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Foreword

Innovations are shaping ourfuture. Experts predict that therewill be more changes in the fieldsof transmission, electronics andsafety of vehicles over the next15 years than there have beenthroughout the past 50 years. Thisdrive for innovation is continuallyproviding manufacturers and sup-pliers with new challenges and isset to significantly alter our worldof mobility.

LuK is embracing these challen-ges. With a wealth of vision andengineering performance, ourengineers are once again provingtheir innovative power.

This volume comprises papersfrom the 7th LuK Symposium andillustrates our view of technicaldevelopments.

We look forward to some intere-sting discussions with you.

Bühl, in April 2002

Helmut Beier

Presidentof the LuK Group

Content

LuK SYMPOSIUM 2002

1 DMFW – Nothing New? . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5

2 Torque Converter Evolution at LuK . . . . . . . . . . . . . . . . . . . . . . . 15

3 Clutch Release Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27

4 Internal Crankshaft Damper (ICD). . . . . . . . . . . . . . . . . . . . . . . . . 41

5 Latest Results in the CVT Development. . . . . . . . . . . . . . . . . . . . 51

6 Efficiency-Optimised CVT Clamping System . . . . . . . . . . . . . . . 61

7 500 Nm CVT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 75

8 The Crank-CVT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 89

9 Demand Based Controllable Pumps. . . . . . . . . . . . . . . . . . . . . . . 99

10 Temperature-controlled Lubricating Oil Pumps Save Fuel . . . 113

11 CO2 Compressors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 123

12 Components and Assemblies for Transmission Shift Systems135

13 The XSG Family . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 145

14 New Opportunities for the Clutch?. . . . . . . . . . . . . . . . . . . . . . . 161

15 Electro-Mechanical Actuators. . . . . . . . . . . . . . . . . . . . . . . . . . . 173

16 Think Systems - Software by LuK. . . . . . . . . . . . . . . . . . . . . . . . 185

17 The Parallel Shift Gearbox PSG . . . . . . . . . . . . . . . . . . . . . . . . . 197

18 Small Starter Generator – Big Impact . . . . . . . . . . . . . . . . . . . . . 211

19 Code Generation for Manufacturing. . . . . . . . . . . . . . . . . . . . . . 225

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500 Nm CVTLuK Components in Power Split

Christian LauingerMartin VornehmAndreas Englisch

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IntroductionFollowing the successful market launch of theAudi multitronic® [1], [2], presently for appli-cations up to 310 Nm, a general demand cur-rently exists for continuously variable trans-missions for more powerful engines. The tar-get here is for around 500 Nm of enginetorque.

When developing a CVT concept (Continu-ously Variable Transmission) for applicationswith considerably higher torque, along withnumerous other aspects, a fuel consumption-optimised transmission structure and the var-iator load are of particular importance duringthe selection and evaluation process. The lim-iting factor for a further increase in the maxi-mum transmittable torque is the transmittingelement.

According to [3], on a non-power-split conven-tional CVT with the further-developed LuKchain, the torque, at a total ratio spread of 6and axis distance of 171 mm, is currently lim-ited to approximately 400 Nm. Smallerspreads would have adverse effects on fuelconsumption.

Figure 1 gives an overview of the total ratiospread and the maximum engine torque of aconventional CVT with various transmissionelements. Currently, the most demanding ap-plication is the LuK variator in the Audi multi-tronic® with a maximum engine torque of310 Nm and a total ratio spread of 6. Potentialfor a torque increase to nearly 400 Nmthrough the further-developed LuK chain [3] ofthe same size is also indicated. Further in-creases in engine torque up to 500 Nm withspreads between 6 and 7 require structuralenhancements such as the arrangement ofthe variator in a power split drive train.

Within the scope of this theoretical study, twopossible concepts for a power split, dual-range CVT were investigated to ascertainwhether the step to the required 500 Nm ispossible within the specified range of total ra-tio spread.

