20076506(JSAE) 2007-32-0006(SAE) Double Clutch Transmission for Motorcycles M. Geieregger, A. Mair, B. Pollak All of AVL List GmbH, Graz Andreas Bilek KTM – Sportmotorcycle AG, Mattighofen Copyright © 2007 Society of Automotive Engineers of Japan, Inc. and Copyright © 2007 SAE International ABSTRACT Until now automatic or automated transmissions are not common in motorcycles whereas in passenger cars a clear trend towards various types of automatic shifting transmissions could be observed during the recent years. For 2-wheelers CVT drives are well known for 50 cc scooters and in the scooter class a trend towards CVTs for bigger displacement vehicles is obvious. Evidence for this is the introduction of scooters with a displacement up to 800 cc which use CVT drives. Reflecting the overall CO2 targets and their impact on individual transportation and the demand for sportiveness in the area of motorcycles dual clutch transmissions (DCT) are certainly a valid alternative. Meanwhile automated manual transmissions (AMT, clutch actuation and gear shifting by actuators) are available in series production. Other than in the most passenger car applications the shifts are initiated by the rider. As in the car area a full “by-wire” interface is utilized. From its driving performance a double clutch transmission (DCT) would fit very well into motorcycle applications. Compared to an automated manual transmission (AMT) a double clutch transmission does not have any interruption of the tractive force during the shifting and a DCT features instantaneous acceleration and the feeling of a “direct” connection of the engine with the rear wheel which is highly appreciated by motorcycle riders. Using a DCT for the motorcycle application it has to be guaranteed that the driver has the same or even better control abilities compared to the manual transmission (MT). To fulfill the driver’s demand of direct response and control of gear shift and clutch operation the motorcycle DCT needs to have driver controlled overrule functionalities of the automation system. Although there are several reasons which make it worth using a DCT for motorcycles there is currently no DCT for motorcycles in production. An investigation to apply passenger car DCT technology to motorcycles is the focus of this paper. The main objective is the study of the installation space requirements and the SETC 2007 packaging of the double clutch into a motorcycle. Special attention is paid to find an arrangement which ensures that there is no negative effect on the riding performance (e.g. no reduction of the achievable lean angle) but also on the visual appearance of the motorcycle. 1. INTRODUCTION & MOTIVATION The Pre-Study for the application of a DCT in a motorcycle focuses on the packaging of the major components. Target is to be able to utilize existing frames and engines. The study is carried out for the KTM LC8 990 cc V2 engine. The major specifications of this engine are shown in Figure 1. KTM LC8 990 Adventure Engine type 2-Cylinder 4-Stroke V 75° 4V DOHC Displacement 999 cc Bore x Stroke 101 x 62.4 mm Power 72 kW @ 8500 rpm Torque 95 Nm @ 6500 rpm Mixture preparation Electronic fuel injection Lubrication system Pressure lubrication Cooling Liquid cooled Transmission 6 gears, dog-clutch engagement Clutch Multi-plate wet in oil sump Clutch actuation hydraulic Figure 1 Main Engine Specifications Following KTM’s “READY TO RACE” philosophy the LC8 990 cc V2 engine is known as one of the most compact and most lightweight engines in the 1000 cc V2 engine category. The LC8 engine may not be the typical target engine for the initial application of a DCT but if it is possible to apply a DCT on this engine it will be no problem to apply this technology on other less sportive motorcycle engines. 1 / 13 Of course it is clear from the beginning of the study that a DCT can never offer the same compactness and weight optimized design as a conventional manually shifted motorcycle transmission since additional components such as the 2ndthe required clutch, an additional shift fork and the actuators for the clutches and the gear shifting will require additional space and will also lead to a certain weight penalty. 2. SYSTEM SPECIFICATION CLUTCH The heart of every DCT is the automatically actuated double clutch. Before the specification of the clutch module is possible considerations on the clutch type (wet vs. dry) and the investigation of possible clutch arrangements and packaging variants on a motorcycle are required. Figure 2 shows the clutch arrangements which are investigated. Version 1: “Packaged clutch” (coaxial or serial) Version 2: “Diagonal clutch arrangement” Version 3: “Right-Left clutch arrangement” As described in section 3 the version no. 1 is finally chosen as the best suitable clutch arrangement. It shall be mentioned that for other engine concepts (e. g. in-line engine instead of a V engine) also other clutch applications are possible but these variants are not covered by this study. Figure 2 Clutch Arrangement 2.1. Clutch Friction Concept One of the major aspects in the definition of a double clutch layout is the friction concept. The clutches can be designed in a wet clutch or a dry clutch layout. In Motorcycle application both are common using multiple discs. The selection of the suitable friction concept depends on several aspects such as packaging, heat capacity (clutch cooling), power train concept and layout as well as industrialization aspects. In terms of packaging motorcycle clutches it can be seen that a multi disc wet and multi disc dry clutch for the same application needs similar space. A disadvantage of the dry clutch in motorcycle applications is an increased noise emission at opened SETC 2007 2 / 13 20076506(JSAE) 2007-32-0006(SAE) clutch conditions caused by missing damping of the oil between the discs. Considering a DCT where one of