kolloquium englisch - schaeffler germany...title kolloquium englisch.qxd author braunwfg created...

2277LLuuKK SSYYMMPPOOSSIIUUMM 22000066

Clutch and release system –Enjoyable clutch actuation!

Matthias ZinkMarkus HausnerRoland WelterRené Shead

At the 1998 symposium, the phenomenon ofincreasing engine torque, primarily with dieselengines, as a result of new technologies (high-pressure injection, supercharging) was dis-cussed in connection with the modulation ability of increased systems [1]. The associatedproblem of clutch pedal load or pedaltravel appeared at that time to have beensolved for the long term by the self-adjustingclutch (SAC). A look at the maximum torquecurve of a 2.0 liter diesel engine, however,shows that engine torques have risen by about40 % since 1998. This is due to the continueddevelopment of charging and injection technol-ogy (see Figure 1).

As a result of this truly “high-performance”work from of our colleagues in engine devel-opment, even new and intelligent tech-nologies such as reduced release-load self-adjusting clutches have been pushed to their limits within a relatively short period oftime.

The SAC, in combination with an over-centerspring on the pedal, sets the standard withregard to good operating load levels for clutch

systems currently used in production. Thepedal load that can presently be achieved isshown in Figure 2 as a function of enginetorque. This analysis is based on a high num-ber of measured vehicles. The range of varia-tion is due, among other things, to differentoperating travels, together with correspond-ingly varying ratios in the release system. Thetarget range for the maximum pedal load is90 N to 110 N. As a result, the SAC, in combina-tion with an over-center spring, can coverapplications up to approximately 300 Nm with-out compromise.

In order to provide the driver with a comfortablepedal load, even with high engine torques, it isno longer possible to restrict ourselves to inno-vations within the clutch alone. New approachesmust be found by considering the entireclutch/actuation system.

Modulating andoperating work –state of the artIn terms of ergonomics and comfort, anyamount of work required from the driver iscontrary to driver desire. Unlike automatic

2288 LLuuKK SSYYMMPPOOSSIIUUMM 22000066

22 CClluuttcchh aanndd rreelleeaassee ssyysstteemm

Introduction

Figure 1 Development of maximum engine torque over time, 2.0 liter diesel

transmissions, this work cannot be arbitrarilyavoided or reduced in a manual transmission,since the driver must be able to modulate the flow of torque between the engine andtransmission, and this modulation is supported,to a certain degree, by the work.

The work on the pedal required from the driver todisengage the clutch is currently, dependingupon application, between six and twelve joules(mean value without friction). For example, anapplication with 300 Nm requires eight joules(Figure 3).

In the first approximation, there is empirically aproportional relationship between the operat-ing work and the maximum pedal load. In orderto identify the optimization potential of theoperating work, its composition must first beanalyzed. The following consideration assumesan infinitely stiff clutch cover and a frictionlesssystem.

The work required to compress the cushionspring corresponds to the area below thecushion spring characteristic curve and willbe considered as the 100 % reference value(Figure 4). At an engine torque of 300 Nm,


CClluuttcchh aanndd rreelleeaassee ssyysstteemm 22

Figure 2 Achievable maximum pedal load in today's systems as a function of engine torque

Figure 3 Operating work on the clutch pedal, based on the300 Nm example application

this work corre-sponds to about1,0 J. This work isrequired so thatthe driver can modu-late the torquetransmitted by theclutch. Therefore the clutch disc mustnot be infinitelystiff, otherwise onlya digital torquetransfer would bepossible.

So that the clutchcannot be engaged,but rather disen-gaged by the driver,the cushion springwork, as is known, isdone by an energyaccumulator in theform of a space-saving diaphragmspring. The workrequired from thedriver to release theclutch disc is, there-fore, the resultantof the diaphragm and cushion springs.Work must also bedone to provideclearance to theclutch disc.

