Patent Abstract:
A powertrain having a torque converter, friction input shifting clutches and shared driving gears has an axially compact design, packages a transmission pump between the torque converter and a clutch hub and achieves seven forward speed ratios utilizing four back-to-back synchronizers.

Full Description:
CROSS REFERENCE TO CO-PENDING APPLICATION 
     This patent application is a continuation-in-part of patent application Ser. No. 11/466,479, filed on Aug. 23, 2006, now U.S. Pat. No. 7,669,497. 
    
    
     TECHNICAL FIELD 
     The invention relates to a powertrain having a power source, a torque converter and a compact seven speed transmission with two friction shifting clutches. 
     BACKGROUND OF THE INVENTION 
     Dual clutch transmissions (DCTs) have been designed with friction launch clutches that connect the output of a torque converter to a lay shaft transmission. Dual clutch transmissions are known for their sporty, performance-oriented characteristics. Dual clutch transmissions typically exhibit good fuel economy due to good gear mesh efficiency and ratio selection flexibility. Synchronizers are typically used to engage gears with the countershafts to complete power flow from the engaged input clutch to an output shaft. The synchronizers have low spin losses; thus, overall operating efficiency is enhanced. However, dual clutch transmissions have some specific design considerations. For example, due to the high heat that can be generated during slip, the shifting clutches must be of a relatively large size. Shudder and oil life durability must also be addressed. Furthermore, cooling circuits for the friction shifting clutches are typically relatively complex due to the heat dissipation requirements of these large clutches. Finally, because lay shaft or countershaft transmissions typically have many sets of axially-aligned, intermeshing gears, the overall axial length of countershaft transmissions may limit there use in some vehicle designs. 
     SUMMARY OF THE INVENTION 
     A powertrain having a torque converter and dual shifting friction clutches connectable to first and second concentric intermediate shafts combines the smoothness and ratio-boosting effects of a torque converter with the low spin losses associated with synchronizers used in dual clutch designs, while preferably providing seven fixed forward speed ratios in an axially compact design. Several aspects of the powertrain contribute to the minimization of axial length. For example, driving gears connected for common rotation with the intermediate shafts intermesh with driven input gears connectable for rotation with each of the respective countershafts, thus functioning as shared driving gears. Preferably, back-to-back synchronizer pairs are supported on the countershafts between adjacent intermeshing aligned gear sets such that only four back-to-back synchronizer pairs are necessary and only four synchronizer selection devices are required to control engagement of the four pairs. Additionally, a parking gear is preferably connected for common rotation with one of the countershafts such that it is radially-aligned with an intermeshing output gear set. (A radial plane is in a plane encompassing radii of the driving or driven gears, perpendicular to the axis of rotation of the input member, output member, intermediate shafts and countershafts in the transmission. Accordingly, as used herein, components that are “radially-aligned” are aligned in a radial plane.) Furthermore, positioning of a transmission oil pump between the torque converter and the first and second friction shifting clutches allows a clutch hub supporting the friction shifting clutches to be configured with clutch hub passages for routing oil delivered from the pump to the friction shifting clutches. 
     Specifically, within the scope of the invention, the powertrain includes a power source and a torque converter that operatively connects the power source with a transmission input member. First and second friction shifting clutches are alternately selectively engagable to operatively connect the transmission input member with first and second concentric intermediate shafts, respectively. A first input driving gear is connected for common rotation with the first intermediate shaft and intermeshes with a first pair of driven input gears that are each connectable for common rotation to a different respective one of the countershafts to selectively transfer torque to the respective countershaft when the first friction input clutch is engaged. Furthermore, a second input driving gear is connected for common rotation with the second intermediate shaft and intermeshes with a second pair of input gears each connectable for common rotation with a different respective one of the countershafts to selectively transfer torque to the respective countershaft when the second friction input clutch is engaged. Thus, when torque is provided through either of the friction clutches to the countershafts, shared driving gears on each of the countershafts transfer the torque to one of the countershafts, depending on synchronizer engagements. Preferably, the first countershaft includes two input driving gears and the second countershaft has two other input driving gears so that four intermeshing aligned gear sets are used for input of torque from the first and second friction shifting clutches. Preferably two output gear sets are utilized, including a first output gear set that has a first output driving gear connected for common rotation with the first countershaft and a first output driven gear connected for common rotation with the output member that continuously intermeshes with the first output driving gear. Similarly, the second output gear set has a second output driving gear connected for common rotation with a second countershaft and a second output driven gear connected for common rotation with the output member that continuously intermeshes with the second output driving gear. Accordingly, the input and output driving gears and the input and output driven gears thereby form six sets of intermeshing aligned gears. By utilizing the four back-to-back synchronizers and selectively engaging the friction shifting clutches, seven forward speed ratios and a reverse speed ratio are achieved. 
