 
pressure and are therefore frequently 
operated under conditions of mixed 
friction. 

- The systems for exhaust gas after-treat¬ 
ment go through a temporary, short¬ 
term cooling phase much more often. 

- The accessory units in the belt drive pass 
through the start process more frequent¬ 
ly (resonances, rotational irregularities). 

- The elements of the air intake system 
(throttle valve, fuel injection system, tur¬ 
bocharger) pass through the start pro¬ 
cess more frequently (mixed friction, 
thermal stresses, pressure fluctuations). 

- The dual mass flywheel often passes 
through the resonant vibration range. 

- Safety-critical auxiliary units (brake ser¬ 
vo unit, power steering pump) must be 
supported using electric drives under 
certain circumstances. 


Number of participants 
(530 people in total) 



Allocation of vehicle brands to participants 


Allocation of participants' complaints 
regarding start-stop system 



■ Complaints ■ Duration of start 

■ No complaints ■ NVH 

■ Failure A/C 

Disturbing during 
stop & go 

Figure 2 Customer criticism regarding start-stop 


Start-stop systems as 
judged by the customer 


A technology and the functions derived from 
it will become accepted on the market on a 
permanent basis only if they meet the needs 
and expectations of customers. Any techni¬ 
cal development must be in line with this. 
Therefore, in the long run it is indispensable 
for a system provider to know the prefer¬ 
ences of end users. So using a test group, 
Schaeffler has looked into whether motorists 
have reservations about start-stop systems 
and if so, why. This test group was not repre¬ 
sentative in accordance with the principles of 
control sampling; rather it consisted of tech¬ 
nically minded Schaeffler employees. The 
answers given were unprompted. After all, 
70 % of respondents were fully satisfied with 
the technology and 30 % expressed criti¬ 
cism. Almost half of the latter criticized the 
length of time between engine standstill and 
restarting (Figure 2). 


Number of complaints about noise / vibration 
related to engine and manufacturer 



Participants wishing to permanently deactivate 
the start-stop system 








Start-Stop 


24 


351 


Another interesting fact is that as many 
as 40 % of customers would like to per¬ 
manently deactivate the start-stop sys¬ 
tem, even though some of these people 
seem to be satisfied with the system 
itself. 


Restart, Change of Mind, acoustics 

Some respondents felt the start time lasted 
too long, even though it exactly corre¬ 
sponded to the time elapsing between 
turning the ignition key and the start of the 
engine. What appears to be normal for a 
key start is obviously found to be uncom¬ 
fortably long for restarts initiated by the 
start-stop system. This is because from a 
subjective point of view, the start process 
is in conflict with the motorist’s desire to 
move off right away. Therefore a start-stop 
system will be evaluated as “good” when 
the restart is subjectively perceived to be 

“Change of Mind” event 
I 1200 ms 

■i r 

800L- - -!- -I - Coventional 

starter pinion 



0 0.4 0.8 1.2 1.6 



Permanently engaged 
starter pinion with 
one-way clutch on the 
crankshaft 


Belt-driven starter 
generator & 
decoupling tensioner 


Time in s 

— Crankshaft speed 

— Starter speed 


Figure 3 Restart times of different start systems, following change-of- 
mind situations 


faster than the key start first thing in the 
morning. 

Since the respondents are not in¬ 
volved in the development and the techni¬ 
cal terms of start-stop systems, they did 
not mention the term “Change of Mind” 
(CoM) in their freely uttered comments. It 
is very likely, however, that a lack of CoM 
ability of the start system was one of the 
major factors that led to their judgment re¬ 
garding poor restart times. “Change of 
Mind” in the context of start-stop systems 
refers to situations where the driver would 
like to go on driving during the phase of 
automated engine shut-off. The engine is 
still coasting at this stage. Yet the conven¬ 
tional starter pinion can only engage the 
starter gear at engine standstill and sub¬ 
sequently initiate the restart. All this 
adds up to a delay which is perceived as 
significant. 

The length of the delay depends 
largely on the technology used. With a 
conventional start¬ 
er pinion, the time 
between the CoM 
event and reach¬ 
ing idle speed 
once again lasts 
up to 1200 ms. 
Using a perma¬ 
nently engaged 
starter of the same 
type, this time is 
reduced by half. 
And with a belt- 
driven starter gen¬ 
erator the time is 
cut by another 
third (Figure 3). 
This technique, as 
well as improved 
starter pinion con¬ 
cepts could there¬ 
fore have a cata¬ 
lyzing function for 
the further spread 


r 















352 


of start-stop technology - especially 
since another one of the testers’ criti¬ 
cisms would be addressed at the same 
time: Approximately one-fifth of those 
who said they were not fully satisfied with 
their start-stop system also complained 
about the vehicle’s noise and vibration 
behavior (NVH) during restart. 


Further challenges 

Furthermore, about one-fifth of the critics 
pointed to the fact that the air conditioning 
system was not in operation during engine 
shutdown, or only for a short time and with 
limited capacity. A further 13 % objected to 
the engine constantly switching on and off 
during stop-and-go traffic. 

For developers, the proportion of over 
40 % of respondents who would like to 
completely shut off the system is a clear 
mandate. There is evidence that quite a 
few of these people did not understand, 
for example, why in certain concrete situ¬ 
ations the engine was not switched off 
even though the vehicle was stationary, so 
that they suspected a malfunction. When 


adding further functions to the start-stop 
system, it therefore seems advisable to 
bear in mind that the “behavior” of tech¬ 
nology must be comprehensible for the 
customer. 

In the overall analysis of the test results 
it is remarkable that the respondents’ over¬ 
riding criticism was to do with lack of com¬ 
fort. Concerns about, for example, in¬ 
creased wear were mentioned just as 
rarely as an appreciation of reduced fuel 
consumption. 


Technology roadmap 


Starter pinions are today by far the most 
common components used for starting 
conventionally powered vehicles with start- 
stop system. Based on this technical level 
and including the benchmark results, the 
following technology roadmap can be sum¬ 
marized (Figure 4). 

The next refinement of the 12 V starter 
pinion will aim to increase the start com- 


Limited electric driving 
Sailing 

Independent air 
conditioning (A/C) 

Improved comfort 




1 

2014 

2017 

2020 

Obligatory 

Desirable 

Market pioneer 


Figure 4 Development of start-stop technologies and functions by 2020 





Start-Stop 


24 


353 


fort (NVH) and the start speed. Improved 
starter pinion technology and variable 
start speeds are obvious measures to en¬ 
sure better pinion engagement. 

Occasionally belt-driven starter gen¬ 
erators (BSG) are already being used 
with 12 V on-board electric systems. They 
offer advantages over the starter pinion 
with respect to noise and vibration be¬ 
havior. However, a bi-directionally acting 
belt drive must be used in order to 
achieve the necessary load reversal in 
the belt drive. This does not reverse the 
direction of rotation, but rather the direc¬ 
tion of the load acting in the belt drive; 
the carrier strand and the return strand 
alternate. 

Start comfort and start speed can be 
further increased if the start-stop system 
can fall back to a second voltage level 
with 48 V. The considerations in this re¬ 
spect have become more urgent be¬ 
cause the energy demand in vehicles has 
grown steadily over the past two de¬ 
cades. This development has been trig¬ 
gered primarily by the increased use of 
driver assistance and multimedia sys¬ 
tems, as well as by more extensive com¬ 
fort and safety equipment. In addition, 
today more components are operated 
with electrical instead of mechanical en¬ 
ergy. If a second lithium ion-based bat¬ 
tery is installed to extend the on-board 
electric system, then the amount of recu- 
perable energy will increase significantly. 
It is very useful for functions with high 
power consumption such as boosting, 
see [1, 2]. 

A separable crankshaft pulley is con¬ 
ceivable as another BSG stage, see [3]. 
The belt is thereby thrown off, as it were. 
The starter-generator can then drive the air 
conditioning compressor when the engine 
is switched off. This functionality can be a 
critical success factor for the further ac¬ 
ceptance of start-stop systems on the 
North American market. In conjunction 


with an electric drive axle, eventually the 
gap to the hybrid drive can be closed with¬ 
out having to bear the cost of a high- 
voltage system: In a 48 V environment, the 
achievable output is large enough to allow 
active sailing and cope with stop-and-go 
traffic without the assistance of the internal 
combustion engine. However, there is still 
no contact protection required. In addition, 
efficiency gains can be achieved during 
recuperating. This is because almost all of 
the kinetic energy can be recovered by re¬ 
generative braking, while until now this was 
absorbed to a greater extent by the drag 
torque of the internal combustion engine, 
see [1]. 

Market development of 
start-stop systems 


Current market situation and outlook 

A specific registration of vehicles, catego¬ 
rized by those with and without start-stop 
system, is not available worldwide. 
Schaeffler has produced the overview be¬ 
low by reconciling data from external mar¬ 
ket research with material from its own re¬ 
search. It is based on all cars featuring a 
conventional powertrain; hybrid and all-electric 
vehicles are excluded. 

Measured by the number of existing ve¬ 
hicles, start-stop systems are found rela¬ 
tively rarely, even in the mature vehicle 
markets in the western world. For the 
EMEA region (Western and Eastern Eu¬ 
rope, Middle East, Africa), however, it is 
becoming apparent that the number of 
new vehicles equipped with this technolo¬ 
gy is steadily increasing. Out of more than 
21 million vehicles with internal combus¬ 
tion engines in 2012, as many as 7.8 million 



354 


30 n 



China Japan North America EMEA 


■ Start-Stop 

■ No Start-Stop 

Figure 5 Number of new vehicles equipped with engine start-stop systems in the most important 
markets 


were already equipped with such a sys¬ 
tem. It is expected that as soon as 2016 
two thirds of all new cars will feature a 
start-stop system (Figure 5). The strongest 
impetus for this is likely to come from 
Western Europe: As from 2019, this system 
will form part of the standard equipment of 
conventionally powered vehicles in most 
segments. 

In North America, however, the pene¬ 
tration rate is still low for two main rea¬ 
sons: Firstly, the fuel savings resulting 
from engines stops in urban traffic calcu¬ 
lated based on the U.S. test cycle are 
much lower than those based on the Eu¬ 
ropean cycle, which means that there is 
less incentive for automakers to install 
such a system. Secondly, the demand in 
the North American market for more fuel- 
efficient technologies is still quite subdued 
due to the comparatively low fuel prices. 
Moreover, for reasons of comfort motor¬ 
ists reject a system with which the air con¬ 


ditioning system cannot be operated dur¬ 
ing engine standstill. 

Apart from China, the only other coun¬ 
try where a significant spread of start-stop 
systems is anticipated is Japan. However, 
in this market a substantial proportion of 
new vehicles are produced as hybrids 
even today. In the rest of Asia, as well as 
in India and South America, according to 
current estimates start-stop technology 
will play little or no role end of this decade 
even though in India the cost of fuel is 
high compared to the average disposable 
income. 

Market expectations for selected 
components deployed in start-stop 
technology 

Schaeffler has identified subsystems of 
powertrains based on internal combustion 



Start-Stop 


24 


355 


Market development 
Conventional starter 



2012 


2016 


2019 


w 
o 

8 0.6 
C 0.5 - 

O 

fO.4-1 

■- 0.3 H 
0 


Market development 
Two-speed starter 


0.2 - 
0.1 - 


re 


2012 


2016 2019 


Market development Market development 



■ EMEA ■ Japan 

■ North America China 


Figure 6 Market development of starter concepts in general and within the four relevant economic 
regions 


engines and assessed the likely market de¬ 
velopment of available technologies up to 
2019 - again broken down to the relevant 
economic regions. 

Starter concepts 

The market potential quantified for these 
components relates to all vehicles that are 
equipped with a start-stop system. The fol¬ 
lowing engine start concepts were taken 
into account (Figure 6): 

- 12 V conventional starter pinion 

- 12 V belt-driven starter generator (BSG) 

- 48 V BSG 

It is difficult to assess the development of 
the market for starter pinion concepts with 


two transmission ratios for cold and com¬ 
fort start. Assuming the success of the 
concept outlined below, up to one million 
units are expected to be sold globally by 
2019. 

Provision of hydraulic pressure 

This market assessment relates exclu¬ 
sively to vehicles that are equipped with 
torque converter automatic transmis¬ 
sions, double clutch and CVT transmis¬ 
sions so that their actuators are depen¬ 
dent on continuous oil pressure. The time 
before such a transmission is ready for a 
restart can be reduced considerably if the 
hydraulic pressure is maintained during 

















356 


Market development 



Market development 



Figure 7 Market development of concepts for 1 
the four relevant economic regions 

engine stoppage. The following options 
are available to ensure this (Figure 7): 

- Reduced leakage 

- Electric auxiliary oil pump 

- Pressure accumulator 

Gear detection 

Gear detection is relevant only for vehicles 
with manual transmission, so the potential 
for this sensor system is correspondingly 
reduced to this configuration, see Figure 8. 
If all gear stages can be detected, then 
sailing operation is possible not only when 
rolling towards a traffic signal, but also at 
higher speeds. Typically a distinction is 
made only between neutral, reverse and 
(any) forward gear. 


Market development 



■ EMEA 

■ North America 

■ China 
Japan 


provision of hydraulic pressure in general and in 

New opportunities and 
approaches 


Starter pinion with two-speed 
transmission 

Based on the findings from the interviews 
with the test group as well as other mar¬ 
ket and technology analyses, Schaeffler 
has intensified its investigations to im¬ 
prove the performance of the starter pin¬ 
ion. The concept that is currently being 
pursued is a starter with two-speed plan¬ 
etary gear. This consists of a double 
planetary gear set with an additional sun 
gear. This sun gear and the planetary 















Start-Stop 


24 


357 


Market development 



■ EMEA ■ Japan 

■ North America ■ China 


Market development 
Full aear detection 



2012 2016 2019 


Figure 8 Market development of gear position detection in general and in the four relevant economic 
regions 


carrier each feature a one-way clutch 
which is designed so that the electric 
motor’s direction of rotation can be re¬ 
versed without changing the direction of 


rotation of the transmission unit’s output 
shaft. Reversing the direction of rotation 
will activate the second gear stage of the 
planetary gear. However, this will happen 



Input 

shaft 



Output 

shaft 


Figure 9 


Two-speed starter with planetary gear 













































358 


only when the engine is warm and is to 
be restarted by the start-stop system. 
The starter can then translate the lower 
friction of the warm engine into higher 
starting speeds using the same power 
input. This not only reduces the start 
time, but also the noise and vibration lev¬ 
els. First gear is used for cold starts only, 
so that the customer perceives a notice¬ 
able difference between cold start and 
restart. This gain in comfort compares 
favorably to the relatively low outlay re¬ 
quired for the mechanical integration. 

Electrified drive for the 
air-conditioning compressor 

When the engine is at a standstill, so too 
is the air-conditioning (AC) compressor. 
The temperature increase in the vehicle 
interior has a detrimental effect on com¬ 
fort after only a short time so that the 
driver will probably switch off the start- 
stop system and immediately restart the 
engine manually. This means that the 
savings potential is not fully utilized. 
While the market is already providing so¬ 
lutions to compensate for the non-avail¬ 
ability of the mechanical drive, these 
bring various disadvantages in their 
wake: The electric air-conditioning com¬ 
pressor cannot be used in price-sensitive 
vehicle segments, and it also has an un¬ 
favorable energy balance. In addition, 
paraffin-based latent heat storage sys¬ 
tems require additional space and can¬ 
not dehumidify the air. Moreover, the dy¬ 
namics in heat input and reclaim 
achievable to date is unsatisfactory. 

Schaeffler has therefore launched a 
project that is expected to result in a 
technically and economically interesting 
alternative. The core of this concept is 
the integration of an electric motor - de¬ 
signed, for example, with a “hollow shaft” 
- between the air-conditioning compres- 



Figure 10 Power split concept for the 

electrification of the air-conditioning 
compressor 

sor and belt pulley. The existing belt drive 
remains unchanged. Instead of a rigidly 
connected or magnetically separable 
pulley, a planetary gear set is used for 
the air-conditioning compressor. It en¬ 
ables a power split and allows the air- 
conditioning compressor to be run con¬ 
ventionally during combustion engine 
operation via the belt pulley. Depending 
on the actual power requirement, a part 
of the belt drive power yielded can be 
converted into generator power using the 
electric motor. When the internal com¬ 
bustion engine is at a standstill, the elec¬ 
tric motor will drive the air-conditioning 
compressor on its own. Mixed modes 
between these two operating points are 
conceivable, too, for example to operate 
the air-conditioning compressor at opti¬ 
mum speed at all times. The concept also 
allows the use of more efficient compo¬ 
nents instead of a piston compressor. 












Start-Stop 


24 


359 


Conclusion and outlook 


Start-stop systems are among the most 
efficient ways of reducing C0 2 emissions 
when cost-performance considerations are 
taken into account. So far they have the 
highest penetration rate in the EMEA coun¬ 
tries, and this is expected to remain so at 
least through to 2019. Outside of these 
four economies, start-stop technology will 
not play a major role from today’s perspec¬ 
tive. While until now start-stop technology 
was based primarily on the conventional 
starter pinion, it seems advisable to pursue 
concepts that offer greater comfort and 
extend the functionality at low additional 
costs. 

For price-sensitive segments this can 
be done using improved starter pinion 
technology. A starter with an integrated 
gear can already achieve improvements in 
acoustic and vibration behavior. A good 
possibility for automobile manufacturers 
to set themselves apart from the competi¬ 
tion is the concept of a permanently en¬ 
gaged starter pinion. Available at moder¬ 
ate costs, this technology is suitable for 
“change of Mind” situations - an impor¬ 
tant functionality to further increase the 
acceptance of the start-stop system in the 
volume market. 

The belt-driven starter generator satis¬ 
fies even higher demands. Restarting is 
more comfortable and faster. In order to 
tap the full potential of this technology, 
however, a second voltage level with 48 V 
is required. In conjunction with an electric 
drive axle, an enhanced start-stop system 
can significantly reduce the gap to the hy¬ 
brid drive because active sailing as well as 
stop-and-go traffic in E mode are possible 
in this constellation. If start-stop technolo¬ 
gy is to gain acceptance on the North 
American market, then the air-conditioning 
system must continue to operate even with 


the engine shut off. Schaeffler can already 
show initial approaches in this regard. 

Derived from Schaeffler’s observation 
of the market and other insights from the 
field, it can be concluded that the focus on 
C0 2 reduction in Europe is not sufficient to 
ensure the market success of start-stop 
technology. Comfort aspects are equally 
important. In addition, the operating strat¬ 
egy of the start-stop system must be de¬ 
signed so that the driver is able to under¬ 
stand the behavior of the system. 


Literature 


[1] Smetana, T.; Sattler, M.: Who’s Afraid of 48 V? 
Not the Mini Hybrid with Electric Axle! Modular 
electric axle drive in a 48-volt on-board electric 
system, 10 th Schaeffler Symposium, 2014 

[2] Reitz, D.: One Idea, Many Applications: Further 
Development of the Schaeffler Hybrid Module. 
10 th Schaeffler Symposium, 2014 

[3] Stuffer, A.; Stief, H.; Schmidt, T.: Introduction 
of 48 V Belt Drive Systems: New Tensioner and 
Decoupler Solutions for Belt Driven Mild Hybrid 
Systems. 10 th Schaeffler Symposium, 2014 

[4] Kirchner, E.; Eckl, T.: Das Automatikgetriebe als 
Bestandteil einer Start-Stopp-Strategie. ATZ, 
June 2013, Volume 115, Issue 6, pp. 500-505 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 




360 


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362 


Introduction 


As explained in [1], a combination of ad¬ 
equate dynamics and performance as 
well as the best possible efficiency define 
the most important requirements for 
modern actuator systems in the pow¬ 
ertrains of automobiles. Transmission ac¬ 
tuators have a significant influence on the 
size, costs and efficiency of transmis¬ 
sions. The latest electromechanical ac¬ 
tuation systems used in transmissions 
from Getrag, Hyundai and Honda have 
demonstrated very impressively that it is 
possible to produce transmission actua¬ 
tors with an average power consumption 
of less than 20 W and excellent controll- 
ablity and dynamics. The Honda sport 
hybrid i-DCD currently sets the bench¬ 
mark for double clutch transmissions 
with an average power consumption of 
12 W [2]. 

The costs of electromechanical actua¬ 
tors increase with the actuation perfor¬ 


mance at an exponential rate. The reduc¬ 
tion of the actuation energy required to 
actuate is therefore very important for an 
efficient actuator system. The above men¬ 
tioned systems have already achieved the 
development targets for a number of 
transmissions applications [2]; the current 
question is: how said systems can be op¬ 
timized further? 

Upsized functionality and performance 
in a reduced design envelope and 
without significant additional complexity 

Since the current systems have achieved 
market satisfaction with regard to power 
consumption, controllability, dynamics 
and durability, the first priority is not to 
improve these characteristics. Custom¬ 
ers would only accept improvements 
without additional costs. Consequently 
an increase in the functionality and pow¬ 
er density without significant additional 
complexity and cost is the goal for many 



Figure 1 Honda transmission with modular actuator system for sport hybrid applications 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_25, © The Author(s) 2014 




Transmission Actuators 


25 


363 


new ideas, which will be presented. Such 
development targets can be mainly 
achieved by increased mechatronic in¬ 
tegration. 

Reduced complexity and costs without 
a significant loss of performance 

The search is on to find simpler and 
more cost-effective actuation systems, 
particularly for the lower vehicle seg¬ 
ments and relevant cost-sensitive mar¬ 
kets. A reduction of performance and 
driving comfort is only accepted to a 
very limited extent by customers. There¬ 
fore, the restrictions in very special and 
infrequently occurring driving situations 
must be discussed. 


Increased modularity 

In order to minimize the development ef¬ 
fort for complex mechatronic systems as 
much as possible, it is advisable to de¬ 
velop actuator modules and associated 
software modules so that they can be 
used for different tasks in the transmis¬ 
sion. It is important that any additional 
costs and disadvantages regarding pack¬ 
aging due to modularity are offset by more 
intelligent architectures. Synergies be¬ 
tween different actuators can also be 
achieved by developing a modular system 
out of components from existing parts. So 
many electromechanical and electrohy- 
draulic components are availiable in the 
Schaeffler group already, which can be 
used for developing new actuation solu¬ 
tions. 


Modular system 
architecture 


Improved p controllers with increased 
computing power, increased memory, 
higher permissible operating tempera¬ 
tures and new data bus systems and 
sensors now allow system architectures, 
which were inconceivable 10 years ago. 
The use of local electronic units can now 
be seen everywhere in the powertrain. 
For example, throttle actuators, water 
pumps, parking locks or four-wheel 
drive systems are equipped with inde¬ 
pendent intelligence. A trend, which has 
also found its way into the transmission 
sector. 

For the first time, it was possible to ob¬ 
tain the most important advantages in a 
double clutch transmission by fitting a hy¬ 
drostatic clutch actuator (HCA) with a 
first-generation local control unit (LCU) 
[2, 3]: 

- Functional safety without mechanical 
self-opening clutch systems 

- Excellent controllability due to im¬ 
proved local sensors 

- Improved EMC characteristics due to a 
reduced requirement for sensor and 
power cables 

- Freedom to optimize actuator mechan¬ 
ics for a low power consumption. 