In a conventionalclutch, the diaphragmspring curve ismade flat to avoidtoo great an increasein work over itsservice life. Thatmeans an operatingwork on the pressureplate in the order of800 % for the newcondition, rising to about 1000 % as wearincreases.

Without the operating point displacementdue to facing wear, the diaphragm spring characteristic can be approximated to the

curve of the cushion spring characteristic,which reduces the disengagement workrequired at the pressure plate. This principlewas put into practice with the SAC, introducedby LuK in 1994. As a result of the load-



Figure 5 Work required to disengage the normally closed clutch (SAC)

Figure 6 Load and travel on the diaphragm spring fingers

Figure 4 Torque transfer and modulating work in a normally open clutch

controlled adjusting mechanism, the operat-ing point remains approximately constant,which can significantly reduce the workrequired (as compared to a conventionalclutch in the new condition) to about 640 %(Figure 5) with no significant increase overthe service life. This notably increases theadvantage of an SAC over a conventionalclutch.

Irrespective of whether a conventional clutchor an SAC is used, levers are required to operate the diaphragm spring. In today'sdiaphragm spring clutches, these are realizedin the form of diaphragm spring fingers(Figure 6). This design is favorable with regard to cost and installation space, but unfa-vorable with regard to stiffness and the asso-ciated travel losses during disengagement.This raises the work required to disengage anSAC to approximately 720 %. An applicationwith 300 Nm thus requires approximately7,2 joules to disengage the clutch.

There are additional stiffness losses in therelease system. The work required from thedriver to disengage an SAC finally increases toapproximately 800 %, with reference to theenergy stored by the cushion spring andrequired to ensure the torque transfer (Figure 7).

As already shown in Figure 3, this corre-sponds to 8,0 joules for an application with300 Nm.



Figure 7 Loads and travels on the clutch pedal

Figure 8 Composition of the operating work in an SAC clutch system

The composition ofthe work required todisengage the SAC asdescribed above is atypical example of thestate of the art. Thesesystems have certain-ly reached a highstandard as a result ofthe developments inthe past severalyears. Nevertheless, itcan be presumed thatthere still remains sig-nificant potential tobe exploited, consid-ering the fact thatstiffness losses stillaccount for some 20 % of the total oper-ating work (Figure 8).

These losses can bereduced by a thor-ough optimization ofthe individual compo-nents on the onehand and by a consid-eration of the systemas a whole on theother.

Reducing lossesthrough componentoptimizationAs shown in Figure 8, approximately half of thelosses can be traced to the stiffness of thediaphragm spring fingers, through which 10 % ofthe work required on the clutch pedal is caused. Thenext obvious step is to study this component on itsown with respect to possible optimization potential.

The introduction of the diaphragm spring clutchin 1962 appeared to bring nothing but benefits.Springs could be designed with low spacerequirements with nearly constant load and highclamp loads.

In the diaphragm spring clutch, the diaphragmspring is used both to generate the clamp loadand for operation. The thickness of thediaphragm spring also essentially defines thestiffness of the spring fingers and thus the stiff-ness of the operation.

That functionality was clearly separated in thecoil spring clutch (Figure 9).

The conflicting aims of clamp load and operatingstiffness can also be resolved with thediaphragm spring clutch. LuK has studied sever-al possibilities in this regard.

Machined load ring By “weakening” the load ring, it is possible to cre-ate diaphragm springs with high finger stiffness.This makes it possible to increase the thickness of



Figure 9 Coil spring clutch

the diaphragm spring by 50 % whilst maintaininga comparable clamp load characteristic (Figure 10).

Reinforced diaphragmspringReinforcing elements can be added to increasethe bending stiffness of the diaphragm spring.Consequent application of this approach yieldsan 80 % increase in stiffness (Figure 11).

Diaphragm spring designHere, the thickness of the diaphragm spring isoptimized with respect to operating stiffness(Figure 12). FE optimization tools are used todefine the geometry of the load ring to generatethe desired clamp load curve. With this method,

the thickness of a diaphragm spring can easilybe doubled whilst maintaining a comparableclamp load characteristic.