     Unique packaging of the transmission oil pump allows simplified routing of clutch and lubrication oil and compact piloting of a clutch hub that supports the dual friction shifting clutches. Specifically, the clutch hub is configured with clutch hub passages for routing oil delivered from the transmission oil pump to the first and second friction shifting clutches. Preferably, a stationary clutch hub support member at least partially supports the clutch hub and is configured with clutch hub support member passages that are in fluid communication with the clutch hub passages so that oil may be routed from the transmission oil pump to the friction shifting clutches through the stationary clutch hub support member. The transmission oil pump is preferably radially-inward of and partially surrounded by the stationary clutch hub support member. Preferably, a stator support shaft connects a stator portion of the torque converter with the clutch hub support member. The oil pump is positioned radially-outward of the stator support shaft. Preferably the stator support shaft is configured with stator support shaft passages that are in fluid communication with the oil pump to route oil delivered from the oil pump to the torque converter. 
     The above features and advantages and other features and advantages of the present invention are readily apparent from the following detailed description of the best mode for carrying out the invention when taken in connection with the accompanying drawings. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a schematic illustration of a first embodiment of a powertrain with an engine, a torque converter, a transmission with dual friction shifting clutches, and an oil pump positioned between the torque converter and the friction shifting clutches; 
         FIG. 2  is a table showing an engagement schedule of the friction shifting clutches and synchronizers in the powertrain of  FIG. 1  to achieve seven forward speed ratios and a reverse speed ratio; and 
         FIG. 3  is a partial schematic fragmentary illustration of the torque converter, oil pump and friction shifting clutches of the powertrain of  FIG. 1 , showing a clutch hub, a clutch hub support member and a stator shaft support member that enable routing of oil from the oil pump to the torque converter and to the friction shifting clutches. 
     
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     Referring to the drawings, wherein like reference numbers represent the same or corresponding parts throughout the several views, there is shown in  FIG. 1  a powertrain  10  for a vehicle (not shown). The powertrain  10  includes a power source or engine  14 , a torque converter  16  and a transmission  18 . The torque converter  16  includes a turbine portion  20 , a pump portion  22 , and a stator portion  24 . An engine output shaft  23  is connected for rotation with a hub member  25  that is connected to the pump portion  22 . The turbine portion  20  is connected with a transmission input member  26 . A fluid coupling between the pump portion  22  and the turbine portion  20  thus operatively connects the engine  14  with the transmission input member  26 . The transmission input member  26  is preferably in the nature of a shaft. Selective engagement of a torque converter clutch  28  allows the engine  14  to be directly connected with the input member  26 , bypassing the torque converter  16 . Preferably, the torque converter clutch  28  is electronically controlled and may be enhanced with a plurality of clutch plates to provide a large clutch torque capacity, thus making the converter clutch  28  able to transmit a large amount of torque. The stator portion  24  is grounded to a stationary member, such as the transmission housing  30 , through a typical one-way clutch  32 . A damper  34  is operatively connected to the torque converter clutch  28  for absorbing vibration. A transmission oil pump  36  is operatively connected for rotation with the pump portion  22 . Support for the pump  36  and fluid communication from the pump  36  to the transmission  18  and to the torque converter  16  will be discussed hereinafter. Briefly, a stator support shaft  38  supports the stator and is located radially-inward of the pump  36 . The stator support shaft  38  operatively supports the stator portion  24  and is connected with a stationary clutch hub support member  40  that is grounded to the transmission housing  30 . A clutch hub  42  operatively connects the input member  26  with first and second concentric intermediate shafts  44 ,  46  by selective engagement of first and second friction shifting clutches CO and CE, respectively. 
