CVT Without Limits

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CVT without limits –Components for commercial vehicle transmissions

Andreas EnglischHartmut FaustManfred HommChristian LauingerMartin Vornehm

The rise in fuel costs and the necessity for fur-ther reductions in emissions require new techni-cal solutions allowing for optimisation of theentire drive train. As a result, the use of chainvariators is also expanding in the passengervehicle sector. Other applications will be addedto the Audi multitronic© [1] VL300 and its newerevolution the VL380 with 420 Nm variator torqueas well as the Ford/ZF CFT30 [2] which arealready in production. Optimisation of the powertrain also makes sense for buses, vans and com-

mercial vehicles as they are responsible for aconsiderable proportion of emissions.

A manufacturer of commercial vehicle transmis-sions applied itself to making the total ratio ofthe transmission so large and variable that theengine can be operated more or less steadily ina lowest consumption and emissions range.Engine optimisation in precisely this range thenoffers additional potential for improvement. Thevariator required for such a transmission waspositively evaluated on the basis of the LuK pro-duction components and development wasbegun.

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55 CCVVTT oouutt ooff lliimmiittss

Introduction

Figure 1 Data and targets comparing passenger and commercial vehicle utilisation of a CVT

Vehicle Audi A6 3.2FSI / 2.7TDI Commercial vehicle

Permitted weight in kg 2 200 ... 4 000 10 000 … 40 000

Max. engine torque in Nm 330 / 380 1 000 … 3 000

Max. engine power in kW 188 / 132 200 … 500

Required lifetime in km 300 000 > 1 000 000

Maximum speed in km/h 250 / 225 120

Period of operation in h > 3 000 > 15 000

Exhaust standards/targets EU4 Increasing requirementforeseeable

Transmission or variator Audi multitronic©

VL300 / VL380Power split CVT

Transmission structure D & Rev. without power split Several ranges + D & Rev.

Fastest complete variatoradjustment in s

1,2 0,65

Max. oil requirement for adjustment in l/min

5 19

Max. variator torque in Nm -60 … +350 / -60 … +420 -350 … +600

Number of chain rotationsduring the period of operationin 106

300 2 000

Center distance variator in mm 171 220

The challengeIf the data of an Audi A6 is compared with thedata of various commercial vehicles, the taskseems almost impossible. With a whole range oftargets, not only is a doubling required but alsoan increase in the size as shown in figure 1.

As the variator data shows, more than one indi-vidual component must be optimised to fulfilthese requirements. The following sections dealwith all these aspects of transmission structuresfrom pulley sets and pump systems to the chain.

In the course of the development, it is becomingincreasingly clear that the challenge of realisinga CVT for commercial vehicles is acceptable.

Transmission architectureOne of the key technologies in this CVT applica-tion is the principle of power splitting which wasalso described in the last LuK Symposium 2002[3, 4] as well as the use of several continuouslyvariable driving ranges. Power splitting allowsan increase of efficiency in combination with a

reduction of variator load. The concepts for cartransmission designs [5] introduced with thistechnology cover capacities up to more than 200kW and corresponding torques, thus leading intothe commercial vehicle segment.

This technological background makes it possibleto develop customised transmission structures,as is also the case, for example, with hydrostat-based transmissions [6]. In comparison with thehydrostat, the chain variator offers efficiencyand acoustic benefits which are of particularnecessity for use in buses. Compared with awholly electrical power conversion, the benefitof the chain converter in cost, efficiency andpower density is even greater.

A common prerequisite of applicable gearboxarchitectures is a speed-up of the variator,because commercial vehicle engines deploytheir power at lower speeds. With some of thetransmission architectures described, this func-tion can take place directly in the planetary gearprovided for input-side splitting. Also planned isa transmission for selection of multiple rangesthat can work where necessary with the conven-tional commercial vehicle dog clutch if a speedsynchronisation is achieved by suitable means.

The options to link aplanetary gear and achain variator togeth-er via adapter stagescan only be dividedinto two classifica-tions, namely with aninput-side or output-side planetary gear.In each of these twoclassifications, thereare alternative config-urations for the multi-range manualtransmission. Severaloptions are describedin figure 2.

In the architectureshown on the left, theplanetary gear is con-figured on the inputside and the multi-range-transmission inthe power path paral-


CCVVTT oouutt ooff lliimmiittss 55

Figure 2 Three alternative transmission architectures with ratio and power load characteristics. Each color represents one of several operation ranges

lel to the variator. If the shaft leading to thetransmission can be locked, a driving range with-out power splitting can also be presented. Belowthe drawing of the transmission architecture isan example of how, with appropriate gear ratios,several continuously variable driving ranges canbe represented. In each case, the bottom dia-gram shows the power percentage to which thevariator is subjected.