Reducing lossesthrough system optimizationAn extensive examination of the entire clutchand operation system can reveal weaknesses,which can then be eliminated through animproved design in future developments.

In addition to the traditional tuning cycleapproach, LuK is increasingly using simulation-supported optimization methods. One methodused is statistical test plan calculation. The procedure used here is described in detail in the article “Simulation Using the DMFW Exam-ple” [1].

Using this method, optimized results can beachieved through many small changes, and gen-eral trends can be deduced. One example of thisis the

“loss reduction through stiffness reduction”

of a hydraulic release system. This is an appar-ent contradiction that requires closer considera-tion before it can be understood.

A common representation of the stiffness ofthe release system components is the connec-tion between volume expansion and pressure.Figure 13 is a schematic representation of the



Figure 10 Diaphragm spring with machined load rings

Figure 12 Geometrically optimized diaphragm spring

Figure 11 Diaphragm spring with additional elements

test setup for deter-mining this charac-teristic and the typi-cal method ofexpression for thisresult, using theexample of a concen-tric slave cylinder(CSC). This procedureis certainly correctfor describing theindividual compo-nents, but is notentirely suitable forthe complete sys-tem. This is becauseonly the relationbetween the travelloss on the pedaland the release loadof the clutch is important here.

The volume expansion of the master and slavecylinder is significantly influenced, among otherthings, by the dynamic seal.

In the depressurized state there is free distancebetween the seal and cylinder wall. This can becalculated from the area marked in red in Figure14 multiplied by the seal length. When thecylinder is subjected to pressure, this free vol-ume is first taken up by the seal. Thus the oildisplaced represents a loss. As the seal itselfcan be considered virtually incompressible, aproportional relationship between the volumeexpansion and the seal length can be pre-sumed.

It follows, therefore, that an increase in thepiston area will lead to an increase in the vol-

ume expansion. Thus the stiffness of the mas-ter and slave cylinders is first decreased in theusual representation and the volume expan-sion raised (Figure 15). It would be a mistaketo conclude that this is also accompanied byincreased loss for the system as a whole, as isshown below.

The initial parameter for the overall system isnot the pressure, but the release load of theclutch. When the surface area is increased, thesystem pressure drops as a result. The advan-tage remains unclear as long as we onlyconsider the volume expansion. If, on theother hand, we do not consider the volumeexpansion as a function of the system pres-sure, but the travel loss as a function of therelease load, we find a significant advantage(Figure 16).



Figure 14 Volume expansion of a concentric slave cylinder (CSC) as an example

Figure 13 Measurement of the volume expansion of release system components

This finding has already been verified usingprototypes. Based on a system that is usual fortoday (master cylinder area = 285 mm², slavecylinder area = 775 mm²) the cylinder areaswere increased by 30 % as an example (mastercylinder area = 380 mm², slave cylinder area =1025 mm²). As a result of this step, the travelloss on the clutch pedal was reduced by 30 %from 25 mm to 17 mm with a release load of2000 N (Figure 16). A remarkable improvementbased on an overall consideration of clutch andoperation.

This highlights a decade-old error – one madeeven by the experts.

The usual surface areas of the master andslave cylinders today arose largely from thestandardized size of the brake cylinder. Those,however, must now be called into questionand possibly redefined in light of this newfinding.

Measures for reducing operatingloadThe first and also obvious approach wasdescribed in the previous section – optimizing the existing system. From what is known today,this potential can be used to reduce the operat-ing loads by about 10 % to 15 %, so that enginetorques up to 350 Nm can be covered withacceptable pedal travels and loads. Over andabove, further measures are necessary in orderto be able to attain operating loads lower than110 N. There are a number of options for doingthis, which are described in the following sec-tions:

• Work redistribution

• Energy accumulators

• Multi-plate clutch

• External energy (active support)

• Clutch-by-wire

Work redistributionIn work redistribution, the total work remainsconstant and is merely distributed more favor-ably over the pedal travel. The principle is simple.The work should be increased where the pedalload is currently low, so that it can be reducedwhere the pedal load is currently high (Figure 17).