     The transmission  18  further includes a first countershaft  50  and a second counter shaft  52  which are axially-spaced from and generally parallel with the intermediate shafts  44  and  46 . 
     Six aligned, intermeshing gear sets are utilized to transfer torque from the intermediate shafts  44 ,  46  via the countershafts  50 ,  52  to an output member  56  (preferably in the form of a shaft) to establish multiple speed ratios between the input member  26  and the output member  56 . The output member  56  is connected to a final drive mechanism  58  which may be connected to vehicle wheels (not shown). 
     A first intermeshing gear set includes gears  60 ,  62  and  64 . The gear  60  is a shared input driving gear that is connected for common rotation with the intermediate shaft  44  and continuously intermeshes with both gears  62  and  64 . The gear  62  is rotatable about the first countershaft  50  and is selectively connectable therewith. The gear  64  is rotatable about the second countershaft  52  and is selectively connectable therewith. 
     A second intermeshing gear set includes gears  66 ,  68  and  70 . The gear  66  is connected for common rotation with the intermediate shaft  44  and continuously intermeshes with both gears  68  and  70 . The gear  68  is rotatable about and selectively connectable with the countershaft  50 . The gear  70  is rotatable about and selectively connectable with the second countershaft  52 . 
     A third intermeshing gear set includes gears  72 ,  74  and  76 . The gear  72  is connected for common rotation with the intermediate shaft  46  and continuously intermeshes with both gears  74  and  76 . The gear  74  is rotatable about and selectively connectable for rotation with the first countershaft  50 . The gear  76  is rotatable about and selectively connectable for rotation with the second countershaft  52 . 
     A fourth intermeshing gear set includes gears  78 ,  80 ,  82  and a gear  84 . The gear  78  is connected for common rotation with the intermediate shaft  46 . The gear  78  continuously intermeshes with both the gear  80  and the gear  82 . The gear  82  continuously intermeshes with the gear  84 . The gear  80  is rotatable about and selectively connectable for common rotation with the first countershaft  50 . The gear  82  is an idler gear supported on a separate axis I. The gear  84  is rotatable about and selectively connectable with the second countershaft  52 . The gears  60 ,  66 ,  72  and  78  are referred to herein as input driving gears. The gears  62 ,  64 ,  68 ,  70 ,  74 ,  76 ,  80 , and  84  are referred to herein as input driven gears. 
     The transmission  18  includes a fifth intermeshing, aligned gear set that includes a gear  86  and a gear  88 . The gear  86  is connected for common rotation with the second countershaft  52  and continuously intermeshes with the gear  88  which is connected for common rotation with the output member  56 . A sixth intermeshing, aligned gear set includes gear a  90  which is connected for common rotation with the first countershaft  50  and a gear  92  which is connected for common rotation with the output member  56  and continuously intermeshes with the gear  90 . The gears  86  and  90  are referred to herein as output driving gears and the gears  88  and  92  are referred to herein as output driven gears. The intermeshing, aligned gear set including the gears  60 ,  62  and  64  may be referred to as a first input gear set. The intermeshing, aligned gear set including the gears  66 ,  68  and  70  may be referred to as a second input gear set. The intermeshing, aligned gear set including the gears  72 ,  74  and  76  may be referred to a third input gear set. The intermeshing, aligned gear set including the gears  78 ,  80 ,  82  and  84  may be referred to herein as a fourth intermeshing gear set. The gear set including the gears  90 ,  92  may be referred to as a first output gear set and the gear set including the gears  88  and  86  may be referred to as a second output gear set. The four input gear sets utilize shared input driving gears: the gear  60 , the gear  66 , the gear  72  and the gear  78 . Each of the input driving gears  60 ,  66 ,  72  and  78  intermesh with gears that are connectable for rotation with each of the countershafts  50 ,  52 . Thus, the shared input driving gears are each able to transfer torque to both of the countershafts  50 ,  52 , the countershaft to which torque is transferred being dependent on the engagement of synchronizers, as will be described herein. 