The modularity and power density can be 
improved in the next generation by further 
development of LCU technology. 

The dedicated sensors for a specific 
application should be directly connected 
to the relevant LCU to enable the direct 
processing of signals. It is advisable, for 
example, to equip the clutch actuator with 
inputs for the transmission input speed 
sensors in addition to the internal sensor 
inputs for travel, angle and pressure. In 



364 


+ 

i 


Input 

Shaft 

Speed 

Sensoi 

© 


i 


2 2 


Input Output 

Shaft Shaft 

Speed OHTempspeed 
Sensor Sen sor Sensor 

© Q © 


i 


2 2 


Connector 

Connector 

LCU1 

LCU2 

Interface 

HCA1 

HCA2 


Software interface: 
each has a torque interface 


2 J-3 M2 

Connector 



Software interface: gear interface 
as well as free resources for the 
higher level transmission strategy 


Figure 2 System architecture 

combination with the improved memory 
size and processing power of new p con¬ 
trollers, all clutch functions such as soft¬ 
ware adaptations or anti-judder control 
can be calculated locally. A torque inter¬ 
face to higher-level strategies is also ad¬ 
visable in the software. Service functions 
for initial operation and diagnosis can also 
be implemented in local control units in 
the future. 

The gear actuator may also be fitted 
with a local electronics unit and have rele¬ 
vant inputs for additional transmission sen¬ 
sors such as oil temperature, output speed 
and parking lock position. The software 
interface is designed as a gear interface. In 
addition, higher-level driving strategies and 
shifting strategies can be implemented in 
the gear actuator. This takes into account 
the current state of the art, whereby all 
transmission functions are installed and 
can be tested on the transmission. 

This Modular actuators not only sim¬ 
plify the transmission system, they also 


increase availability 
as limp-home strat¬ 
egies can be imple¬ 
mented into each 
individual actuator. 
In addition, the suit¬ 
ability of the actua¬ 
tors for other appli¬ 
cations is signifi¬ 
cantly increased. The 
requirements for 
clutch or gear actu¬ 
ators for different 
applications are very 
similar. For exam¬ 
ple, clutch actua¬ 
tors are also used in 
four-wheel drive or 
hybrid drives. Intelli¬ 
gent gear actuators 
can also be used 
in hybrid transmis¬ 
sions for shifting 
synchronized shifting elements and the 
parking lock device. 

The new actuators are equipped with 
four connector pins for bus systems. 
The can be configured in different ways. 
If an actuator is inserted into an existing 
CAN bus, the bus can be looped (daisy 
chain). Additional pins are not required 
on other control units. If actuators are 
used in transmission systems, it is better 
to use several CAN buses so that the 
transmission remains available if a CAN 
bus fails. 


Increased integration 


The objectives of downsizing complexity 
and upsizing performance are consistent¬ 
ly achieved through further development 
of the hydrostatic clutch actuator. 


CAN1 
4 *— CAN 2 
-Wake-up 

































Transmission Actuators 


25 


365 



► Connector with larger pin assignment 

► More powerful electronics 

► Fully integrated electronics 
Synchronized spindle drive 

► Plausible reference stop 
Extensometer technology omitted 

Figure 3 Further development of the 

hydrostatic clutch actuator (HCA) 

A new pi controller and output stages with 
higher performance as well as a connector 
with an increased number of pins increase 
the functionality, performance and modu¬ 
larity of the clutch actuator without signifi¬ 
cant additional outlay. Some mechanical 


interfaces can be eliminated due to the in¬ 
creased level of integration, saving space 
and reducing costs. The use of a new syn¬ 
chronized planetary roller spindle (SPWG) 
and a verifiable reference position with as¬ 
sociated software intelligence negates the 
requirement for a travel sensor. 

Reduced number of 
power drives 


If a value analysis of the sensors is con¬ 
ducted, it will be found that the brushless 
drives including the relevant electronics, 
sensors and cables form the largest por¬ 
tion of the costs. Although it has been pos¬ 
sible to reduce the costs of the drives by 
making great efforts to reduce the number 
of expensive magnetic components and by 
using LCU technology, the drives still re¬ 
main the cost drivers for electromechanical 
actuators due to increased magnet prices. 
There is however no alternative to EC mo¬ 
tors due to the superior power density, dy¬ 
namics and operating life required for high- 
ly-dynamic actuators. In addition to the 
selection of the most economical and suit¬ 
able design for the drives, the question 
also arises as to whether a double clutch 
or hybrid transmission requires a drive for 
each actuator. 


One actuator for two clutches 

For reasons relating to the design envelope 
and costs, consideration can be given as to 
whether one HCA can supply more than 
one engagement system, for example two 
clutches or one clutch and a sub-transmis¬ 
sion. This may be possible if the two en¬ 
gagement systems do not require actuation 
energy at the same time. 









366 


Valve assembly 



Valve ▼ CO Hydrostatic 

control Cl clutch actuator 


MM— 


A/W\-SfeHmH 

lyTju 

lIi 




E~ 


EC motor + 
transmission 


Figure 4 HCA with seat valves and two 
consumers 

Even with a double clutch transmission it 
is justifiable to ask the question whether 
the transmissible torque must be in¬ 
creased on both clutches at the same 
time. If anything, this is only required in a 
few special cases, which can be solved 
by suitable calibration, such as tip-in dur¬ 
ing an overlapping phase, or the genera¬ 
tion of drag torque on the inactive clutch 
for resolving balked gears or for damping 
the powertrain. This approach is more 
suitable for other applications, such as in 
hybrid vehicles with a CO clutch for en¬ 
gine decoupling and a Cl clutch for 
starting. The stroke volume can be di¬ 
vided between a number of consumers 
by means of valves. Long actuation peri¬ 
ods are critical due to the limited volume 
in one stroke if the system cannot be 
sealed sufficiently. However, the required 
actuation time must be ensured because 


so-called seat valves are also not totally 
leak-proof. “Replenishment strategies” or 
a double stroke HCA could improve func¬ 
tionality. 


Bi-rotational pump 

It can be seen that even highly efficient ac¬ 
tuator systems such as the HCA reach 
their limits due to limited stroke volumes. 
Especially if there is leakage in the engage¬ 
ment system or multiple slave cylinders are 
supplied by one actuator. In such cases, 
consideration should be given to actuator 
systems, which provide a continuous vol¬ 
ume flow. 

The Bi-rotational pump is the current 
state of the art [4, 5]. However, this type of 
system is only advisable if the intrinsic dis¬ 
advantages, in terms of the power con¬ 
sumption, are accepted. With this tech¬ 
nology a power consumption smaller than 
20 W can not be reached in a double 
clutch transmission. Leakage in the clutch 
system, for example on the feedthroughs 
for rotating pressure pistons, can be com¬ 
pensated by using low-pressure pumps, 
although this has a negative effect on the 
space requirements for the hydraulic lines 
and slave pistons. It is more advisable to 
eliminate leakage in the clutch system in 
order to facilitate further efficiency in¬ 
creases in transmissions. Reducing pow¬ 
er consumption and leakages will achieve 
the optimum in terms of efficiency, costs 
and package requirements. The use of a 
local control unit for the Bi-rotational 
pump is effective for the above mentioned 
reasons. 

An additional design example for a 
Bi-rotational pump is the variable dis¬ 
placement pump actuator, which allows 
a force-dependent nonlinear ratio. The 
core of this actuator is an electrically- 
driven pump with a variable delivery 
stroke. Direct coupling of the delivery 













Transmission Actuators 


25 


367 



Figure 5 Variable displacement pump actuator 

stroke varation mechanism with the gen¬ 
erated load ensures needs-based actua¬ 
tion: 

- Long delivery stroke at low pressure for 
bridging free travel. 

- Reduced delivery stroke and thus re¬ 
duced pump torque at higher pressure 
for torque transmission. 

A pump with the lowest possible level of 
leakage, for example in the form of an 
axial piston pump, and the incorporation 
of specific load-dependent friction char¬ 
acteristics, show promising results in ini¬ 
tial simulations. The limit of the possible 
actuation force can be significantly in¬ 
creased without any requirement for 
adapting the electric motor and its elec¬ 
tronics with only moderate disadvantages 
with regard to energy consumption com¬ 
pared to an HCA. This type of actuator 
could be used in conjunction with seat 
valves for various loads as a hydrostatic 
actuator with limited stroke volume. 


Active interlock transmission 
actuator with one power drive 

The electromechanical active interlock gear 
actuator, which LuK launched on the mar- 



Figure 6 Active interlock actuator for up to 
10 shifting elements with actuation 
of the parking lock 














































368 



EC motor 


One-way clutch 


Integrated local 
transmission control unit 


Cam mechanism 


Figure 7 Single-motor gear actuator 


ket in the Hyundai 6-speed DCT in 2011, is 
already a power-on-demand system with a 
very low energy consumption [6, 7]. 

The actuator system has been further 
developed so that five shift rails instead of 
the previous four can be actuated. In total 
ten shifting elements can be actuated with 
the new actuator. The active interlock gear 
actuator is already used for actuating the 
parking lock in some volume-production 
applications [2]. The functionality of the ac¬ 
tuator was significantly increased without 
considerable additional comlexity and 
costs. 


LuK presented a gear actuator in 2006, 
which was equipped with only one drive 
motor instead of a select motor and a shift 
motor [8]. A new single-motor gear actuator 
was developed on the basis of previous 
findings from the successful double-motor 
gear actuator and the new targets for in¬ 
creased modularity from the system archi¬ 
tecture. 

Two simple mechanical elements - a 
one-way clutch and a cam mechanism 
- are used to realize the full select and 
shift functionality with only a single mo¬ 
tor. 



Transmission Actuators 


25 


369 


5 directional control valves 2 directional control valves 



Figure 8 Comparison of hydraulic gear actuators 

Clever use of these two elements allows 
the movement for engaging gears to be 
assigned to the motor’s first direction of 
rotation. The unwanted gears are firstly 
disengaged automatically in the active in¬ 
terlock actuator. The second direction of 
rotation is responsible for returning into 
the neutral position and subsequent se¬ 
lection movements. 

The elimination of one motor frees up 
design space that is used for integrating a 
local transmission control unit. 

Both the single-motor gear actuator and 
the new hydrostatic clutch actuator (HCA) 
show how increased integration can reduce 
costs while maintaining or increasing the 
level of functionality. 

The hydraulic active interlock 
transmission actuator (HGA) 

Hydraulically-actuated double clutch trans¬ 
missions use gear actuators for actuating 
shifting elements in the transmission, which 
shift two gears at a time by means of indi¬ 
vidual gear actuator pistons. If there are 


more than nine shifting elements, five single 
gear actuator pistons each with one travel 
sensor are required, which on hydraulic sys¬ 
tems must be controlled by a directional 
control valve in order to engage the relevant 
gear. 

If the active interlock principle of the 
electric motor drive is used, it would be pos¬ 
sible to reduce the number of shift axes 
from five to two, i.e. one shift and one selec- 



Magnet for angle sensor 

Figure 9 Hydraulic active interlock gear 
actuator (HGA) in detail 


370 


tor axis. Instead of using electromechanical 
drives, these two axes can still be actuated 
hydraulically, for example, using an axial 
and a swivel unit. The complexity could 
be halved while maintaining a similar level 
of performance. The 
HGA is therefore 
an excellent exam¬ 
ple of how com¬ 
plexity and costs 
can be significantly 
reduced through in¬ 
novative ideas. 

The main advan¬ 
tages of this trans¬ 
mission actuator are 
the compactness 
and the modularity 
compared to large 
conventional gear 
actuator units. The 
parameters for dif¬ 
ferent actuation val¬ 
ues, for example, 
the gearshift force 
or dynamics, can be 
set by changing the 
pressure and vol¬ 
ume flow. 

Of course, it 
would also be con¬ 
ceivable to use a dif¬ 
ferent system than 
the central hydraulic 
system for actuation 
of the HGA. Two 
separate hydrostatic 
clutch actuators 
(HCA) or two bi- 
rotational pumps 
could also be con¬ 
sidered. The above 
mentioned solutions 
with only one actua¬ 
tor and a number of 
seat valves are also 
feasible, because 


selecting and shifting movements never take 
place at the same time, and no continuous 
force must be applied. Especially if there is 
no central hydraulic system for controlling the 
HGA, for example, in a hybrid transmission. 


EBEEB m 


Power Unit 


«Tasb 

Actuate 1 


«Tasb 

Actuate 2 


Gearbox 


l 1 

eTasb 


l i 
l 1 

Shift 


1 ■ 



1 l 

! 


J 1 

aTasb 



Select 






J 



Dual Clutch 


cTasb 

Engage 1 


sTasb 

Engage 2 

— m — 


——* - 


L . . r 


111 "Aill-Ul 


Power Unit 


«Tasb 

Actuate 1 


«Tasb 

Actuate 2 


Gearbox 


sTasb 

Shift 


«Ta$b 

Select 


Dual Clutch 


«T3sb 

Engage 1 


((Tasks 

Engage 2 





Power Unit 


<(Tasb 

Actuate 1 


«Tasb 

Actuate 2 


Gearbox 


«Task» 

Shift 


«Tasb 

Select 


Dual Clutch 


sTasb 

Engage 1 


«Tasb 

Engage 2 

- m - 




Power Unit 


flTasb 

Actuate 1 


ffTasb 

Actuate 2 


Gearbox 


eTa$b 

Shift 


eTa$b 

Select 


Dual Clutch 


slash 

Engage 1 


*Tasb 

Engage 2 





Figure 10 Excerpt from the functional analysis of a double clutch 
transmission 
























































































































































































Transmission Actuators 


25 


371 


Double clutch 

transmission actuated with 
two actuators 


The double-motor DCT (2M DCT) 


clutch transmission, it is highly advisable to 
assign one drive to the active clutch and 
the other drive to preselect the gear in the 
inactive sub-transmission. Both actuators 
are subsequently available for transferring 
the torque from one clutch to the other. 
One sub-transmission with the associated 
clutch can therefore be actuated by only 
one drive. 


As previously shown, there are solutions for 
actuating two clutches with one actuator or 
performing selection and shifting opera¬ 
tions with one power drive. Only two elec¬ 
tromechanical drives are therefore required 
for a double clutch transmission. To ensure 
that the restrictions in terms of time are min¬ 
imized, it is important to carry out a system¬ 
atic analysis of when each drive is required 
and what disadvantages could occur if in 
exceptional cases actuator functions are 
performed in series instead of in parallel. 

Because more than two drives are nev¬ 
er required at the same time in a double 


Double-motor DCT with combined 
shift and clutch actuation drums 

The basic structure of this DCT comprises 
two sub-transmissions, which are operat¬ 
ed by two sub-shift drums connected to 
each other by means of a gear stage. The 
gear stages are each driven by an electric 
motor, which either generates a selection 
movement or shifting movement with the 
subsequent sequential clutch actuation for 
the relevant sub-transmission depending 
on the direction of rotation. Each clutch is 



Figure 11 Structure of the 2M DCT with combined shift and clutch actuation drums 



































































372 



Figure 12 2M DCT pin control system and angle-controlled one-way 
clutch 


connected to the relevant shift drum drive 
via an angle-controlled one-way clutch, 
which enables clutch actuation according 
to the position on the circumference of the 
shift drum. 

The shift drums can rotate freely in the 
direction of selection movement by means 
of a three-dimensional ramp geometry and 
a sprung gearshift fork pin until the circum¬ 
ferential position of the required gear is 


reached (selection 
operation). If the 
shift drum is rotated 
in the opposite di¬ 
rection, the shift 
fork pin slides along 
the diagonal gear 
ramp surface and 
engages the gear 
(gearshift opera¬ 
tion). 

In this moment, 
the angle-controlled 
one-way clutch 
reaches its locking 
position and moves 
the clutch actuating 
lever with the gear 
engaged by means 
of an additional ro¬ 
tation of the shift drum. The one-way clutch 
function with angle control is realized in the 
clutch actuating lever by means of a pin that 
is free to move radially and is preloaded in 
the shift drum shaft and a one-way clutch 
ramp that is matched to the shift drum pro¬ 
gram. The pin cannot be moved in the shift 
drum shaft in the direction of locking and 
therefore transmits the torque to the clutch 
actuating lever. 


Reservoir 




Slave cylinder with 
release bearings 

Integration of 2M DCT actuators into a transmission 


Master cylinder 


Clutch lever 


Figure 13 


Sub-gearbox 1 


Sub-gearbox 2 








Transmission Actuators 


25 


373 


The clutch actuating levers press on the 
push rods of two hydrostatic master cylin¬ 
ders, which, in turn, operate two slave cylin¬ 
ders arranged coaxially relative to the trans¬ 
mission input shafts. In conjunction with the 
integrated engagement bearings, the mas¬ 
ter cylinders and the slave cylinders form a 
compact engagement system, which is 
completely enclosed within the clutch hous¬ 
ing. The shift drums and gear train are lo¬ 
cated in the normal positions for transmis¬ 
sions equipped with shift drums. Both 
drives with the integrated electronics and 
the hydraulic fluid reservoir are located in a 
suitable position on the transmission. 

Integration variants of 2M DCT 
actuators 

2M DCT actuators have already been imple¬ 
mented for different DCT gear sets with only 
minor changes to the gearshift forks and 
housing. 


However, it must be stated that an opti¬ 
mum solution for this type of system can 
only be found in very close collaboration 
with the transmission manufacturer. Com¬ 
promises must frequently be found with 
regard to the design envelope, costs and 
functions. 

2M DCT with bi-rotational pumps 
and HGA 

High-pressure bi-rotational pumps can be 
installed in conjunction with an HGA as 
another system for using the two direc¬ 
tions of rotation of an electric motor for 
two different subfunctions in the trans¬ 
mission. One of the two bi-rotational 
pumps is assigned to each clutch. This 
means the clutches can be actuated inde¬ 
pendently. The second pressure connec¬ 
tion (in a reversing direction) is connected 
with the HGA using a simple valve logic. 
This arrangement ensures a very high 


Gear shifter 



Gear shifter 



x 

r 


i 


j 


X 

r 


IX 


u 


o 



Reserving pump 1 




Reserving pump 2 




Figure 14 2M DCT variant with bi-rotational pumps and HGA 













































































374 


level of functionality of the transmission 
with the reduced number of electric 
drives. The HGA divides the available en¬ 
ergy via two valves into the required stroke 
for selection and the swivel movement for 
gearshifts. 

Modular system of 
components 


Control units, electric motors and 
sensors 

The Schaeffler Group now has a wide range 
of modules for electronic components due 
to its many years of intensive development 
of electromechanical actuators for trans¬ 
mission systems. 


Mechanical components 

With the expertise and experience gained 
with regard to power screws and nuts 
made of different materials as well as 
screw drives with optimized friction be¬ 
havior such as KGT, PWG and SPWG, 
Schaeffler has innovative drive compo¬ 
nents for actuators, which can be used for 
specific applications. 

The latest developments are: 

- The synchronized planetary roller spin¬ 
dle (SPWG), a spindle drive with a con¬ 
stant pitch and very high power den¬ 
sity. 

- A verifiable reference position, which 
is detectable by the software and can 
be clearly differentiated from the stiff¬ 
ness or blocking of the system. 

- The integrated, friction spring band, 
which prevents the actuator snapping 
open in case of a fault (2r| band). 



Control unit for 
2 BLDC motors, 
chassis attachment 



LCU for EC motor, 
actuator integrated 



Universal control unit 
for EC/DC motor, 
transmission attachment 


Universal 
BLDC motors, 
hall sensors 



EC motor rotor/stator modules 


Drive modules; 
LCU + 

EC motor 




LCU stator unit 
' for EC engine 



Angle sensors (inductive, hall, 

(GMR, Hall, AMR) LVDT > 




Figure 15 Control units, electric motors and sensors for transmission actuators 



Transmission Actuators 


25 


375 


Angular contact 
bearing spindle 
drive 


Planetary roller drive 

i ; __ —, (PWG) 


rwnn 

I / i 


Screw-type 
drive (multiple) 



Roller screw drive 
(RGT) 



Screw-type drive 

sealed, 

self-locking 



Planetary roller drive 
synchronized (SPWG) 



Wrap-spring 
eccentric disc 
drive, 

self-locking 



Formed spring 
ball drive 



Ball screw 
drive (KGT) 



Screw-type drive 

Plastic, 

self-locking 



Figure 16 Mechanical components 


Planetary roller drive 
synchronized (SPWG) 
direction-based friction 
(2n belt) 



Eccentric 
spindle drive 


New compact actuator for the 
automation of clutches 

Precise and individual new actuators can 
be developed from the modular compo¬ 
nents by using new design and simulation 


tools. Reliable statements can be made at 
an early stage with regard to the system 
behavior and operating life. 

Such an example is the new compact 
actuator. Innovative mechanical compo¬ 
nents such as the synchronized planetary 



Figure 17 Compact actuator for the automation of clutches 




























376 


roller spindle (SPWG), 
the verifiable refer¬ 
ence stop and the 2r\ 
band were integrated 
into a compact and 
modular actuator 
with new electronic 
components. This 
actuator seems par¬ 
ticularly suitable for 
MTs, AMTs, hybrid 
drives and electric 
clutch applications. 


Actuators 
for hybrid 
applications 


Electric central release bearing (EZA) Electric axle actuator (EAA) 



Formed spring band 


Push-pull 
Release bearing 


Rotor 


Stator 


Sensors 


BLDC 

motor 


Pinion with 
crown wheel 


Spindle 
with KGT 


Gear shift 
fork 


Different actuators Figure 18 Electric clutch actuator (ECA) and electric axle actuator (EAA) 


are also required in 

vehicles with new types of hybrid and elec¬ 
tric drives. In contrast to conventional 
transmissions, electromechanical power- 
on-demand actuator systems are usually 
required in these applications. Many appli¬ 
cations can be operated with available ac¬ 
tuators. The lever actuator [6, 8, 9] or HCA 
[2, 3, 6] is used in such powertrains for 
clutch actuation or a gear actuator is used 
for activating shifting elements and the 
parking lock. For special applications, for 
which available actuators are not suitable, 
new actuators can be quickly and eco¬ 
nomically developed for the special re¬ 
quirements from the modular system of 
components mentioned above. 

The electric clutch actuator (ECA) for 
controlling disconnect clutches in hybrid 
modules can be completely integrated with 
the clutch into the hybrid module [10, 11]. 
The associated power electronics (ACU = 
actuator control unit) are also mounted on 
the hybrid module. A highly-integrated hy¬ 
brid module with simple interfaces is cre¬ 
ated. 


The electric axle actuator (EAA) is used for 
shifting into the neutral position and shift¬ 
ing gears in electric drives. The electric 
motor and ball screw drive (KGT) are com¬ 
ponents from the modular system. The 
EAA can also be controlled by the actuator 
control unit (ACU) mounted on the electric 
axle. 