A dual-range CVT has, compared to a con-ventional CVT, the advantage that both thevariator torque and the variator ratio spreadrequired for a given total ratio spread can bereduced. Both result in a reduction of variatorload, which allows for an increase of the trans-mission torque capacity and consequently theengine torque and total ratio spread.

In the following chapter, the various designpossibilities for a dual-range transmission willbe considered with regard to their suitabilityand subsequently evaluated.

Thereafter, the particular requirements re-garding the clamping system and the hydrau-lics will be discussed. Finally, a possible con-trol strategy will be illustrated. The focus herewill be on the development of a strategy for acomfortable mode change.

Fig. 1: Maximum Permissible Engine Torque and Spread of Standard CVT with Various Transmitting Elements

Mechanical

Dual-range CVTIn CVT's currently on the market, the varia-tor is arranged in a conventional drive train,a so-called single-range CVT. The torquecapacity has already been described in theintroduction. Therefore, an increase in en-gine torque up to 500 Nm with simultaneousincrease of the total ratio spread requiresstructural transmission enhancements suchas a dual-range CVT.

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In its simplest form, a dual-range CVT has asecond fixed-ratio shaft arranged in parallelwith the variator. The two branches arebrought together at the output shaft through asummation gear set (e.g. a simple planetarygear set). The drive train, together with thevariator, can be driven either directly or in apower split mode by disengaging or engaginga clutch in the fixed-ratio section. A dual-rangeCVT is therefore more elaborate than a single-range CVT from both a mechanical and con-trol point of view.

Dual-range CVT designs are also possiblewhere the variator is permanently located ina power split drive train; however, to reach thesame total ratio spread the designs of thesummation gear sets are more complex.

The literature contains references to otherknown multi-range concepts [4], which werenot included in this study.

The illustration in figure 2 shows the designand operation of a dual-range CVT. The sum-mation gear set is a simple planetary gear setwith negative set ratio i1 (negative gear set).On a negative gear set, when the carrier isheld, the sun and the ring gear turn in oppositedirections [5]. The clutch in the fixed-ratio splithas the designation KH. In order to limit the de-

grees of freedom of the planetary gear setwhen operating in unsplit mode, a furtherclutch KL, which connects the sun gear andthe carrier, is necessary. The planetary gearset then rotates locked-up and therefore hasa fixed ratio of 1. For simplicity the schematicdrawing does not show the drive-off element(clutch or hydrodynamic torque converter)and the axle differential. An additionalbrake B, which facilitates braking of the planetcarrier, is required for the reverse gear. Theoutput shaft is connected to the ring gear.

In unsplit mode, KL is closed and KH is open.In power split mode, the relationship is re-versed. The point where switch-over of theclutches takes place and which separates thetwo ranges is designated as the mode changepoint.

The concept from figure 2 makes it possibleto achieve power split either during drive-off(higher driving range is then unsplit) or in thehigher driving range (drive-off is then unsplit).To do so, the fixed ratios and the planetarygear set ratios must be changed and the ac-tuation of the clutches KH and KL inter-changed. Additional clutches are not required.In the following, the transmission from figure 2

Fig. 2: Schematic Layout of a Dual-Range CVT Showing Power Flows Unsplit (Left) and Power Split (Right)

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is shown unsplit in the lower driving range(from now on called LOW range) and powersplit in the higher driving range (HIGH range).

Figure 2 additionally shows a schematic viewof the power flow through the variator in unsplitand power split modes. In the unsplit mode,figure 2a, the entire engine power flowsthrough the variator. When switching over topower split, a reactive-power flow forms in thesplit transmission section as shown infigure 2b. Consequently, the variator torquechanges its sign, i.e. the variator is driven fromthe secondary pulley set. The extent of thepower flow through the variator is dependenton the total ratio and can be analysed as fol-

lows. When the drive-off clutch / torque con-verter clutch is engaged, the engine and var-iator input speeds are identical. The power iscalculated from the product of speed andtorque with the result that the ratio betweenvariator torque and engine torque is identicalto the ratio between variator power and enginepower. On a single-range CVT, this ratio isequal to 1. In power split applications, this isno longer the case.