the clutches is nearly always opened this effect can have a considerable contribution to the over all noise emission. Regarding heat flow and cooling of the clutch the wet type has an advantage over the dry clutch in long time load as the major heat arising from the friction when the clutch is slipping can be transferred into the oil which again can be cooled by an oil cooler. Thus the cooling power can be controlled by the oil flow rate. However the cooling of the wet clutch by the oil has the disadvantage of a limited peak heat load due to the cracking of the oil molecules. For example the limiting temperature of a standard automatic transmission oil is approximately 200 to 300°C[1]. The limiting temperature of the dry system is only determined by the friction material. Usually the materials used for dry clutches are capable to withstand temperatures up to at least 300°C. The decision on the clutch type for a motorcycle application is mainly determined by the usual power train layout of a motorcycle. One objective of this study is to define a DCT for a motorcycle which is very similar to the existing manual shifted version. Therefore a wet clutch and a common lubrication for the engine and clutch and transmission is chosen. Additionally the smaller outer diameter of a wet clutch can be easier packaged than the big diameter dry type. 2.2. Packaging In terms of packaging of the clutch and transmission the possible arrangements of the two clutches are investigated. There are several possible combinations which could be used in a motorcycle application in terms of clutch position (see Figure 2). On the one hand side it has to be decided whether to mount both clutches on one position as shown in version1 or to mount one clutch one each side of the engine as shown in version 2 and 3. For version1 the type of arrangement of both clutches has to be decided. Under consideration of packaging, actuation and type of clutch (wet/dry) it has to be decided to mount the clutches in a serial way or coaxial position relative to each other (see Figure 4). 20076506(JSAE) 2007-32-0006(SAE) KTM Adventure current clutch 200 100 132 111 11 100 100 100 VW DCT launch clutch 320 279 193 163 4 93 120 175 VW DCT shift clutch 320 386 138 103 5 109 210 125 Torque Nm Clamping Force [%] Outer Diameter Friction Disc mm Inner Diameter Friction Disc mm Number of Discs - Clutch Safety (µ= 0.09) [%] Pressure Load of Friction Disc [%] Maximum relative velocity at open clutch [%] Figure 3 2.3. Clutch Specification The definition of the clutch size for the concept layout is mainly determined by the required safety of the clutch against slipping, the maximum pressure of the discs and the maximum relative velocity between discs and plates at open clutch position. In this study the references for limiting values are derived from the existing series production clutch of KTM Adventure as well as from the VW series production DCT. The information on the VW DCT is taken from a benchmark study done at AVL. In Figure 5 an overview on the major influencing parameters which are considered is shown. Figure 4 Coaxial vs. Serial Clutch [2] Figure 5 Major clutch parameters The parameters regarding clamping force, clutch safety, pressure load of friction disc and maximum relative velocity at open clutch are shown as relative values whereas the Adventure clutch represents 100%. Beside the clutch safety value the values for maximum pressure of the plates and the value of the maximum relative velocity of the friction disc and plate are influencing the clutch specification and the clutch size. Clutch safety against slipping Comparing the minimum values for clutch safety against slipping it can be seen that the value of the VW DCT launch clutch is slightly smaller compared to the KTM clutch. For this study the clutch safety values of the KTM application are set as limiting values for the clutch layout. Maximum pressure of the plates In terms of pressure of the plates it has to be distinguished between launch clutch (odd gears 1,3,5) and shift clutch (even gears 2,4,6). The clutch used for vehicle launch is usually more critical for clutch shutter when launching the vehicle. A major influence on shutter is the maximum pressure of the friction plates. The target for the pressure within this study for the launch clutch is set to the values as used in the KTM MT transmission. Comparing the values of the VW DCT it can be seen that much higher pressure can be applied compared to the KTM values. The reason for this is seen in a different friction material, disc design as well as in the fact that the VW DCT is using special transmission oil compared to the standard motorcycle The arrangement of both clutches on the same position according to version 1 in Figure 2 is most similar to the usual layout of a motorcycle. Furthermore this clutch arrangement is also the common solution on passenger car DCT applications. The clutches can be arranged in a coaxial or serial concept as shown in Figure 4. These arrangements differ significantly in terms of required packaging space. Comparing two types of the same friction concept (wet/dry) the coaxial version needs less space in axial direction while the serial type requires less space in diameter. Compared to passenger car applications the DCT application on a motorcycle offers further possibilities for the clutch arrangement. As the power train of a typical motorcycle is usually designed in a way that the engine and transmission are combined in one housing where the transmission is located parallel to the engine (crankshaft) it would be possible to mount one clutch on the left and the other one at the right side of the engine. (See version2 and 3 in Figure 2). Using such a concept for the clutch arrangement would ndrequire an additional (2) primary drive which has to be integrated into the engine / transmission housing. SETC 