Reduced-load clutches have in principle astrong “drop-off” in the release load character-istic because of the approximation of thediaphragm spring to the cushion spring curve. A



Figure 15 Increased volume expansion with increased piston area

Figure 16 Relationship between pedal travel loss and release load

dual effect can thus be achieved by this redistri-bution: the realization of a harmonic load char-acteristic and the reduction of the maximumpedal load.

The technical solution is a variable ratio forclutch operation. At low pedal loads the ratiois reduced and at high pedal loads it is raised.A clutch system with clutch, hydraulics andpedal provides three possibilities for varia-tion. Installation space and tolerance sensitiv-ity argue against a design implementationinside the clutch. At LuK, therefore, the “vari-able hydraulic ratio” and “variable pedalratio” concepts are being pursued.

Variable hydraulic ratioThe hydraulic ratio is calculated from the slave to master cylinder area ratio. Thismeans that the variability can be achieved by changing one of the two areas as a func-tion of the piston stroke. Because of the toler-ance and wear situation of the clutch system,

implementation in the slave cylinder is notfeasible. This is because the slave cylinder piston position does correspond to any specif-ic clutch position. For this reason, LuK isdeveloping a master cylinder with a variablepiston surface (Figure 18). A design with amoving primary seal and variable cylinderdiameter is preferable, since the greatestpressure therefore occurs at the smallest sealgap. The risk of gap extrusion is therefore min-imized.

The advantages of the variable master cylinderinclude a relatively simple design with no addi-tional components and neutral installationspace requirements. Since the variability isachieved through the variable sealing gap, thedesign is limited. The spread (differencebetween the largest and smallest ratio) current-ly being tested is 14 %. Since the total workmust remain constant, there is potential for aload reduction of about 7 %. Combined with theoptimization measures described above, aswell as a modified over-center spring, this solu-tion can cover engine torques up to 400 Nmwith pedal loads less than 110 N.

Variable pedal ratioFor engine torques greater than 400 Nm, a variable pedal ratio (VPR) is a very promisingapproach. The principle is the same as that inthe previous section, but more work can beredistributed with a mechanical design with apossible spread of up to 60 %. The potential forload reduction is thus approximately 30 %,because of which, applications up to 500 Nmcan be covered.



Figure 17 Work redistribution in clutch operation

Figure 18 Master cylinder with variable cross section

Figure 19 shows a production drawing for thissolution. The system consists of two rollersmounted on the pedal and one piston rod per-manently connected to a coulisse track. Whenthe clutch pedal is operated, the rollers follow

the coulisse track, with each of the two rollersable to support a load that is perpendicular tothe track (Figure 20). The function lines fR1 andfR2 of these two loads and the function line fK

of the piston load intersect in the load center of the system. The parallel distancelH(sP) from function line fK to the pedal rota-tional axis determines the lever arm and thusthe ratio of the clutch pedal as a function ofpedal travel sP.

Through the design of the coulisse, any ratiocurve is possible as a function of pedal travel.There are, however, design limitations in theform of the piston cross forces, surface pres-sure and system stiffness.

Figure 21 shows a measurement of a function-al example. The original maximum pedalload of 200 N and drop-off of 100 N could bereduced by this system to 160 N and 40 Nrespectively.

A favorable side effect of the flatter pedal loadcurve is a likewise flatter clutch torque charac-teristic curve between the engagement pointand approximately 100 Nm. This is because agreater ratio acts in this range like one of acomparable non-variable system. The result isimproved clutch modulation capability in thelower torque range (traffic jam or maneuveringoperations).