Figure 3 shows the ratio of variator torque toengine torque depending on the total ratio. Inunsplit mode, i.e. between total transmissionunderdrive (UD) and mode change, the ratiois constant and equal to 1.

Fig. 3: Ratio of Variator Torque to Engine Torque and Variator Ratio (Top) and Drive-Time Distributionin MVEG (Bottom) in Relation to the Total Ratio

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With mode change into power split, the signof the variator torque changes due to the re-active power present, see figure 2b. As the to-tal ratio moves further towards total transmis-sion overdrive (OD), the ratio of variator torqueto engine torque continually reduces. On thearrangement shown here, the amount of var-iator torque in total transmission overdrive(OD) is only around 10% of the engine torque.Consequently, a reduction of the mechanicalload of the variator is achieved at these oper-ating points.

The reduction of power flow through the var-iator in power split mode results in an improve-ment of the transmission efficiency with a con-sequent positive effect on fuel consumption.Figure 3 shows the drive-time distribution inMVEG. In the out-of-town segment (EUDC),the variator operates predominantly in powersplit mode. The variator is then only subjectto low torque levels with consequently lowlosses. This results in an overall improvementin fuel consumption. The hydraulic pressurefor the clamping force and consequently forthe pump [6], [7] can be reduced in proportionto the applied variator torque.

If a dual-range CVT is used, the variatorspread can be reduced. Figure 3 also showsthe variator ratio over the total ratio. Betweendrive-off and mode change, the variator ratiochanges from variator UD (greatest variatorratio) to variator OD (smallest variator ratio).As the total ratio moves further from modechange towards total transmission OD, thevariator adjusts in the opposite direction backtowards variator UD. Because the ratio rangeof the variator is run through twice, the variatorspread can be reduced in comparison with asingle-range CVT of the same total ratiospread. In this illustration, the reductionamounts to 34%.

The reduced variator spread in the dual-rangeCVT results in lower chain loads. This is illus-trated in figure 4 where the normalised chainrunning radii on pulley set 2 are compared forconventional and dual-range CVT.

Fig. 4: Comparison of Standard Chain Running Radii on Single and Dual-Range CVT

By limiting the variator spread, the minimalchain running radius is increased by roughly20 - 25% with a corresponding reduction inchain pulling forces. Consequently, the mini-mum number of rocker pins engaged in thevariator is correspondingly increased and theindividual rocker pins have a lower stress leveldue to a decrease in bending. This results ina more constant force distribution in the linkplates [8] with less load on the chain.

This reduction in the variator torque in HIGHrange and the lower load on the chain due tothe reduced variator spread means that thevariator is subject to less load in a dual-rangeCVT. Using a load cycle with a maximumtorque of 500 Nm, the variator in a dual-rangeCVT has a life time around 10 times longerthan in a single-range CVT.

The results shown so far are for summationgear sets arranged as negative gear sets. Inprinciple, a dual-range CVT can also beachieved with a planetary gear set with a pos-itive set ratio [5] (positive gear set). On a pos-itive gear set, when the carrier is held, the sungear and the ring gear turn in the same direc-tion, e.g. reverse gear set from a standardtransmission. If a positive gear set is used, thevariator split, fixed-ratio split and output shaftmust be connected to the planetary gear set

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in a different combination. In this case, the sunwould be the output, which would be unfavour-able for design reasons. A further disadvan-tage of positive planetary gear sets is a low-ered efficiency due to the larger number of ro-tating gears [5] and the resultant higher fuelconsumption. Furthermore, due to the highernumber of planet gears the costs and com-plexity of positive gear sets are greater. Forthese reasons, only negative gear sets are con-sidered in this study.