Depending on the application, the benefits ofone architecture category or another may pre-vail. The transmission depicted on the left pro-duces low driving ranges with very low variatorload – beneficial for an almost steadily usedunit. When changing from one driving range tothe next, the variator resets.

The center column shows a transmission whichhas a manual transmission assembly with twonon-coaxial input shafts. Incontrast to the transmissionshown on the left, it is possi-ble to change between rangeswithout resetting. To do this,the variator must transfer anaverage of 50 % of enginepower – more than with thetransmission on the left, butthis is sufficient for lightcommercial vehicles or vans.

With the transmission shownon the right, the cast of partsis simply switched betweeninput and output. In the exam-ple at hand, this primarily influences the torqueand speed ranges – but not the performance.

In order to depict a 'geared neutral' transmission,the planetary gear should fundamentally beplaced on the output side. If the planet is config-ured on the input side, the opposite can beshown: very long ratios up to a 'geared zero' ratiowhich allows continuous start/stop of the engine.

When driving, the benefit of all the architecturesshown lies in the, for commercial vehicles, com-paratively low number of gears and thereforerange changes. For example, with four drivingranges after start-up only one shift is made andcity centre driving can continue virtually withoutany range changes. Another range change is nec-essary on leaving the city and a final one on themotorway.

Variator developmentbeyond 500 NmThe main dimension: Center distanceIn order to increase the torque capacity, it isnecessary to enlarge the main dimensions. Withthe center distance of 220 mm selected herecompared to the 150 ... 190 mm feasible for pas-senger cars, a whole series of aspects arerelaxed. In part, the effect is clearly even greaterthan the 25 % enlargement in the center dis-tance, as the following list illustrates. Thechanges are indicated by arrows, the number ofwhich reflects the relationship to the center dis-tance enlargement:

The combination strengthens several strainelimination effects so that the maximum 600 Nmstarting from the 420 Nm already realised in themass-produced VL380 is very possible withclean design of all such aspects.

Efficiency measurementup to 600 NmTo confirm the above considerations as well asthe measurement results [7] and simulations [8]published by research institutions, efficiencymeasurements were carried out at LuK on a newhigh-performance test stand. As a completecharacteristic diagram comprising several ratios,speeds and torques (and to some extent clamp-ing forces) was to be used as the basis, this alsoposed considerable challenges for the test



Change Benefit

Chain radii ↑ Chain tractive forces ↓, Clamping forces ↓

Number of chain links ↑ Forces per rocker pin (RP) or joint (RJ) ↓↓

Chain rocking angle ↓ Better load distribution on chain link cross section

Force per rocker joint ↓↓ Deflection of the RJ ↓, Link load on chain edge ↓↓

Shaft diameter ↑ Bending stiffness ↑↑, Efficiency ↑

Crowning of the surface ↓ Stress ↓, Wear ↓, Wear per RJ ↓↓

department. For example the electrical generatorengine of this test stand has a nominal powerrating of 720 kW.

The results confirm that chain variators achieveefficiency figures of over 97 %. Figure 3 showsas an example the 600 Nm efficiency character-istic diagram with a constant drive speed of2500 min-1.

Shown are the raw measurements data (exceptsmoothing for noise suppression) from the test

transmission including the losses of the selfaligning pulley bearings. Due to its design withreplaceable pulley discs, the stiffness of the testtransmission is below the target stiffness. Inreality, the efficiency could therefore be evenhigher.

The center distance of 220 mm thus shows theexpected positive effect.

Continuously variablehydromechanical torquesensorThe pulley technology in use on the Audi VL380is the basis for use at even greater torques [9]. Ofcentral importance is the space-saving continu-ously variable torque sensor (VTS) [14] imple-mented inside the pressure chamber, shown infigure 4 in production design with stampedsheet metal parts.

The VTS provides the indispensable propertiesfor reliable continuous operation in commercialvehicles:

• Precise clamping force proportional to theactual torque for all ratios

• Prompt clamping to prevent damage evenwith jumps in the torque



Figure 3 Chain variator efficiency characteristic diagram

Figure 4 Driving pulley set with VTS continuously variable torque sensor for Audi VL380 with 420 Nm variator torque

The function of the continuously variable sensoris to convert the torque introduced along thecomponents shown in blue in the figure via ballramps into an axial force. These balls shown inyellow are arranged inside the mechanism. Thetorque then enters the brown component on theshaft via the opposite ramp and from theremoves on to the fixed sheave or via the teeth tothe moveable sheave components coloured ingreen.