2007 3 / 13 engine oil used in the KTM application due to the common oil circuit of engine and transmission. The shutter influence can also be seen at the VW application. The maximum pressure of the even gear clutch is much higher compared to the clutch for odd gears (= launch clutch). Maximum relative velocity of the clutch discs One additional effect which has to be considered for the definition of the clutch size is the relative velocity between friction disc and clutch plate at open clutch. Assuming an operating condition where the vehicle is in standstill while gear 1 and 2 are engaged and the maximum input speed is applied the outer discs are moving towards the inner friction plates with the highest possible relative velocity. If this operating condition can really appear in the motorcycle depends on the chosen control strategy. As the clearance between disc and plate is in the range of 0.1 mm to 0.3 mm a high shear stress within the oil occurs. With increasing velocity the shear stress increases and as a result of this stress forces the oil is heated up. To avoid overheating of the oil which would result in cracking of the oil molecules the maximum level of the stress and therefore the max relative velocity has to be limited. This effect is most critical in a clutch arrangement where the clutch is located directly on crankshaft of the engine (version 2 in Figure 2) since such an arrangement leads too a higher relative velocity because the clutch rotates with engine speed which is approximately double the speed of the transmission input speed which is equal to the clutch speed in a usual motorcycle transmission. Of course the higher relative speed of the clutch on the crankshaft would be partly compensated by a smaller clutch diameter due to the lower torque at the crankshaft. Within this study the limiting velocity is set to a level of the VW DCT application as this effect is more critical in this application due to higher speed (engine speed) and bigger clutch diameter. 3. TRANSMISSION LAYOUT SELECTION Within this study several different transmission layout configurations are analyzed (e.g. arrangement of 3 shafts in a triangle vs. 2 shaft axes whereas 2 of the 3 shafts are concentric, shaft in hollow shaft vs. bearing in the middle, vertical split engine cases vs. horizontal split cases,…). Some of the concepts are rejected in the early stage of the study either because they are not applicable for a motorcycle at all or because they are not suitable for the V2 engine concept of the LC8 engine. Finally a feasibility investigation of three different layouts as mentioned in section 2 is carried out to identify the most promising layout for this motorcycle application. In the following the investigated transmission layout concepts are described and discussed in terms of required clutch size and packaging. SETC 2007 4 / 13 20076506(JSAE) 2007-32-0006(SAE) 3.1. Version 1 “Packaged Clutch” As already mentioned version 1 according to Figure 6 is most similar to usual motorcycle transmission layouts. Furthermore this configuration where both clutches are located at the same position is comparable to passenger car applications in terms of clutch arrangement. Figure 6 Transmission Layout Version 1 To keep the transmission dimension in axial direction as small as possible a coaxial arrangement of the clutches is chosen. The total width of the transmission would be the width of the current manual transmission plus the additional space requirement for a fourth shift fork (for details see section 4). The two transmission input shafts are designed in a hollow shaft system. As shown in Figure 6 either the outer or inner clutch can be used as vehicle launch clutch. The position of the launch clutch whereas the 1stleads to different gear arrangements and 2nd gear are positioned close to the shaft bearing as this is usually done. Inner clutch used for launch (1.1) Gears 1, 3, 5 are positioned on the inner shaft Gears 2, 4, 6 are positioned on the hollow outer shaft Outer clutch used for launch (1.2) Gears 1, 3, 5 are positioned on the hollow outer shaft Gears 2, 4, 6 are positioned on the inner shaft The advantage of using the outer clutch for the odd gear group (version 1.2) is a bigger clutch diameter which results in higher safety against slipping using the same clamping force compared to the inner clutch. In the VW double clutch transmission the outer clutch is used for the odd gears and respectively as vehicle launch clutch. When the outer clutch is used for the odd gears of the motorcycle DCT the first gear has to be located on the hollow shaft. The inner diameter of the hollow shaft is too big in relation to the root diameter of the 1st gear. Thus the material between root diameter and hole of the shaft is not sufficient which would yield into tooth brakeage. Therefore such a layout would only be possible by increasing the root diameter of 1st gear which again results into an increased center distance of the transmission shafts. Assuming the same gear ratio and module as in the existing MT application an axis distance increase of approximately 20% would be required. This however is not acceptable in terms of packaging. In the arrangement shown in version 1.1 the inner clutch is working Thus the 1sttogether with the odd gear group. gear is located at the inner shaft which is smaller the 1stin diameter. Thus the tooth root conditions of gear are the same as in the MT version. The disadvantage of the smaller diameter of the launch clutch in this arrangement is compensated by a balanced definition of the number of friction plates, friction area and clamp force. A more detailed description of the clutch definition is given in section 4. Advantages of version 1: Clutch- / transmission arrangement similar to usual motorcycle layout Also typical optical appearance of a motorcycle can be maintained Only one primary drive required No modification of alternator No major effect on center distance of transmission shafts Finally version 1.1 is chosen as the optimal concept. 