Figure 19 Clutch pedal with variable ratio

Figure 20 Relationship between coulisse track and pedal ratio

The complexity of the variable pedal ratio sys-tem is higher than that of the variablehydraulic ratio system. Additional compo-nents are unavoidable and the installationspace requirements at the pedal interfaceincreased. In order to keep these require-ments to a minimum, the version developedby LuK is largely integrated in the mastercylinder. In the design shown in Figure 19,only a second fastening pin is required on thepedal side.

Energy accumulatorEnergy accumulators are charged at low pedalloads and release their energy in areas of highpedal loads. In today's systems, this principle isalready frequently used on the pedal in the formof an over-center spring. The advantage of over-center springs is that they are already chargedand pre-stressed before the system is actuated.This not only redistributes the operating workrequired from the driver, but also reduces it.Over-center springs always have a symmetricalcharacteristic curve with respect to the cross-over point, which means that limitations arealready placed on the pedal load correction.Two-stage characteristic curves, especially incombination with the SAC characteristics, are afurther development (Figure 22). It has enabledeven critical applications to be configured com-fortably.

These systems reach their limits with increasedengine torque. As mentioned at the start, a pedalload of less than 110 N can hardly be realizedwith an engine torque of 350 Nm and above.

LuK is therefore studying alternatives. Againhere, the clutch system offers three possi-ble installation positions: the clutch, thehydraulic system and the pedal assembly.Integration into the hydraulic system wouldonly be possible with an extensive reconfigu-ration. LuK is therefore pursuing the followingsolutions: “Clutch with servo spring” and“Pedal assembly with leaf spring, roller andcoulisse.”

Clutch with servo springThe integration of an independent servo springin the clutch provides a potential pedal loadreduction of up to 20 N. However, all the toler-ances of the diaphragm spring finger heightand clutch wear lead to significant variation inthe level of load and load characteristic.



Figure 22 Pedal load curves with one- and two-stage over-center spring

Figure 21 Pedal load measurement with and without vari-able pedal ratio

Pedal assembly with leaf spring,roller and coulisseIn this system, a leaf spring, one end of whichhas a fixed connection to the pedal box, servesas an energy accumulator (Figure 24). The freeend has a roller bearing that presses on acoulisse fixed to the pedal. The required torqueabout the pedal rotation point is therefore gen-erated.

This system offers a number of advantages com-pared to the conventional over-center spring:

• Any assistance characteristic is possible

• Preload independent of assistance load

• Push-pull-push possible

With the over-center springs used today, onlyone reversal of load direction is possible.Before the cross over point, the over-centerspring generates a positive load to return thepedal and after the cross over point a negativeload to reduce the pedal load. In the spring sys-tem presented here, the load direction can inprinciple be reversed any number of times(push-pull-push function, Figure 24). So in caseof insufficient return load with the pedal in thedown position, the load can be increasedaccordingly.

Measurements with functional samples pro-duced good results with a clear gain in perform-



Figure 23 Design and operating principle of the servo spring

Figure 24 Energy accumulator with leaf spring, roller and coulisse

ance compared to the over-center springs usedtoday. In the example shown in Figure 24, themaximum pedal load is reduced by 65 N whilstachieving the desired return load of 15 N. It isworth noting that the operation is virtually fric-tionless (low hysteresis increase).

This concept can be used to cover applicationsup to 500 Nm with a desired pedal load of 110 N.Design limits are set by the permissible stressesin the leaf spring and the surface pressurebetween the roller bearing and coulisse track.

Multi-plate clutchBy multiplying the number of friction surfaces,the multi-plate clutch is able to reduce the oper-ating load. In a twin-plate clutch, for instance,the operating load can be lowered by 40 % withthe same moment of transmission. Some of thebenefit is lost in ensuring the separation of thetwo clutch plates.

In addition to the load benefits, the multi-plateclutch also provides improved thermal capacityand the potential to reduce the diameter of theclutch. About 20 mm more axial installationspace is required.

LuK now has several twin-plate clutches in standard production for engine torques above500 Nm. These clutches all take advantage ofthe proven SAC technology.