Outlook 


The actuation future lies in even more highly- 
integrated and intelligent modules [9]. If actua¬ 
tors act directly on the pressure plate of a rotat¬ 
ing clutch, they can be designed much smaller 
because the required actuation energy is very 
low due to the significant reduction in the ac¬ 
tuation path. There is still a long way to go until 
actuators and electronics are available for the 
challenging conditions in a clutch. LuK has set 
itself this specific target. The basic function of 




Transmission Actuators 


25 


377 


transmitting electrical energy and signals and 
the function of small electric motors or piezo 
actuators has already been proven. 

The current state of the art for double 
clutch actuation uses four actuators. This 
paper has presented many solutions with 
two actuator drives for a double clutch 
transmission. The provocative question is: 
can the number of drives be further re¬ 
duced to one, or even zero? One idea to 
obtain such an actuation solution has 
shown promising initial results by tapping 
into the energy directly from the rotating 
powertrain. 

Of course, there are a range of technical 
challenges to overcome before such ideas 
can be implemented. However, these ideas 
must be investigated in more detail in the 
near future. 


Summary 


Currently known electromechanical and 
hydraulic transmission actuation systems 
are being continuously developed while tak¬ 
ing into account new boundary conditions, 
but above all using new technologies in me¬ 
chanics and electronics. The discussed de¬ 
velopments and ideas show that both a 
significant reduction in complexity and the 
provision of increased functionality are pos¬ 
sible. The watchwords here are intelligent 
system architectures, modularity, and in¬ 
creasing integration on the basis of a versa¬ 
tile modular system of components. 


Literature 


[1] Wagner, U.; Zink, M.; Feltz, C.: Double Clutch Systems 

- Modular and Highly Efficient for the Powertrain 
of Tomorrow, 10 th Schaeffler Symposium, 2014 

[2] Mueller, B.; Ubben, H.; Gantner, W.; Rathke, G,: 
Efficient Components for efficient Transmis¬ 
sions. CTI Symposium, 2013 

[3] Mueller, B.; Kneissler, M.; Gramann, M.; Esly, N.; 
Daikeler, R.; Agner, I.: Smaller, Smoother, Smarter- 
Advance development components for double clutch 
transmissions. 9 th Schaeffler Symposium, 2010 

[4] Faust, H.: Future Requirements for Transmission 
Benchmarking, GETRAG Drivertrain Forum 2012 

[5] EP 1 236 918 B1, Clutch system with a pump 
actuated clutch, ZF Sachs AG 2001 

[6] Wagner, U.; Buhrle, P; Muller, B.; Kimmig, K.-L; 
Kneissler, M.: Dry double clutch systems - In¬ 
novative components for highly-efficient vehicle 
transmissions, ATZ 11/2009, pp. 826-833 

[7] Kimmig, K.; Rathke, G.; Reuschel, M.: The Next 
Generation of Efficient Dry Double Clutch- 
Systems, VDI Congress, 2013 

[8] Wagner, U.; Berger, R.; Ehrlich, M.; Homm, M.: Elec- 
tromotoric actuators for double clutch transmissions 

- Best efficiency by itself; 8 th LuK Symposium, 2006 

[9] Wagner, U.; Muller, B.; Henneberger, K.; 

Grethel, M.: What makes a transmission oper¬ 
ate - Tailored actuation systems for double 
clutch transmissions; CTI Symposium, 2011 

[10] Stopp, R.: P2 Hybrid Module - Modular concept 
to hybridize all modern automatic transmission, 
VDI Tagung, Innovative Fahrzeugantriebe, 2012 

[11] Mitariu, M.: Herausforderungen und innovative 
Losungen der nachsten Hybridmodulgenera- 
tion, VDI Tagung, Kupplungen und Kupplungs- 
systeme in Antrieben, 2013 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 




378 


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379 


Light, Compact and Efficient 

Schaeffler differential systems set the pace 


Thorsten Biermann 


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380 


Introduction 


The Schaeffler lightweight differential was pre¬ 
sented for the first time at the 2010 Schaeffler 
Symposium in Baden-Baden, Germany [1]. At 
the time, the innovative aspect of the design 
focused on the reduced weight and smaller 
mounting space required for the ground¬ 
breaking differential concept. Since then, the 
lightweight differential has been further opti¬ 
mized in order to overcome the last of the 
concept’s drawbacks in comparison to the 
existing bevel gear differential. During the 
optimization phase, the key focus has been to 
improve the rigidity of the differential, and re¬ 
duce frictional losses in the main bearing sup¬ 
port. A further aim was to reduce the produc¬ 
tion costs. Today, this means that there is 
virtually no effect on costs, at least when the 
differential is operated within high torque 
ranges. For those who have not yet come 
across the Schaeffler lightweight differential, 



Figure 1 The father of the spur gear 

differential “Alexander T. Brown” 


what follows is a brief explanation of how the 
component came into existence. 

Based on the most recently available 
sources, Alexander Timothy Brown can be de¬ 
scribed as the father of the “spur gear differen¬ 
tial”, a category of differential that includes the 
Schaeffler lightweight differential. Born in 1854, 
Brown quickly developed into a technical all- 
rounder as well as someone “with consider¬ 
able inventive talent”. Yet Brown is not known 
for his early designs for guns or even typewrit¬ 
ers - instead, he is known for inventing the first 
pneumatic tire for automotive vehicles, which 
he patented on December 20,1892. 

The tire wear caused by driving around 
bends - a problem that Brown faced when 
designing his new tires - may be the reason 
behind his invention of a new type of differen¬ 
tial. US patent no. 691591 was granted on 
January 21, 1902 (Figure 2). The patent relat¬ 
ed to a design variant for a new differential as 
an alternative to the existing bevel gear differ¬ 
ential. Brown’s solution was to omit bevel 
gears in favour of spur gears - a concept 
made possible by designing the differential as 
a planetary transmission. In contrast to a con¬ 
ventional planetary transmission consisting of 
a sun gear, planet gears and a ring gear, 
Brown’s spur gear differential did not feature 
ring gears. Instead, the transmission featured 
two output sun gears. Pairs of planet gears 
were then arranged around the circumference 
of the output sun gears. At all times, one plan¬ 
et gear would be meshed with the left-hand 
sun gear while the other planet gear would be 
meshed with the right-hand sun gear. The 
planet gears meshed with one another in the 
free area that remained between the gearing 
of the two sun gears. 

The new design ensured that torque was 
distributed symmetrically to the wheels while 
still using the same number of teeth on the 
sun gears and planet gears. The symmetrical 
design also boasted a high proportion of iden¬ 
tical parts. The closed design of the differen¬ 
tial suggests that it was intended for use as a 
differential arranged coaxially to the rear axle. 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_26, © The Author(s) 2014 



Differential Systems 


26 


381 


Ho. 69l,«i. Plttntifl Jill. 21, 1902, 

A. T. BStOWH. 
flUHHL 



WITNESSES; 


BY 

\ ATTOflNEYS 


Figure 2 Excerpt from the patent specification 
for the first spur gear differential 
from 1902 

The Schaeffler lightweight 
differential 


The Schaeffler lightweight differential shares 
some common features with Brown’s origi¬ 
nal design. Just as Brown did, the develop¬ 
ers at Schaeffler decided to use three pairs 
of planet gears in the basic variant (Figure 3). 

This design ensures that forces are dis¬ 
tributed evenly across the individual contact 
points of the gearing, regardless of the manu¬ 
facturing tolerances. The only issue was 
that the axial mounting space required for 


the original design led to problems when 
trying to integrate the differential into mod¬ 
ern transmissions. With most present-day 
transmissions that are arranged transverse¬ 
ly at the front of the vehicle, there is less 
mounting space directly next to the drive 
gear due to the design of the bearing seat 
for the output shaft or due to the position of 
the transmission gears. In comparison to 
Brown’s spur gear differential, this problem 
meant that a more axially compact solution 
had to be found. The engineers were able to 
find the solution by further developing the 
compensation gearing of the differential. In 
contrast to Brown’s design, the Schaeffler 
lightweight differential uses two sun gears 
in different sizes [2]. As one planet gear from 
each pair of planet gears is now arranged 
around a larger pitch diameter, it has been 



Figure 3 Lightweight differential 















382 


US 691,591 US 6,480,532 B2 




Figure 4 Transmission diagrams of the 

differential variants by Brown and 
Schaeffler 

possible to shift the contact point between 
the gearing of the planet gears to the small¬ 
er sun gear. In contrast to Brown’s patent, 
the Schaeffler lightweight differential there¬ 
fore features only two levels of gearing in¬ 
stead of three. This design means a signifi¬ 
cantly smaller axial mounting space is 
required. The differential remains in the 
mounting space of the drive gear and can 
replace the previously used bevel gear dif¬ 
ferential without damaging any surrounding 
structures (Figure 4). 

Figure 5 shows the gearing as viewed 
from the side and reveals the idea behind 
the new design for the sun gears. 



Figure 5 Side view of the compensation 
gearing 


There is a difference in size between the 
two sun gears, despite the fact that they 
have the same number of teeth. This dif¬ 
ference is due to the differing profile dis¬ 
placement of the gearings. The smaller 
sun gear has an extremely negative profile 
displacement. The area of the involute 
found directly at the base circle of the 
gearing is used. The larger sun gear has 
an extremely positive profile displace¬ 
ment. Here, the area of the involute that is 
furthest away from the base circle is used. 
The teeth of the larger sun gear have a 
pyramid-shaped cross section with a 
wide tooth root. On the smaller sun gear, 
the tooth root is comparatively narrow. 
This design leads to a higher load being 
placed on each tooth root on the smaller 
sun gear. The smaller sun gear must 
therefore be slightly wider than the larger 
sun gear. As both sun gears have the 
same number of teeth as well as the same 
module, both gearings have the same 
base circle diameter. The same tangential 
forces are applied, meaning identical 
torques are produced at the two sun 
gears. Despite the asymmetric design of 
the compensation gearing, the torque is 
therefore distributed symmetrically to the 
side shafts. This being the case, equation 1 
applies to the internal transmission of the 
differential: 

z n z 79 z 0 z 9 

• pi 1 _ pi 2 . su 2 su 2 

z , z n z z , 

su 1 pi 1 pi 2 su 1 

Nevertheless, the number of teeth on the 
planet gears does not necessarily need to 
be identical, as they cancel each other out 
in the equation. In fact, the number of teeth 
on the narrower planet gears can therefore 
be slightly larger in order to optimize the 
contact point between the gearing of the 
planet gears. For example, a larger number 
of teeth can increase the contact ratios 
without the radial mounting space having to 
be enlarged. 




















Differential Systems 


26 


383 


Alternatively, there is also a solution that 
uses different numbers of teeth on the sun 
gears. In this case, the difference in the 
number of teeth must be calculated such 
that the system can still be mounted. In con¬ 
trast to profile displacement, a change in 
the number of teeth on the sun gears actu¬ 
ally has an impact on torque distribution, 
meaning that any adjustment to the number 
of teeth must be matched by a correspond¬ 
ing transmission between the planet gears. 
This is only made possible by using a 
stepped planet gear. 

In equation 2, only the number of teeth 
on the second planet gear is canceled out. 
A possible solution for a differential with 
three pairs of planet gears would be to have 
one sun gear with 36 teeth, a second sun 
gear with 33 teeth, and a stepped planet 
gear with either 11 or 12 teeth. 


i= _ 

7 

^ p/2 

.Z-2. 

= hsL 


Z S u2 _ 

Z S ul 

7 

^pllb 

Zpi 2 

Z S ul 

Zpllb 

Z su ] 


Another possibility would be to use different 
modules in the compensation gearing of the 
differential. This would be possible using 
the same or a different number of teeth on 
the sun gears. The latter variants perform 
worse than the former variants at least in 
terms of costs at the present time due to the 
stepped planet gears. As a result, the latter 
variants are not currently being pursued. 

What all of these variants have in com¬ 
mon, however, is their extremely narrow de¬ 
sign in comparison to the existing gearings 
of bevel gear differentials. This narrow de¬ 
sign is primarily a result of the increased 
number of gearing contact points. On the 
bevel gear differential, there are four gearing 
contact points between the differential pin¬ 
ions and the output bevel gears as stan¬ 
dard. In contrast, torque on the lightweight 
differential featuring three pairs of planet 
gears is transferred to the two sun gears via 
three contact points each, creating a total of 
six gearing contact points. 


In addition, a fundamental mechanical law 
also has an effect on the lightweight dif¬ 
ferential: Torque equals force times the 
length of the lever arm. On the spur gear 
differential and the bevel gear differential, 
the distance of the gearing contact point 
to the center of the differential is equal to 
the length of the lever arm. As the sun 
gears on the lightweight differential have a 
significantly larger gearing diameter than 
the bevel gears, the gearing forces are 
significantly reduced while maintaining the 
same torque. In conjunction with the num¬ 
ber of gearing contact points, the relation¬ 
ship between the two diameters allows for 
a comparatively delicate gearing design. 
Such an optimum layout and design for a 
compensation gearing with a high level of 
power density is an essential aspect of 
the new differential design variant from 
Schaeffler. 

Another key focus that required sev¬ 
eral development loops was the design of 
the differential housing and the bearing 
support. It was important to design the 
housing such that a high level of rigidity 
could be achieved at the gearing contact 
point of the drive gear, as well as a signifi¬ 
cant reduction in the amount of friction at 
the main bearing support in comparison 
to a bevel gear differential. At the same 
time, the reshaped bracket of the differen¬ 
tial housing must not be exposed to high 
levels of stress. However, a fundamental 
issue stood in the way of these objectives: 
a significantly reduced distance between 
the bearings in comparison to the bevel 
gear differential. The following application 
examples show how it was possible to 
take this problem - which at first appeared 
to be a serious disadvantage - and trans¬ 
form it into an advantage. 



384 



Figure 6 CVT with and without lightweight differential 


Current developments 


Optimizing a CVT 

Figure 6 shows a continuously variable 
transmission (CVT) before and after replac¬ 
ing the traditional bevel gear differential with 


Angular contact 
ball bearing 
(left) 


0 55.0 x 0 83.0 x 15.5 


C = 28.0 kN 
CO = 25.5 kN 


a lightweight differential. Despite the rela¬ 
tively low torque capacity totaling a rated 
2750 Nm at the axis, the lightweight differ¬ 
ential boasts a weight saving of approxi¬ 
mately 1.1 kg. The lightweight differential has 
a total weight of 5 kg yet the strength of the 
housing and gearing has been increased. 

The main difference is the change in 
bearing type and bearing support in com¬ 
parison to the bevel gear differential. Instead 
of an X arrange¬ 



Angular contact 
ball bearing 
(right) 

0 55.0 x 0 83.0x15.5 

C = 28.0 kN 
CO = 25.5 kN 


Friction 


M diff. in Nm 


2 kN @ 80 °C 

50 

100 

500 

750 

1,000 

s 

o 

2,000 


40 

-0.297 

-0.293 

-0.221 

-0.220 

-0.192 

0.001 

0.386 

£ 

E 

c 

100 

-0.319 

-0.316 

-0.244 

-0.241 

-0.213 

-0.021 

0.356 

200 

-0.347 

-0.345 

-0.271 

-0.271 

-0.243 

-0.057 

0.311 

400 

-0.393 

-0.392 

-0.315 

-0.317 

-0.293 

-0.118 

0.229 

*= 

600 

-0.429 

-0.428 

-0.351 

-0.354 

-0.333 

-0.166 



1,000 

-0.488 

-0.488 

-0.404 

-0.410 

-0.402 



£ 

1,400 

-0.536 

-0.534 

-0.447 

-0.453 





1,900 

-0.588 

-0.586 

-0.493 






M_diff.: Moment of differential n_diff.: Speed of differential 

Figure 7 Reduced drag torques through the use of angular contact ball 
bearings in an O arrangement 


ment using tapered 
roller bearings, two 
angular contact ball 
bearings are used in 
an O arrangement. 
As such, it has been 
possible to design 
the bearing support 
on the lightweight 
differential in an 
extremely efficient 
manner with regard 
to friction. At the 
same time, a long 
service life as well as 
a high level of rigidity 
have also been 
achieved. In Figure 7, 
the torque-depen¬ 
dent frictional power 
values of an opti¬ 
mized bearing sup- 




























Differential Systems 


26 


385 


port are shown in comparison to a bevel gear 
differential. The green areas indicate the load 
scenarios in which the bearing support of the 
lightweight differential performs better in 
terms of friction in comparison to the bevel 
gear differential. 

In the most common load scenarios, fric¬ 
tion savings of up to 80 % can be achieved, and 
the ball bearing support achieves an extremely 
high level of efficiency, even in the partial load 
range. This partial load range represents a key 
focus of conventional fuel consumption cycles. 
So in this application example based on the 
New European Driving Cycle (NEDC) it is theo¬ 
retically possible to achieve fuel consumption 
savings of up to 0.35 g of C0 2 /km in addition to 
the weight saving. 

Optimizing a manual front transverse 
transmission 

The second application example (Figure 8) 
shows a manual transmission arranged 
transversely at the front of the vehicle, with 



Figure 8 Manual transmission with lightweight 
differential 


a rated torque at axle of approximately 
6500 Nm. Even at high torques upwards of 
6000 Nm, thanks to the massive weight sav¬ 
ings despite the greater number of compo¬ 
nents it is possible to eliminate any impact on 
costs in comparison to many existing bevel 
gear differentials. This remains true as long 
as similar volumes are produced. 








Figure 9 The bevel gear differential versus the lightweight differential: The red lines show the contact 
angle of the bearing support 



























































































386 


The reason behind this cost benefit lies in 
the fundamentally similar production meth¬ 
ods used for the two differentials. The com¬ 
pensation gearing of the differential is ex¬ 
truded and the housing parts are 
deep-drawn while avoiding any machined 
rework wherever possible. In addition, the 
cold metal sheet forming techniques in use 
entail relatively low levels of energy con¬ 
sumption in comparison to traditional cast¬ 
ing techniques. 

Another reason behind this cost benefit 
is that a higher number of components may 
be required for the bevel gear differential 
in some cases: At high torques, two 
differential pinions are often no longer suffi¬ 
cient to transfer the gearing forces. Accord¬ 
ingly, the number of differential pinions is 
increased, which, in turn, requires a larger, 
circumferentially closed housing design. 

In the present example, the weight of 
the bevel gear differential including the ta¬ 
pered roller bearing support and drive gear 
is equal to 13.4 kg. The differential housing 
must be divided to facilitate the assembly of 
four differential pinions. At maximum torque 
peaks, the differential gearing generates ex¬ 
pansion forces of more than 100 kN, which, 

Bevel gear vs. Spur gear 

differential differential 

in X arrangement in O arrangement 



BGD SGD 


■ Bearing 

■ Differential 

Figure 10 Rigidity measurement 


in addition to the torque, act upon the screw 
connection between the differential cage 
and the axle drive gear. Despite the high op¬ 
erating weight, there is therefore no real po¬ 
tential to increase the life of the overall sys¬ 
tem. As a result, it is difficult to imagine 
increasing the torque or even reducing the 
weight. 

Despite the high torque, the developers 
at Schaeffler increased the number of pairs 
of planet gears in the lightweight differential 
to a total of four in order to keep the differ¬ 
ential under the mounting space of the drive 
gear. Both halves of the housing are pressed 
completely into the drive gear and are rivet¬ 
ed at four points between the pairs of planet 
gears in order to optimally support the drive 
gear (Figure 11). 

In addition to stabilizing the bearing 
support, the flanges on the differential 
housing are used to center and guide the 
output sun gears and side shafts. Hard¬ 
ened sleeves are pressed into the flanges. 
These sleeves are fitted with correspond¬ 
ing oil reservoirs. Both the sun gears and 
the bevel gears are extended beyond the 
housing. As the sun gears are also fitted 
with internal sealing caps, the stub shafts 
can be disassembled without the risk of 
losing any oil. 

Thanks to a combination of roller bear¬ 
ings and axial needle roller bearings, plus a 
new type of flange bearing, the bearing sup¬ 
port offers an extremely high level of rigidity. 
The flange bearing relies on manufacturing 
technology similar to that used for clutch re¬ 
lease bearings or strut bearings. 

Figure 10 shows a comparison of results 
for current prototypes. The lightweight differ¬ 
ential has a more deflected shape, yet this is 
then offset by a more rigid bearing support. 
The results are generally at the same level for 
both transmission variants, yet show the level 
of displacement of a drive gear riveted to the 
differential. The next step in the development 
process represents a departure from this 
principle towards the use of a laser welded 






Differential Systems 


26 


387 



Initial situation 13.4 kg Variant A (welded) 9.1kg Variant B (riveted) 9.6 kg 

Figure 11 Weight savings at high torque classes 


connection, as is already used in the volume 
production of various bevel gear differentials. 
Using a welded connection creates an addi¬ 
tional weight saving of approximately 500 g 
in comparison to the riveted variant. 

Furthermore, the design of the drive 
gear is significantly simplified and the rigidi¬ 
ty of the system is further enhanced by the 
circumferential weld seam. These charac¬ 
teristics mean that the drive gear in the 
lightweight differential has a reduced level 
of displacement in comparison to that of the 
drive gear in the bevel gear differential. 

The modified drive gear also offers the op¬ 
tion to change the technology used for manu¬ 
facturing the blank. Instead of classic forging 
now ring rolling can be used, which can po¬ 
tentially contribute to a significant cost reduc¬ 
tion in the production of the drive wheel. Dis¬ 
pensing with riveting also opens the possibility 
of using variable three or four pairs of planet 
gears, depending on the torque requirement. 
This flexibility is very congenial for a modular 
system. 

The alternative design for the transmis¬ 
sion with a lightweight differential therefore 
creates a weight saving of around 4.3 kg 
without taking into account any optimiza¬ 
tions made to the transmission housing it¬ 
self (Figure 11). 

In conjunction with an optimum design 
for the bearing support, the weight saving 


corresponds to a maximum saving of 
around 0.6 g of C0 2 /km. The use of the 
axial needle roller bearing support has rela¬ 
tively little effect when it is positioned on the 
coast side (Figure 12). 

Both application examples show how 
the lightweight differential from Schaeffler 
can help to reduce weight and fuel con¬ 
sumption. The weight saving of almost more 
than 4 kg can also help to significantly re¬ 
duce the total weight of the transmission. 
By considerably reducing the mounting 
space required, new free space is created 
for the design of the bearing support, which 
enables reductions in friction of up to 80 % 
in the partial load range. The lightweight dif¬ 
ferential is also increasingly attractive when 
it comes to costs. In principle, similar costs 
to those associated with previous solutions 
can therefore be expected for high-volume 
manufacturing. 

One disadvantage of the described 
solutions is simply the scope of applica¬ 
tion limited to transverse transmissions. 
For this reason, both new and familiar so¬ 
lutions for cost-optimized differentials or 
even those with an extremely high level of 
power density are considered in the fol¬ 
lowing section. The purpose of taking a 
closer look at these solutions is to expand 
the portfolio of differentials offered by 
Schaeffler. 






