Abstract:
A vehicle dual clutch transmission is provided which includes a first clutch housing driven along a first axis extending generally aligned with the output axis of an engine. A second clutch housing is provided which is chain driven from the first clutch housing and extends along a second generally parallel axis. First and second input shafts are respectively driven by the first and second clutches. An output shaft or shafts is provided that is driven by the input shafts and has selective meshed torsional force transferring gear contact therewith. A differential input gear is driven by the output shaft(s).

Description:
FIELD OF THE INVENTION 
       [0001]    The present invention relates to dual clutch transmissions for automotive vehicles. 
       BACKGROUND OF THE INVENTION 
       [0002]    Generally speaking, land vehicles require a powertrain consisting of three basic components. These components include a power plant (such as an internal combustion engine), a power transmission, and wheels. The power transmission component is typically referred to simply as the “transmission.” Engine torque and speed are converted in the transmission in accordance with the tractive-power demand of the vehicle. Presently, there are two typical transmissions widely available for use in conventional motor vehicles. The first and oldest type is the manually operated transmission. These transmissions include a foot-operated start-up or launch clutch that engages and disengages the driveline with the power plant and a gearshift lever to selectively change the gear ratios within the transmission. When driving a vehicle having a manual transmission, the driver must coordinate the operation of the clutch pedal, the gearshift lever, and the accelerator pedal to achieve a smooth and efficient shift from one gear to the next. The structure of a manual transmission is simple and robust and provides good fuel economy by having a direct power connection from the engine to the final drive wheels of the vehicle. Additionally, since the operator is given complete control over the timing of the shifts, the operator is able to dynamically adjust the shifting process so that the vehicle can be driven most efficiently. One disadvantage of the manual transmission is that there is an interruption in the drive connection during gear shifting. This results in losses in efficiency. In addition, there is a great deal of physical interaction required on the part of the operator to shift gears in a vehicle that employs a manual transmission. 
         [0003]    The second and newer choice for the transmission of power in a conventional motor vehicle is an automatic transmission. Automatic transmissions offer ease of operation. The driver of a vehicle having an automatic transmission is not required to use both hands, one for the steering wheel and one for the gearshift, and both feet, one for the clutch and one for the accelerator and brake pedal in order to safely operate the vehicle. In addition, an automatic transmission provides greater convenience in stop and go situations, because the driver is not concerned about continuously shifting gears to adjust to the ever-changing speed of traffic. Although conventional automatic transmissions avoid an interruption in the drive connection during gear shifting, they suffer from the disadvantage of reduced efficiency because of the need for hydrokinetic devices, such as torque converters, interposed between the output of the engine and the input of the transmission for transferring kinetic energy therebetween. In addition, automatic transmissions are typically more mechanically complex and therefore more expensive than manual transmissions. 
         [0004]    For example, torque converters typically include impeller assemblies that are operatively connected for rotation with the torque input from an internal combustion engine, a turbine assembly that is fluidly connected in driven relationship with the impeller assembly and a stator or reactor assembly. These assemblies together form a substantially toroidal flow passage for kinetic fluid in the torque converter. Each assembly includes a plurality of blades or vanes that act to convert mechanical energy to hydrokinetic energy and back to mechanical energy. The stator assembly of a conventional torque converter is locked against rotation in one direction but is free to spin about an axis in the direction of rotation of the impeller assembly and turbine assembly. When the stator assembly is locked against rotation, the torque is multiplied by the torque converter. During torque multiplication, the output torque is greater than the input torque for the torque converter. However, when there is no torque multiplication, the torque converter becomes a fluid coupling. Fluid couplings have inherent slip. Torque converter slip exists when the speed ratio is less than 1.0 (RPM input&gt;than RPM output of the torque converter). The inherent slip reduces the efficiency of the torque converter. 