3.2. Version 2 “Diagonal Clutch Arrangement” In Figure 7 the basic packaging layout of the clutch arrangement according version 2 is shown. Version 2 is based on the idea of mounting one clutch on each side of the transmission. Clutch 1 is located directly on the crankshaft at the alternator side of the transmission. Clutch 2 is located at the same position as on the original LC8 transmission. SETC 2007 5 / 13 20076506(JSAE) 2007-32-0006(SAE) Figure 7 Transmission Layout Version 2 The additional clutch on the crankshaft (clutch 1) requires a relocation of the alternator. For example the alternator has to be moved to the front of the engine and has to be driven by an additional pair of gears. The drive gear of the second primary drive can not be used to drive the generator since this gear only rotates when the clutch is closed. Alternatively it can be checked if the alternator can be moved to the right engine side where the drive gear of the primary drive may be used to drive the alternator. When the alternator is moved away from the crankshaft the starting system has to be modified since in the original design the one way clutch of the starting mechanism is positioned in the alternator as this is usual for motorcycles. Furthermore a second primary drive is required. The principle design of this second primary drive including damping mechanism would be the same as in a usual manual motorcycle transmission. Two separate primary drives for the gears 1,3,5 and 2,4,6 offer the potential for a more flexible definition of the gear ratios. The second primary drive would lead to an increased engine width. The clutch on the crankshaft (clutch 1) could be designed very compact since the torque is approximately half of the usual transmission input torque. Problems may occur with the higher maximum relative speed between the clutch discs. But as mentioned before the higher relative speed of the clutch on the crankshaft would be partly compensated by a smaller clutch diameter due to the lower torque at the crankshaft. Compared to version 1 the clutch on the transmission input shaft (clutch 2) can be designed in a more compact way since there is only one clutch located at this position instead of two nested coaxial clutches. The clutch dimensions (e.g. outer diameter) could be optimized according to the clutch load parameters. Compared to the current manual transmission clutch 2 could be packed into approximately the same radial and a smaller axial installation space since this clutch would have only 6 discs instead of 11 (see section 4) The total width of the transmission would be the width of version 1 plus the additional space requirement for the second primary drive and the drive for the alternator. The major reason for disqualification of version 2 is the collision of the additional primary drive with the final drive (chain sprocket). When the gear ratio of the primary drive is the same as on the current manual transmission, the overlap is bigger than 30 mm. Some improvement of the situation can be achieved when a different gear ratio is chosen for this primary drive but this is not sufficient to solve the problem completely. To avoid the collision a gear ratio of the primary drive of 1 would be required which leads to a high 20076506(JSAE) 2007-32-0006(SAE) transmission input speed which then causes the following problems. Too high gear ratio of the lowest gear (1 or 2) of this input shaft required. Possible compensation potential of final drive is not sufficient. High speed may cause problems with the bearings. stnd A possible solution would be to increase the center distance of the transmission shafts but the required 30 mm are not acceptable in terms of packaging. For other motorcycle engines where an increased center distance of the transmission shafts can be accepted a variant with adapted primary gear ratio, slightly increase center distance of the transmission shafts and higher final drive ratio may be acceptable. Another theoretical solution is to arrange the additional primary drive outside of the chain sprocket. This version is unacceptable due the significant increase in engine width, the very complicated service (e.g. changing of sprocket) and the collision with the frame. Finally version 2 is rejected due to the following main reasons: Not acceptable requirement for increased center distance of transmission Additional primary drive (more components, increase in transmission width) Figure 8 Transmission Layout Version 3 The additional clutch on the left engine side (clutch 1) collides with the alternator. Therefore a relocation of the alternator as discussed for version 2 is required too. Again this would lead to a modified starting system. Reference Version 1.1 Shift Clutch 200 153 160 145 6 104 153 157 Launch Clutch 200 204 129 93 6 102 117 127 Version 1.2 Launch Clutch 200 1 160 135 6 125 118 157 Shift Clutch 200 192 119 95 6 92 171 117 Torque Clamping Force Outer Diameter Friction Disc Inner Diameter Friction Disc Number of Discs Clutch Safety (µ= 0.09) Pressure Load of Friction Disc Maximum relative velocity at open discs Figure 9 Clutch Dimensioning Nm [%] mm mm - [%] [%] [%] KTM Adventure 200 100 132 111 11 100 100 100 Significant modifications for relocation of alternator and new starting system 3.3. Version 3 “Right – Left Clutch Arrangement” In Figure 8 the basic packaging layout of the clutch arrangement according version 3 is shown. In version 3 one clutch is positioned on each side of the transmission but different to version 2 both clutches are located on the transmission input shafts. Furthermore also version 3 requires a second primary drive which offers the same potential in terms of a more flexible definition of the gear ratios but would again lead to an increased engine width. Same as for version 1 both clutches have to transmit the transmission input torque. Nevertheless the clutches would have a different size in order to optimize the clutches in terms of the different limits regarding maximum pressure for the launch and shift clutch. Following the argumentation for version 2 the major reason for disqualification of version 3 is the collision of the additional primary drive with the final drive (chain sprocket). 