However, the axial force produced by the ballsdoes not have a direct effect on the cone pulleybut, with the blue component, closes a hydraulicoutflow orifice. This mechanism very dynamical-ly adjusts a hydraulic pressure proportional tothe torque. The pressure then generates theactual clamping force for the green moveablesheave via large clamping areas. The same pres-sure is also supplied to the driven shaft forclamping.

The dependency of the ratio is thus achievedinsomuch as the ball ramp mechanism has dif-ferent ramp slopes at different radii. The rampangle appropriate for the respective ratio isselected through the radial positioning of theballs by means of the guiding surfaces, whichslide axially ratio-dependent with the moveablesheave (also shown in green in figure 4).

The detailed enlargement of the blue componentwith the various slopes ramps, figure 5 is used toillustrate the ratio dependence of the continu-ously variable torque sensor.

In the result, the pressure related to the torque isthe greatest in underdrive, declining continuallytill overdrive. The three-dimensional shape ofthe components is optimally adapted to theclamping requirement determined under manyloads.

The promptness achieved through the directhydromechanical principle is illustrated usingthe measurement in figure 6.

In the measurement shown, a step in the drivetorque excites a decaying drive-train oscilla-tion on the test stand. In a real vehicle this cor-responds, for example, to a sudden, jerkyacceleration. Even in this situation there is vir-tually no delay between the measured torqueand the measured pressure adjusted by thetorque sensor. Quantitatively, the adjustedpressure at each instance is also congruentwith the target pressure calculated from themeasured torque.



Figure 5 VTS ramp contour with ratio-dependent slopesand resulting clamp characteristic

Figure 6 Promptness of the VTS clamp system whenexposed to a step in the torque

HydraulicsThe double piston principle already used in themultitronic© makes it possible to hydraulicallyoperate all required quick changes of ratio evenwith small installed pump capacity. The result-ing benefit in consumption is also to be strivedfor with commercial vehicles. Due to theextremely rapid adjustments required with theexisting multi-range transmission structure,the double piston principle is actually indispen-sable for efficient hydraulics. It forms the basisfor the following hydraulic and pump develop-ment.

Independent actuation ofthe adjustment pressuresDespite upscaling the entire system and the pis-ton surfaces, the considerably higher coasttorque due to engine brake assemblies, as onlyone reason cause an increase in the requiredpeak pressure up to 100 bar. This requires high-er pressure amplification in the correspondingvalves for the adjustment pressure chambers.Due to independent actuation of both valves,this high valve amplification is stable with vol-ume flows up to 19 l/min and the hydraulics gainin actuation precision.

The tasks of the clamping and adjustment sys-tem and the solution implemented by thishydraulics are clearly illustrated in figure 7. Forsteady operation, the variator requires twoforces on the pulley sets which are in a particularproportion dependent on the ratio, the so-calledforce-balance Zeta ζ or also Kp/Ks. In drivingmode, the force-balance value for the LuK CVTchain is 1.05 (UD) to max. 1.6 (OD). In coastingmode, it is the reciprocal force-balance-valuewith inverse ratio, i.e. around 0.95 (OD) to 0.6(UD). Respective to the required force-balance,the pivot point of the rocker shown in grey maybe thought of as displaced. The task of theclamping pistons shown in red is to generate thebasic clamping forces on both pulley sets,whereupon high-pressure oil is exchangedbetween the pistons during the variator adjust-ment. The task of the small adjustment pistonsshown in orange is to generate the additionalresidual forces and adjustment forces requiredfor equalisation.

This hydraulic system offers not only a highdegree of stability and precision in the control ofthe pressures, but also permits other advancedfunctions due to the independence of the pres-sure control. Examples of these functionsinclude an increase in clamping safety on poorroad surfaces or a slight reduction in clampingforce with appropriate design of the torque sen-sor, e.g. for compensation of residual centrifugaloil pressure forces. The result is optimised oper-ating efficiency.

When adjusting the ratio, the forces balance isleft in a controlled manner. Each adjustmentpressure may be optionally reduced and/or theother adjustment pressure increased in combi-nation. This degree of freedom is beneficialespecially with the rapid adjustments of a multi-range transmission.

