Abstract:
A power transmission includes a rotary input member adapted to receive drive torque from a source of torque, a rotary output member and a torque transmitting mechanism selectively operable to transfer drive torque between the input member and the output member. The torque transmitting member includes a selectively operable one-way clutch including a rotatable member for switching the one-way clutch between a released mode and a locked mode. The torque transmitting mechanism also includes a friction clutch and an actuator. The friction clutch is operable to transmit torque to the rotatable member and control the mode of operation of the one-way clutch. The actuator is operable to apply a clutch and brake engagement force to the friction clutch.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application is a divisional of U.S. patent application Ser. No. 10/295,120 filed Nov. 15, 2002 now U.S. Pat. No. 6,827,664 that claims the benefit of U.S. Provisional Application Ser. No. 60/336,126, filed Nov. 15, 2001. The entire disclosure of Ser. No. 60/336,126 is incorporated by reference herein. 
    
    
     FIELD OF THE INVENTION 
     This invention relates to automatic shifting power transmissions and more particularly to power transmissions having a single torque transmitter operable to be selectively actuated to launch both forward and reverse drives. 
     BACKGROUND OF THE INVENTION 
     In automatic shifting power transmissions it is common practice to install a hydrodynamic fluid drive, such as a torque converter, between the power source (engine) and a multi-speed gear configuration, such as a planetary gear arrangement. The torque converter provides a torque multiplier between the engine and the gearing to improve the vehicle launch performance. The torque ratio of the torque converter is generally in the range of 1.60 to 3.3 depending on the particular application. As is well-known, the torque converter is a slipping device that has a high efficiency loss at vehicle launch. This loss decreases, but continues, as the torque converter approaches a 1.0 to 1.0 speed ratio at high speed and low torque. In recent times, a torque converter lock-up has been added to most transmissions to effectively remove the torque converter from the power path and thereby improve the overall efficiency of the transmission. 
     Other considerations have been given to improving the overall efficiency of the transmission. Specifically, newer automatic transmissions, with their higher number of gear ratios and higher overall ratio, have less need for torque converter multiplication. This reduced need for torque multiplication has lead to the use of a “friction launch” device in place of a torque converter within stepped gear automatic transmissions. For example, the use of a starting clutch in lieu of the torque converter has been proposed. The advent of electronic controls improves the operation of a starting clutch as a vehicle launch device. The starting clutch is, however, a rotating torque transmitting device which must still deal with all of the complexities associated with such a vehicle launch device. Thus, the control needs considerable accuracy to insure consistent fill times, and to compensate for variable fluid leaks at the rotating shaft seals. This requires accurate hydraulic flow volumes and pressure control over a wide range of operating requirements. Also the use of a starting clutch merely replaces one rotating mechanism with another, albeit a more efficient mechanism. There is only slight axial space saving and perhaps more complex control features. 
     As noted, if friction launch is applied to a conventional automatic transmission, the starting clutch must fulfill the same requirements as the torque converter. These requirements include, for example, shift and launch quality, NVH and driveline isolation, mass, peak acceleration, and durability. Many of these and other requirements apply to both forward and reverse launches. 
     The most common implementation of friction launch is simply to locate the starting clutch between the engine and the input to the transmission. This implementation has been used in production with manual transmissions, automated manual transmissions, and continuously variable transmissions. One advantage of using a stand-alone starting clutch assembly is that it can be designed as a drop-in replacement for the torque converter of an existing automatic transmission. Stand-alone starting clutches also have the ability to isolate the transmission from torsional engine inputs, which is an advantage compared to other embodiments which locate the starting clutch further downstream in the powerflow. Another advantage is that the starting clutch can be designed for a lower torque capacity than some of the other implementations. Another variation of a dedicated launch is to locate the launch clutch on the output of the transmission. One of the benefits of an output clutch is the ability to use the clutch as a “fuse” between the transmission and road inputs, thus allowing the clutch to slip at certain load levels and protect the remainder of the transmission. 
     When applying friction launch to a stepped gear automatic transmission, greater benefits in cost, mass, and packaging can be achieved if an existing clutch can be re-used as the launch clutch. Preferably, the powerflow will allow the same input clutch to be used for both forward and reverse launches. Greater benefits in cost, mass, and packaging can be achieved if the same clutch can be used for both forward and reverse launches. 
     Another way to apply friction launch to a stepped gear automatic transmission would be to re-use one or more existing reaction clutches for launch. The use of a reaction clutch offers possible cooling advantages and eliminates the need for centrifugal compensation compared to a rotating clutch. Again, the greatest benefits can be achieved if the same reaction clutch can be used for both forward and reverse launch. It is also possible to re-use a combination of an existing input clutch and an existing reaction clutch for forward and reverse launch. One example of such an automatic transmission is disclosed in commonly-owned U.S. Pat. No. 6,471,616. 
     SUMMARY OF THE INVENTION 
     It is an object of the present invention to provide a power transmission with an improved friction launch mechanism. 