4. DETAILED INVESTIGATION VERSION 1 SETC 2007 6 / 13 Version 1.1 is chosen as the most promising concept for the application of a DCT on the KTM LC 8 engine. The following sections show the results of the detailed investigations. 4.1. Clutch dimensioning As basis for the packaging investigations the clutch dimension are specified. Figure 8 shows the main dimensions and most important parameters for the clutch dimensioning for the current Adventure transmission in comparison to the calculation results of version 1.1 and 1.2. The parameters regarding clamping force, clutch safety, pressure load of friction disc and maximum relative velocity at open clutch are shown as relative values whereas the Adventure clutch represents 100%. As described in section 2 the current KTM clutch and the VW DCT are used as reference to set the limits for the major parameters during the clutch calculation. The detail calculation of the clutch is carried out for the versions 1.1 and 1.2. In principle the version 1.2 where the outer clutch is the launch clutch would be the preferred version but as explained in section 3 version 1.1 is finally chosen so that the 1st gear can be positioned on the smaller inner shaft. For the finally chosen version 1.1 a clutch safety against slipping of 102% for the launch clutch and 104% for the shift clutch in relation to the original clutch of the manual transmission is achieved whereas the friction value for the wet oil system is set to 0.09. Since the clutch is not any longer actuated by hand (see 4.2) it is possible to increase the clamping force and therefore reduce the number of clutch discs from 11 to 6. To avoid clutch shutter the pressure load of the launch clutch is only moderate increased (117%) compared to the LC8 clutch. This is done by an increased contact area of the clutch discs. An increase of the pressure by 17% is accepted to keep the clutch size small. For the maximum relative velocity the VW clutch is set as limit whereas the maximum of 157% relative to the LC8 is still significantly lower than the 175% of the VW clutch. 4.2. Clutch Actuation The clutches are actuated by hydraulic pistons via a lever spring system. The required oil pressure can be either generated by an electro motor connected to a piston by a spindle gear or with a direct driven oil pump. The electro motors or hydraulic valves are controlled by a transmission control unit (TCU). The principle arrangement of the lever spring system for the clutch actuation is shown in Figure 10. The outer clutch (shift clutch) is actuated by piston 1 via lever spring 1. Via this clutch the transmission input torque (Tin) is transmitted to the gears 2, 4, 6. The inner clutch (launch clutch) is actuated by piston 2 via SETC 2007 7 / 13 20076506(JSAE) 2007-32-0006(SAE) lever spring 2. Via this clutch the transmission input torque (Tin) is transmitted to the gears 1, 3, 5. The lever ratios are between 3 and 4. This reduces the actuation force of the hydraulic pistons but increases the required piston travel. The maximum required piston travel including compensation of the clutch wear is 13.5 mm for the outer piston and 17 mm for the inner piston. Figure 10 Example Clutch Actuation Principle The lever springs rotate with the speed of the clutch cage and are connected to the actuation pistons via axial bearings. An exploded view of the clutch inclusive actuation lever springs and hydraulic pistons is shown in Figure 11. The clutch actuation force is additionally loading the transmission bearings all the time during operation whereas in a normal motorcycle transmission the clutch release force only occurs during the clutch opening. This load occurs independent from the clutch type (normally open / normally closed) as one clutch is clamped and the other is open beside of the situations of shifting or stopped vehicle. Figure 12 Additional Bearing Load Due to the additional bearing load the life time of the bearings have to be checked. A first investigation shows that the additional axial force would only have a negligible able effect on the bearing load at maximum torque operating point. At operating points with lower torque (e.g. 50%) the relative reduction of the bearing life time would be considerable (e.g. life time only 40% to 50% of the manual transmission) but the absolute lifetime in these operating points is that high that even such a high relative reduction is not critical for the overall life time of the bearings especially when a usual load cycle is taken into consideration. The investigations regarding bearing lifetime are carried out for the bearings used in the original manual transmission. By changing the bearing type or introducing an additional axial trust bearing the bearing life time can be increased significantly and brought up and above the level of the manual transmission. Assuming comparable regulations as in passenger car applications another issue which has to be considered caused by the “normally open” clutch control strategy is an additional device which secures the motorcycle against rolling away when parking. However this is not seen to be a most critical issue. Usually a motorcycle is parked in a plane and even area where uncontrolled rolling away is not expected. If special applications require a parking device several solutions are possible. When using the electro hydraulic clutch actuation system with the