Thus the benefits of several systems are com-bined here: The continuously variable torque



Figure 7 Double piston principle and independent adjust-ment pressure actuation to balance the requiredforces in the variator

sensor contributes robustness and prompt-ness and the independent pressure control ofthe adjustment chambers offers the freedom ofsoftware-controlled free clamping force. Acombination with a slip control [10] of theclamping force is also possible with this sys-tem, without losing the benefits of the torquesensor.

Pumps and cooling systemThe lubricating oil and cooling requirements ofthe clutch, variator, gears and bearings make theuse of a low-pressure pump advisable in trans-missions with a power classification in excess of400 kW. Following identification of the design-relevant operating point (fully loaded hill startwith maximum clamping force and clutch cool-ing), the result for the intended gearbox-archi-tecture is a low-pressure, gerotor-style pumpwith 29 cm3 delivery volume.

The low-volume high-pressure pump required,for example, for the clamping force of the varia-tor is designed as a symmetrically divided, dual-

flow, fully compensated vane pump with a deliv-ery volume of 10 cm3 in total. A similar pumpdeveloped by LuK is also in production in theAutotronic© from DaimlerChrysler [11].

Combined with the low-pressure pump, itforms a tandem pump as a unit, figure 8, on ashaft which is overdriven by the engine. Thedelivery of low-pressure oil guarantees a cavitation-free supply to the high-pressurepump, permitting the compact design of theintake system as well as an efficient filter con-cept.

Pump efficiency due tointelligent control of thepump flowsIn comparison with a single-flow high-pressurepart, the hydraulic power requirement is con-siderably reduced through intelligent control ofthe second pump flow. For this purpose, anelectronic control valve is included in thehydraulic controls. Its function is explained infigure 9.

The position of the flow control valve illustratedon the left shows that the second flow of thevane-cell pump is switched to circulation. Thepump's drive torque is thus drastically reducedin the majority of driving situations. Note thatthis oil is not lost to the low-pressure consumersfor lubrication and cooling.

In the section of the diagram repeated on theright, the flow control valve is electrically con-trolled (disconnected) such that the second flowis united with the first flow via a one-way valve.This ensures that there is also enough high-pres-sure oil for the fastest adjustments of the varia-tor, e.g. when starting up or when changing driv-ing range.

Thanks to the optimised design, the pump loss-es of this transmission are low, as is also thecase for the multitronic© and Autotronic©. Analternative concept with only one pump for allconsumers would have caused a threefoldpower requirement with no cost benefitbecause the costs of a high-pressure pump arescaled unfavourably to those of a low-pressurepump.



Figure 8 Hydraulic tandem pump from the low-pressuregerotor assembly (blue) and dual-flow high-pressure vane pump (red)

Precautions for long-term useConsistent use of the surface technologies forwear protection positively tested in the car appli-cation, e.g. hard anodizing or chemical nickelcoating of the pistons, ensures smooth hydraulicfunction even with the high life expectancy of acommercial vehicle.

The CVT chainStrengthThe increasing experience and process optimisa-tion with the strength-optimised light-link geom-etry have increased the torque capacity of the 37mm wide LK3708 chain to such an extent that,according to the first trials with the 220 mm cen-ter distance, no enlargement of the chain seemsnecessary for 600 Nm variator torque. All theresults described in this article have beenachieved with this 37 mm wide chain. The con-nection between the center distance, strength-

ening and torque capacity described in LuK Sym-posium 2002 [12] is even exceeded with largecenter distance.

The range of chains in figure 10 is completed bynarrower chains and chains with reduced pitch inthe lower torque range. To allow for torquesgreater than even 600 Nm, a chain with anexpanded pitch, i.e. LK10 links with stronger linkcross section is also under development. Thanksto stronger rocker pins, the forces from the chainedge are evenly distributed on the adjacent links.

The suitability in terms of strength for theintended application was proven with commer-cial vehicle load cycles using damage calcula-tions. A comparison of the force strokes for pas-senger and commercial vehicles (including thereduction achieved through power splitting) isshown in figure 11. The more extensive quantityof force strokes due to the mileage is at a similarforce level for both commercial and passengervehicles. The maximum force strokes of bothgroups which are only slightly increased despitethe significant rise in maximum torque, occurcomparably seldom.