         [0005]    While torque converters provide a smooth coupling between the engine and the transmission, the slippage of the torque converter results in a parasitic loss, thereby decreasing the efficiency of the entire powertrain. Further, the torque converter itself requires pressurized hydraulic fluid in addition to any pressurized fluid requirements for the actuation of the gear shifting operations. This means that an automatic transmission must have a large capacity pump to provide the necessary hydraulic pressure for both converter engagement and shift changes. The power required to drive the pump and pressurize the fluid introduces additional parasitic losses of efficiency in the automatic transmission. 
         [0006]    In an ongoing attempt to provide a vehicle transmission that has the advantages of both types of transmissions with fewer of the drawbacks, combinations of the traditional “manual” and “automatic” transmissions have evolved. A type of combination type transmission is commonly referred to as a dual clutch transmission. 
         [0007]    Examples of dual clutch transmissions and control methods can be found by a review of U.S. Pat. Nos. and Patent Application Publications 5,711,409; 6,966,989; 6,887,184; 6,909,955; 2006/0101933A1; and 2006/0207655A1 commonly assigned. 
         [0008]    Dual clutch transmissions can be utilized in front wheel drive engines. When utilizing a dual clutch transmission in a transverse mounted engine, it is desirable to make the width of the transmission as short as possible. An example of a dual clutch transmission for a front wheel drive vehicle is shown in Patent Application 2008/004288. It is desirable to provide a dual clutch transmission suitable for a transverse mounted front wheel drive vehicle or other vehicle which is axially shorter than that described in Patent Application 2008/004288. 
       SUMMARY OF THE INVENTION 
       [0009]    To meet the above noted and other desires, a revelation of the present invention is brought forth. 
         [0010]    In a preferred embodiment, the present invention brings forth a vehicle dual clutch transmission which includes a first clutch housing driven along a first axis extending generally aligned with the output axis of an engine. A second, clutch housing is provided which is chain driven from the first clutch housing and extends along a second generally parallel axis. First and second input shafts are respectively driven by the first and second clutches. An output shaft or shafts is provided that is driven by the input shafts and has selective meshed torsional force transferring gear contact therewith. A differential input gear is driven by the output shafts. 
         [0011]    Further areas of applicability of the present invention will become apparent from the detailed description provided hereinafter. It should be understood that the detailed description and specific examples, while indicating the preferred embodiment of the invention, are intended for purposes of illustration only and are not intended to limit the scope of the invention. Figs. 
     
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         [0012]    The present invention will become more fully understood from the detailed description and the accompanying drawings, wherein: 
           [0013]      FIGS. 1A and 1B  are sectional laid out views of a six speed version vehicle dual clutch transmission according to the present invention; 
           [0014]      FIG. 2  is a front schematic view partially sectioned of the transmission shown in  FIGS. 1A and 1B  looking from the vehicle engine towards the transmission; and 
           [0015]      FIGS. 3A and 3B  are sectional views similar to that of  1 A and  1 B of a five speed version of a vehicle dual clutch transmission according to the present invention. 
       
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
       [0016]    The following description of the preferred embodiment(s) is merely exemplary in nature and is in no way intended to limit the invention, its application, or uses. 
         [0017]    Referring to  FIGS. 1 and 2 , a transverse mounted engine  10  for a front wheel drive vehicle powers a six speed dual clutch transmission  7  of the present invention. The engine  10  has an output shaft  12 . The output shaft  12  is torsionally connected with a damper  14 . The damper  14  is torsionally connected with a first clutch housing  16 . Clutch housing  16  is operatively associated with a first clutch friction pack  18 . Friction pack  18  is actuated by a clutch piston  17 . Friction pack  18  is operatively associated with a hub  20 . The hub  20  is torsionally connected with a first input shaft  24 . First input shaft  24  provides the odd gear ratios for the transmission  7 . First input shaft  24  mounts an integral or fixably connected first gear  26 , a third gear  28  and a fifth gear  30 . A fifth/third synchronizer  31  is provided to selectively torsionally engage the gears  30  or  28  to the shaft  24 . A mid-portion of the first input shaft  24  is rotatably mounted by a bearing  32  and a rearward end by a bearing  34 . Fifth gear and third gear  30  and  28  are both rotatably mounted on needle bearings  36  and  38  respectively. 