388 


NEDC driving cycle 

- Maximum reduction in fuel consumption: 
0.025 1/100 km 

- Maximum reduction in C0 2 emissions: 
0.58 g/km 


Differential 

Type 

Bearing support Arrangement 

BGD 

1 

TRB 

X 

SGD 

2 

ACBB 

O 

SGD 

3 

ACBB/AX + RH 

O 

SGD 

4 

TRB/AX + RH 

O 


TRB - Tapered roller bearing 
ACBB - Angular contact ball bearing 
AX - Axial needle roller bearing 
RH - Drawn cup roller bearing 
TBB - Tandem ball bearing 


■a 

0 ) 

DC 


80- 


40- 



0.7 

0.6 

0.5 

0.4 

0.3 

0.2 

0.1 

0.0 


2 3 

Types 


■ Reduction in friction 

■ C0 2 emissions 


Figure 12 Reduction in fuel consumption at 
high torque classes 


The search for tomorrow’s 
innovations 


As we have demonstrated by examining 
Alexander Brown’s invention, an occasional 
look into the past can indeed be worthwhile. 
Sometimes, inventions from bygone eras 
can even highlight one approach or another 
that could once again prove useful with the 
help of modern manufacturing technology. 


The Wildhaber-Novikov differential 

The idea for the “Wildhaber-Novikov” differ¬ 
ential was hit upon a few years ago when 
looking at the involute compensation gear¬ 
ing of the lightweight differential, which had 
only just entered into development. The 
project description for this differential is 
based on the type of gearing used for the 
differential pinions, which deviates from the 
conventional involute gearing. 

An alternative circular-arc gearing had 
already been developed in 1926 by Dr. Ernst 
Wildhaber. With this gearing design, the 
convex teeth meshed with concave gaps 
and the radius of the contact points were 
approximately the same. In 1956, this gear¬ 
ing design was revisited and refined by 
Dr. Mikhail Novikov, a Soviet developer and 
military officer. In general, a higher level of 
power density is attributed to this gearing 
design than to comparable involute gear¬ 
ings, and its use in various military vehicles 
not only in the former Soviet Union has cer¬ 
tainly contributed significantly to the reputa¬ 
tion of the gearing design. 

The developers at Schaeffler then hit 
upon the idea of accommodating the size dif¬ 
ference between the sun gears - while keep¬ 
ing the same number of teeth - by using one 
convex sun gear and one concave sun gear 
instead of achieving this via profile displace¬ 
ment, as is the case on the lightweight differ¬ 
ential. This principle is explained in Figure 13. 

Using this solution, it is possible to shift 
the gearing contact point between the con¬ 
cave and convex differential gears via the 
smaller concave sun gear, as is also possible 
on the lightweight differential. The aim is to 
create a differential gearing that exhibits an 
extremely high level of power density as well 
as narrow radial dimensions. A design of this 
type could provide a solution for a differential 
featuring bevel gearing, for example. How¬ 
ever, in this case the radial dimensions for the 
compensation gearing are limited, meaning 
the traditional gearing featured in the light- 









Differential Systems 


26 


389 



Figure 13 Asymmetric gearing on the basis of the involute and the circular-arc profile 


weight differential can no longer be used. 
Although this idea appears to be pioneering 
at first glance, the engineers were not able to 
confirm that the Wildhaber-Novikov differen¬ 
tial has a higher level of power density in pre¬ 
vious investigations. For this reason, this ap¬ 
proach has not proven successful to date, 
meaning it has been necessary to look for 
alternative solutions. 


Oliver Saari’s differential 

In 1966, various solutions for differentials us¬ 
ing a spur gear design were published under 
US patent no. 3,292,456; these differentials 
once again demonstrated a significant in¬ 
crease in performance in comparison to so¬ 
lutions already in existence (Figure 14). Inven¬ 
tor Oliver Saari designed the gearings for 
these differentials such that the compensa¬ 
tion planet gears were not arranged in pairs 
— instead, all planet gears meshed with one 
another. As a result, the load on the gearing 
contact point between the planet gears is 
significantly reduced and the overall axial 
length of the gearing is shortened. Thanks 


Dec. 20, 1966 o. r saari 3,292,456 

SFIK LIlHTtMQ DIFFERENTIAL 

Fll*d April 90, 19M 9 9 



Figure 14 Excerpt from the patent specification 
submitted by Oliver Saari 





































390 


to the high number of gearing contact points, 
the gearings of the sun gears were also kept 
sufficiently narrow, despite narrow radial di¬ 
mensions. As a result, it is possible to create 
a relatively compact design. 

This design was of such interest to the 
engineers at Schaeffler that they began de¬ 
velopment of a new differential variant based 
on Oliver Saari’s invention and in conjunction 
with the asymmetrical design for the sun 
gears. The result was the heavy-duty differ¬ 
ential in addition to the lightweight differential 
from Schaeffler. 

The Schaeffler heavy-duty 
differential with all-wheel 
drive disconnect system 


As the provisional title would suggest, the 
developers at Schaeffler are currently work¬ 
ing on a heavy-duty differential with the aim 
of creating a differential that has a higher 
level of power density than that of the exist¬ 
ing bevel gear differential. When designing 
this differential, the radial mounting space 
requirements must still be fulfilled - unfortu¬ 
nately, this was not achieved with the first 



Figure 15 Heavy-duty differential with AWD 
disconnect system 


attempt with the Wildhaber-Novikov differ¬ 
ential design. The weight and the axial 
mounting space required for use are less 
important in this particular scenario. 

The development is based on the idea 
that the customer should not have to resort 
to the next largest bevel gear differential 
when the torque of the powertrain is in¬ 
creased. In this case, the customer can in¬ 
stead continue to use the compact heavy- 
duty differential from Schaeffler. It is therefore 
also possible to indirectly create a weight 
saving. In addition, it is also possible to inte¬ 
grate features such as a differential lock or an 
all-wheel drive disconnect system into the 
extra axial mounting space. 

Figure 15 shows a cross section of a 
heavy-duty differential featuring an addi¬ 
tional all-wheel drive disconnect system. 
The all-wheel drive disconnect system in 
the rear-axle differential shown is used to 
reduce the drag torques in the powertrain 
by immobilizng the cardan shaft. To do this, 
it is not sufficient to simply interrupt the 
power flow to the cardan shaft in the trans¬ 
fer gear. Instead, it is also necessary to fur¬ 
ther separate the rear-axle differential and 
the wheels, as otherwise the powertrain is 
dragged by the rear axle. 

The engineers at Schaeffler decided to 
perform this separation between the differ¬ 
ential drive gear and the differential itself. To 
this end, the differential housing - compris¬ 
ing a single unit up to this point - is divided 
into two housings arranged coaxially to one 
another. The outer housing holds the axle 
drive gear, and the inner housing incorpo¬ 
rates the compensation gearing. Although 
the differential itself is still dragged along 
when the disconnect system is used, it may 
be possible to achieve reductions in fuel 
consumption in the range of 5 % according 
to a technical publication from 2009 [3]. 

Another distinctive feature of the Schaeffler 
solution is the design of the clutch unit. 
Although the clutch unit shown in Figure 16 
may at first glance look like a conventional 



Differential Systems 


26 


391 



Figure 16 Cross section of the axle transmission 
with a self- reinforcing clutch unit 

actuated wet clutch, it is in fact a one-way 
clutch unit actuated via additional multi-disk 
plates. 

The use of one-way clutches on the rear 
axle to immobilize the powertrain is well 
known, and engaging the rear axle is a rela¬ 
tively straightforward procedure. As soon 
as engine speed is applied to the cardan 
shaft and the wet clutch is actuated, the 
rear axle is “overtaken” and the one-way 
clutch is locked. On a simple one-way 
clutch, the manner in which the rear axle is 
engaged by the prior actuation of the wet 
clutch is unfavourably abrupt. However, in 
the case of the Schaeffler solution, engage¬ 
ment of the rear axle is damped and only 
possible if the one-way clutch is being actu¬ 
ated by the wet clutch. The clamping func¬ 
tion of the one-way clutch facilitates a sig¬ 
nificantly higher torque capacity than that of 
a comparable wet clutch. In forward direc¬ 
tion the wet clutch is only being used as an 
actuation system. 

As soon as the engine speed at the car¬ 
dan shaft falls under the specified speed, 
the one-way clutch is disengaged when the 
wet clutch is not actuated and the drive 
wheel comes to a stop based on the friction 
at the cardan shaft. As the high torques in 
traction mode are absorbed by the integrat¬ 
ed “sprag plates” and only the torques in 
reverse gear or in overrun mode need to be 


dealt with by the wet clutch, the clutch unit 
can be designed for an extremely high level 
of power density. Both the axial and radial 
mounting spaces are therefore compact 
enough to allow the distance between the 
bearings in the original bevel gear differential 
to be retained. 


Conclusion 


Sometimes, the key to the future lies in the 
past. An in-depth examination of the ideas 
and concepts from the pioneering age of 
automobiles and their relationship to today’s 
state of the art can provide a starting point 
for innovations that can solve present- 
day problems. Current developments at 
Schaeffler show the way forward, helping to 
increase customer benefits and find an¬ 
swers to pressing questions relating to as¬ 
pects such as lightweight construction, 
costs and C0 2 emissions. 


Literature 


[1] Smetana, T.; Biermann, Th.; Hoehn, B.; Kurth, 
F.: Schaeffler lightweight differentials. 

9 th Schaeffler Symposium, 2010, pp. 94-105 

[2] Smetana, T.; Biermann, Th.: Kompakte Leicht- 
bau-Differenziale (Compact lightweight dif¬ 
ferentials). ATZ (Automobiltechnische Zeitschrift 
- Automotive journal) 2/2011, pp. 108-113 

[3] Gassmann, T.; Schwekutsch, M.: Verringerung 
des allradbedingten Mehrverbrauchs durch 
dynamische Allradabschaltung. ATZ 9/2009, 
p. 672 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 




392 


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393 


The Chassis of the Future 


Markus Baeuml 
Florin Dobre 
Harald Hochmuth 
Manfred Kraus 
Hartmut Krehmer 
Roland Langer 
Dominik Reif 


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394 


Introduction 


When it comes to developing chassis, to¬ 
day's challenges go far and above the tradi¬ 
tional conflict of having a comfort-based and 
sportive set-up. Replacing hydraulic systems 
with electromechanical actuators in chassis 
technology is particularly progressing at 
quite a rate, with scores of functions are al¬ 
ready being realised using electromechani¬ 
cal means. In terms of steering, the last hy¬ 
draulic systems are currently being replaced 
with electromechanical systems in the 
D segment. Electric and hybrid vehicles are 
the driving force behind this application of 
electro-hydraulic brake boosters. However, 
these boosters continue to be based on a 
hydraulic brake with a mechanical safe state. 
Gradual conversion of the anti-roll system is 
expected from 2015 onwards. Only the ac¬ 
tive chassis (Active Body Control, ABC) is 
currently still designed as a hydraulic system, 


but it can also be replaced with an electro¬ 
mechanical version. 

A whole host of benefits is associated 
with electrification of the chassis. Thus, the 
principle of on-demand actuation results in 
lower energy consumption. New features, 
such as the Continuous Damping Control 
(CDC), have also been developed in parallel 
with this benefit. CDC dampers already 
make up the extra specifications list in the B 
and C segments. Figure 1 shows the tech¬ 
nologies and their penetration in the indi¬ 
vidual vehicle segments. 

Requirements of chassis of 
the future 


Stringent requirements regarding C0 2 re¬ 
duction also mean that chassis technology 
will have to utilise the potentials provided by 


Characteristic Function Segment 


A B C C/D D 

Sub A B-SUV C-SUV CD-SUV D-SUV 

Lateral Electric steering s S S S S in future 

dynamics Anti-roll system 0 0 

Rear-wheel steering 0 0 

Superimposed steering 0 0 

Torque vectoring 0 

Vertical dynamics Variable dampers 0 0 0 S 

Air springs 0 S/O 

Level control 0 2) 0 2) 0 2) 

ABC (active body control) S/O 

Longitudinal Electronic parking brake S/O S S 

dynamics Electronic brake booster S 1 * S 1 * S 1 * s in future S in future 

Driver assistance Lane departure warning OOO 

system Emergency braking assist 0 0 0 0 

Traffic jam assist 0 0 0 

Self-driving vehicles 2017/18 3) 


S = standard feature ^ will be standard feature on electric vehicles 3 > Semi-self-driving 

0 = optional feature 2) SOP = 2017 onwards 

Figure 1 Chassis technologies and their penetration in various vehicle segments 

Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 

DOI 10.1007/978-3-658-06430-3_27, © The Author(s) 2014 














Chassis 


27 


395 


Urbanisation 


Product differentiation 


Trend 

Reduction in 

C0 2 emissions 

Affordable 

travel 

Comfort and safety 


e-mobility/ 

hybridisation 

Platform strategy 

Self driving vehicles 

(/> 

£ 

Friction 

reduction 

New chassis 
layouts/concepts 

Network/ 
connected driving 

3 

</> 

CO 

CD 

2 

Lightweight 

design 

Cost optimised 
solutions 

New vehicle 
concepts 


Demand-based 

control 

Car sharing 

New chassis 
applications 


Energy 

recuperation 

Technology aimed at 
older drivers 


Driving pleasure 


Extension of platform 
strategy functions 


New chassis 
applications 



Figure 2 Trends in chassis technology 

lightweight construction, friction reduction 
and more efficient actuators [1]. This is ac¬ 
companied by the use of new materials or 
existing materials with optimised character¬ 
istics in terms of rigidity and strength. 
What’s more, many chassis systems are 
also used as a way of making vehicles stand 
out within a platform. Figure 2 shows an 
overview of the current trends. 

Over the next few years, buzzwords 
such as connectivity, autonomous or semi- 
autonomous driving will have a consider¬ 
able bearing on chassis development [2]. 
Related to this development is, ultimately, a 
modified safety strategy; for instance ex¬ 
tended latency periods requiring the basic 
mechanical function to be protected. This 
protection may also necessitate enhanced 
or additional redundancy/safety state. In 
light of these possibilities, new requirements 
will be demanded of existing actuators. 

What’s more, actuators, sensors and 
systems are increasingly networked to gen¬ 
erate new overarching functions, to increase 
availability and to improve safety. This could 
be achieved, for instance, by a mutual plau¬ 
sibility in the context of a safety concept ac¬ 
cording to ISO 262622. Key elements of the 
future thus include cameras, sensors, an¬ 
tennas, as well as corresponding software 


for networking in the vehicle and with the 
environment [3] 

Of key importance is the increase in the 
use of camera and radar-based as well as 
laser-based systems. These systems in¬ 
clude polarising and infra-red cameras, in 
addition to stereo ones. Used in combina¬ 
tion with information regarding temperature 
and humidity, it is possible to detect aqua¬ 
planing and black ice. 

Current Schaeffler 
solutions 


Products for reducing weight 

In the wheel bearing area, the market has 
seen a gradual introduction of lightweight 
construction solutions with face spline 
and weight-optimised flange design. The 
technology is becoming increasingly pop¬ 
ular and is well on the way to setting a 
new industry standard in the foreseeable 
future - a standard that Schaeffler will 
have created. 


















396 



Current design 


10 % Weight 
reduction 



Figure 3 Wheel bearing with face spline design compared to dominant design to date with internal 
gear teeth 


Figure 3 shows a comparison of a third- 
generation wheel bearing in its previous de¬ 
sign and one with face spline. 

The benefits from this technology, such 
as 10 % rigidity increase, 10 % weight re¬ 
duction, 50 % higher transferable torque 
as well as a reduction in unsprung mass 
yet still with simple assembly process, 
have been utilised in series applications 
since 2009. 

An additional measure for reducing 
weight comes about by cutting the bear¬ 


ing flange weight while maintaining its ri¬ 
gidity. By applying numerical procedures, 
it has already been possible to make 
weight reductions of 20 % without com¬ 
promising the axial rigidity. Figure 4 shows 
a wheel bearing with a weight-optimised 
flange compared with a conventional 
bearing flange. 

The result is optimised tension curves, 
which have also been used as a basis for an 
enhanced fibre flow of the flange. It is feasi¬ 
ble to use driven and non-driven axles. 


Current design 


Lightweight solution 





Figure 4 Comparison of a current wheel bearing with a wheel bearing with weight-optimised flange 




Chassis 


27 


397 


Friction reduction products 

Seal friction determines wheel bearing friction 
to a great extent, which is why it makes sense 
to start there with measures designed to re¬ 
duce friction. The wheel bearings offered by 
Schaeffler can be fitted with low-friction seals, 
which reduce friction by around 50 % com¬ 
pared to seals offered by competitors. This 
50 % reduction thus makes it possible to cut 
C0 2 emissions by around 1 g/100 km. It is 
worth mentioning that the sealing effect is still 
the same compared with today’s convention¬ 
al two and three-lip seals (Figure 5). 

Mechanical actuators with ball screw 
drive for chassis applications 

Many linear actuators are equipped with a 
ball screw drive as a mechanical actuating 
element. Schaeffler launched a ball screw 
drive for electromechanical power-assisted 
steering on the market as far back as 2007. 



New: Previously: 

seal with seal with 

seal lip and three seal lips 

labyrinth seal 

Figure 5 Comparison of conventional seal 
with a friction-reduced seal 


Ball screw drives for electrically 
assisted steering systems_ 


Ball screw drives for 
-parking brakes 



Ball screw drives for 
clutch release systems 


Ball screw drives for 
brake boosters 


Figure 6 Overview of ball screw drive applications 




















398 



v-M f j OHO(K : ' i H —s 

Torque 

Gearbox and Motor and ECU _ 

. sensor 

decoupling unit 


Stabiliser halves 


Figure 7 Design of the anti-roll system 

This steering ball screw drive is designed 
along the lines of the principle of modular 
design and can cover a wide range of ap¬ 
plications. It provides a virtually constantly 
high degree of efficiency of more than 90 % 
over the entire temperature range and is 
supplied together with a four-point support 
bearing. Ball screw drives and support 
bearings designed to meet customer re¬ 
quirements of minimized backlash can be 
provided. 

In parallel to this, a compact ball screw 
drive with a pitch diameter of up to 4 mm 
has been developed; this compact version 
has been used by Continental in its electric 
parking brake since 2011. Other applica¬ 
tions based on this design are currently in 
the development phase — for instance, ap¬ 
plication in the electromechanically operat¬ 


ed brake booster. Figure 6 shows other po¬ 
tential applications for the compact ball 
screw drive. 


Electromechanical anti-roll system 

Over the last few years, Schaeffler has 
played its role in driving the replacement of 
hydraulic with electromechanical systems 
thanks to developing an electromechanical 
anti-roll system. The plan is for series pro¬ 
duction of this system to start in 2015. The 
benefits offered by the system are: 

- Little or no tilting of the vehicle when 
cornering as a function of the present 
lateral acceleration 

- More accurate steering behaviour, im¬ 
proved agility and stability 


1 


A 


Mechanics 
(planetary gear 
train and 
decoupling unit) 


Vehicle 

Power 

bus 

supply 



SPI (serial peripheral interface) 


Figure 8 Actuator system architecture 
















Chassis 


27 


399 


- Enhanced system dynamics compared 
to hydraulic systems 

- Simple installation and easy maintenance 

- Reduction in the number of field com¬ 
plaints by up to 30 % compared to hy¬ 
draulic systems 

- Installation in hybrid vehicles possible 

- Reduction in fuel consumption of up to 
0.3 litres compared to hydraulic anti-roll 
systems, and 

- Weight neutral compared to hydraulic 
systems 

The system comprises a brushless direct 
current motor with control system, trans¬ 
mission, torsion bars and a decoupling unit 
(Figure 7). The E/E architecture is shown in 
Figure 8. 

To complement a pure rotary actuator 
and to enhance comfort, the Schaeffler 
solution features a decoupling element, 
which enables one-sided disruptions in 
the road surface to be absorbed. Trans¬ 
mitting pulses to the body is thereby also 
reduced as well as strong vertical motion 
caused by one-sided disturbance excita¬ 
tion. Design and 
function of the anti¬ 
roll system are ex¬ 
plained in detail in 
[4] and [5]. The ef¬ 
fect of the decou¬ 
pling unit for small 
disturbance excita¬ 
tions is shown in 
Figure 9. 

The decoupling 
unit demonstrates 
excellent efficiency 
particularly for 
small disturbance 
excitations with an 
amplitude of up to 
5 mm. Larger dis¬ 
turbance excita¬ 
tions can be cor¬ 
rected by the 
disturbance con- 



— without decoupling unit 

— with decoupling unit 


Figure 9 Dynamic stiffness as a function of the 
frequency of one-sided disturbance 
excitation for systems with and 
without a decoupling unit 

trailer. As the input parameter, this con¬ 
troller requires different functions, includ¬ 
ing the torque in the anti-roll system and 
the vertical displacement of the wheels. 
The overall controller structure is shown in 
Figure 10. 



Vehicle signals, e.g. lateral acceleration, speed, steering-wheel 
angle etc. 


Figure 10 Block diagram of the anti-roll system 











400 



Figure 11 Sensor layer for measuring the wheel force at the wheel bearing 
(on the left) and for measuring the steering moment in the 
steering gear 


The interference 
can be corrected 
up to a frequency of 
approximately 8 Hz. 

The maximum fre¬ 
quency depends on 
the amplitude. If the 
information about 
the road surface 
collected by a ste¬ 
reo camera is avail¬ 
able as the input 
signal and informa¬ 
tion from the navi¬ 
gation system about 
the route can be 
used, the distur¬ 
bance controller can be improved still fur¬ 
ther by means of anticipation. 

Alternatively, the body tilt and the effect 
of one-sided disturbance excitation on the 
body can also be prevented by hydraulically 
adjustable struts on each wheel. In addition 
to the anti-roll motion, this kind of system 
also prevents a pitching motion during brak¬ 
ing and accelerating. However, this does not 
apply to air-sprung systems on account of 
the compressibility of air. 

Future Schaeffler 
solutions 


and thus record the forces acting on the 
wheel, including the brake forces generated 
during braking. These forces can be used 
as an input signal for various chassis control 
systems. The wheel force measurement be¬ 
ing developed at Schaeffler also enables 
accurate recording of the vehicle weight, 
which may be of interest for light commer¬ 
cial vehicles. 

The measurement principle is based on 
the arrangement of strain gauges on a two- 
dimensional or three-dimensional tensioned 
surface. The strain gauges are attached us¬ 
ing thin-film technology. The basic layer de¬ 
sign is shown in Figure 12. 

The geometry of the strain gauges is 
“cut” into the sensor layer using laser, with a 
top cover attached to protect the sensor 
layer. To illustrate the technology, Figure 13 


Sensor layer for measuring 
wheel force 

Schaeffler is currently developing a sensor 
layer for measuring wheel force; this layer 
can be applied to two or three-dimensional 
components such as bearing components. 
Figure 11 shows several examples of appli¬ 
cations. Application to the wheel bearing 
enables the wheel force to be measured 


Sensing layer 

Sensory 
layer 0.2 Mm 

Insulation 
3-5 Mm 

Cr 

Bonding agent Substrate 
0.2 Mm 

Figure 12 Sensor layer design 







Chassis 


27 


401 



Figure 13 Sensor layer on a bearing outer ring 

shows an applied sensor layer using a bear¬ 
ing outer ring as an example. 