     A parking gear  94  is radially-aligned with the output gear set  90 ,  92 . Thus, the parking gear  94  is situated in what may otherwise be unused, empty space and does not require any addition to the axial length of the transmission  18 . 
     The transmission  18  includes four pairs of back-to-back synchronizers: A, B, C and D. The back-to-back synchronizer pair A includes synchronizer A 1  and a synchronizer A 3 . The synchronizer A 1  is selectively engagable to connect the gear  64  for common rotation with the second countershaft  52 . The synchronizer A 3  is selectively engagable to connect the gear  70  for common rotation with the second countershaft  52 . A single synchronizer selection device  96  is operable to engage either the synchronizer A 1  or A 3 . Specifically, a synchronizer selection device  96  is shiftable to the left to engage the synchronizer A 1  and shiftable to the right the engage the synchronizer A 3 . 
     The back-to-back synchronizer pair B includes a synchronizer B 2  and a synchronizer BR. The synchronizer B 2  is selectively engagable to connect the gear  76  for common rotation with the second countershaft  52 . The synchronizer BR is selectively engagable to connect the gear  84  for common rotation with the second countershaft  52 . A single synchronizer selection device  98  is operable to control engagement of both the synchronizer B 2  and the synchronizer BR. Specifically, the synchronizer selection device  98  is shiftable to the left to engage the synchronizer B 2  and shiftable to the right to engage the synchronizer BR. 
     The back-to-back synchronizer pair C includes a synchronizer C 5  and a synchronizer C 7 . The synchronizer C 5  is selectively engagable to connect the gear  62  for common rotation with the first countershaft  50 . The synchronizer C 7  is selectively engageable to connect the gear  68  for common rotation with the first countershaft  50 . A single synchronizer selection device  100  is operable to control engagement of both the synchronizers C 5  and C 7 . Specifically, the synchronizer selection device  100  is shiftable to the left to engage the synchronizer C 5  and shiftable to the right the engage the synchronizer C 7 . 
     The back-to-back synchronizer pair D includes a synchronizer D 6  and a synchronizer D 4 . The synchronizer D 6  is selectively engagable to connect the gear  74  for common rotation with the first countershaft  50 . The synchronizer D 4  is selectively the synchronizers D 6  and D 4 . Specifically, the synchronizer selection device  102  is shiftable to the left to engage the synchronizer D 6  and shiftable to the right to engage the synchronizer D 4 . 
     Referring to  FIG. 2 , the engagement schedule of the shifting friction clutches CO and CE as well as the synchronizers is shown. As indicated in  FIG. 2 , seven forward gears (i.e., speed ratios) and a reverse gear (i.e., speed ratio) are achieved. Those skilled in the art will recognize that the gears shown in  FIG. 1  may be designed with various tooth counts that, when the clutches and synchronizers are engaged according to the truth table of  FIG. 2 , will result in seven forward speed ratios and a reverse speed ratio corresponding with the seven forward gears and the reverse gear. For example, the following gear ratios may have different numerical values: First gear: tooth count gear  64 /tooth count gear  60 ; Second gear: tooth count gear  76 /tooth count gear  72 ; Third gear tooth count gear  70 /tooth count gear  66 ; Fourth gear: tooth count gear  80 /tooth count gear  78 ; Fifth gear: tooth count gear  62 /tooth count gear  60 ; Sixth gear: tooth count gear  74 /tooth count gear  72 ; Seventh gear: tooth count gear  68 /tooth count gear  66 ; Reverse gear: tooth count gear  84 /tooth count gear  78 ; First countershaft ( 50 ) output: tooth count gear  90 /tooth count gear  92 . Second countershaft ( 52 ) output: tooth count gear  86 /tooth count gear  88 ; The tooth counts are selected to achieve desired speed ratios, ratio steps, and overall speed ratio. 
     To establish the reverse speed ratio, the clutch CE and the synchronizer BR are engaged. By engagement of the clutch CE torque is transferred from the input member  26  to the intermediate shaft  46 . By engagement of the synchronizer BR torque is transferred from the intermediate shaft  44  to the second countershaft  52  via intermeshing gears  78 ,  82  and  84 , with the gear  82  acting as an idler gear so that the gear  84  rotates in the same direction as the gear  78 . Torque is transferred from the second countershaft  52  to the output member  56  via the intermeshing gears  86  and  88 . 