Range conceptsGenerally, on a dual-range CVT, three differ-ent range concepts are possible:

� Unsplit in drive range LOW and power splitin HIGH (as in figure 2)

� Power split in LOW and unsplit in HIGH(generally geared-neutral-capable)

� Power split in both drive ranges

The last mentioned variant is significantlymore elaborate than the first two concepts dueto the greater number of clutches and/or plan-etary gear sets. It was also explained in theprevious section that a reduction in variatorload can be achieved in a dual-range CVT dueto a smaller variator ratio spread and due toreduced variator torque through the powersplit. In view of these properties only the firsttwo concepts listed will be considered.

The negative ratios of the planetary gear setsin the two concepts are different. The ratioswere selected with the aim of minimising thevariator load for each variant at the specifieddimensions.

Figure 5 shows the normalised chain running ra-dii on pulley set 2 depending on the total ratio.Figure 5a shows the curve for the concept withpower split in HIGH. These are the same valuesas in figure 4. Figure 5b shows the same curvefor the concept with power split in LOW.

As figure 5b shows, during drive-off with pow-er split in LOW the variator adjusts from vari-ator OD towards variator UD. The adjusting di-rection of the variator is therefore opposite tothat of the concept with power split in HIGH.The illustration shows that the severely dam-age-prone smallest running radius in the con-cept with power split in LOW can be increasedby up to 35%.

Figure 6 shows the ratio of variator torque to en-gine torque depending on the total ratio for bothconcepts. The concept with power split in HIGHis shown on the left (figure 6a), power split inLOW is shown on the right (figure 6b). In totaltransmission UD the variator torque with powersplit in LOW is several times greater than thatof the other concept. This results in considerablyhigher loading for the variator and chain than onthe concept with power split in HIGH.

Fig. 5: Comparison of Standard Chain Running Radii for the Concepts with Power Split in HIGH (a) andPower Split in LOW (b)

a) b)

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Fig. 6: Relationship of Variator Torque to Engine Torque Depending on Total Ratio for the Concepts withPower Split in HIGH (a) and Power Split in LOW (b)

The damage accumulation calculations forboth concepts show a higher chain load for theconcept with power split in LOW. This meansthat the benefits resulting from the reducedvariator spread are overcompensated for bythe high variator torque.

The variant with power split in LOW also hasdisadvantages with regard to fuel consump-tion. In the main driving ranges of MVEG, thevariator operates conventionally; i.e. the var-iator torque is identical to the engine torque.The variator loss is consequently higher andthe efficiency respectively lower. Further-more, the hydraulic pressure for the clampingforce and the pump cannot be lowered, whichis possible with the concept power split inHIGH. Overall this results in improved fuelconsumption with power split in HIGH.

The results show that out of the numerouspossible dual-range CVT variants, the oneswith power split in HIGH represent the mostfavourable concepts with regard to variatorload and fuel consumption.

Optimum variants for the summation gear setThe following section deals with the questionof what are the other options for the summa-tion gear set with power split in HIGH and

which is the most favourable. Previous con-siderations assumed a summation gear set ar-ranged in the form of a simple negative gear set.

With a coupled planetary gear set (consistingof two planetary gear sets), it is possible to re-duce the variator spread even more while si-multaneously limiting the variator torque tolevels comparable with those in figure 6a withpower split in HIGH and a simple negativegear set as a summation gear. Consequently,it is possible to halve the variator wear com-pared to the case of a simple negative gearset. Additionally, this arrangement allows forhigher efficiency levels.

Therefore, considering both fuel consumptionand variator load, a coupled planetary gear setis the most favourable variant.

Selection of the planetary gear ratiosThe following illustrates how to determine thegear ratios i1 and i2 of the two coupled plan-etary gear sets. These ratios have a direct ef-fect on the variator load and the fuel consump-tion. Through appropriate selection of both ra-tios, the jump in variator torque at the modechange point (figure 3) and the variator torquein total transmission OD are minimised. Thespecification of total transmission UD and OD

a) b)

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along with the permissible range for the re-verse gear ratio defines the main parametersfor the arrangement. The relationship be-tween total ratio and variator ratio is then fixed,if the ratio at the mode change point is also de-fined.

Figure 7 shows the ratio of variator torque andengine torque, TVariator/TEngine, at the modechange point in power split depending on setratios i1 and i2. Along the continuous black linethe ratio of TVariator/TEngine is constant.