spindle gear the self locking behavior of the spindle combined by the closed hydraulic system a permanent closing of one clutch even at standstill can be realized. Using a hydraulic actuation system with constant driven oil pump and hydraulic valves a SETC 2007 8 / 13 20076506(JSAE) 2007-32-0006(SAE) locking of one clutch at standstill can not be guaranteed due to oil leakage in the valves. Here a special park brake device as used in passenger car applications could be implemented. That most likely additional device is required which secures the motorcycle against rolling away. There are several ways how this can be realized and this not seen to be a critical issue. In Figure 13 the main parameter for the clutch actuation are summarized. Reference Adventure V1.1 Outer Inner Clamping Force [%] 100 153 204 Actuation force [%] 100 51 62 Disc Clearance total [%] 100 55 55 Wear total [%] 100 55 55 Actuation travel [%] 100 1 179 Figure 13 Clutch Actuation Parameters The clamping force at the clutch discs is 1% for the outer and 192% for the inner clutch compared to the Adventure clutch where the clamping force is limited by the actuation by the rider. Although the clamping force is significantly higher the actuation force of the hydraulic piston can be kept relatively low due to the lever spring actuation system but the required piston travel is increased accordingly. Due to the higher clamping force the number of clutch discs is reduced which reduces the total disc clearance and the total wear. 4.3. Clutch Lubrication Comparing the clutch lubrication and cooling concept of the MT version with the DCT it has to be investigated if an increased forced cooling oil flow is required. As the weight and torque which are mainly influencing the heat load during launch are the same in MT and DCT version it is expected that the heat load on the clutch discs is the same. An additional effect which has to be considered in the coaxial clutch arrangement of the DCT is that the inner clutch is surrounded by the outer one which results in different lubrication conditions compared to the MT version. A difference in terms of heat load on the clutch arises by the fact of overlapping clutch during shifting under load. Here it has to be investigated if additional cooling is required. In the motorcycle application the boundary conditions are less critical than in cars as the motorcycle has less weight and gear spread which are mainly determining the heat load on the clutch during shifting. Figure 14 Clutch Lubrication / Cooling If detailed investigations show that additional lubrication of the clutches is required the forced oil flow controlled by the oil pump has to be adapted. The principle of the oil path via the inner shaft would be similar to the MT version and can be seen in Figure 14. However this topic is strongly depending on the final vehicle type and operating conditions. Comparing an off-road motorcycle with an on road application different load caused by different driving profiles will be applied to the clutch. Off-road applications usually show much higher launch cycles compared to on roads and often higher load on the clutch during launch. Depending on the application the necessity of additional oil cooling of the clutch will be decided later. 4.4. Gear arrangement Compared to the MT version the gear arrangement of the gears has to be changed as a DCT needs 2 separate gear groups; one for the odd gears 1, 3 and 5 and one for the even gears 2, 4 and 6. Each group is located at one shaft to ensure a pre-selection of next gears. The location of which group is at which shaft is depending on the layout of the clutch. As already mentioned in section 2 the definition of the clutch size is depending whether the clutch is used for launch or only shifting. Respectively the gear group for odd gears containing the 1st gear has to be mounted to the launch clutch. According to section 3 the arrangement where the outer clutch is used for launch would be the preferred version in terms of clutch safety and shutter behavior. However this arrangement is limited as the root diameter of the first gear is too small for the hollow SETC 2007 9 / 13 20076506(JSAE) 2007-32-0006(SAE) shaft. Thus the arrangement of inner clutch used for launch is used in this study. Figure 15 Gear arrangement In Figure 15 the gear arrangement is shown. For the DCT configuration the position of the gears 1 2 and 5 6 is exchanged compared to the MT version; gear 3 and 4 stay at the same position. In a detailed investigation it has to be considered that the bearing loads will change due to the rearrangement of 1st and 2nd gear. Furthermore the bearing concept for the transmission has to be changed. Due to the hollow shaft system additional bearings are required (see Figure 16). The inner shaft is supported in the outer shaft by two needle bearings. Currently needle bearings with a needle diameter of 2 mm are foreseen in the concept and the bearing width is adjusted to a value which will be capable to achieve the required bearing lifetime. The bearing size has to be validated by detail bearing calculations. Beneficial in terms of bearing life time is the relative low bearing speed which is the differential speed between inner and outer shaft. As the shift system is designed in a way where only sequential gear engagement is possible the difference in shaft speed is only between two neighboring gears; e. g. considering a gear spread of 1.36 and a transmission input speed of 4000 rpm this results in a difference speed of 1400 rpm which is uncritical. For the design of the 1st gear it has to be decided whether to use a pinion shaft or separate gear with spline connection to the shaft (see Figure15). On top a 20076506(JSAE) 2007-32-0006(SAE) variant with spline is shown and the lower picture of Finally the variant with pinion shaft is chosen since the Figure 16 shows