As proof of the measurement accura¬ 
cy, it is helpful to compare this layer with a 
laser extensometer. Experiments with pla¬ 
nar samples, which were stretched on a 
traction engine and their elongation in 
synchronously recorded with the sensor 
layer as well as using the laser extensom¬ 
eter, have provided fairly good correlation 
(Figure 14). 

The past few years have seen that the 
process reliability of the individual process 
steps has been systematically demonstrat¬ 
ed and increased. Currently, preparations 
for winning projects and customers are be¬ 
ing ramped. 


Comparison of measured extensions 
Laser extensometer vs. 
thin-film torque sensor 



Laser extensometer extension 
in pm/m 

Figure 14 Comparing the elongations of planar 
samples with the sensor layer 


Level adjustment 

In today’s vehicles, air suspension is 
used to adjust the ride height to various 
driving and load conditions. This suspen¬ 
sion system can inherently absorb very 
poor lateral forces and is therefore not 
well-suited to McPherson strut axles. In 
addition, the costs for air springs are an¬ 
other reason the system has not become 
established in the B and C segments. 
Hydraulic height adjustment systems are 
used in the sports car sector, in particu¬ 
lar on the front axle to make it easier to 
drive over ramps [6]. The tendency of 
markets towards potentially failure sensi¬ 
tive hydraulic actuators is to oppose fur¬ 
ther proliferation of this technology. 
There is therefore a need for electrome¬ 
chanical systems designed to adapt the 
ride height. 

The following functions can be supported 
by this kind of system. 

- Lowering the vehicle to reduce aerody¬ 
namic drag either on all four wheels or 
only on the front axle to bring a laden 
car back into the trim position 

- Raising the vehicle to make it easier to 
get in 

- Raising a sports car to protect the 
spoiler when driving over car park 
ramps 

- Raising vehicles for light off-roading, as 
well as 


















402 




Anti-twist protection 

Mounted 
locking ring 

Locking sleeve 
Locking contours 
Spring seat 
Cam contour 
Spindle 
Belt drive 

Ball screw drive nut 
with belt wheel 

Electric motor 




Figure 15 Actuator for the level adjustment on the front axle 


- Lowering the vehicle to make it easier 

to load the luggage compartment 
The solution developed by Schaeffler is 
shown in Figure 15. 

The actuator essentially comprises a 
ball screw drive, a belt drive, an electric 
motor and a locking assembly. In this case, 
the vehicle load is not supported on the 
ball screw drive but on the locking assem¬ 


bly, which locks the vehicle’s ride height. 
The ball screw drive itself is only used to 
adjust the different heights. Figure 16 
shows a detailed view of the locking 
assembly. 

The spindle is fixed on the damper 
to raise and lower the vehicle, while 
the nut is driven by a belt. The nut rotat¬ 
ing leads to an axial displacement 



Figure 16 Locking assembly in detail 

























































Chassis 


27 


403 



Power flow when lifted and lowered Power flow when locked 


Figure 17 Power flow during raising, lowering and locking 


of the unit com¬ 
prising the nut, con¬ 
trol contour, motor, 
housing and spring 
seat, and this is 
what changes the 
ride height. 

To lock the 
height, the locking 
ring engages in dif¬ 
ferent locking con¬ 
tour grooves de¬ 
pending on the 
position when low¬ 
ering. This action 
maintains the vehi¬ 
cle at the required 
level. As the vehicle 
is offset in any position on the locking ring, 
the drive and spindle lock remain load-free 
in the locked state (Figure 17). 

To aid a better understanding, the three 
different ride heights and resulting design 


positions of the actuator are summarised in 
Figure 18. The number of grooves deter¬ 
mines the possible ride height. A third 
groove means that a central position can 
also be realised. 


Bottom position Central position 


Top position 




Figure 18 Position of the actuator at different ride heights 















































404 


Connection according to customer requirements 



8 


Stroke 
(according to 
customer 
requirements) 


0 175 

(according to customer 
requirements) 


Installation position of the actuator on the rear axle 


The current engi¬ 
neering knowl¬ 
edge enables ad¬ 
justment ranges 
of 40 mm, which 
can be extended 
even further de¬ 
pending on the 
available space. 

The selected de¬ 
sign also allows 
installation on the 
rear axle, where 
dampers and 
springs are often 
arranged sepa¬ 
rately. The only 
action needed to 
accommodate this 
installation is to 
merely rotate the 
motor by 180° (Fig¬ 
ure 19). Figure 19 

For E/E imple¬ 
mentation, E/E components are already 
available on the market. Selected ECU 
includes two power stages, they can 


control two electric motors simultane¬ 
ously. The resulting system architecture 
is shown in Figure 20. 



Figure 20 System architecture of the level adjustment 


(with 40 mm stroke) 






































Chassis 


27 


405 


The proposed sys¬ 
tem configuration 
can be seen in Fig¬ 
ure 21. 

By virtue of the 
actuator design, 
selected system 
architecture and 
proposed system 
configuration, the 
market is filled 
with diverse and 
promising applica¬ 
tions. Preparations 
are currently un¬ 
derway to con¬ 
struct test vehicles 
this year. 



Vehicle signals, e.g. vehicle speed, steering wheel angle etc. 


Figure 21 System configuration for the level adjustment 


Actuator camber and toe-in actuation 

The approach taken by Schaeffler for cam¬ 
ber and toe-in actuation is based on an ec¬ 
centric drive, which is mounted to the rear 


axle carrier, that can be designed as an 
individual wheel actuator [7]. Figure 22 
shows the mechanical concept. 

The axle-side actuator provides actua¬ 
tion of the toe-in and/or support arm. The 




^3 mm eccentricity 
"^(6 mm travel) 


BLDC Coil Slipping 

motor spring clutch 

lock 


Elastomeric bearing 



Eccentric shaft 
Travel 


Figure 22 Design of the eccentric actuator for use on the rear axle carrier 




























406 


actuation speed and force are based on 
the power of the selected drive. The actua¬ 
tion travel is a function of the underlying 
eccentric feature. The E/E architecture 
uses the E/E components familiar from the 
level adjustment system with two integrat¬ 
ed power stages to control two electric 
motors. This results in the following actua¬ 
tor characteristics: 

- Actuation travel = 6 mm in the case of 
this eccentricity of 3 mm, 

- Maximum actuation time < 2 s 

- Maximum actuation load 5 kN 

- Actuator diameter < 65 mm 

To reduce the engine speed, a worm 
wheel or planetary gear train can be 
used. Another feature of the drive is its 
overload clutch, as well as mechanical 
short circuit to protect the bearings. Fur¬ 
thermore, the actuator can be integrated 
into an elastomer metal cartridge on 
request. 

Previous customer feedback indi¬ 
cates that the market is looking for an 
alternative to the linear actuator on the 
rear axle. This alternative does not al¬ 
ways need highly dynamic actuation. The 
stated actuation time of 2 seconds for 
toe-in actuation with a noticeable reduc¬ 
tion in turning circle is usually sufficient. 
Current plans are to kit out a prototype 
vehicle this year. 


Developing the anti-roll system 

In the course of developing the anti-roll 
system further, a split stabiliser is opened 
for driving in a straight line and closed 
when cornering. Thus, a quasi-static ten¬ 
sion state is produced when cornering. 
When driving in a straight line, however, 
the stabiliser is open and rolling move¬ 
ments of the bodywork for the reciprocal 
disturbance excitation through the road 
to the opposite side of the vehicle, are 
suppressed. 


In order to significantly reduce the vehicle’s 
rolling angle when cornering, the stabiliser 
rigidity is increased by more than 20 % 
compared to a passive stabiliser. The de¬ 
sign for this type of anti-roll system is shown 
in Figure 23. 

In this design, the clutch is actuated via 
electromechanical linear actuator (consist¬ 
ing of electric motor, ball screw), such as 
depending on the steering angle and vehi¬ 
cle speed and other vehicle status parame¬ 
ters. The functional principle of the clutch 
is based on a locking device developed at 
Schaeffler, the design of which is also 
shown in Figure 23. 

The current engineering knowledge has 
a weight of 3.5 kg without stabiliser halves. 
Compared to the design used in series pro- 


Rotational Switchable Elastomer 

Cable set angle locking coupling 
\ sensor mechanism / 




Figure 23 Split anti-roll stabiliser 

















Chassis 


27 


407 


duction, this equates to a weight reduction 
of more than 50 %. If the stabiliser halves 
are not designed as steel pipes, but in glass 
fibre reinforced plastic, this produces a po¬ 
tential total weight of the entire actuator of 
around 4 to 4.5 kg. 


“Switchable" wheel bearings 

Schaeffler has developed a triple row 
wheel bearing to reduce friction compared 
to the tapered roller bearings used in gen¬ 
eral and for higher wheel loads. This bear¬ 
ing features two equally tensioned rows of 
balls. To further reduce friction, the bearing 
can be designed such that only the outer 
rows of balls are used when driving in a 
straight line, and the central row is en¬ 
gaged when cornering. This is done by 
specifically changing the bearing preload, 
as shown in Figure 24. 

The balls with their spring deflection are 
shown as springs. 

Only the outer rows of balls are loaded 
when driving in a straight line; the central 
row is not loaded. When cornering, the cen¬ 
tral row (which is designed a four-point con¬ 
tact bearing) is engaged in order to support 
the drive performance in the bend by pro¬ 
viding the required high level of rigidity. To 
this end, only a few oversized balls are fitted 
in the four-point contact bearing, which 
means that the remaining balls in the cage 
have clearance and reduce friction when 
driving in a straight line. When cornering, 
these balls are in contact and then absorb 
the required forces. Initial simulation results 
show an additional reduction in friction of 
more than 25 %. 




Figure 24 Switchable wheel bearing with offset 
outer rows of balls 


Active electromechanical damping 

One possible approach of realising an ac¬ 
tive, or at least partially active, chassis is 
produced by using an electromechanical 


damper; this damper simultaneously works 
as an actuating element and actively feeds 
forces into the chassis. The idea of being 
able to utilise the lost energy of vehicle 
damping has been explored for over 




























408 



20 years; the result is to use a brushless di¬ 
rect current motor using a ball screw drive 
to transfer the vertical motion of the wheel in 
a rotational motion of the rotor, thereby re¬ 
cuperating the damping energy [8]. 

What’s more, this kind of damper pro¬ 
vides the prerequisite for optimising the 
damping characteristic curves beyond the 
options offered by the hydraulic system [9]. At 
the same time, it forms the basis for realising a 
(partial) active suspension. Previous solutions 
show an unfavourable cost-benefit ratio and 
are also difficult to integrate into the space 
available. In addition, other requirements, 
such as overload capability or the response 
characteristic for small excitations, have pre¬ 
vented further development in this field. 

Schaeffler is continuing to develop an ac¬ 
tuator, which will fit as far as possible in the 
existing space of a hydraulic damper, that of¬ 
fers a better cost-benefit ratio than previous 
solutions as well as improved overload ca¬ 
pacity. The basic configuration of the damp¬ 
er comprises a brushless direct current 
motor, a ball screw drive with bearing ar¬ 
rangement and a damper pipe (Figure 25). 

The wheel module with McPherson strut 
is excited vertically through the road surface. 
This translation is converted in the damper to 
a rotation and dampened by the regenera- 



Figure 25 Design of active electromechanical 
damping 


Figure 26 Characteristic curve and application 
area of an electric damper 

















Chassis 


27 


409 


tively operated electric motor. A centrifugal 
brake is used to slow down the rotor rotation 
in the electric motor in the event of large 
pulses. The design of the electric damper is 
based on the characteristic curve of the 
damper during a suspension and rebound of 
a hydraulic damper as well as being based 
on the physical limits of the electric motor in 
generator mode (Figure 26). 

To obtain basic findings, Schaeffler de¬ 
signed an electric damper (identical to the 
one seen in Figure 25) and tested it on the 
test rig. The findings for four different road 
surfaces (A, B, C, D) are shown in Figure 27; 
the amplitude and speed increase in alpha¬ 
betical order. Significant regenerative power 
is achieved with excitation profile C and D, 
but is more likely to be achieved on poor 
roads or when off-roading. If one assumes 
“normal” amplitudes of 10 to 30 mm in ac- 



Figure 27 Measured power generated a function 
of damping force and speed 

cordance with profile A and B, the resulting 
regenerative power ranges from 20 to 30 W. 
This is too little power to justify high volume 
production purely on the grounds of energy 
regeneration. Another option is if the damp¬ 
er can also be used in the chassis as an 



Figure 28 Characteristic diagram of the active electromechanical damping with generator mode and 
active adjustment range 








410 


actuating element [11[. The derivation of the 
underlying function equations of the damp¬ 
er is performed using the quarter vehicle 
model [10]. 

The installed electrical output of around 
1.9 kW per wheel enables active engage¬ 
ment in the chassis. The characteristic dia¬ 
gram of the electromechanical damper is 
shown in Figure 28. The overload capability 
is a result of the centrifugal brake function. 

With the active electromechanical damp¬ 
ing, the entire range [12] of a possible influ¬ 
ence on the driving dynamics can be ex¬ 
tended, thereby significantly boosting the 
benefit for customers. The series produc¬ 
tion use of technology now depends on 
customer acceptance, which is to be stud¬ 
ied over the coming months. 


Outlook 


The range of the chassis applications offered 
by Schaeffler requires a multi-pronged ap¬ 
proach when developing new products. First¬ 
ly, customers in an extremely cost-driven and 
competitive market should be provided with 
added value when it comes to bearing appli¬ 
cations; this can be achieved by offering in¬ 
novative developments. Secondly, mechani¬ 
cally oriented innovations form a sound basis 
for designing new mechatronic chassis sys¬ 
tems. In addition, the task for Schaeffler engi¬ 
neers is also to create and realise added with 
new and trend-setting concepts. The objec¬ 
tive of all these efforts is to generate function 
added value particularly in terms of power 
density, energy efficiency, weight and func¬ 
tional integration as well as to create cost ben¬ 
efits compared to today's technology. To do 
this, the broad knowledge and experience 


held within the Schaeffler Group as well as 
that experience of selected cooperation part¬ 
nerships will be used in a specific manner. 


Literature 


[1] Ammon, D.: Herausforderung Fahrwerkstechnik, 
Tagungsband Chassistech 2009, pp. 1-24 

[2] Kruger, M., Kraftfahrzeugelektronik, 2008, 

2. Auflage, S. 21 ff 

[3] von Hugo, C., The next step towards autono¬ 
mous driving. 22 nd Aachen Symposium, 2013, 
pp. 751-765 

[4] Krimmel, H.; Deiss, H.; Runge, W.; Schurr, H.: 
Elektronische Vernetzung von Antriebsstrang 
und Fahrwerk. ATZ 108, 2006, no. 5, 

pp. 368-375 

[5] Beiker, S.; Mitschke, M.: Verbesserungsmoglich- 
keiten des Fahrverhaltens von Pkw durch zusam- 
menwirkende Regelsysteme. ATZ 103, 2001, 

no. 1, pp. 38-43 

[6] Hohenstein, J.; Schulz, A.; Gaisbacher, D.: Das 
elektropneumatische Vorderachsliftsystem des 
Porsche 997 GT3. ATZ 112, 2010, no. 9, 

pp. 622-626 

[7] Kraus, M.: Actuators for Challenging Chassis. 
8 th LuK Symposium, 2006 

[8] US patent 5091679, 1990 

[9] Kraus, M.: Chassis Systems — Schaeffler Can 
Do More Than Bearings. 9 th Schaeffler Sympo¬ 
sium, 2010 

[10] http://web.mit.edu/newsoffice/2009/shock- 
absorbers-0209.html 

[11] Willems, M.: Chances and Concepts for 
Recuperating Damper Systems. 21 st Aachen 
Symposium, 2012 

[12] Rau, M.: Koordination aktiver Fahrwerk-Re- 
gelsysteme zur Beeeinflussung der Querd- 
nynamik. Dissertation University of Stuttgart, 
2007, pp. 91-122 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



412 


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413 


Electric Driving without 
Range Anxiety 

Schaeffler’s range-extender transmission 


Andreas Kinigadner 
Dr. Eckhard Kirchner 



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414 


Introduction 


Battery electric vehicles offer the option of 
emission-free local mobility. The range of 
these vehicles will remain limited in the fore¬ 
seeable future due to high battery costs and 
the increased weight associated with the 
limited energy storage density. This has 
lead to the increasing development of 
range-extender drive systems during the 
last few years. These concepts in most cas¬ 
es use a serial hybrid drive, in which the in¬ 
ternal combustion engine is operated solely 
as a generator. These are usually internal 
combustion engines specially developed for 
this application or sometimes stationary op¬ 
erated engines, for which a number of vari¬ 
ants and even Wankel type engines have 
been proposed. However, the implementa¬ 
tion of these special engines is associated 
with large investments and is frequently not 
feasible due to high cost pressures. In addi¬ 
tion to the technical and commercial chal¬ 
lenges of implementing this technology, se¬ 
rial hybrid drives have a poor tank-to-wheel 
efficiency on long distance routes. 

The focus must therefore be placed on 
developing alternative solutions, particu¬ 
larly for electrification in the compact vehi¬ 
cle segments. Schaeffler’s range-extender 
concept is based on adding a special 
transmission to an existing internal com¬ 
bustion engine to produce a full hybrid. A 
simple automatic spur gear transmission 
and an electric motor are used instead of a 
conventional automatic or double clutch 
transmission. The typical range of driving 
conditions for an electric vehicle can be 
completely covered at low system costs. A 
powertrain architecture with a direct me¬ 
chanical linkage of the internal combustion 
engine improves the efficiency balance of a 
vehicle over long distances. In addition, 
Schaeffler’s range-extender transmission 
allows automobile manufacturers to imple¬ 


ment a modular drive strategy without car¬ 
rying out fundamental changes to the ve¬ 
hicle architecture. 


Concept 


What is a range extender? 

A range-extender vehicle differs from a hy¬ 
brid vehicle in that it can be operated with 
the electric motor only during day-to-day 
operation. This also includes acceleration 
and high-speed driving. There is no clear 
distinction between range extenders and 
plug-in hybrid vehicles, whose batteries can 
be charged from a power socket. The range 
extender is sometimes even described in 
technical literature as a type of plug-in hy¬ 
brid [1]. 

Most range extender vehicles were orig¬ 
inally designed as serial hybrids, i.e. the in¬ 
ternal combustion engine is operated only 
as the drive for an electric generator. The 
most prominent example from the pioneer 
age of the automobile is the Mixte car devel¬ 
oped by Ludwig Lohner and Ferdinand 
Porsche in 1902. The overall efficiency of 
this type of system architecture is not only 
dependent on the efficiency of the engine 
and generator, but also on the losses during 
charging and discharging of the battery. 

At high driving speeds, a serial hybrid 
drive has a lower overall efficiency than a 
direct drive by means of an internal com¬ 
bustion engine due to conversion losses [2]. 
This is why some range-extender vehicles 
are already equipped with a power-splitting 
hybrid drive system. A selectable mechani¬ 
cal fixed drive ensures optimum overall effi¬ 
ciency in this case. 

A comparison between the powertrain 
concept of the Opel Ampera, which has a 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3 28, © The Author(s) 2014 



Range-Extender 


28 


415 


Internal combustion engine torque 
Installed electric power 
Mass of electric system 
Mass of mechanical system 
C0 2 emissions combined*,** 
Range km/I gasoline (100 % RE) * 
Range km/kWh electric (100 % plug-in) * 



115 

111 

109 


Range plug-in = 50 km 
Range RE mode = 500 km 
* Mean (NEDC, CADC, LuK CUP) 
** Power mix Germany 2010 


■ First generation 
Second generation 


100 % 


Figure 1 Comparison of performance criteria for a serial hybrid and a vehicle of identical perfor¬ 
mance with a powertrain similar to the Ampera. 


mechanical fixed drive, and a conventional 
serial hybrid shows the significant advan¬ 
tages with regard to C0 2 emissions (-9 % in 
combined mode, Figure 1). At the same 
time, the range of electric operation also in¬ 
creases by 9 % [3]. This type of configura¬ 
tion does however have disadvantages: An 
additional clutch and a more complex oper¬ 
ating strategy are required. In addition, the 
spatial arrangement of the generator unit 
comprising the internal combustion engine 
and generator can no longer be freely se¬ 
lected in the vehicle. 

A range extender for the B and C 
segment 

Range-extender vehicles with on-demand 
mechanical drive are particularly suitable for 
vehicles that are mainly driven in short-run 
operation, but are occasionally also used 
for longer interurban journeys. This makes 
the range extender particularly attractive for 
the B and C segment, the more so since 
this segment accounts for high quantities 
worldwide. If such a drive can be produced 
in line with market requirements, it would fill 
a gap between 


- expensive range-extender vehicles, which 
can cover large distances with an inter¬ 
nal combustion engine after draining 
the battery (for example, the Opel 
Ampera) 

- and battery electric vehicles in the 
A segment, which have no range prob¬ 
lems in urban traffic despite having a 
low battery capacity (for example, the 
VW up!). 

The B and C segments are under a high de¬ 
gree of pressure from competitors interna¬ 
tionally so it is necessary to produce a 
range-extender solution at very low costs. 
Schaeffler therefore aimed to simplify the 
conventional range-extender concept dur¬ 
ing development as follows: 

- Re-utilization of the internal combus¬ 
tion engine and its characteristics in 
terms of function and interfaces 

- No change in the design envelope, no 
change of vehicle architecture in con¬ 
ventional front transverse powertrain 
platforms 

- Use of only one electric motor 

- Use of a single electromechanical ac¬ 
tuator if possible 

- Simplification of the transmission by 
using three or even only two gear steps. 


416 



Figure 2 Schematic diagram of Schaeffler’s 
range-extender module with three 
mechanical gears 

Basic concept 

Preliminary considerations led to the sche¬ 
matic diagram of the powertrain shown in 
Figure 2. 

The range-extender module is connect¬ 
ed to the internal combustion engine using 
an one-way clutch (OWC), which can, for 
example, be designed as a roller clutch. The 
electric motor is 
also connected in a 
selectable manner 
by means of a sep¬ 
arate input shaft. 

The input shaft can 
drive the front axle 
differential (FD) and 
thus the wheel di¬ 
rectly via the gear 
step freewheel (S4). 

Three additional, 
fully independent 
selectable speed 
gears (SI, S2 and 
S3) can be used to 
connect the internal 
combustion engine 
to the output shaft 
and to manage 
gearshift opera¬ 
tions. The signifi¬ 


cant simplification of the transmission and 
the associated reduction in costs compared 
to current hybrid vehicle designs are imme¬ 
diately apparent. 


Design 

Schaeffler’s range-extender transmission is 
a current advanced development project. 
The following information does not there¬ 
fore refer to a specific transmission design 
but describes the ideas on which the de¬ 
sign of the prototype is based. Figure 3 
shows the prototype design, which has not 
yet been optimized for specific vehicle and 
powertrain dimensions. 