Fig. 7: Ratio TVariator/TEngine at the Mode Change Point in Power Split depending on Set Ratios i1 and i2

Figure 8 shows progressively the variator ra-tio, the variator torque and the calculatedchain damage for a time portion taken from aWOT acceleration. In the illustrated time por-tion, the vehicle speed increases from around160 km/h to 220 km/h. The maximum chaindamage occurs directly at the mode changepoint. One cause for this is the small runningradius of the chain on pulley set 2 at the modechange point as shown in figure 6. On the oth-er hand, the chain speed is at its maximum atthe mode change point due to a constantlyhigh engine speed. The large chain centrifugalforce is therefore a further reason for the highchain damage in the vicinity of the modechange point. However, the amount of dam-age in power split is already minimised by se-lecting the set ratios so that the absolute val-

ues of the variator torque, |TVariator/TEngine|, isas small as possible. Subsequently, it will beshown that this ratio cannot be further re-duced, because other criteria (overall size, re-verse gear ratio) must also be fulfilled. Conse-quently, there is a minimum |TVariator/TEngine|.The range with the undesirably high|TVariator/TEngine| values is shown in blue infigure 7. To calculate the time period shown infigure 8, optimal gear set ratios i1 and i2 werechosen. Other set ratios with higher values of|TVariator/TEngine| are shown in figure 8 to runinto the blue zone and would result in higherchain damage.

As already mentioned above, there are othercriteria which limit the value ranges for i1 andi2 besides the total ratio spread and jump sizein the variator torque. The desired total ratiospread, the target value for the reverse gearratio, packaging and the resulting limitationsfor the overall size of the planetary gear setfurther limit the selection of the set ratios.

The shaded area in figure 9 shows the criteriaand the associated non-permissible ranges inthe i1-i2 level. The permissible i1-i2 values inthis application are taken from the overlap ofall non-shaded areas. This value range for i1and i2 is marked in green.

The maximum permissible engine torque fora dual-range CVT, depending on total ratiospread and axis distances, was calculated us-ing the optimising measures described in theprevious sections. The calculations are basedon the further-developed LuK chain [3].Figure 10 shows the results. Additionally, thevalues for a single-range CVT with a total ratiospread of 6 are shown. The values for both thecurrent standard chain and the further-devel-oped LuK chain are plotted. According tofigure 10, a CVT with a total ratio spread of 6for 500 Nm is possible, with an axis distancebetween 170 mm and 190 mm. To achieve atotal ratio spread of 7, the maximum permis-sible engine torque has to be reduced byroughly 5%.

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Fig. 8: Variator Ratio, Variator Torque and Chain Damage versus Time during a WOT Ac-celeration in the Vicinity of the Mode Change Point

Fig. 9: Exclusion Criteria for Determination of Set Ratios i1 and i2 of Coupled Planetary Gear Set

Fig. 10: Maximum Permissible Engine Torque of a Dual-Range CVT in Comparison with a Single-Range CVT for Various Axis Distances and Spreads

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HydraulicsIt was shown in figure 3 that the variator of thedual range CVT favoured in this case is oftenin overrun condition. This is significantly dif-ferent than a single-range CVT and has con-sequences for the clamping system. For ex-ample with a dual-stage hydro-mechanicaltorque sensor [6], [7], this would then requirea layout compromise for both the overrun anddrive conditions. Consequently, a dual-stagetorque sensor would no longer provide fuelconsumption benefits as on the multitronic®.For this reason, an electronically controlledclamping is envisioned for the power split CVT,which leaves all options for the different oper-ating conditions open [10].

The following control functions must be providedon the dual-range CVT by the hydraulic system.

� Clamping� Adjustment� Forward and reverse clutches� Mode change� Cooling of clutches

Further optimisation is possible due to the factthat the hydraulic system is designed for par-ticularly low back pressure. In power splitmode, as mentioned in the previous chapter,the variator is driven with only a fraction of theengine torque. Therefore, the clamping forcecan also be reduced. If the hydraulic backpressure is low, the pump would also be sub-ject to reduced pressure resulting in lowerpump drive torque.