External energy to reduce pedal loadsPassive systems are limited based on theirefficiency. LuK currently believes that a pedalload of 110 N can be achieved with enginetorques up to 500Nm using the current pas-senger car sector pedal travels (120 ... 160 mm)and clutch dimensions. For applications above500 Nm for which a twin-plate clutch cannotbe used, there is the option of actively sup-porting the clutch pedal in a manner similar topower-assisted steering. In order to make theexpense of an active system pay off, LuK cur-rently envisages the following applicationareas:

• Required pedal loads below 110 N for applications between 400 Nm and 500 Nm

• Applications above 500 Nm

• Shortening pedal travel/redefiningergonomics

• Applications outside the automotive sector (e.g., commercial vehicles, tractors)

• Retrofits

• Optional equipment



Figure 25 Design and characteristic curves of a twin-plate SAC

LuK has studied a number of possible versionsand configurations for active support with thefollowing development goals:

• Self-contained, easily adaptable unit

• Low interface requirements (add-on)

• Function independent of other components(e.g., combustion engine)

• Differentiation from clutch-by-wire

• Maintaining a direct connection betweenclutch and pedal

These requirements are met in the electro-hydraulic CSA (Clutch Servo Assistance) systemdeveloped by LuK. This is a pump unit driven byan electric motor positioned directly betweenthe slave and master cylinder (Figure 26).

This unit consists of an electric motor, electron-ics and hydraulic system. The electronics are formonitoring only and protect the system fromoverload (temperature, current). They are notrequired for the actual function of the unit.

The hydraulic system (Figure 27) consists of aninternal gear pump, a regulating valve and a

safety valve, which ensures clutch operationunder all conditions.

The system has five possible operating modes:

The system is in sleep mode when, for example,the ignition key is out. In this state, the entiresystem is inactive and without power. The sys-tem therefore has no energy requirements andno functionality.

When there is a possible intention to operate,the system is in stand-by mode. This state maybe defined by the presence of the ignition keyin the ignition. The electronics are now active,but there is still no power to the electricmotor.

If the clutch pedal is operated from stand-bymode, the pump is driven by the electricmotor. Up to a freely definable pressurethreshold pS in the release system, there is nosupport, since the regulating valve is pre-loaded by a spring. The control edge is stillcompletely open and the pump cannot buildup pressure.

When the pressure threshold pS is exceeded,pressure-proportional support is generated andthe release system is divided into pressureranges: high pressure (pslave) and low pressure



Figure 26 Clutch servo assistance (CSA) – arrangement inthe overall system

Figure 27 Hydraulic system of the CSA

Sleep Electronics and electricmotor not powered

Stand-by Electronics powered, elec-tric motor not powered

Pump turning Regulating valve open

Pump turning Regulating valve atoperating point

Emergency Safety valve operation

(pmaster) (Figures 28and 29). The propor-tionality is expressedas the reduction fac-tor k and is deter-mined by the arearatio A2:A1 of the reg-ulating valve. If, forexample, the goalis to bisect the pres-sure level above the start-up thresh-old pS in the mastercylinder, then theresult is a regulatingvalve surface ratio ofA2:A1 = 1:2.

In summary, the following equation character-izes the layout of the regulating valve:

Through the pressure threshold pS and reduction factor k parameters, the pedal load

curve can be accurately corrected to two tar-get values. Figure 29 shows an example ofthis. The target was to reduce the original250 N pedal load to 120 N with a subsequentdrop in pedal load (drop-off) of 20 N. The target could be achieved with a switch-onthreshold of 30 N (corresponds to a pressurethreshold pS of 5 bar) and a reduction factorof 0,4.



Figure 29 Measurement of the pedal load with and without CSA

Figure 28 Pump turning, regulating valve at operating point (support)

A rational alternative to the CSA would be, forexample, a pump driven directly via the acces-sories. The benefit of this would be the elimina-tion of the electric motor and electronics, mak-ing it possible to, amongst other things,significantly lower the system costs. The disad-vantages include higher interface demands onother systems and the lack of functionality whenthe engine is not running.