a variant with pinion shaft. Figure 16 Hollow Shaft System Arrangement The difficulty using a spline connection configuration is that the integration of the spline to fix the 1st gear is critical while maintaining a minimum gear diameter but ensuring that the shaft diameter at 1stthe position of the gear is bigger than the diameter of the left transmission shaft bearing. Theoretically the spline connection has the potential to increase the diameter of the needle bearings and therefore the shaft diameter in the middle since the gears can be assembled from the left side. This potential can not really be utilized since the maximum possible needle bearing diameter is also limited by the gears which are mounted on the hollow shaft which should be kept as small in diameter as possible while maintaining the required wall thickness of the hollow shaft. In the design variant using a pinion shaft configuration the assembly of the gears of 3rd and 5th gear has to be done from the right side (“over the running surface of the needle bearings”). Therefore it has to be ensured that the root diameter for the sliding gear spline needs to be bigger than the inner diameter of the needle bearing of the hollow shaft system. This may require a reduction of the needle bearing diameter which reduces the shaft diameter in this area which would result in a weakened shaft design in terms of strength and deformation. Comparing the outer diameter of the VW DCT inner shaft with a diameter in the range of 23 mm with the diameter of the inner shaft in this concept with pinion gear of 24 mm this feature is not seen to be critical. SETC 2007 10 / 13 theoretical potential of a bigger shaft diameter with the spline connection can not be utilized due to other constraints but would cause problems to fix the 1st gear. The gear arrangement in two groups (odd group and even group) requires an additional (4thDCT concept 3rd and 4th) shift fork. In the gear require a separate actuation which needs two individual shift forks for the gears 3 and 4. In the original configuration one sliding gear for gear 3 and 4 using only one shift fork is required. This has an impact on the packaging size of the transmission. As shown in Figure 15 the axial width of the DCT gear box will increase by approximately 10 mm compared to the original MT version. 4.5. Shift system The principle shift system concept of the DCT using sliding gears which are actuated by a shift drum via shift forks is the same as at the existing MT version of the KTM LC8 engine. The transmission actuation and shift system of the DCT is shown in Figure 17. Different to the MT the actuation of the shift drum is done by an electric motor via a spindle drive. As the spindle drive can be designed to be self locking no additional detent system is required. The control of the electric motor is done by a transmission control unit (TCU). The necessity of an additional positioning sensor for the detection of the shift drum position is depending on the control system and electric motor layout. Figure 17 Transmission Actuation and Shift System The additional shift fork (see section 4.4) and the shift strategy of a DCT where two gears are engaged at the same time require a new shift drum. In Figure 18 the principle groove layout of the shift drum is shown. First one additional grove is required for the 4th shift fork and second the grooves itself have to be adapted to the DCT shift strategy. In order to keep the same length of the shift drum although an additional groove is required it is necessary to change the shift drum material from Aluminum to steel and to skip the rollers 20076506(JSAE) 2007-32-0006(SAE) on the shift forks in order to reduce the width of the grooves. In Figure 18 the gear sequence of the DCT shift drum can be seen. Starting from neutral position no gear is stengaged. Together with engagement of the 1 gear the nd2 gear is engaged. The following step keeps gear 2 stengaged while the odd gear set is changed from 1 to rd3 gear. This strategy of keeping one gear engaged while changing the gear of the other group is also done thfor the following gears up to 6 gear. This results in 6 positions of the shift drum. Figure 19 DCT vs. LC8 MT The main objective of the packaging was to keep the outer dimensions of the DCT as close as possible to the current manual transmission and at least to ensure that necessary modifications of the outer contour are kept a small as possible in those areas where extended dimensions of the transmission cases would cause negative effects on the riding performance (e. g. reduced lean angle) or on the riders ergonomics. Though one of the major limits in this study is keeping the width and diameter of the clutch cover the same as in the MT version. As shown in section 4.1 the definition of the clutch size requires an increased outer diameter of the friction discs compared to the MT version but this can be partly compensated by a more compact design of the clutch housing. Nevertheless the clutch outer diameter increases by approximately 14% and slightly exceeds the existing clutch cover surface. The length of the clutch is reduced according to the reduced number of clutch plates. This provides space for the hydraulic pistons for the clutch actuation. Figure 18 Shift Drum Layout In Figure 20 a dummy model of for the actuator is shown on top of the transmission. Generally the packaging of the clutch actuator is not very critical as the connection between clutch actuation piston and actuator is done by hydraulic lines. Thus depending on the vehicle packaging situation different positions for the actuator are possible. The position of the primary drive is kept at the same position as on the MT version. The transmission itself has the same center distance of the transmission shafts as the current manual transmission. As described in section 4.4 the total width of the gear group increases by approximately 10 mm mainly due to the additional shift fork. This requires modifications of the transmission case in the area of the chain sprocket of the final drive and possibly a rearrangement of the chain line. During further optimization loops this can be partly compensated by shifting the transmission a little bit to the primary drive side by modification of the primary drive. During the optimization of the shifting drum groove layout the DCT offers a higher freedom than the MT since the actuation forces can be higher. There is no limitation of actuation force by the riders feel engaging and disengaging by foot actuation. Furthermore it would be possible to realize different rotation angles of the shift fork e.g. for N 1/2 compared to 1/2 2/3 to optimize the gradient of the groove. 