Cascade hydraulics

Figure 11 shows the design of a cas-cade hydraulic system, which is ar-ranged for these particular require-ments.

The cascade operates a priority sys-tem: the clamping force, together withthe clutches, has the highest priority;the adjustment elements are operat-ed secondarily.

The number of functional elementshas risen moderately in comparisonto the multitronic®:

Loss reductionIn order to keep the flow demand dur-ing adjustments as low as possible,thereby minimising the pump drivetorque for the transmission concept il-lustrated here, a double-piston sys-tem as on the multitronic® is required.

Fig. 11: Schematic Diagram of Hydraulics: Clamping and Adjustment System

cascade-hydraulics

multitronic®

electro-valves

6 3+ torquesensor

valves 11 9

pumps vanesize 122%

gearsize 100%

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Because of the use of three proportionalvalves (clamping, adjustment 1 andadjustment 2) to control the clamping forceson two pulley sets, there is an additional de-gree of freedom. The control strategy can ex-ploit this in order to minimise the system pres-sure (i.e. minimise the maximum value of thethree required pressures). The advantage ofthe double-piston principle is thereby furtherincreased with regard to loss minimisation.

Control strategyThe same demands with regard to drivingcomfort are placed on a dual-range CVT as ona single-range CVT. This does not only applywithin a mode range, but also in particular dur-ing mode change. The control strategy istherefore an extremely important factor inachieving the comfort requirements.

Strategy aim

Because of the good fuel consumption, com-fort and driving performance characteristics ofthe multitronic® [1], [2], a similar strategy is as-sumed here: The driver expresses the desiredperformance through the pedal position andthe transmission ensures that this perform-ance is delivered at an optimum consumptionlevel of the engine (low engine speed). Accel-eration takes place evenly and comfortably.

This comfort demand also exists during pedalchanges or ratio changes: the output powermust change comfortably. It must not be no-ticeable that there are several ranges, or thata change between the ranges is occurring.Simply ‘pausing briefly at the mode changepoint, shifting, and then continuing to adjust’is not sufficient.

The behaviour during ratio changes in thiscase has proven to be particularly relevant tocomfort, as internal rotational masses in thetransmission act as flywheel accumulators,which absorb and release power.

Figure 12 shows an example of a fast UD ad-justment, which means: the transmissionshould rapidly raise the engine speed (at con-stant output speed). The ideal speed progres-sion of the engine then shows no comfort-re-ducing characteristics. The speed progres-sion of the internal pulley set 2, however,shows curve sections of differing steepnessand even a kink at the mode change point. Itis shown in figure 13 that this results in a lossof comfort. Immediately before the modechange in power split mode, the pulley set 2has to accelerate even more than the enginedue to the variator adjustment. After the modechange, in the unsplit range, the speed of pul-ley set 2 is nearly constant.

Because an increase in speed always re-quires torque and consequently power, thetractive force changes in the shown speedprogression of pulley set 2. This change in thetractive force is not included in figure 12, butis assumed in the simulation in figure 13.

Fig. 12: Schematic Drawing of Engine Speed and Pulley Set 2 Speed over Timeduring an Adjustment at Constant Vehicle Velocity

In view of this, the strategy aim is not only tocomfortably change modes at the right time,but an overall more comfortable progressionof the output performance. This is achievedthrough engine intervention and a dynamicgradient control for the engine speed.

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Simulation resultsFigure 13 shows the simulation results for ademanding situation: a UD adjustment afterkickdown with a mode change during a speedgradient. Illustrations are of the time-depend-encies of variator ratio, engine speed and ve-hicle acceleration.

Fig. 13: Simulation of Mode Change with an Opti-mised Strategy (Bold Lines)

For Comparison, the Curves for a Strate-gy without Engine Intervention and with-out Gradient Adjustment are Shown (Thin Lines).