It can be seen that the majority of com¬ 
ponents used in the range-extender trans¬ 
mission are components currently used in 
volume production. This means it was pos¬ 
sible to use synchro ring packages operated 
by shift sleeves from manual transmissions 
[4]. The actuator driven by an electric motor 
with an interlock function, which Schaeffler 


















Range-Extender 


28 


417 


developed for double clutch transmissions, 
can, in principle, be used for clutch actuation 
[5]. In contrast to conventional transmissions 
such as a manual or double clutch transmis¬ 
sion, a separate reversing gear is not re¬ 
quired for reverse gear in the range-extender 
transmission. The design offers a high level 
of freedom for the shaft arrangement due to 
elimination of multiple tooth meshes. This 
has major advantages with regard to the 
packaging space and integration. 


Function 


Power transmission 

Under the above mentioned premises of a 
vehicle that is mainly driven in electric 


mode, the hybrid transmission shown in 
Figure 3 enables the use of three gears 
with a total of only five gear meshes for 
both the electric motor and the internal 
combustion engine. The internal combus¬ 
tion engine can only be used above a 
speed of 10 km/h due to the omission of a 
launch device, which does not cause any 
restrictions because the electric motor 
covers these operating conditions. 

The design of the gear set enables the 
tractive force to be increased by the internal 
combustion engine in electric mode and 
vice versa, i.e. both drives assist each other 
reciprocally. In this regard, it is important to 
select the shift point so that the engage¬ 
ment of the internal combustion engine is 
not perceived as an impairment of comfort. 
Figure 4 shows a schematic sawtooth dia¬ 
gram for vehicle operation with a well 
charged battery. The shift point for engag¬ 
ing the electric motor can be freely selected 
from a large range. 



-EM/Gen - VKM 
— optional optional 


V veh in km/h 


Figure 4 Vehicle operation with the battery in a high state of charge 












418 



- EM/Gen - ICE 

— optional optional 


V„ ph in km/h 


Figure 5 Vehicle operation with the battery in a low state of charge 


Figure 5 shows the sawtooth diagram with 
the battery almost fully discharged. If the 
battery has an insufficient state of charge, it 
must be charged for a short time while the 
vehicle is stationary before starting. The 
battery can be recharged during vehicle op¬ 
eration at high speeds in the power-splitting 
mode and also if the driver wishes to accel¬ 
erate slowly to moderately while driving. The 
possibility of a breakdown due to a flat bat¬ 
tery can therefore be eliminated by using an 
intelligent charging strategy in conjunction 
with the maximum possible load point shift 
if the internal combustion engine during ve¬ 
hicle operation. 


Operating conditions 

Six different power flows, which each cor¬ 
respond to an operating condition, can be 
selected with the three ratio stages and 
three shifting elements that are indepen¬ 
dent of each other. 


Condition 1: 

Generator mode 

The parking lock can be activated in gener¬ 
ator mode. The internal combustion engine 
drives the generator via the gear wheel S3, 
Figure 6. 


•i h *2 



























Range-Extender 


28 


419 


Condition 2: 

Vehicle launch and reverse driving 

Vehicle launch is only possible in electric 
mode due to the selected ratios and the 
working ranges of the internal combustion 
engine and electric motor. The internal com¬ 
bustion engine is switched off, shift element 
S3 is open and S4 is closed. The drive func¬ 
tion both in a forwards and reverse direction 
is now taken over by the electric motor only, 
Figure 7. 



Figure 7 Power flow in electric mode 

All driving conditions can be overcome 
during urban operation in all-electric 
mode provided that the battery has a 
sufficient state of charge. Shifting is not 
necessary until approximately 50 km/h. 
Reversing with the internal combustion 
engine powertrain is not possible with 
the selected design, but is also not 
necessary. 

Condition 3: 

Hybrid city driving 

If the battery is in a low state of charge or 
when driving uphill, the internal combus¬ 
tion engine can even be used in first gear 
at speeds between 5 and 10 km/h de¬ 
pending on the specific design. The ratio 
enables crawling at slow speeds. The elec¬ 
tric motor can be used as a generator or a 



drive depending on the battery’s state of 
charge. First gear is mainly used by the in¬ 
ternal combustion engine, SI is closed and 
the electric motor can be connected via S3 
or S4. 

Condition 4: 

Hybrid drive at moderate speeds 

If the battery is in a low state of charge, 
the internal combustion engine can be 
operated in second gear at driving 
speeds above the speed range of first 
gear. Shift element S3 is closed again 
and SI or S2 is opened for this purpose, 
Figure 9. 



Figure 9 Power flow at moderate speeds in 
hybrid mode 







































420 


By designing the operating strategy appro¬ 
priately, it is also possible in condition 4 to 
use some of the torque produced by the in¬ 
ternal combustion engine for operating the 
electric motor as a generator via gear wheel 
S3 if the battery is in a low state of charge. 

Condition 5: 

Accelerating to high speeds 

If a vehicle equipped with Schaeffler’s 
range extender leaves the urban zone, the 
internal combustion engine can be en¬ 
gaged in order to rapidly reach high 
speeds. The internal combustion engine is 
then engaged via the third gear by closing 
the shifting element S2, Figure 10. Presyn¬ 
chronization is carried out by matching the 
speed of the internal combustion engine. 



The electric motor also provides accelerat¬ 
ing power via second gear so that high 
torque and good acceleration values can be 
achieved. 

Condition 6: 

Driving at high speed 

After the vehicle reaches the required 
speed, it is also advisable to direct the pow¬ 
er flow of the electric motor via third gear. 
Shifting element S3 is opened and S4 is 
closed at the same time for this purpose, 
Figure 11. 



Power split operation, in which the internal 
combustion engine is used to charge the 
battery via the generator, is also possible in 
this shifting condition when no electric ac¬ 
celeration power is requested. 

Figure 12 shows a summary of the pos¬ 
sible power flows and the required actuator 
positions. It is clear that due to the sequen¬ 
tial gearshift system, the torque flow during 
gearshifts is not interrupted because one of 
the two torque paths in the transmission 
is always closed. Furthermore, generator 
mode is possible at any time due to the con¬ 
nection between the electric motor and the 
wheel. 


Operating strategy 

The operating strategy for the range-ex¬ 
tender transmission is mainly dependent on 
three parameters: 

- The battery’s state of charge (SOC) 

- The torque required by the driver (posi¬ 
tion of the accelerator pedal) 

- Current speed range (urban/rural 
roads/highway). 

One possible operating strategy enables 
all-electric mode within a speed range of 
70 to a maximum of 120 km/h if the bat¬ 
tery is sufficiently charged. The internal 




























Range-Extender 


28 


421 


Operating condition 


S, 

S 2 

S 3 

S 4 

Neutral/generator 

1 

0 

0 

1 

0 

EM i/reverse 

2 

0 

0 

0 

1 

Hill mode 

3 

1 

0 

0 

1 

EM/EM + ICE i. 

4 

0 

0 

1 

1 

Power shift 

5 

0 

1 

0 

1 

EM/EM + ICE i 2 

6 

0 

1 

1 

0 


Figure 12 Shift pattern of the range-extender transmission (0 = open, 1 = closed) 


combustion engine can intervene and 
provide assistance during high accelera¬ 
tion if this is not prevented by an operating 
strategy which is aimed at ensuring emis¬ 
sion-free local mobility. The internal com¬ 
bustion engine is engaged above a de¬ 
fined speed, for example, 50 km/h. At 
higher speeds, particularly during opera¬ 
tion on highways or for long distances, the 
internal combustion engine is always 
switched on in order to achieve optimum 
overall efficiency. 

The internal combustion engine is also 
engaged in urban areas if the battery has 
a low state of charge. This is in accor¬ 
dance with current design criteria for se¬ 
rial hybrid drives. The share of power gen¬ 
erated electrically is greatly reduced and 
is completely switched off at high speeds. 
If the driver wishes to accelerate strongly, 
the electric power output is limited de¬ 
pending on the condition of the battery. 
The internal combustion engine provides 
the missing torque in order to fulfill the 
torque requirements of the driver. Firstly, 
this means the required driving perfor¬ 
mance can be achieved and secondly the 
electric motor can also serve as a genera¬ 
tor. This always ensures the battery is in a 
state of charge, which enables vehicle 


launch using only electric power. It is pos¬ 
sible to warn the driver via a signal that 
charging of the battery is urgently re¬ 
quired. With the selected design, this can 
be carried out at charging stations but 
also when the vehicle is stationary with 
the engine running (operating condition 1, 
see above). Alternatively, it is also possible 
to charge the battery using the engine’s 
generator if this was not omitted for cost 
reasons. Charging power of up to 3.6 kW 
can be achieved with this type of solution, 
which is equal to a normal AC power sup¬ 
ply connection. 

The internal combustion engine can still 
be started using the low-voltage battery if 
the high-voltage battery is fully depleted. 
However, launching is not possible immedi¬ 
ately because the vehicle must initially pro¬ 
duce sufficient power while stationary to 
continue the journey. 


Simplification to two gears? 

The initial approach of using a simple 
transmission with three gears can be 
further simplified by omitting the first 
gear, Figure 13. The internal combustion 
engine and electric motor can be oper- 




















































422 



Figure 13 Simplification to a two-speed design 

ated in both gears and gearshifts with¬ 
out an interruption of the tractive force 
are still possible. The design envelope, 
the mass as well as the complexity of 
the gearshift system can be minimized 
due to the reduced structure. The back¬ 
ground for this simplification is the opti¬ 
mized cost-benefit ratio of the system 
because a vehicle equipped with a two- 
speed solution or the three-speed de¬ 
sign must always launch using electric 
power only, although the internal com¬ 
bustion engine can only be engaged 
above 10 to 20 km/h. If the battery is in 
a high state of charge, a vehicle 
equipped with this variant would be 
driven in one gear using electric power 
only as far as possible and the internal 
combustion engine would not be en¬ 
gaged until high speeds are reached. All 


the operating conditions of hybrid driv¬ 
ing can also be realized. 

A vehicle breakdown due to the sys¬ 
tem-related necessity of a purely elec¬ 
tric launch and the use of only one elec¬ 
tric motor is unlikely due to the operating 
strategy. In addition, the recuperation 
characteristics can be designed so that 
the ease of electric launch is always en¬ 
sured. The minimalist approach with 
only two gears therefore offers a more 
cost-effective but still functional alter¬ 
native to the three-speed variant 
presented. 


Simulation 


The range-extender transmission devel¬ 
oped by Schaeffler has already under¬ 
gone initial testing in different simula¬ 
tions. It was important to determine the 
potential for reducing C0 2 and to test 
the behavior under extreme driving con¬ 
ditions. The focus is placed on the two- 
speed variant in order to show the pos¬ 
sibilities offered by Schaeffler’s concept 
with regard to the reduction in fuel con¬ 
sumption that can be achieved. 

The vehicle model designed in Matlab 
Simulink corresponds with typical values 
in the C segment. The assumed values 
were a vehicle weight of 1,450 kg and a 
four-cylinder naturally aspirated engine 
with a nominal power of 62 kW at 
5,000 rpm and maximum torque of 
130 Nm at 3,500 rpm. The electric motor 
has a nominal power of 60 kW and a 
torque of 200 Nm (continuous) or 300 Nm 
(peak). 

The battery size of 9 kWh was selected 
so that a guaranteed range of electric op¬ 
eration of 30 km can be achieved. This is a 
conservative assumption based on an op- 






Range-Extender 


28 


423 


Gearshift and NEDC 

Base ratio EM 

Long ratio EM 

Vehicle speed at EM shift in km/h 

45 

90 

45 

120 

Fuel consumption in g C0 2 /km 

58 

60 

58 

61 


Figure 14 Initial simulation results 

erating strategy in which a battery with a 
SOC of 40 % is regarded as “almost fully 
discharged”. 

In addition, two ratios for the second 
gear were modeled. With the base ratio 
of 3.8, the gearshift takes place at a 
speed of 90 km/h in operating condition 
6, while with the longer ratio of the elec¬ 
tric motor the gearshift is not made until 
120 km/h. Alternatively, a significantly 
lower shifting point of 45 km/h was 
used for the simulation. Figure 14 shows 
the simulation results for the NEDC test 
cycle. 

The results achieved in the initial simula¬ 
tion are encouraging. Firstly, the assump¬ 
tions, for example, with regard to the inertia 
class and the useable battery capacity are 
very conservative and these could be sig¬ 
nificantly more favorable in a lighter vehicle 
with optimized battery management. Sec¬ 
ondly, the consumption levels of the internal 
combustion engine could be reduced if a 
smaller engine with a higher power density 
is used as is increasingly the state-of-the- 
art. Thirdly, the SOC of the battery was 
higher after running the cycle than at the 
start, which is not a requirement in the cer¬ 
tification regulations. 10 to 12 g C0 2 /km 
alone could be saved by making a corre¬ 
sponding adjustment to the operating 
strategy. 

From the current perspective, it is 
likely that Schaeffler’s range-extender 
concept can achieve a certified emission 
level of 50 g C0 2 /km for the assumed 
C segment vehicle. A comparison shows 


what influence the useable battery ca¬ 
pacity and vehicle mass has on fuel con¬ 
sumption. A shorter distance can be 
driven using electric power only with a 
useable battery capacity of 60 % than if 
the useable battery capacity is increased 
to 75 %. For the cycle consumption, this 
increase in the battery capacity means 
approximate 14 % reduction in the fuel 
consumption or C0 2 emissions with the 
above assumptions. 

In addition to the benefits with regard 
to fuel consumption, the acceleration val¬ 
ues of 0 to 100 km/h in less than 11 sec¬ 
onds show that driving pleasure is not 
sacrificed either in combined or all-elec¬ 
tric mode. It was also important to verify 
the functional capability of the range- 
extender transmission during extreme 
driving maneuvers, particularly on steep 
gradients. The results are also encouraging 
here: 

- The electronically limited maximum 
speed of approximately 150-160 km/h 
is safely reached on a typical highway 
gradient of 6 %. 

- All gradients of practical relevance can 
be overcome at the low speeds in ac¬ 
tual road traffic. 

Even challenging requirements such as ac¬ 
celerating uphill can be carried out with the 
available battery charge either in combined 
or all-electric mode. 
















424 


Summary and outlook 


The Schaeffler range-extender concept 
shows potential for realizing a range- 
extended electric vehicle with significantly 
reduced system power and costs. The use 
of only one electric motor and a very simple 
transmission allows this concept to be inte¬ 
grated into a conventionally driven vehicle 
cost-effectively. 


Literature 


[1] Hofmann, P.: Hybridfahrzeuge: Ein alternatives 
Antriebskonzept fur die Zukunft. Wien: Springer, 
2010 

[2] Kirchner, E.: Leistungsubertragung in Fahrzeug- 
getrieben. Berlin: Springer, 2007, p. 559 ft. 

[3] Najork, R. et al.: What’s the Transmission 
Content in E-Mobility? 10 th International CTI 
Symposium, 2011 

[4] Hirt, G.; Massini, S.: Gearshift systems and 
synchronization: At the threshold of mecha- 
tronics. 9 th Schaeffler Symposium, 2010 

[5] Bruno M.: Transmission actuators. 

10 th Schaeffler Symposium 2014 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 



426 


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427 


One Idea, Many Applications 

Further development of the 
Schaeffler hybrid module 


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Dierk Reitz 
Willi Ruder 
Uwe Wagner 


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428 


Introduction 


Hybrid vehicles permitting one to two kilo¬ 
meters of driving using electric power - so- 
called full hybrids - are primarily found in 
upscale vehicle segments at present. These 
vehicles were equipped with automatic 
transmissions even before electrification, 
and the bell housing has prevailed as the 
installation location for the electric drive unit 
since this does not require the existing ve¬ 
hicle architecture to be fundamentally 
adapted for the hybrid versions. A module 
consisting of an automated disconnect 
clutch and an electric motor is incorporated 
between the internal combustion engine 
and the transmission. 

As early as 2010, Schaeffler was supply¬ 
ing integral components for such drive sys¬ 
tems; generally referred to as “P2 hybrids.” 
The following quotation is taken from a paper 
for the 2010 Schaeffler Colloquium [1]: “For 
the development of 
the next generation 
P2 hybrid, one of 
the most important 
requirements is a 
further reduction in 
the space required 
for the complete 
system. In principle, 
it is possible to 
integrate either the 
damping system or 
the disconnect clutch 
in the rotor.” 

The purpose 
of this paper is to 
demonstrate what 
stage of develop¬ 
ment Schaeffler has 
attained to date. The 
next step planned is 
to make use of the 
high fidelity control 


of an electric motor incorporated in the pow¬ 
ertrain in order to cancel out undesired tor¬ 
sional vibrations from the internal combustion 
engine. Finally, we will show that the chosen 
hybrid module design is also suitable for use 
with a 48-volt on-board electric system in 
combination with a manual transmission. 

The new generation of the 
high-voltage hybrid module 


Complete system 

A marked increase in electric power require¬ 
ments can be observed due to the trend to¬ 
wards plug-in vehicles, and hybrid vehicles 
will be able to meet the entire New European 
Driving Cycle (NEDC) in the future. A primary 
development goal for the next generation of 


Module 

housing 


Actuator 


Cooling channel 
(water) 



One way clutch 
Central bearing 


Figure 1 Cross-section of a second generation hybrid module 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3_29, © The Author(s) 2014 






Hybrid Modules 


29 


429 



Figure 2 Installation dimensions of the hybrid modules: on the left, the first generation of 2010; on 
the right, the current stage of development 


the Schaeffler hybrid module has therefore 
been to increase the power and torque den¬ 
sity, while at the same time reducing the 
design envelope required. The installation lo¬ 
cation - between the internal combustion 
engine and the transmission - is also to re¬ 
main unchanged. Moreover, as with the first 
generation, no modifications should be nec¬ 
essary to the hardware of the internal com¬ 
bustion engine or the transmission other 
than perhaps the addition of an electric 
pump for the transmission oil. 

The second generation of the Schaeffler 
hybrid module (Figure 1) falls in line with this 
market trend and allows very high torques of 
up to 800 Nm to be transferred. Transferring 
such high torques is made possible by a 
patented bifurcation of the power flow [2]. 
The torque of the internal combustion engine 
is channeled towards the transmission both 
via the closed disconnect clutch as well as 
via a parallel one-way clutch. 

In each instance, the torque passes to an 
intermediate shaft via a vibration damper. This 
shaft has a double bearing support: a ball 
bearing in the area of the clutch and, at the 


front end, a pilot bearing that can be integrat¬ 
ed into either the crankshaft or the damper. 
The output comes from the reaction plate of 
the disconnect clutch, which could be incor¬ 
porated completely into the rotor. 

The actuation and sensor elements that 
are functionally necessary have been fully 
integrated into the module. An electrome¬ 
chanical central release bearing optimized 
for this module takes care of actuating the 
clutch. A permanent magnet motor with an 
external rotor design is used to drive the ac¬ 
tuator. 22 magnets have been glued inside 
the bore of the rotor, while the raceway for 
the ball screw is mounted on the outside. 

The module has been designed in such a 
way that the disconnect clutch both starts 
the combustion engine and transfers the 
subsequent traction torque for powertrains 
with low torque requirements of up to 300 Nm. 
In order to achieve this torque, not only is 
the actuator’s ability to engage the clutch 
utilized but also its ability to pull the dia¬ 
phragm springs once the clutch has closed 
(push-pull principle). In order to use combus¬ 
tion engines with torques of over 300 Nm, a 











































430 



one-way clutch has 
been added to 
transfer the traction 
torque. This clutch is 
connected in paral¬ 
lel with the discon¬ 
nect clutch, thereby 
offering two advan¬ 
tages: For one thing, 
this arrangement al¬ 
lows high clutch dy¬ 
namics to be main¬ 
tained along with 
very good control 
quality; for another, 
the time-consuming 
adjustment proce¬ 
dure after the 
combustion engine Figure 3 
starts is no longer 
necessary since its speed is automatically 
matched by the one-way clutch when speed¬ 
ing up and coupling to the electric motor. 

Thanks to the layout chosen, it was pos¬ 
sible to downsize the installation dimen¬ 
sions considerably (Figure 2). The module’s 
outside diameter has been reduced by 
12 mm to 303 mm, while the overall length 
has been shortened from 152 mm to 135 mm 
depending on the performance of the elec¬ 
tric motor. 

The integral components of the hybrid 
module used should be standardized re¬ 
gardless of the application in order to keep 
system costs as low as possible. Among 
other things, this applies to: 

- the actuating elements 

- the central bearing support 

- the rotor support 

- the clutch 

- the rotor position encoder (resolver). 
The modular layout remains unchanged 
irrespective of whether a conventional 
stepped automatic transmission, a double 
clutch transmission (wet/dry), or a contin¬ 
uously variable transmission (CVT) is in¬ 
volved. Even manual transmissions can 


be hybridized with the layout chosen. The 
housing and rotating components on the 
engine and the transmission are the inter¬ 
faces that are specific to the customer or 
application. In other words, depending on 
the application, the vibration damper is 
adapted to the characteristics of the com¬ 
bustion engine used, and the drive plate 
to the transmission input. 

In order to achieve maximum system 
efficiency, the Schaeffler hybrid module 
has been developed as a dry system. Rel¬ 
evant to the cycle, the input power of the 
actuator is under 10 W. The bearing con¬ 
cept was optimized to such an extent that 
the drag torque is < 0.5 Nm during elec¬ 
tric-powered driving. The most important 
subsystem in this optimization process is 
the electric motor, the efficiency of which 
was able to be optimized to peak levels 
greater than 95 %. Any lost heat is con¬ 
veyed by special thermally-conductive 
potting and a cold water jacket. Concern¬ 
ing the losses of the individual compo¬ 
nents, a thermal model provides confir¬ 
mation of the maximum temperatures 
realized during operation. 
























Hybrid Modules 


29 


431 


If the continuous output required by the ve¬ 
hicle is very close to the required peak out¬ 
put of the electric drive, it is also possible to 
design the module with a “wet” layout. In 
doing so, better heat dissipation is provided 
by oil volume flow around the rotor and the 
coil ends. Since dividing up the electric mo¬ 
tor and the disconnect clutch into wet and 
dry regions would increase design com¬ 
plexity, having a wet clutch as well would be 
advantageous. The benefits of higher-per¬ 
formance cooling are offset by the efficiency 
disadvantage due to the added energy re¬ 
quired for supplying the cooling oil as well 
the increase in drag losses in the gap be¬ 
tween the rotor and the stator. Figure 3 
shows the complete module in the wet in¬ 
stallation space with an optional allowance 
(dotted line) for a damper on the output 
side. The clutch can be optionally actuated 
via a hydrostatic actuator [3] or through the 
transmission. 


Clutch system 

Since the traction torque of the combus¬ 
tion engine is transferred via the one-way 
clutch, the basis for the clutch design 
mainly depends on the torque that is re¬ 
quired for combustion engine re-start. In 
doing so, it is necessary to use a high de¬ 
gree of precision to set a torque of approx. 
110 Nm (depending on the combustion en¬ 
gine) in the < 100 ms that is required to ac¬ 
celerate the crankshaft. Motor speeds dur¬ 
ing electric-powered driving of up to 
4,000 rpm result in a load cycle during 
which the crankshaft is accelerated to the 
corresponding differential speed. The wear 
reserve required by the linings is based on 
this cycle and a total number of 800,000 start¬ 
ups over the operating life of a plug-in 
hybrid. 