Figure 9 Control of the high and low-pressure volume flows for different devices

Durability of the pulleysurfacesIn consideration of an enormous expected life-time of more than one million km, particularattention must be paid to the subject of wear.However the high efficiency values already indi-cate low wear values: As a rule, wear requires

energy loss. For a CVT as a friction transmission,it is not only the material wear (quantifiable inweight per friction energy) that counts, but alsothe quality wear of the friction surfaces (quantifi-able for example as a change in friction value orchange in roughness).

Based on the materials and test experiences [13]compiled, both aspects are provided in intensive

and successful testingon several high-per-formance durabilitytest stands. Figure 12shows a few interme-diate results relatingto the long-term sta-bility of the metallictribological system.The change in theforce-balance value ζover several thousand(!) hours is depictedhere. The force-bal-ance value is a goodindicator for changesin the friction values.

The upper half of fig-ure 12 shows theresults, in black, of a



Figure 11 Summary of the link force strokes for passenger and commercial vehicles

Figure 10 Torque capacity of the variator with different chain types

reference car application with a center distanceof 171 mm. The discernibly low change in theforce-balance value ζ shows that the frictioncoefficient µ has only minimally changed. Thetargeted service life of 3000 h for a passengercar can be checked using a condensed test pro-cedure within approx. 150 h. The tribological sys-tem is very stable and offers noticeably highreserves.

Shown in blue are the results of the same assem-bly with a center distance of 171 mm, but whichhas been exposed to the scaled loads of thecommercial vehicle application. This scalingmeans that the torques were reduced in such away that the resulting pin-specific forces corre-spond to a 220 mm system. This intermediatestep contingent upon the trial method alreadysupports the service life potential of a center dis-tance enlargement. The targeted service life ofmore than 15000 h of a commercial vehicle

could be achieved using a condensed test proce-dure of 1000 h duration. The test was evenextended to a running time of 2500 h as a safe-guard. The different symbols thereby indicatedifferent test procedures, both of which havebeen run. In the procedure represented by asquare, the ratio is changed in several stages sothat a high load concentration is exerted on therocker pin but not on the pulley. The test proce-dure represented by a diamond takes place witha fixed ratio so that the contact loads are concen-trated both on the rocker pin and on the pulley(however both friction partners can thus also bewell adjusted to each other). The results of thevariator tested with a fixed ratio once again con-firm that mixed cycles represent the most rigor-ous and therefore most efficient test method.

Shown in green are the results of a true 220 mmvariator which is being exposed to the unscaledcommercial vehicles load cycle applying up to



Figure 12 Top: Change of the force-balance of different variators in durability tests. Bottom: Width wear of the rocker pins in these durability tests.

600 Nm. The initially fast change in force balanceas well as the high width wear at the beginning(lower half of the figure) are the result of thehighest load points being tested first in the testprogram. The stable behaviour is thereby alsothe result of a pitch sequence optimisation of thechain to be described.

The tribological system of chain, pulley setsurface and appropriate oil used here in thechain variator demonstrates overall stablebehaviour of the friction values and confirmsthe achievability of the ambitious service lifetargets.

Durability of the rocker pinend facesThe second endurance aspect is the wear of theend faces on the chain side, which leads to adecrease in the chain width. In the long-termtests shown above, this end face wear was deter-mined with periodic inspections. These resultsare shown in the lower half of figure 12 using thesame colour and symbol selection.

The overall low end face wear is a characteristicof optimised heat treatment.

The design of the rocker pin end faces ensuresthat stresses relevant to the wear, such as Hertz-ian stress, do not exceed the permitted level.Figure 13 shows an analysis of the contact pointsand characteristics on these end surfaces, calcu-lated using the three-dimensional chain calcula-tion program 'CHAIN', which also takes intoaccount all elastic deformations from the shaftsto the rocker pins when doing so.

The coloured areas are the contact ellipses, thestress of which is visualised using the colour.This calculation using LuK's CHAIN calculationprogram takes into account all elastic anddynamic effects on pulleys and chain as well asthe joint kinematics of the rocker pins. Thus,load details are traceable and can be taken intoaccount in the chain design.

With regard to any further increase in the run-ning time, optimisation of the pitch sequencecan also make a contribution, figure 14. Pitchsequences of long and short links arefavourable for acoustic priority, with the directsequence of two long links specifically exclud-

ed. The reason for this is that the greatest endface loads occur experimentally and in calcula-tions at precisely these locations. Chains opti-mised in such a way do not just show lower wearrates but can also withstand more overall wearbecause the width reduction in the chain occursuniformly.