Clutch-by-wireIn modern engines, the gas pedal is no longermechanically linked to the throttle valve or theinjection system. Instead there is just a sensoron the gas pedal that forwards the driver's com-mands to an actuator via a control device (elec-tric gas pedal). This creates the possibility toadapt the engine characteristic as required tothe driving situation. This is now a basic prereq-uisite for making optimal use of the potential oftoday's engines and ensuring a comfortabledrive.

A clutch-by-wire system promises comparablepotential for the clutch (CBW, Figure 30). Thisis the entry-level version of the family ofautomated shift transmissions and repre-sents the most complex version for clutchoperation for the manual shift transmission.The CBW system eliminates the release systemas a fixed connection between the clutch and clutch pedal and uses a pedal travel sen-sor, a controller and an actuator to operate theclutch (Figure 30). The pedal operating load

can be chosen freely using a pedal spring orcomparable mechanism. This makes an opti-mal layout possible regardless of the enginetorque.

In order to justify the high expenditure onhardware and software, a CBW system offersnumerous possibilities for improving drivingcomfort. These include, in particular, slip con-trol for vibration isolation between the engineand transmission, which is dealt with in thearticle “Software for Automated Transmis-sions” [1].

In a similar manor to how the electric gas pedaladapts the engine characteristics, CBW can beused to adapt the torque characteristic of theclutch to the specific driving situation. Animportant example in this context is a launchprocess, which is influenced by the combina-tion of engine torque and clutch torque. Figure31 shows how the full-load curve of a 2.0 literdiesel engine has changed over the years. Inthe idle speed range, the engines operate boththen and now in induction mode, with the max-imum engine torque practically unchanged atlow speeds. By contrast, the maximum enginetorques, and therefore also the clutch torques,have tripled over the years (see also Figure 1).As a result, a weak engine is combined with astrong clutch at low engine speeds (red area inFigure 31).

This circumstance can be rectified with a CBW.Using information such as the engine and



Figure 30 Clutch-by-wire (CBW)

Figure 31 Development of the full load characteristic of a2.0 liter diesel engine

transmission speeds, the clutch torque as afunction of the pedal travel can be varied tobest suit each and every driving condition.

LuK has already installed CBW prototypes inseveral different vehicles and tested them. Thelaunch performance has improved remarkablycompared to conventional clutch systems as aresult of the function just described.

SummaryDespite the large number of technical solutionspresented here, the application areas of the different concepts can be clearly defined.

The controlling parameters are the maximumengine torque and the desired maximum pedalload as shown in Figure 32. This should beviewed as a recommendation, which can varydepending on the philosophy of the vehiclemanufacturer.

Smaller steps can be achieved through the sepa-rate optimization of individual components. The

correct technical solutions for the torqueincreases of the coming years can only be foundwith the correct dimensioning of the completesystem.

The complexity of the solutions for high torqueapplications may appear high at first. Howev-er, if the intension is to offer manual transmis-sions in this class, there is no alternative but tostart with the total system development assoon as possible.

In order to support this, LuK has put the pre-requisites in place with a complete switchfrom a component supplier to a system sup-plier. This requires close cooperation betweenthe clutch and chassis departments at theautomobile manufacturers. Together with the automobile manufacturers, LuK will con-tinue to present innovative solutions, whichoffer the end consumer optimal operatingcomfort despite significantly increased enginetorques.



Figure 32 Options and potential for reducing operating load

Literature[1] Zink, M.; Shead, R.: Clutch and Actuation as

a System, 6th LuK Symposium 1998.

[2] Fidlin, A.; Seebacher, R.: Simulation Technol-ogy Using the DMFW Example, 8th LuK Sym-posium 2006.

[3] Küpper, K.; Serebrennikov, B.; Göppert, G.:Software for Automated Transmissions, 8thLuK Symposium 2006.



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