As was already described above, the 
clutch is actuated by an electromechanical 
central release bearing (ECRB). The neces- 



Figure 4 Interface of the electromechanical 
central release bearing (ECRB) in 
installation position 


sary axial motion is generated via a ball 
screw drive directly linked to the rotor mo¬ 
tion by means of a carriage running along a 
corresponding track (Figure 4). The actual 
release bearing that actuates the dia¬ 
phragm spring is mounted on the carriage. 
The stroke from the touch point to the 
100 Nm point of the clutch can be traveled 
in less than 100 ms, and the release force of 
the actuator is at most 1,800 N in the cho¬ 
sen design. Unlike with a hydraulic design, 
the electromechanical actuator can transfer 
forces in two directions. The fact that the 
clutch has been designed as a so-called 
“push-pull clutch” makes it possible to in¬ 
crease the transferable torque considerably. 
When not actuated, the clutch is closed. 


Electric motor 

The Schaeffler hybrid module uses a per¬ 
manently excited synchronous motor that 
possesses high reluctance, thereby reduc¬ 
ing the quantity of rare earth elements re¬ 
quired. Since the clutch sitting inside the 
rotor is standardized, the electric motor al¬ 
ways has the same inside diameter; thanks 
to the modular design, the outside diameter 
is adjusted depending on the required ca- 


432 



41 kW Motor 

80 kW Motor 

Type 

PSM 

PSM 

Torque peak (10 s) 

180 Nm 

280 Nm 

des. 

100 Nm 

160 Nm 

Speed operation 

7,000 rpm 

7,000 rpm 

burst 

> 10,200 rpm 

> 10,200 rpm 

Power peak (10 s.) 

41 kW 

80 kW 

des. 

25 kW 

48 kW 

Efficiency 1,500 - 2,500 rpm 

> 95 % 

> 95 % 

Dimensions 

D 270 mm, d 182 mm 

D 270 mm, d 182 mm 


L 86 mm 

L 115 mm 

Design voltage 

264 V 

264 V 


Figure 5 Technical data of the EM-H-270 

pacity as well as on the available radial 
space. The available diameters are 260 mm, 
270 mm, and 290 mm. Further adjustment 
of the length of the electric motor results in 
an almost infinitely variable matching of mo¬ 
tor performance to the application require¬ 
ments. 

The following table shows the designs of 
two electric motors of this type for plug-in 
hybrid vehicles of the B/C and D segments; 
(Figure 5). 



0 2,000 4,000 6,000 


Speed in rpm 

Figure 6 Reluctance percentage of electric 
motor EM-H-270-86 


Following the electric motor design all the 
way to vehicle testing is part of developing 
the complete hybrid module at Schaeffler. 
As described above, the electric motor is 
designed with a high degree of reluctance. 
This initially results in the advantage that 
peak output can be provided up to high 
speeds. Furthermore, the efficiency of the 
upper speed range is clearly improved, 
and self-heating is reduced by cutting ed¬ 
dy-current losses in the magnets, thereby 
simplifying rotor cooling. The interdepen¬ 
dence of torque and speed is represented 
in Figure 6. 


Power electronics 

Power electronics which, in future gen¬ 
erations, will be slated as a hybrid mod¬ 
ule component, are still in the preliminary 
development stage (Figure 7). By using 
new electronic components, it is possi¬ 
ble to achieve dimensions that are con¬ 
siderably more compact, thereby en¬ 
abling them to be integrated in the 
module despite being positioned below 
the powertrain. This way the disadvan¬ 
tages involved with external wiring of 
power electronics and motors (costs, 
EMC, etc.) can be avoided. 

















Hybrid Modules 


29 


433 


Active vibration 
damping 



Figure 7 Hybrid module with directly mounted 
power electronics 

The goal of production development is 
standardization on the functional level. The 
following advantages result when control 
and power elements are separated: 

- The one control unit, once developed, 
can be reused again and again, since, 
as a rule, the basic functions remain 
nearly the same for all engine-power 
classes. 

- The control and power units can thus 
be joined as flexibly as required by 
each existing installation space. 

- The power output stages are freely 
scalable, and can be integrated into the 
system; at present, their designs range 
from 300 W to 100 kW. 

Besides reduced cabling expenditures, 
the linked cooling and one-piece housing 
design result in further cost and weight 
savings at the system level. Moreover, the 
overall installation space required is less 
than with discrete components. 


Active damping of speed fluctuations in 
the powertrain by using an electric motor 
is an idea that was already pursued in the 
1990s. Solutions aiming to provide the 
crankshaft with complete damping mainly 
failed due to the power demands involved. 
A basic requirement for effectively using 
this function is connecting a damper up¬ 
stream of the electric motor. This pow¬ 
ertrain layout is given by the module and 
enables designers to focus on the cancel¬ 
ing function by targeting the main order of 
the combustion engine. Any additional 
energy requirements are thus limited to 
recording (sensor elements) and pre¬ 
cise phase regulation in the range up to 
approx. 80 Hz. 

The basic function of such a system 
is portrayed schematically in Figure 8. In 
the process, the electric motor only 
smoothes out the already damped main 
order on the secondary side of the 
damper, which is possible with an ex¬ 
traordinarily low amount of energy. Thus, 
the torque required for cancelation (de¬ 
pending on the damper design) drops to 
about one tenth of the value required by 
the electric motor mounted directly on 
top of the crankshaft. 

Active vibration cancelation is being 
developed with the goal of achieving ide¬ 
al comfort and efficiency in an available 
installation space by means of mechani¬ 
cal damping, active vibration cancel¬ 
ation, and damping through starting ele¬ 
ment micro-slip. In the actual design 
process, an ideal compromise is struck 
between these two objectives that is 
primarily oriented to the required NVH 
vehicle behavior and the energy input 
required. 



434 


ICE 


Clutch + 
hybrid eMot 


Drivetrain 


Vehicle 




Figure 8 Functional diagram of active vibration damping in a hybrid powertrain 


Particularly in the event that any pow¬ 
ertrain resonances occur, the electric mo¬ 
tor can be used to actively reduce them in 
a narrow speed range. In some cases, 
depending on the rigidity of the transmis¬ 
sion, this approach will allow a second 
damper positioned downstream of the hy¬ 
brid module to be eliminated. 

By means of a simulation, it was pos¬ 
sible to show that interplay between ac¬ 
tive damping via the electric motor and 
starting clutch micro-slip offers ideal en¬ 
ergy conditions coupled with a high de¬ 
gree of vibration comfort (Figure 9). While 
the slippage generated in the clutch at 
1,200 rpm results in power losses of 
700 W, the electric motor operates at 
350 W in this range. For speeds greater 
than 1,500 rpm, however, slippage regulation 
is more energy-efficient, while the power 
requirements placed on the electric motor 
continue to climb. Nevertheless, this ideal 
depends on the specific application and 
can therefore vary. What must be kept in 
mind is that these power losses refer to 
full-throttle operation of the internal com¬ 
bustion engine. In relevant cycles, these 
power requirements are much smaller. 

For a long time, active vibration can¬ 
celation and the associated rapid chang¬ 
es in discharge current required for this 
strategy appeared to have a negative ef¬ 
fect on the operating life of the battery. 


Since, however, the overall battery size 
became significantly smaller with the in¬ 
troduction of plug-in hybrid vehicles, the 
energy throughput for cancelation is re¬ 
duced accordingly, down to less than 2 %. 
What is more, it has since been empiri¬ 
cally proven that cell damage due to cyclic 
micro-discharge events is much less than 
originally feared. This is especially true 
when there is no ion conversion in the 
battery, i.e. if the current is regulated with¬ 
in the generator or drive mode. [4] 

The development of special control al¬ 
gorithms for active vibration cancelation is 
currently being tested on internal combus¬ 
tion engine test rigs and in vehicle tests. 


The 48-volt hybrid module 


Motivation 

The first steps with hybridization can natu¬ 
rally be taken using lower power systems. 
For one thing, this approach makes it pos¬ 
sible for the voltage to stay below the safe¬ 
ty-critical value of 60 V. What is more, the 
expenditure for the complete system can be 
decreased considerably. In particular, the 













Hybrid Modules 


29 


435 



— Engine — Clutch slip control 

— Transmission input shaft with Arc spring damper — active damping 

— Transmission input shaft with Arc spring damper + CPA 
-- target 


Figure 9 Power losses from combined active vibration damping of an electric motor and slippage 
monitoring, depending on the engine speed 


energy storage device is reduced in size by 
a factor of three, with a useful capacity of 
approx. 300 Wh. If the new voltage level 
is used, equipping vehicles with a mildly 
hybridized drive is all that is necessary to 
make substantial consumption savings 
possible. Simulations show, for instance, 
that a 12 kW electric motor with an asyn¬ 
chronous design can lower the consump¬ 
tion in the standard European driving cycle 
by around 10 %. 

When a hybrid module with an inte¬ 
grated transmission is used, this system 
is more efficient due to the fact that the 
gear ratio can also be used to operate 
the electric motor with ideal efficiency. 


Compared to HV systems, ideal efficien¬ 
cy comes at lower speeds since the 
combustion engine runs more of the 
time, thereby determining the speed of 
the electric motor. 

A further and significant improvement in 
fuel consumption can be achieved by re¬ 
placing today’s conventional asynchronous 
motor with a synchronous motor with a 
higher power density. The layout and the 
effect on consumption are explained in 
greater detail in the chapter on the 48-volt 
electric motor. 











436 


Combination with a manual 
transmission 

As a rule, the structural design of a hybrid 
module employs the same concepts when 
used with a 48-volt application as with a 
high voltage application. One particular 
challenge stems from the fact that manual 
transmissions are still frequently used today 
in price-sensitive compact and mid-sized 
segments. 

Using function matrices, Schaeffler has 
chosen four designs from a host of possible 
topologies, studying the specific advantag¬ 
es and disadvantages that distinguish them 
when it comes to linking the combustion 
engine to the fixed-transmission hybrid 
module: 

- Impulse clutch 

- Adaptation of the existing hybrid mod¬ 
ule for 48 V without further changes 

- One-way clutch combined with a lock¬ 
up clutch - coaxial 

- One-way clutch combined with a lock¬ 
up clutch - axially parallel 

Due to this module’s limited capacity, it is 
not feasible to start the engine via the 
disconnect clutch. A basic distinction 
was therefore made initially between 
continuing to start the combustion en¬ 
gine via the conventional starter or by 
means of the rotating masses. This iner¬ 
tia is utilized by using an impulse clutch 
(Figure 10), and the combustion engine is 
brought up to speed solely by closing the 
clutch. Involved here is a very rapidly ac¬ 
tuating clutch that has to be able to 
transfer very high fluctuating torques of 
up to 1,500 Nm. This clutch is not modu¬ 
lated, but rather is either completely 
opened or closed. An important require¬ 
ment for this system is reducing the 
crankshaft related inertia to a minimum. 
The complete hybrid module is installed 
along with the electric motor on the side 
of the crankshaft and is supported by 


e 



Figure 10 Structural design of an impulse 
clutch 

two rows of ball bearings. The clutches 
are actuated by two release bearings 
which are controlled via a diaphragm 
spring (startup clutch) and a lever spring 
(disconnect clutch). The expenditure in¬ 
volved in the design is similar to that of a 
double clutch in a double clutch trans¬ 
mission. 

Other combustion engine concepts in¬ 
volve a separate starter system in order to 
re-couple the engine after a coast/drive 
phase. In this way, the use of a one-way 
clutch as a low-cost alternative to the 
standard clutch is conceivable. With this 
topology, the combustion engine is start¬ 
ed by the conventional starter and me¬ 
chanically coupled once it reaches the 
speed of the electric motor. 

The disadvantage of such a solution is 
that a vehicle parked in first gear would no 
longer have a “gear brake.” Since the one- 




Hybrid Modules 


29 


437 


way clutch does not block the powertrain 
in one direction of rotation, the vehicle 
would start to roll if the parking brake is 
not set on a hill. 

This disadvantage can be avoided by 
equipping the one way clutch with a lock¬ 
ing function. To this end, for instance, a 
shift sleeve can be used that provides a 
form-fit connection between the second¬ 
ary damper side and the rotor holder. 
This spline connection is initially closed 
and can be opened via a tie-rod linked to 
the starting clutch. A hydrostatic clutch 
actuator (HCA) - produced by Schaeffler 
recently for double clutch transmissions 
- is used as for actuation [5]. Moreover, 
particularly with a small energy storage 
device and a high state-of-charge (SOC), 
it is necessary to be able to re-couple the 
combustion engine in order for it to take 
up the driving torque. This function is fa¬ 
cilitated by one-way clutch locking as 
well. Since actuation of the already exist¬ 
ing starting clutch can also be used for 
the shift sleeve, no additional actuator is 
necessary. 

Locking the one-way clutch also en¬ 
sures that the driving feel does not change 
for the driver during the drive phase when 
the energy storage device has a high 
SOC. The combustion engine then takes 
up the driving torque again. 

With respect to time and comfort, a 
warm start from a stop-start situation can 
be initiated directly via the 48-volt electric 
motor. 

The configuration in the installation 
space can be either axially parallel or co¬ 
axial. An axially parallel layout permits the 
use of an asynchronous motor which al¬ 
ready exists due to the development of 
the belt-driven starter generator. In this 
system, torque is transferred via a belt 
with two-fold to three-fold ratio. 

An essential requirement for realizing 
such a layout is that the installation space 
above the clutch bell must be able to be 


used for the electric motor (Figure 11). 
This appears to be feasible, especially 
with front-transverse powertrains. This 
layout results in the least amount of pow¬ 
ertrain lengthening. 



Figure 11 Typology and structural design of 
the hybrid module with an 
asynchronous motor in parallel 
arrangement 
















438 



Figure 12 Topology and structural design of the 
hybrid module with a locking one-way 
clutch in coaxial arrangement 

In comparison, the coaxial model is shown 
in Figure 12. The one-way clutch, one-way 
clutch locking mechanism, and, in part, the 
starting clutch are radially nested under the 
rotor. Thanks to good thermal coupling of 
the electric motor, the stator can be air¬ 
cooled. 


48-volt PSM electric motor 


For cost reasons, asynchronous motors are 
primarily being introduced to the market for 
mild hybrid applications; a fact that is also 



2,000 4,000 6,000 

Speed in rpm 


90 

80 

70 



Figure 13 Efficiency range of a permanently- 

excited synchronous motor employed 
in the 48-volt hybrid module 


recommended in Figure 11. In order to take 
advantage of the available installation 
space, especially with the coaxial layout 
shown in Figure 12, it is preferable to use 
a permanently-excited synchronous motor 
(PSM), the power density of which is up to 
30 % greater depending on the demand. In 
addition, the greater efficiency of the PSM in 
conjunction with optimizing the speed range 
relevant to the cycle (Figure 13), yet again re¬ 
sults in a markedly improved C0 2 balance. 

The increased efficiency of the PSM 
compared to the ASM results in an added 
consumption benefit of up to 3.5 %. This 
delta is due to improved efficiency - and 
also to the fact that any recuperated energy 
that cannot be directly reused in the on¬ 
board electric system “flows through” the 
electric motor multiple times. 


Operation strategy with a manual 
transmission 

Driver acceptability is vital for successfully 
launching mild hybrid vehicles with manual 
transmissions in the market. An essential 
component for this is that the powertrain 
always delivers the acceleration required 
by the driver. The power distribution be¬ 
tween the combustion engine and the 
electric motor must be configured in such 
a way that it is practically imperceptible to 
the driver. 

The combustion engine not only 
switches on automatically when a lot of 
power is required, but also for high electric 
motor rpm levels when the driver does not 
upshift. For acoustics reasons as well, the 
switch-on point is at about 3,500 rpm. In 
order to maintain good vehicle drivability, 
the constant speed of the electric motor 
in the consumption cycle is limited to 
50 km/h despite reduced C0 2 -emission 
potential. At higher speeds, the combus¬ 
tion engine is only decoupled during the 
drive phase. 




















Hybrid Modules 


29 


439 


As has already been recommended for pure 
start-stop systems, due to the sailing func¬ 
tion and improved market acceptance, the 
starting clutch has been automated for this 
hybrid module. This configuration makes it 
possible to use creep in the starting clutch 
in order to be able to realize a very comfort¬ 
able motor startup feel. 


Outlook 


At present, B-model testing is being con¬ 
ducted on the next generation of high- 
voltage models of the Schaeffler hybrid 
module. The current projects cover all 
conventional automatic transmissions. In 
the process, it has become clear that the 
hardware for the various configurations 
can indeed be designed with standard¬ 
ized basic components. 

With a P2 hybrid module, it is already 
possible to realize consumption benefits of 
around 10 % based on a 48-volt system 
with an asynchronous motor. Additional 
potential of up to 3.5 % is available by us¬ 
ing a synchronous motor. Combined with 
the possibility of moving the vehicle at 
low speeds of up to approx. 15 km/h us¬ 
ing only electricity, this module makes for 
entry-level electrification that is ideal. 

Thanks to mild hybridization, manual 
transmissions are fit for the next generation. 
Vehicle testing will show whether drivers ac¬ 
cept the extra functions without noticing 
any sacrifices in comfort. Adjusted acoustic 
factors and automated clutch action will 
help with this. 


Literature 


[1] Wagner, U.; Reitz, D.: The future comes 
automatic: Efficient automatic transmissions 
provide a basis for hybrid capable drive trains. 
9 th Schaeffler Symposium, 2010 

[2] Schutzrecht DE 102012207941A1: 

Hybridmodul fur einen Triebstrang eines 
Fahrzeuges. 

[3] Mueller, B.: Transmission Actuation: Downsiz¬ 
ing Complexity, Upsizing Performance. 

10 th Schaeffler Symposium, 2014 

[4] Wang, J.: Cycle-life model for graphite-LiFeP04 
cells. Journal of Power Sources, 2011 

[5] Mueller, B.; et. al.: Smaller, Smoother, Smarter: 
Advance development components for 
double-clutch transmissions. 9 th Schaeffler 
Symposium, 2010 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 




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More Agile in the City 

Schaeffler’s 
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Dr. Raphael Fischer 


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442 


Introduction 


Wheel hub drives offer a high theoretical po¬ 
tential for designing completely new vehicle 
architectures. They are particularly attractive 
for small, highly maneuverable city vehicles 
with battery-electric drive [1]. The demand 
for these vehicles will continue to rise in the 
future against the background of advancing 
urbanization worldwide and stricter environ¬ 
mental protection specifications. The target 
markets are particularly the rapidly growing 
cities in Asia and North and South America. 

The use of a wheel hub drive has various 
advantages for drivers: 

- Usable space is gained in the vehicle 
body. No “engine compartment” is re¬ 
quired, which means new body de¬ 
signs are possible. 

- The wheel turning angle can be in¬ 
creased because drive shafts are not 
required. Maneuverability is significantly 
improved from the customer’s perspec¬ 
tive. This also applies when the vehicle 
has a driven rear axle because targeted 
assisted steering with torque vectoring 
can be operated on road surfaces with a 
low friction coefficient. 

- Driving pleasure and safety are increased 
because the control quality of the drive is 
higher than that of central drive systems 
because power is transmitted directly 
without a transmission and side shafts. 
These conventional target values of auto¬ 
mobile development will be decisive for 
achieving customer acceptance of small 
city cars. In our opinion, electric vehicles 
will not be marketable on solely rational 
grounds - small traffic area and a good 
C0 2 footprint. 

- Driving will be significantly simpler: For 
example, when starting on ice only the 
maximum transmissible torque is ap¬ 
plied even if the accelerator pedal is 
fully depressed. 


- Last but not least, passive safety is also 
increased because conventional drive 
units with high masses fitted in the en¬ 
gine compartment will no longer enter 
the vehicle interior if a frontal impact 
occurs [2]. 

However, the design of wheel hub drives 
means a radical break with previous design 
criteria. Currently, it is not advisable to equip 
a “general purpose vehicle” with an electric 
wheel hub drive because, due to the torque 
characteristics of an electric motor, a choice 
must be made between a high starting 
torque and a limited final speed or a high fi¬ 
nal speed and a low starting torque. Electric 
vehicles with a center drive solve this con¬ 
flict of objectives due to the installation of a 
suitably sized electric motor with a trans¬ 
mission, which is not advisable in a wheel 
because of the restricted space. This article 
therefore only covers drives for city vehicles 
which reach a maximum speed of 130 km/h 
and are suitable for short interurban routes, 
but not for vehicles used by frequent driv¬ 
ers. 

Since 2007, Schaeffler has been work¬ 
ing intensively to realize the theoretical ad¬ 
vantages of this drive principle. Initially, the 
achievable torque was only 84 Nm, which 
was far from adequate for most driving con¬ 
ditions. However, even at the time it could 
be shown that in principle an electric motor 
can be integrated into a wheel. A prototype 
was subsequently built based on an Opel 
Corsa, which already had a continuous 
torque of 200 Nm (maximum 530 Nm) per 
wheel. The electric motor but not the power 
electronics were fitted in the wheel on this 
prototype. 

Schaeffler took into consideration the 
experience gained from this pre-production 
model during further development of the 
wheel hub drive. The main focus was on ful¬ 
filling customer requirements for higher 
torque and achieving a higher level of inte¬ 
gration. Since 2013, tests have been carried 
out on a further wheel hub drive jointly de- 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3 30, © The Author(s) 2014 



Wheel Hub Drives 


30 


443 


signed with Ford. We are reporting about 
the design and initial results from the driving 
dynamics tests in this article. 

Concepts for an electric 
wheel drive 


In principle, the drive force produced by 
one or more electric motors can be trans¬ 
mitted in different ways. Traction motors 
integrated into the transmission dominate 
in the hybrid and electric vehicles currently 
produced. Schaeffler has developed a hy¬ 
brid module to volume production readi¬ 
ness, which will be presented in a separate 
article [2]. Schaeffler is also currently test¬ 
ing the potential of a range extender trans¬ 
mission [3]. There are a number of different 
topologies for wheel drives (Figure 1). 

Conventional electric drives are currently 
designed as center drives. Schaeffler is 


currently developing a volume production- 
capable center drive solution to volume pro¬ 
duction readiness [5]. The electric motor 
can be used in combination with a light¬ 
weight differential to control the distribution 
of torque to individual wheels. This type of 
electric axle is particularly suitable for sporty 
electric vehicles and vehicles suitable for 
covering long distances with a plug-in hy¬ 
brid drive. 

Wheel hub drives have been used until 
now in small-volume production in com¬ 
mercial vehicles, urban buses and in the 
military sector. Wheel hub drives are cur¬ 
rently only found as prototypes in passen¬ 
ger cars. There are a large number of so¬ 
lutions, in which two separate motors 
fitted in a center housing arrangement 
each drive a wheel via side shafts. This 
solution is not preferred by Schaeffler be¬ 
cause the significant advantages offered 
by the wheel hub drive (such as the very 
efficient use of space and high level of 
maneuverability) cannot be completely 
utilized. 