         [0018]    Selectively meshing in a torsional force transferring manner with the gears of the input shaft  24  is first output shaft  40 . First output shaft  40  is mounted by a front bearing  42  and a rear bearing  44 . The first output shaft  40  has a gear  48  for meshing with the fifth gear  30 , and a gear  50  for meshing with the third gear  28 . Both gears  48  and  50  are spline connected to the first output shaft  40 . The first output shaft  40  also mounts a one way clutch  52 . A parking gear  89  used for the parking function is connected with the gear  48 . In an alternative embodiment, not shown, the parking gear is placed within the differential. The one way clutch  52  mounts a gear  54  to mesh with first gear  26  to provide the first gear ratio. The first output shaft  40  has an integrally formed output pinion  53  to drive a differential ring input gear  56 . 
         [0019]    The clutch housing  16  is also fixably engaged with a first sprocket  60 . The sprocket  60  also has attached a hydraulic pump gear  61  that drives a hydraulic pump  63 . To conserve axial space, sprocket  60  axially overlaps the clutch piston  17 . The sprocket  60  drives a chain  62 . Chain  62  drives a second sprocket  64  ( FIG. 1B ). Sprocket  64  has a larger diameter than the sprocket  60 . Sprocket  64  is fixably connected with a second clutch housing  68 . Clutch housing  68  rotates on a second axis  70  which is parallel with an axis  72  that the first clutch housing  16  and the engine output shaft  12  rotate about. A second clutch friction pack  74  is also operatively associated with the second clutch housing  68 . Second friction pack  74  is operatively associated with a hub  76  which powers a second input shaft  80 . The second input shaft  80  is rotatably mounted by front bearing  82  and a rear bearing  84 . Rotatably mounted on the second input shaft  80  on needle bearings are sixth gear  86 , fourth gear  88  and second gear  90 . To selectively torsionally connect the gears,  86 ,  88  and  90  to the shaft a six/four synchronizer  92  and a second/reverse synchronizer  94  are provided. A needle bearing mounted reverse driving gear  95  meshes with an idler shaft reverse driven gear  98 . 
         [0020]    To mesh with the gears of the second input shaft  80  in a torsional force transferring manner, there is a second output shaft  100 . The second output shaft  100  has spline connected gears  102 ,  104  and  106  to selectively torsional force transferring manner mesh with sixth, fourth and second gears  86 ,  88  and  90  respectively. The gear  106 , when the transmission  7  is placed in reverse, meshes with reverse idler shaft driving gear  110  to provide the reverse gear function. Reverse idler shaft driving gear  110  and reverse idler shaft driven gear  98  are both fixably connected on reverse idler shaft  111 . A pinion (final drive) gear  108  meshes with the ring gear  56  (shown in both  FIGS. 1A and 1B ) of differential  116  which in turn powers front wheel drive shafts  118  and  120 . 