     In one aspect of the present invention, a selectively operable torque transmitting mechanism is engaged to initiate both forward and reverse operation in a power transmission. Another aspect of the present invention, the torque transmitting mechanism controls the torque transmitting operation of one planetary gear member of a ratio planetary gearset in a multi-speed power transmission. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a schematic representation of a multi-speed planetary gear arrangement in a power transmission incorporating the present invention; 
         FIG. 2  is a lever diagram of the planetary gear arrangement shown in  FIG. 1 ; 
         FIG. 3  is a truth table chart describing the engagement pattern for the torque transmitting mechanisms associated with the power transmission shown in  FIG. 1 ; 
         FIG. 4  is a schematic representation of an alternative embodiment of the multi-speed planetary gear arrangement shown in  FIG. 1 ; 
         FIG. 5  is a lever diagram for the planetary gear arrangement shown in  FIG. 4 ; 
         FIG. 6  is a chart describing the gear ratios and torque transmitting mechanism engagement schedule for planetary gear arrangement shown in  FIG. 4 ; 
         FIGS. 7 and 8  are schematic stick and lever diagrams for a multi-speed gear arrangement in a power transmission according to another alternative embodiment of the present invention; 
         FIGS. 9 and 10  are schematic stick and lever diagrams for a mutli-speed gear arrangement in a power transmission according to another alternative embodiment of the present invention; 
         FIG. 11  is a chart describing the engagement schedule for the torque transmitting mechanisms associated with the planetary gear arrangement shown in  FIG. 9 ; 
         FIGS. 12 and 13  are further alternative versions of a planetary gear arrangement according to the present invention; and 
         FIGS. 14 through 27  relate to and illustrate various embodiments of selectively controllable bidirectional clutches which find particular application in the transmissions of the present invention. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     A powertrain  10  has a power source, such as a conventional internal combustion engine  12 , a multi-speed planetary transmission  14 , and a conventional final drive mechanism  16 . Planetary transmission  14  includes an input shaft  18  connected directly with engine  12 , a multi-speed planetary gear arrangement  20 , and an output shaft  22  connected directly with final drive mechanism  16 . Planetary gear arrangement  20  includes a compound planetary gearset  24 , two simple planetary gearsets  26  and  28 , three selectively engageable rotating torque transmitting mechanisms  30 ,  32  and  34  and two selectively engageable stationary torque transmitting mechanisms  38  and  40 . 
     The rotating torque transmitting mechanisms  30 ,  32  and  34  are conventional fluid-operated multi-plate clutch assemblies, the construction of which is well-known in the art of power transmissions. Likewise, the stationary torque transmitting mechanisms  38  and  40  are conventional fluid-operated brake assemblies of either friction plate type or band plate type. The torque transmitting mechanisms are controlled in engaged and disengaged states by a conventional electro-hydraulic mechanism, not shown, which includes a hydraulic valving arrangement and an electronic control unit (ECU) that incorporates a conventional programmable digital computer. The torque transmitting mechanisms are engaged and disengaged in accordance with performance and operating signals such as, for example, engine speed, vehicle speed, and engine torque to name a few. Those familiar with the art of transmission control will be familiar with the many features and functions that are available with electronic controls. 
     Referring to  FIG. 1  of the drawings, first planetary gearset  24  is shown to include a sun gear  42 , a ring gear  44 , and a planet carrier assembly  46 . Meshed pairs of pinion gears  50  and  52  are rotatably supported on pinion shafts  54  and  56 , respectively, that extend between laterally-spaced carrier segments of carrier assembly  46 . Pinion gears  50  mesh with sun gear  42  while pinion gears  52  mesh with ring gear  44 . 
     Second planetary gearset  26  includes a sun gear  58 , a ring gear  60 , and a plurality of pinion gears  62  that are meshed with both sun gear  58  and ring gear  60 . As seen, pinion gears  62  are rotatably supported on pinion shafts  64  that also extend between the laterally-spaced carrier segments of carrier assembly  46 . Thus, carrier assembly  46  is common to both first planetary gearset  24  and second planetary gearset  26 . A ring gear assembly  66  is defined by ring gear  44  of first gearset  24  and ring gear  60  of second planetary gearset  26  being connected together to rotate as a unitary component. Third planetary gearset  28  is shown to include a sun gear  68 , a ring gear  70 , and pinion gears  72  in meshed engagement with both sun gear  68  and ring gear  70 . Pinion gears  72  are rotatably supported on shafts  74  extending between components of a carrier assembly  76 . In addition, sun gear  68  is shown to be held stationary due to its direct connection to a stationary housing portion  78  of transmission  14 . 
     With continued reference to  FIG. 1 , ring gear  70  is shown to be continually drivingly connected to engine  12  through input shaft  18 . The output of planetary gear arrangement  20  is ring gear assembly  66  which is continually drivingly connected to final drive unit  16  through output shaft  22 . Torque transmitting mechanism  30  is operably arranged to control selective engagement of carrier assembly  76  with sun gear  42 . Likewise, torque transmitting mechanism  32  is operably disposed to control selective engagement of carrier assembly  76  with sun gear  58 . In addition, torque transmitting mechanism  34  is operably disposed between ring gear  70  and common carrier assembly  46 . Torque transmitting mechanism  38  is operably disposed to selectively control braking of common carrier assembly  46  and, as will be detailed, is further arranged to act as the low/reverse starting or “launch” clutch for automatic transmission  14 . Finally, torque transmitting mechanism  40  is operably disposed to control selective braking of sun gear  58  and carrier assembly  76 . 