Figure 1 Topologies for electric wheel drives in road vehicles 














444 


Schaeffler is using a highly-integrated de¬ 
sign of wheel hub drive for future city vehi¬ 
cles, in which the wheel becomes a power 
module. At the same time, drive systems 
positioned close to the wheel are also being 
tested as part of research and advanced 
development projects. Two examples of 
projects are mentioned, which pursue differ¬ 
ent approaches with the drive positioned 
close to the wheel. 

In the “FAIR” project [6] carried out joint¬ 
ly with BMW and the Deutschen Zentrum 
fur Luft- und Raumfahrt (National Aeronau¬ 
tics and Space Research Center of the Fed¬ 
eral Republic of Germany) a gear system 
was integrated into the wheel to reduce the 
speed of the electric motor mounted on the 
vehicle in front of the wheel and to decouple 
the vertical motion of the wheel from the 
drive (Figure 2). 



Figure 2 Cross-section through the drive 

positioned close to the wheel from 
the FAIR project 


Electric motor 
bearing support 



Split suspension spring 


Figure 3 Design of a wheel drive positioned 
close to the wheel with a transmis¬ 
sion integrated into the wheel 

As part of an advanced development proj¬ 
ect, a variant was tested at Schaeffler, in 
which the drive motor is supported using a 
split suspension spring and connected to a 
transmission via a short side shaft (Figure 3). 
This means the motor only moves in step 
with the spring motion of the wheel to a re¬ 
duced extent. 

The following sections of this article fo¬ 
cus on wheel hub drives, which in our opin¬ 
ion are the most suitable drive arrangement 
for electric city vehicles. 

A new generation of 
wheel hub drive 


Design and construction 

In 2010, Schaeffler set itself the goal of de¬ 
signing a wheel with a highly-integrated 
drive, which incorporates the electric motor 
and power electronics in addition to the 
conventional wheel components such as 



















Wheel Hub Drives 


30 


445 


the service brake. For the first time there is 
no requirement for all the pulsed cables laid 
through the vehicle, which is also advanta¬ 
geous with regard to electromagnetic com¬ 
patibility. This can also be managed with 
other arrangements of system components, 
but results in additional coordination require¬ 
ments. 

The high level of integration is, of 
course, a major challenge for develop¬ 
ment engineers: The total available design 
envelope is only 16 liters. The task of ac¬ 
commodating the complete drive in a 
wheel with a 16-inch diameter was solved 
by carrying out a large number of individ¬ 
ual optimization measures (Figure 4). A 
width of approximately 200 mm gives a 
conventional tire dimension (195 or 205 
tire). The tire corresponds to a volume- 
produced tire both with regard to the di¬ 
mensions and the design. However, the 
steel rim is a special design, because the 
wheel disk is connected to the rim shoul¬ 
der instead of the drop center as in a nor¬ 
mal wheel (semi full face). Forged steel 
rims could be used in a later volume-pro¬ 
duced design, which would combine an 
elegant design with a high load carrying 
capacity. The hole circle required for 
screw mounting and centering is compat¬ 
ible with current standard connections. 
The entire design has been selected to 
ensure that a tire change does not cause 
additional work. 

The magnetic gap with a diameter of 
278 mm and a width of 80 mm must be 
maintained within very close tolerances to 
ensure optimum function of the electric mo¬ 
tor. The air gap has a difference in radius of 
1 mm. Any tilting, which would allow the 
stator and rotor to rub against each other, 
must be prevented in order to avoid corrod¬ 
ing surfaces. The wheel bearing is therefore 
of very rigid design. The rigidity is approxi¬ 
mately twice as high as in a conventional 
wheel bearing. Locking of the wheel if con¬ 
tact occurs between the stator and rotor 


Electronics 



Wheel 


bearing 

Brake 


Figure 4 Design of Schaeffler’s wheel hub 
drive 


has been ruled out by carrying out in-house 
tests. Contrary to frequently expressed as¬ 
sumptions, rubbing does not cause unsta¬ 
ble driving dynamics. 

Wear does not occur because the 
dimensions of a rolling bearing do not 
change significantly during an operating 
period of 200,000 km. A wheel bearing is 
also volume production technology, which 
can be manufactured with current tools and 
machines. 

The modified support plate of the 
standard drum brake is used to integrate 
the electric motor into the wheel. The wa¬ 
ter-cooled stator is supported by the 
brake anchorplate, which is extended 
and made thicker in the direction of the 
outer side of the wheel. The rotor is locat¬ 
ed on a flange with the brake drum. How¬ 
ever, the brake is not omitted but is avail¬ 
able as a redundancy level and parking 
brake. Previous operation of the prototype 
has shown that even during trips in the 
mountains (long descents with an 18 % 
gradient) braking can be carried out with 
the electric motor only. 





446 


The drive unit is sealed with a contact lip 
seal, which was derived from an indus¬ 
trial application. The seal integrity is in 
accordance with the standard for wheel 
bearings so that no moisture can enter 
even if there is exposure to a high-pres¬ 
sure cleaner. The precondition for this is 
a corresponding seal design so that it is 
even protected against high water pres¬ 
sure. 


Electric/electronic components 

A permanently excited synchronous mo¬ 
tor was integrated into the available de¬ 
sign envelope, which produces 350 Nm of 
continuous torque even under unsuitable 
temperature conditions. The maximum 
achievable torque is 700 Nm per wheel, 
i.e. 1,400 Nm for the axle. The starting 
torque is even high enough to enable the 
tested prototype with four occupants to 
start on a 25 % gradient. The electric mo¬ 
tor design was selected to ensure a uni¬ 
form torque output up to high speeds. 
The total continuous torque is available up 
to a traveling speed of 100 km/h. The out¬ 
put of the electric motor is 33 kW (contin¬ 
uous) and 45 kW (peak), whereby this 


value should not be overestimated and 
can be calculated from P = m ■ w directly. 
The specified values apply for operation 
with a voltage of 360/420 V. 

Prior experience with the prototype 
built in 2010 has shown that air cooling is 
not sufficient to produce the high contin¬ 
uous torque required. This is particularly 
true if typical automotive worst case sce¬ 
narios are taken into consideration in the 
design - for example, a hill start with a 
low speed at a high outside temperature 
(40 °C). The decision to design a water- 
cooled unit was made at an early stage for 
this reason. Cooling is carried out with 
conventional coolant based on glycol. The 
coolant firstly flows through the power 
electronics and electric motor stator and 
then reverses and is returned in a coun¬ 
terflow process. The drive contains a low 
volume of coolant. The normal air/water 
heat exchanger, which is also fitted in the 
front end of vehicles with an internal com¬ 
bustion engine, is used as a heat ex¬ 
changer. 

The electronic components required 
for control are also fitted in the wheel. This 
applies for the high-voltage power elec¬ 
tronics as well as the low-voltage motor 
control system. The arrangement of the 




Figure 5 Test stand run with the new drive (right); efficiency data map of the electric motor with 
operating points from the ARTEMIS cycle (left) 







Wheel Hub Drives 


30 


447 


power electronics has been selected so 
that only a very small distance must be 
covered by pulsed cables to the electric 
motor. 

The drive is only controlled locally in the 
wheel to a limited extent. The torque re¬ 
quirement is passed from the general con¬ 
trol unit, on which the driving strategy is 
stored, to the controller in the wheel, which 
is responsible for controlling and monitoring 
the electric motor. Requirements with re¬ 
gard to driving dynamics are placed by the 
vehicle’s safety computer and also imple¬ 
mented in the wheel. 


Test stand runs 

The drive was firstly put into operation on 
a test stand before it was integrated into 
the vehicle. The control system is initially 
adjusted to match the electric motor and 
power electronics. At the same time, 
characteristics such as efficiency, con¬ 
tinuous and peak torque as well as the 
thermal behavior of the system are de¬ 
fined (Figure 5). 

After initial operation, strength and ri¬ 
gidity tests were carried out on an internal 
drum test stand specially developed for 
this purpose at the Fraunhofer LBF in 
Darmstadt (Figure 6). The drive is mount¬ 
ed on a hexapod and placed on the inter¬ 
nal surface of a rotating drum. The drum 
is provided with lateral thrust ribs, which 
can be used to apply lateral loads to the 
wheel similar to the contact with a curb¬ 
stone or when cornering sharply. The aim 
of the tests was to check the lateral rigid¬ 
ity in order to ensure no rubbing occurs 
between the stator and rotor even under 
extreme lateral loads. 

The tests have also shown that no 
rubbing occurs between the stator and 
the rotor even with increased air pressure 
and a load that causes destruction of the 
tire. 



Figure 6 Design of a test stand for testing the 
function of the wheel hub drive 
acting under mechanical forces 

Vehicle integration 

A Schaeffler wheel hub drive at the current 
level of development was fitted in a Ford 
Fiesta used as a test vehicle in collabora¬ 
tion with the Ford Research Center Aachen 
(Figure 7). The high-voltage battery is inte¬ 
grated into what was previously the engine 
compartment. In addition to the fitting of 
high-voltage components, the adjustment 
between the engine and vehicle control 
system involved a significant outlay. In par¬ 
ticular, the restbus simulation, i.e. the sim¬ 
ulation of signals for omitted components 
such as the internal combustion engine by 



Figure 7 Mounting position of the wheel hub 
drive in a test vehicle 






448 


means of software is very challenging. In 
addition, essential chassis components 
such as the suspension and damping were 
adjusted to match the characteristics of 
the drive. 

The system weight for the complete 
wheel hub drive is 53 kilograms per wheel. 
It must be taken into consideration that the 
total vehicle weight is not increased com¬ 
pared with an identical vehicle fitted with a 
diesel engine (1,290 kg empty). This in¬ 
cludes a lithium-ion battery with a nominal 
capacity of 16.2 kWh. The axle load distri¬ 
bution is also the same as the volume-pro¬ 
duced vehicle. 

A variety of driving dynamics tests were 
carried out with the test vehicle at a testing 
site. The tests showed that the prototype 
was at least equal to a comparably driven 
volume-produced vehicle up to a speed of 
130 km. 


Figure 8 shows the results of driving dy¬ 
namics analyses with the preceding pro¬ 
totype (Schaeffler Hybrid) because the 
front axle was included in the tests along 
with the assessment of the rear axle. The 
driven maneuvers are plotted on the 
x-axis, the assessment determined for the 
vehicle is plotted on the y-axis. Zero 
stands for “unsalable”, the top mark ten 
for the perfect vehicle. The original vehicle 
is in the range 6.5 to 9. 

The criteria used refer, in particular, to 
vertical and lateral dynamics as well as the 
steering reactions. All assessments are 
within the range of results for the volume- 
produced vehicle. In this context, it must 
be emphasized that this driving behavior is 
only achievable with a spring damping sys¬ 
tem that is adjusted for higher masses. The 
modifications were carried out both in the 
Schaeffler Hybrid and in its successor 



■ Schaeffler HybridRear axle withwheel hub drives 

♦ Schaeffler HybridRear axle withwheel hub drivesFront axle withadditional mass (2*39 kg) 


Figure 8 Results for a comparison of driving dynamics 




























Wheel Hub Drives 


30 


449 


vehicle, a Fiesta, using volume-produced 
components. The results are significant 
under the aspect that the drive system 
operates on a twist beam - an axle design 
which was not originally designed for this 
drive. 

Significant increases in performance 
were noticeable during some maneuvers 
which use the potential of torque vectoring. 
For example, the speed was increased by 
10 km/h during a standardized double lane- 
change maneuver test with the cones 
spaced at 18 meters. 

After carrying out initial operation of the 
drives, a slip control system was applied as 
a basis for adjusting the torque vectoring 
and ESP functionalities. The high torque 
output of the drive system is actively used 
by a suitable control system for stabilizing 
the driving behavior. 

Winter testing was also carried out in 
North Sweden during February and March 



Figure 9 Adjustment of chassis, driving 

dynamics control and torque vectoring 



Figure 10 Wheel hub drive during winter testing 
















450 


2013 (Figure 10). The tests showed that 
the function of the drive is even ensured in 
wet and adhesive snow and at tempera¬ 
tures down to -33 °C. The vehicle also 
benefits from the selected concept, which 
does not require a hydraulic system and 
transmission. 


Future developments 


Further development of electric/ 
electronic components 

Development of the wheel hub drive is cur¬ 
rently underway at Schaeffler. A modified 
electric motor from the industrial sector is 
used in the current prototype, which is pri¬ 
marily optimized to produce a high torque 
output. Schaeffler is developing an electric 
motor specially matched to the require¬ 
ments of the wheel hub drive for the next 
development stage. 

A continuous torque of 500 Nm per 
wheel is required for a vehicle with the total 
weight of the presented prototype (approx¬ 
imately 1.5 tons) in order to transmit suffi¬ 
cient drive force in all driving situations. A 
further increase in torque density is there¬ 
fore the objective of Schaeffler’s develop¬ 
ment. The next generation of the wheel 
hub drive will be designed to fit inside an 
18-inch wheel, which is a conventional size 
for the vehicle class under consideration. 

The efficiency of the motor must still be 
increased at the operating points relevant 
for the driving cycle. The acoustic proper¬ 
ties for vehicle applications are also in need 
of improvement. Work will be carried out on 
these specific points for the next evolution¬ 
ary stage. 

The situation with regard to power 
electronics is similar. The electronics de¬ 


veloped for the prototype in collaboration 
with the Fraunhofer Institute for Integrated 
Systems and Device Technology USB have 
operated so far without failure. However, 
these electronics would not fulfill a typical 
specification used in automotive manufac¬ 
turing. 

Schaeffler is following a modular strategy 
for the further development of electric com¬ 
ponents for hybrid and electric drives, so 
that other drive variants such as the hybrid 
module or electric axle can be delivered 
with designs that are as similar as possible 
in order to rapidly achieve a significant unit 
cost degression with future volume produc¬ 
tion orders. 


The MEHREN research project 

In the MEHREN research project (Multi¬ 
motor electric vehicle with highly efficient 
use of space and energy, and uncompro¬ 
mising driving safety), Schaeffler is already 
working on the next generation of wheel 
hub drives in conjunction with Ford and 
Continental as well as the RWTH Aachen 
University and the University of Applied 
Sciences in Regensburg [7]. The focus of 
this project is on the implementation of a 
new software architecture matched to the 
requirements of wheel hub drives. This 
should, in particular, allow optimized coop¬ 
eration between the electric motor and 
service brake. 

The importance of functional safety is 
also taken into account in a special sub- 
project. 

The MEHREN project should also show 
for the first time what potential exists for 
new vehicle architectures if the wheel hub 
drive is used as a standard drive from the 
start of development. A virtual prototype of 
a purpose-built vehicle will be developed by 
2015. 



Wheel Hub Drives 


30 


451 


Summary Literature 


The development work carried out by 
Schaeffler on the wheel hub drive since 
2007 has proved that this drive can be 
successfully implemented in electric city 
vehicles. The torque density is in its sec¬ 
ond stage of development and almost at 
the required level. Schaeffler has dis¬ 
proved the counter-arguments frequently 
used in discussions about wheel hub 
drives, in particular, the negative influence 
of higher wheel mass on driving dynam¬ 
ics. Future development work will focus 
on further improving the electrical and 
electronic components as well as optimiz¬ 
ing the control quality and functional safe¬ 
ty. Ultimately, it will be important to actu¬ 
ally design and test new vehicle concepts 
made possible by the newly available 
space. 


[1] Gombert, B.: From the intelligent wheel bearing to 
the “robot wheel”. 9 th Schaeffler Symposium, 2010 

[2] Heim, R. et. al.: Potenzial von Radnabenant- 
rieben fur StraBenfahrzeuge. ATZ 114, 2012, 

No. 10, p. 752ff. 

[3] Reitz, D.: One Idea, Many Applications: Further 
development of the Schaeffler hybrid module. 
10 th Schaeffler Symposium, 2014 

[4] Kinigadner, A.; Kirchner, E.: Electric Driving without 
Range Anxiety: Schaeffler’s range-extender trans¬ 
mission. 10 th Schaeffler Symposium, 2014 

[5] Smetana, T.: Who’s afraid of 48 V? Not the 
Mini Hybrid with Electric Axle! 10 th Schaeffler 
Symposium, 2014 

[6] http://www.dlr.de/dlr/presse/desktopdefault. 
aspx/tabid-10310/473_read-6724/year- 
all/#gallery/9297, 10-24-2013 

[7] Wielgos, S. et al.: Reseach Project MEHREN: 
Potentials of Highly Integrated Wheel Hub Units for 
New Vehicle Concepts. 22 nd Aachen Colloquium 
Automobile and Engine Technology, 2013 


Open Access. This chapter is distributed under the terms of the Creative Commons 
Attribution Noncommercial License, which permits any noncommercial use, distribution, 
and reproduction in any medium, provided the original author(s) and source are credited. 




452 


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454 


Introduction 


The discussion concerning electrification 
of vehicle powertrains has caused an 
abrupt rise in the number of potential drive 
concepts. In the past, the question re¬ 
volved around whether a diesel or petrol 
engine, an automatic or manual transmis¬ 
sion was the right choice; today’s offer¬ 
ings include a huge variety of new archi¬ 
tectures, with an electric motor added to 
the combustion engine or used as a single 
drive. 

Most of the concepts are supported 
by the tangible benefits of the respective 
model. Certain arrangements and combi¬ 
nations seem to be beneficial depending 
on the weight given to advantages and 
disadvantages. This places significant ad¬ 
ditional burden on automotive manufac¬ 
turers and developers as a mainstream 
has yet to emerge. 

This article cannot and will not clarify 
this issue. Rather, its aim is to solidify in¬ 
teresting individual aspects of the vari¬ 
ous powertrain concepts on offer and to 
consider how a property of this kind can 
be transferred to a completely different 
drive. Viewed in this light, this approach 
has much in common with genetic engi¬ 
neering, which is based on removing in¬ 
dividual genes from a living being and 
then inserting them into another. The 
procedure is based on the understanding 
that all species ultimately have genomes 
with similar structures. Similarities can 
also be seen in technical products, even 
if they are based on different technolo¬ 
gies. When we get down to the basics, 
we realise that all products are based on 
a few physical principles. 

The laws of conservation for energy, 
momentum and charge lead to a far- 
reaching analogy between mechanics 
and electrics. This analogy is presented 


in the first section and then the analogy 
principle is expanded. The aim is to 
transfer certain objectives, thought pro¬ 
cesses or procedures to other technolo¬ 
gies. What new findings, perspectives 
and ideas may we discover? These are 
presented in the following using exam¬ 
ples. 

Mechanics and electrics 
- are they really two 
different worlds? 


Literature on physics or the engineering 
sciences contains a great many analogy 
analyses between mechanics, electrics, 
acoustics and hydraulics. In this article, 
we are restricting ourselves to electric 
and mechanical correlations and are us¬ 
ing the physical laws of conservation as 
the basis. We will then draw up parallels 
between the disciplines. In addition to 
the key laws of conservation, we will be 
using the first law of thermodynamics 
along with the law of conservation of mo¬ 
mentum in mechanics and charge in 
electrics. 

Based on these two physical vari¬ 
ables, the total of which always remain 
constant (an experiment you may also 
want to conduct), it is feasible to view 
momentum and charge as analogous to 
one another. It follows directly that their 
time derivatives also correspond: Force 
(F=*jj) and current (1 = ^-). By gradually 
expanding this analysis, it is also possi¬ 
ble to establish analogies for other 
mechanical and electrical variables [1] 
(Table 1). 

From this, it follows that accelerating a 
mass corresponds to charging a capacitor 
(Figure 1). 


Schaeffler Technologies GmbH & Co. KG, Solving the Powertrain Puzzle , 
DOI 10.1007/978-3-658-06430-3 31, © The Author(s) 2014 



What Powertrains Could Learn from Each Other 


31 


455 


Mechanics 

PnAKnn 

Electrics 

energy 

Power P 

1 mm i lo 1 

energy 

Power P 

impuis i 

Force F= — 

vsiiarge u 

niirrpnt 1 — 

dt 

Speed v 

dt 

Electrical Potential U 


Mass C 

Spring^^^^^^^^^^^^^^^^SSolenoid 

Reciprocal spring constant jL Inductivity L 
Damper Resistor 

Mechanical resistor Electrical resistor 

Viscosity Electrical conductivity 


Table 1 One of the possible analogies 

between mechanical and electrical 
variables. This example uses an 
accurate circuit analogy based on 
the laws of conservation of 
momentum and charge 


If we look at electrical oscillating circuits, 
the same analogous mechanical trans¬ 
ducers can be obtained by replacing the 
capacitors with mass, coils with springs 
and electrical with mechanical resistance. 
If this approach is taken, these analogous 
systems can then be described using 
analogous equations. 

Unfortunately, mechanical and electri¬ 
cal engineering were developed at differ¬ 
ent times and by different scientists, 
which is why individual variables have dif¬ 
ferent names. This makes it difficult to 
spot analogies in the equations at first 
glance. 

Of course, the exact analogy also has 
its limits, namely when certain physical 
conditions differ from each other. On ac¬ 
count of this naming convention, correla¬ 
tions for the acoustic and optical Doppler 
effect are not exactly the same; this is be- 


Mechanics 



Electrics 



charge 

capacitor 



Figure 1 Analogy of mechanical and electrical oscillating circuit 






456 



Speaker Antenna 


Figure 2 A comparison can be drawn 

between a speaker and an antenna. 
Both emit output in the form of 
waves 


cause acoustic waves are bound to the 
media, while electromagnetic waves are 
not. The laws of relativity apply to the latter 
with a speed of light that is the same for all 
observers, regardless of their speed. This 
restriction does not apply to a “mechani¬ 
cal” sound wave observer. 

Figures 2 and 3 show further examples 
of analogies. Speakers and antennas cor- 


Transmission and Transformer 


Mechanics 


Electrics 



Pulley 



Transmission 



Transformer 


Figure 3 A direct comparison can be drawn 
between the transformer and a 
pulley or a gearbox when rotation is 
involved. This group of transformers 
also includes converters, inverters 
and power converters. 


respond directly. Transformers find an anal¬ 
ogy in a pulley or gearbox. 

A transformer converts electrical en¬ 
ergy within the energy form so that the 
product comprising voltage and electrical 
current remains constant. For a pulley, 
force and travel are analogous, whilst for a 
transmission the variables are torque and 
rotation angle. 

On the electrical side, the group of 
transformers also includes converters, in¬ 
verters and power converters in general 
terms. Strictly speaking, these devices 
perform the same task as a transformer. 
They convert voltage and current into a 
different voltage, while taking the law of 
conservation of energy into account. 

This last example in particular shows 
that an analogy can also be understood in 
a wider sense when the components can¬ 
not be accurately described using the 
same physical, basic equations. The cru¬ 
cial aspect for this kind of extended anal¬ 
ogy is to use the same or at least a similar 
physical fundamental idea. The following 
example does just that. 


Electric gearbox or 
mechanical transformer? 


An initial comparison of combustion en¬ 
gines and electric drives shows that, in 
addition to a clutch, the combustion en¬ 
gine needs a shiftable gearbox with a lot 
of gears as a start-up device, while the 
electric motor can cope without any of 
these elements. 

The combustion engine therefore re¬ 
quires multi-gear gearboxes with the larg¬ 
est possible spread angle, as