         [0021]    In operation, transmission  7  is powered from the output shaft  12  of the engine  10 . Rotation of the engine output shaft  12  causes rotation of the damper  14  which dampens torsional vibrations provided by the reciprocating piston nature of the engine  10 . The damper  14  rotates clutch housing  16  and sprocket  60 . The sprocket housing  60  drives chain  62  which drives the larger sprocket  64  associated with the second clutch housing  68 . The friction pack  18  is engaged by clutch piston  17  therefore torsionally connecting the first clutch housing  16  with the hub  20  and the first input shaft  24 . A first gear is provided by gear  26 . Gear  26  engages a corresponding gear  54  of the first output shaft  40  (via one way clutch  52 ). While the above is happening, according to parameters of the electronic controller (not shown), the friction pack  74  is disengaged and the synchronizer  94  is actuated leftward as shown in  FIG. 1  to torsionally connect the second gear  90  with the second input shaft  80 . When the controller signals for the transmission to engage in second gear, the friction pack  18  is released and the friction pack  74  is engaged. The transmission  7  is now in second gear. The above noted sequence of operation is commonly referred to as a pre-selecting sequence. In a non pre-selecting sequence, the second synchronizer  96  does not connect the second gear  90  to the second output shaft  80  until friction pack  18  is released. Although the remainder of the operation of the transmission  7  is described in a pre-selection mode, it is apparent to those skilled in the art that a combination of both pre-selection and non pre-selection modes can be utilized. After the shift is complete to second gear, to prepare for the shift to third gear, synchronizer  94  is brought to a neutral position allowing the second gear  90  to be torsionally disengaged from the second input shaft  80 . Simultaneously, synchronizer  31  is moved rightward engaging third gear  28  with the first input shaft  24 . This movement is primarily completed while the friction pack  18  is disengaged. Upon re-engagement of the clutch pack  18 , to connect the hub  20  with the clutch housing  16 , the rotation of the first input shaft  24  moves third gear  28 . Third gear  28  now has torsional force transferring mesh contact with first output shaft  50  which in turn causes pinion  53  to rotate differential gear. To disengage the third gear  28 , the fifth/third synchronizer  40  is moved leftward simultaneously the friction pack  74  is engaged and synchronizer  92  is engaging fourth gear  88  with second input shaft  80 . As mentioned previously, to switch to the fifth gear  130 , the friction pack  74  is disengaged. The friction pack  18  is engaged turning the first input shaft  24  and rotating the fifth gear  50 . The fifth gear  50  meshes with gear  48  of the first output shaft which in turn turns the ring gear  56  powering front wheel drive axles  118  and  120  through the differential  116 . For a switch to the sixth gear, synchronizer  92  is moved leftward engaging gear  86  with the second input shaft  80  while the friction pack  18  is disengaged. Synchronizer  31  is moved rightward disengaging the fifth gear  30  from the first input shaft  24 . After engagement with the shaft  80  and upon engagement of the friction pack  74 , sixth gear  86  has torsional force transferring mesh contact with gear  102  of the second output shaft  100 . Second output shaft  100  pinion gear  108  torsionally meshes with input gear  56  of the differential  116 . 
         [0022]    The sprocket  64  on the even gear input shaft  80  is larger so that the reverse gear  95  can be larger on the even input shaft  80  with second reverse synchronizer  94 . The size of the reverse gear  95  does not get too small, therefore the gear ratio for reverse gear  95  can be larger for packaging considerations. Another advantage of transmission  7  design is that it has only two input axes  72  and  70 . Therefore, the chain  62  does not need as many tensioners to ensure the maximum amount of torsional input for each sprocket. There is less chance of chain jump and also better torsional force transferring efficiency. 
         [0023]    By using a reverse idler driving gear  110  that reaches back to the second output gear  106  for reverse hook-up, there is a reduction in the length of the input shaft  80  in the axial direction. The above allows placement or the input shaft reverse driving gear  95  in the same axially bisecting transverse plane as the final drive gear mesh between the second output shaft pinion  108  and the differential input gear  56 . Normally, transmissions are not able to utilize the radial space on the input shaft  80  in the same axially bisecting transverse plane across from the final drive pinion (because the final drive pinion is in the way of the output gear in such a position), but since with the present design of the transmission  7 , the reverse idler shaft driving gear  110  inputs into the second output gear  106 , transmission  7  is able to put input shaft reverse driving gear  95  in the shown axial position. 
         [0024]    The step reverse idler shaft reverse gears  98 ,  110  also provides opportunity increase the gear ratio which to make it generally equal to the first gear ratio is important for launch in reverse. 
         [0025]    Another advantage of the present invention is that all of the synchronizers  31 ,  92  and  94  are on the input shafts  24 ,  80  and therefore avoids a need for large diameter or multi-cone synchronizers on the output shafts (to handle a larger rotational inertia of both the input and output shaft)  40  and  100  can be avoided. 