     To establish the first forward speed ratio, the input friction clutch CO and the synchronizer A 1  are engaged. Torque is transferred from the input member  26  to the intermediate shaft  44  via engagement of the input shifting friction clutch CO. Torque is transferred from the intermediate shaft  44  to the second countershaft  52  by engagement of synchronizer A 1  through the intermeshing gears  60  and  62 . Torque is transferred from the second countershaft  52  to the output member  56  via intermeshing gears  86  and  88 . 
     To establish the second forward speed ratio, input friction shifting clutch CE and the synchronizer B 2  are engaged. Torque is transferred from the input member  26  to the intermediate shaft  46  via engagement of clutch CE. Torque is then transferred from the intermediate shaft  46  to the second countershaft  52  via engagement of the synchronizer B 2  through the intermeshing gears  72  and  76 . Torque is transferred from the second countershaft  52  to the output member  56  via the intermeshing gears  86  and  88 . 
     To establish the third forward speed ratio, the input friction shifting clutch CO and the synchronizer A 3  are engaged. Torque is transferred from the input member  26  to the intermediate shaft  44  via engagement of the clutch CO. Torque is transferred from the intermediate shaft  44  to the second countershaft  52  via engagement of the synchronizer A 3  through the intermeshing gears  66  and  70 . Torque is transferred from the second countershaft  52  to the output member  56  via the intermeshing gears  86  and  88 . 
     It will thus be appreciated that all of the four lower speed (higher numerical reduction ratios) gears (1 st , 2 nd , 3 rd  and Reverse) are disposed on the second countershaft  52 . 
     To establish the fourth forward speed ratio, the input friction shifting clutch CE and the synchronizer D 4  are engaged. Torque is transferred from the input member  26  to the intermediate shaft  46  via engagement of the clutch CE. Torque is transferred from the intermediate shaft  46  to the first countershaft  50  via the engagement of synchronizer D 4  through the intermeshing gears  78  and  80 . Torque is transferred from the first countershaft  50  to the output member  56  through the intermeshing gears  90  and  92 . 
     A fifth forward speed ratio is established by engagement of the input friction shifting clutch CO and the synchronizer C 5 . Torque is transferred from the input member  26  to the intermediate shaft  44  via engagement of the clutch CO. Torque is transferred from the intermediate shaft  44  to the first countershaft  50  via engagement of the synchronizer C 5  through the intermeshing gears  60  and  62 . Torque is transferred from the first countershaft  50  to the output member  56  through the intermeshing gears  90  and  92 . 
     A sixth forward speed ratio is established by engagement of the input friction shifting clutch CE and the synchronizer D 6 . Torque is transferred from the input member  26  to the intermediate shaft  46  via engagement of the clutch CE. Torque is transferred from the intermediate shaft  46  to the first countershaft  50  via engagement of the synchronizer D 6  through the intermeshing gears  72  and  74 . Torque is transferred from the first countershaft  50  to the output member  56  through the intermeshing gears  90  and  92 . 
     A seventh forward speed ratio is established via engagement of the input friction shifting clutch CO and the synchronizer C 7 . Torque is transferred from the input member  26  to the intermediate shaft  44  via engagement of the input friction shifting clutch CO. Torque is transferred from the intermediate shaft  44  to the first countershaft  50  via engagement of synchronizer C 7  through intermeshing gears  66  and  68 . Torque is transferred from the first countershaft  50  to the output member  56  through the intermeshing gears  90  and  92 . 
     It will thus also be appreciated that all of the four higher speed (lower numerical reduction ratio) gears (4 th , 5 th , 6 th  and 7 th ) are disposed on the first countershaft  50 . This arrangement of higher speed gears on the first countershaft  50  and lower speed gears on the second countershaft  52  is facilitated by the difference in center to center distances between the axes of the first and second countershafts  50 ,  52  and the common axis of the two intermediate shafts  44  and  46 . In  FIG. 1 , this greater distance “X” between the axis of the second countershaft  52  and the common axis of the two intermediate shafts  44  and  46  is illustrated and contrasted with the smaller distance “Y” between the axis of the first countershaft  50  and the common axis of the two intermediate shafts  44  and  46 . This arrangement also assists maintenance of reasonable torque and speed ratios on the ratio gears. Distributing the overall gear state ratio, i.e., that overall gear ratio between the input member  26  (or the intermediate shafts  44  and  46 ) and the output member  56 , between the ratio and transfer gearsets allows for smaller gear diameter packaging than an arrangement that develops all the gear state ratios in one gearset. 