4.6. Packaging Figure 19 shows the CAD models of the complete DCT concept including shift system, actuation, clutches and clutch actuation and the current manual transmission of the KTM LC8. These models are used for the packaging investigations of the DCT. SETC 2007 11 / 13 20076506(JSAE) 2007-32-0006(SAE) The shift drum and shift fork can be packaged into the given space. The actuation of the shift drum by an electric motor via a spindle drive a relatively compact solution and offers several possible positions for the packaging whereas for definition of the final version it is principally possible to arrange it at both ends of the shift drum under different angles. Finally the position of alternator, oil pump drive, starter, balancing gears and water pump are kept at the same position as on the MT version. For multiple applications a DCT will enhance the riding comfort and the continuous torque flow over a shift will also enable a high acceleration performance. This study was the first step to a successful application of a DCT on a motorcycle. In a next step a more detailed transmission design needs to be prepared particularly regarding actuation and control unit. In terms of actuation it has to be investigated if a system using a permanent driven oil pump in combination with electric valves is more suitable for such an application compared to an electromotor system. Considering packaging it is expected that the constant driven oil pump is less critical. Here it also has to be investigated if a separation of the oil circuits between engine oil and transmission/hydraulic oil is possible and beneficial. In the existing MT version the same oil is used for both engine and transmission lubrication. Using a different oil circuit for the transmission and engine would have the benefit of using a special oil with better lubrication behavior compared to the engine oil for the transmission. This would increase the lubrication condition for the bearings and gears. Thus the life time of both could be increased. Engine and transmission structure as well as the electric system shall be adapted for DCT: an additional transmission control unit (TCU) will be required which communicates with the engine control unit (ECU); also an electronic throttle body shall be considered to utilize the whole potential of this transmission technology in terms of riding performance and comfort. With focus on safety it has to be investigated more detailed if an electronic throttle needs to be equipped with a mechanical connection to the driver throttle to guarantee closing of the intake manifold in case of electronic malfunction. Another important feature of DCT will be to develop engine and transmission control strategies adequate to the desired motorcycle application (sportive acceleration / cruising comfort). Here it has to be investigated more detailed if the final functionality needs to be designed in a way where the driver of the motorcycle has the possibility to overrule the electronic control system in terms of gear shift and clutch operation. It is expected that especially in driving conditions like wheelie or drift a driver controlled clutch operation is required. Here it has to be investigated if a driver controlled clutch actuation which is overruling the electrical system can be realized on a mechanical way or if it can be done via the electric actuation system by an electronic clutch lever sensor. The final key factor for the market success of a DCT in a motorcycle will be a refined ECU and TCU calibration. Further to the system study presented it is therefore clearly recommended to verify the DCT performance in a complete motorcycle demonstrator. Figure 20 Packaging of DCT The overall impact of the double clutch transmission on the over all package can be seen in Figure 20. Most significant is the increase in clutch diameter and the increased width of the gear group. The packaging of a DCT will lead to slightly extended dimensions of the base engine and transmission cases and requires the packaging of additional components (e. g. clutch actuators) on other locations on the motorcycle. 5. CONCLUSION AND OUTLOOK The objective of this study is to investigate the installation space requirements and packaging with special focus on lean angle of the double clutch transmission into a motorcycle. The study shows that the DCT can be packaged into a motorcycle. Of course the application of a DCT will require compromises in terms of compactness and weight compared to a manual motorcycle transmission. Furthermore it shall be mentioned that also an increased inertia of the transmission has to be taken into account. SETC 2007 12 / 13 6. REFERENCES [1] OMV-Aktiengesellschaft: http://www.omv.com/smgr/portal/jsp/index.jsp?p_site=AT [2] Dipl. –Ing. Dipl.-Wirt.-Ing. Gerd Jäggle, Dipl.-Ing Karl Ludwig Kimmig, Dr.-Ing. Reinhard Berger, Dr.rer.nat Julien Boeuf, LUK GmbH & Co. oHG, Bühl: Systemauslegung von Doppelkupplungen für große und kleine Fahrzeuge; VDI Berichte Nr.1987, 2007 SETC 2007 13 / 13 20076506(JSAE) 2007-32-0006(SAE)