In the left part of the time record, the acceler-ator pedal is actuated by 30% and the engineis held by the transmission at an optimum con-sumption level of 1100 min-1. The vehicle ac-celerates slightly and the variator is adjustedslowly in HIGH. At the time t � 25 s, a kick-down actuation is carried out, followed by an

adjustment from HIGH to LOW. The centerpart of the illustration shows how the variatoradjusts towards mode change (here at varia-tor ratio 0.5), where the mode change to LOWtakes place with simultaneous acceleration ofvehicle and engine. A further acceleration of thevehicle can be seen in the right part of the illus-tration. The engine is running with considerationto speed tracking – similar to that in multitronic®

[1], [2] – at near optimum performance.

A comfort evaluation according to the vibra-tion amplitudes in the vehicle accelerationwas performed. With an optimised strategy,the peak to peak distance in the area of themode change is approx. 0.4 m/s2. In compar-ison, a gear change of 1.25 in an automatictransmission equates to approx. 0.5 m/s2.However, it has to be emphasised that thespecified accelerations differ in slope. Anotherpoint is that the comfort demand varies ac-cording to the driving situation.

The thinly drawn curves in the figure are acounter-example of the time dependence witha control strategy, which includes neither en-gine intervention nor any adjustment of speedgradients. Here the largest distance from peakto peak in the vibration amplitudes of the ve-hicle acceleration is approx. 3 m/s2. These vi-brations decline after 4 seconds.

Two questions were theoretically answeredthrough this strategy development:

1. Can a power split CVT be controlled witha reasonable effort?

Answer: Yes, the above strategy is suitable forembedded control.

2. Can the mode change be realised com-fortably?

Answer: It is believed that the mode changecan be realised without any impact on comfortand vehicle performance.

The basis of the work is the knowledge of theoperational characteristic of the variator (mul-titronic®), the know-how regarding clutch ac-tuation and shifting (Easytronic® [11]) as wellas the development and simulation tools avail-able within LuK.

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SummaryPower split transmission structures present apromising concept for future applications ofCVT’s with torques up to 500 Nm and total ra-tio spreads between 6 and 7. With a dual-range CVT, the limitation of torque transfer ca-pability of the chain can be raised, as the pow-er flow through the variator at many loadpoints is decreased. Within the scope of thepresent theoretical investigation, it is shownthat with LuK components, a 500 Nm CVTwith a power split can be realised. The avail-able potential of a dual-range CVT can onlybe fully exploited if the concept is optimisedwith regard to efficiency along with variatorand chain load.

On the hydraulic concept, the LuK double-pis-ton principle was combined with speciallyadapted cascade hydraulics. A further devel-opment of the clamping system is integrated,which has advantages regarding fuel con-sumption and component load.

Computer simulations with a control strategyspecially developed for this transmissionstructure show that high driving comfort isachievable even across the mode change.This is possible without any significant limita-tion in driving dynamics.

References[1] Nowatschin, K.; Fleischmann, H.-P.;

Gleich, T.; Franzen, P.; Hommes, G.;Faust, H.; Friedmann, O.; Wild, H.:multitronic® – Das neue Automa-tikgetriebe von Audi, ATZ 102 (2000)7/8 and ATZ 102 (2000) 9.

[2] Gesenhaus, R.; Nowatschin, K.;Hommes, G.; Deimel, A.: Wie erlebt derFahrer die neue Getriebegeneration

multitronic® von Audi? VDI reports no.1610, Getriebe in Fahrzeugen 2001.

[3] Indlekofer, N.; Wagner, U.; Teubert, A.;Fidlin, A.: Latest Results from the CVTDevelopment, 7th LuK Symposium 2002.

[4] Förster, H. J.: Stufenlose Fahrzeug-getriebe, publisher TÜV Rheinland,1996, and included references.

[5] Looman, J.: Zahnradgetriebe Springer-Verlag, 1996.

[6] Faust, H.; Linnenbrügger, A.: CVT De-velopment at LuK, 6th LuK Symposium1998.

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