         [0026]      FIG. 2  is a view, as mentioned previously, of transmission  7  looking from the engine  10 . The view of  FIG. 2  is a view from right to left in  FIGS. 1A and 1B . Referring to  FIG. 2  in particular, a shift fork  124  is slidably or axially mounted on a rail  126 . The shift fork  124  is used to control movement of the fifth/third synchronizer  31 . The shift fork is connected with a bracket  128 . Bracket  128  is connected with a piston  130 . The piston  130  has its extreme end slidably mounted within opposing cylinders (not shown). The opposing cylinders can be filled or evacuated by hydraulic solenoids (not shown) control by an electronic transmission control module (not shown). These solenoids attach to a valve body  132 . Another shift fork  134  is slides on a rail  136  and is connected with a bracket  138 . Bracket  138  is connected with a piston  140  which has its opposing ends slidably mounted within opposing cylinders. Fluid lines within the valve block  132  are filled or relieved to control the movement of the shift fork  134 . Another shift fork  144  is mounted on a rail  146  and is connected with a bracket  148  which is connected with a piston  150 . Shift fork  144  is provided for controlling sixth/fourth synchronizer  92  and shift fork  134  is provided for controlling synchronizer  94 . Advantage of the inventive transmission  7  is that the valve block  132  bridges over a valley between the first and second input shafts  24  and  80  and additionally, the shift rails  146 ,  136  and  126  are all positioned within the valley adding to the compactness of the design of the transmission  7 . 
         [0027]    Referring to  FIGS. 3A and 3B  five speed transmission  207  according to the present invention is shown. Its operation is substantially similar to the six speed transmission  7  previously described. The five speed transmission  207  has a damper  14  connected with the output shaft  12  of an engine  10 . The damper  14  is connected with a clutch housing  202  which is operatively associated with a first friction pack  214  which when engaged torsionally connects the first clutch housing  202  with a hub  206  that powers a first input shaft  208 . First input shaft  208  that is on a common axis  210  with the engine output shaft  10  carries fourth and second gears  212  and  214  respectively. A reverse driving gear  216  is also provided which meshes with an idler shaft driven reverse gear  218 . Fourth gear  212  and second gear  214  mesh with gears  220  and  222  that are affixed with the output shaft  224 . Gears  212  and  214  are synchronized by fourth/second synchronizer  213 . Output shaft  224  has a geared section  226  which engages with a ring gear  228  of a differential  230  powering front axle  118  and  120 . 
         [0028]    Transmission  207  has an odd input shaft  260  which carries fifth gear  262 , third gear  264  and first gear  266 . The fifth and third gears  262  and  264  are synchronized by fifth/third synchronizer  268 . The shaft  260  rotates on an axis  270  generally parallel with axis  210  and is driven by a hub  272  operatively associated with a second clutch housing friction pack  274  which in turn is operatively associated with a second clutch housing  276  which is affixed with a sprocket  278  driven by chain  280  which is in turn driven by the first sprocket  203 . The configuration of transmission  207  has both of the output shafts  208  and  260  being selectively torsionally engaged with a common output shaft  224 . The first gear  266  is operatively associated with a one way clutch  280  which is connected with the output shaft  224 . To engage with a third gear  264 , the output shaft has a gear  284 . Output shaft gear  220  engages fifth input gear  262 . The output shaft  224  also has a gear  290  which is used to engage the transmission  7  in park. 
         [0029]    The operation of transmission  207  is essentially similar to that of transmission  7  with the exception that there are only five forward gear ratios and that the even input shaft is aligned with the axis of the engine output shaft  12 . Additionally, the reverse idler shaft driving gear  218  can be selectively joined on idler shaft  250  by a synchronizer  252 . An idler shaft reverse driving gear  254  meshes with the second output gear  222  provided on the output shaft  224 . Again, the differences of diameter of the gears  218  and  254  increases the gear ratio allowing reverse gear to have high torque capabilities similar to that of the first gear. 
         [0030]    The description of the invention is merely exemplary in nature and, thus, variations that do not depart from the gist of the invention are intended to be within the scope of the invention. Such variations are not to be regarded as a departure from the spirit and scope of the invention.