SummaryFollowing the successful production launch of theAudi VL380 with 420 Nm variator torque, achievedwith the LuK CVT components (pulley sets with thenovel, continuously variable torque sensor, opti-mised LK3308 chain and hydraulic control withvane cell pump) ways are being sought in whichthis technology can also be utilised for commer-



Bild 13 CHAIN screenshot with visualisation of the contactproperties ratio in underdrive

Bild 14 Correlation of the local width wear with thesequence of long and short links. The optimisedchain avoids the direct sequence of two long links.

cial vehicle applications with power splitting andvariator torque up to 600 Nm. The challenges withrespect to structural development, variatordesign, pump and hydraulic development as wellas chain design required careful preliminary con-sideration and the courage to act which is reward-ed by the positive test results.

The interplay of a great deal of detailed work hasnow resulted in a promising situation for theapplication of power split cvt in commercial vehi-cles as well as in cars with an increased torque.

The first prototypes of a special transmissionstructure are being built in collaboration with anestablished customer in the commercial vehiclesector.

References[1] Fleischmann, H.-P., Gutz, H.; Kumpf, G.; Mar-

tin, F.; Schöffmann, M.: Die neuen Getriebeim Audi A6. ATZ 106(2004), Sonderheft AudiA6, p. 128-138.

[2] Wagner, G.; Remmlinger, U.; Fischer, M.;CFT30 – A Chain Driven CVT for FWD 6 Cylin-der Application. SAE Technical Paper Series2004-01-0648.

[3] Lauinger, C.; Vornehm, M.; Englisch, A.: Das 500Nm CVT. 7. LuK Symposium (2002), p. 91-106.

[4] Englisch, A.; Lauinger, C.; Vornehm, M.; Wag-ner, U.: 500 Nm CVT – LuK Components inPower Split. CVT 2002 Congress, Munich,7/8 October 2002, VDI Reports No. 1709(2002), p. 147-163.

[5] Tenberge, P.; Müller, J.; Sewart, J.: CVT fürhöchste Drehmomente – CVT mit Umschlin-gungsvariator und Leistungsverzweigung.Getriebe in Fahrzeugen 2004, Friedrichs-hafen, VDI Reports No. 1827 (2004), p. 669-712.

[6] Schumacher, A.; Harms, H.: Potential vonleistungsverzweigten Getrieben in leichtenund schweren Nutzfahrzeugen. Nutz-fahrzeuge 2005, Böblingen, VDI Reports No.1876, p. 63-78.

[7] Sattler, H.: Abschlußbericht Forschungs-vorhaben No. 221 “CVT Wirkungsgrad”;Forschungsvereinigung Antriebstechnik e.V.(Publ.), Frankfurt.

[8] Lebrecht, W.; Ulbrich, H.: Vergleich von CVT-Umschlingungsgetrieben. 4th InternationalCTI Symposium of Innovative Vehicle Trans-missions, Berlin 5.12.2005, Lecture/Article E6.

[9] Englisch, A.; Faust, H.; Homm, M.; Teubert,A.; Reuschel, M.; Lauinger, C.: Entwick-lungspotentiale für stufenlose Getriebe. ATZ103(2003), Issuet 7/8, p. 676-685.

[10]Faust, H.; Homm, M.; Reuschel, M.: EfficiencyOptimised CVT Hydraulic and Clamping System.CVT 2002 Congress, Munich, 7/8 October2002, VDI Reports No. 1709 (2002), p. 43-58

[11] Greiner, J.; Kiesel, J.; Veil, A.; Strenkert, J.:Front-CVT Automatikgetriebe (WFC 280) vonMercedes-Benz. Getriebe in Fahrzeugen2004, Friedrichshafen, VDI Reports No. 1827(2004), p. 421-445.

[12] Indlekofer, N.; Wagner, U.; Fidlin, A.; Teubert,A.: Neueste Ergebnisse der CVT-Entwicklung.7. LuK Symposium (2002), p. 63-72.

[13] Linnenbrügger, A.; Baumann, M.; Endler, T.:High Performance Chain CVTs and their Tri-bological Optimisation. Tribology of VehicleTransmissions 2005, Tsukaba/Japan 16/18February 2005, Proceedings p. 14-19.

[14]Englisch, A.; Faust, H.; Homm, M.; Teubert,A.; Vornehm, M.: Hochleistungs-CVT-Kompo-nenten. Getriebe in Fahrzeugen 2004,Friedrichshafen, VDI Reports No. 1827(2004), p. 649-668.



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