     Referring now to  FIG. 3 , the torque converter  16 , the pump  36 , the stationary clutch hub support member  40 , the clutch hub  42  and the stator support shaft  38  are shown in greater detail. The transmission oil pump  36  is operatively connected to the pump portion  22  of the torque converter  16  and to the engine output member  23  via pump portion hub member  25 . The stator portion  24  is operatively connected to the stator support shaft  38  which is connected with the stationary clutch hub support member  40 . The transmission housing  30  is bolted or otherwise connected with the stationary member clutch hub support member  40 . Thus, the pump  36  is radially-outward of the stator support shaft  38  and is radially-inward of and supported by the stationary clutch hub support member  40 . 
     The turbine portion  20  is operatively connected for rotation with the input member  26 . The input member  26  in turn is operatively connected for rotation with the clutch hub  42 . The clutch hub  42  supports a portion of the friction input shifting clutches CO and CE. The friction input shifting clutch CO has another portion operatively connected for rotation with the intermediate shaft  44 . The friction input shifting clutch CE has another portion operatively connected for rotation with intermediate shaft  46 . Engagement of the clutch CO connects the input member  26  and the clutch hub  42  for rotation with the intermediate shaft  44 . Engagement of the clutch CE connects the input member  26  and the clutch hub  42  for rotation with the intermediate shaft  46 . 
     Transmission oil from pump the  36  is routed to the torque converter  16  and to the friction shifting clutches CO and CE. Torque converter clutch apply oil is routed from the pump  36  to the torque converter  16  via a sleeve passage  101  which is an annular passage between the pump portion sleeve shaft  103  and the stator support shaft  38 . The pump portion sleeve shaft  103  operatively connects the pump portion  22  to the pump  36 . Torque converter clutch release oil is also routed to the torque converter  16  through a stator support shaft passage  104  in the stator support shaft  38 . The oil makes its way from the pump  36  to the stator support shaft passage  104  through a horizontally running crevice  108  between the stationary clutch hub support member  40  and the stator support shaft  38 . Lubrication oil for gears supported on the intermediate shafts  44  and  46  is delivered via a lubrication passage  106 . 
     The stationary clutch hub support member  40  also has clutch hub support member passages  110  and  112  formed therein to direct clutch apply oil from the pump  36  via the horizontal crevice  108  to clutch hub passages  114  and  116 . The clutch hub  42  also has additional passages (not shown) in fluid communication with one another for delivering dam oil to the clutches CO and CE. These additional dam oil passages are located in a different radial plane than the passages  114  and  116 . The clutch hub passage  114  directs oil to the input friction clutch CE and the clutch hub passage  116  directs oil to the input friction clutch CO. 
     The passage  118  is an oil supply to the pump  36  in the housing  30  and is fluidly communicable with a filter (not shown) through which oil is supplied to the pump  36 . Multiple valves, represented by a valve  120 , control oil flow through passages  101 ,  104 ,  106 ,  108 ,  110 ,  112 ,  114  and  116  and communicate with one or more valve bodies (not shown). 
     Thus, the unique packaging of the oil pump  36  adjacent the clutch hub  42  enables a relatively simple cooling circuit for the friction shifting clutches CO and CE through the stationary clutch hub support member passages  110 ,  112  and the clutch hub passages  114 ,  116 . During assembly of the transmission  18 , the clutch hub  42  is piloted over the stationary hub support member  40 , which serves to partially support both the clutch hub  42  and the pump  36 . 
     While the best modes for carrying out the invention have been described in detail, those familiar with the art to which this invention relates will recognize various alternative designs and embodiments for practicing the invention within the scope of the appended claims.

Technology Classification (CPC): 8