     Referring to  FIG. 2 , a lever diagram  10 A is shown which depicts powertrain  10  in a modified form. The corresponding numbers of lever diagram  10 A will be given the same numeric designation as the components of powertrain  10  with an “A” suffix. For example, lever arm  28 A represents planetary gearset  28  and node  68 A represents sun gear  68 . Planetary gearsets  24  and  26  are combined as a single lever with the designation  24 A– 26 A. Portion  24 A consists of nodes  42 A,  44 A and  46 A while the portion  26 A consists of nodes  58 A,  60 A and  46 B. 
     During operation of transmission  14 , a neutral condition is established by disengaging all of the torque transmitting mechanisms. To establish the reverse drive ratio, torque transmitting mechanism  32  is fully engaged and torque transmitting mechanism  38  is engaged under controlled conditions dependent on signals received by the ECU. To provide a controlled launch, torque transmitting mechanism  32  is fully engaged prior to initiation of engagement of torque transmitting mechanism  38 . Thus, torque transmitting mechanism  38  is the launch device for the reverse drive ratio. Controlled engagement of torque transmitting mechanism  38  results in controlled launch or acceleration of the vehicle in which powertrain  10  is installed. This ratio is referred to as the reverse launch ratio. 
     To establish the first forward drive ratio from the neutral condition, torque transmitting mechanism  30  is fully engaged and torque transmitting mechanism  38  is engaged under controlled conditions depending on the commands from the vehicle operator, as interpreted by the ECU. The ratio thus established is also designated as the forward launch ratio. To provide a controlled launch, torque transmitting mechanism  30  is fully engaged prior to initiation of engagement of torque transmitting mechanism  38 . Thus, torque transmitting mechanism  38  is the launch device during the first forward drive ratio. As torque transmitting mechanism  38  is engaged, the vehicle will accelerate in a controlled fashion. 
     To sequentially establish the second forward drive ratio, torque transmitting mechanism  38  is disengaged and torque transmitting mechanism  40  is engaged while torque transmitting mechanism  30  remains engaged. The third forward drive ratio is established by releasing torque transmitting mechanism  40  and engaging torque transmitting mechanism  32  while torque transmitting mechanism  30  remains engaged. The fourth forward drive ratio is established by releasing torque transmitting mechanism  32  and engaging torque transmitting mechanism  34  while maintaining engagement of torque transmitting mechanism  30 . The fifth forward drive ratio is established by releasing torque transmitting mechanism  30  and engaging torque transmitting mechanism  32  while maintaining engagement of torque transmitting mechanism  34 . Finally, the sixth forward drive ratio is established by releasing torque transmitting mechanism  32  and engaging torque transmitting mechanism  40  while engagement of torque transmitting mechanism  34  is maintained. The sequence of controlled engagement and release of the various torque transmitting devices is shown in  FIG. 3 . 
     With reference now to  FIGS. 4 and 5 , a modified version of planetary gear arrangements  20  is shown and identified by reference number  20 ′. In addition, due to the similarities of many components, like reference numerals are used to identify components of planetary gear arrangement  20 ′ that are similar to those of planetary gear arrangement  20 .  FIG. 4  schematically illustrates a compact six-speed transmission  14 ′ that is well-suited for use in high torque capacity applications. Transmission  14 ′ includes five clutching elements (six with 1–2 overrunning clutch “OWC” added) and three planes of gears. Two of the planes of gears are incorporated into common carrier  46  for compact packaging. In order to handle high-torque capacity applications, the combined gearset uses long ring gear assembly  66  to bridge the two planes of planetary pinions. This long ring gearset has better scalability for high torque applications compared to conventional long pinion ravigneaux arrangements. By using this arrangement with the common carrier and long ring gearset, a family of six-speed automatic transmissions can be developed with varying torque capacities. A significant advantage relates to the ability to commonize many transmission designs using similar and/or the same components. 
     Referring again to  FIG. 4 , planetary gear arrangement  20 ′ is shown to include a modified low/reverse starting clutch, now identified by reference numeral  38 ′. In particular, torque transmitting mechanism  38 ′ utilizes a band-type brake assembly in place of the multi-plate arrangement shown in  FIG. 1 . In addition, an OWC  80  is shown operably installed between stationary housing  78  and common carrier assembly  46 . The lever diagram of  FIG. 5  is generally similar to that of  FIG. 2  with the exception that OWC  80  has been included. Furthermore,  FIG. 6  is a truth table chart describing the gear ratios and torque transmitting mechanism engagement schedule for planetary gear arrangement  20 ′. 
     Referring now to  FIGS. 7 and 8 , corresponding schematic stick and lever diagrams are provided to illustrate another embodiment of the present invention. In particular, transmission  114  includes a planetary gear arrangement  120  which uses a low torque/high-speed grounded ratio changing torque transmission mechanism as the starting clutch for both forward and reverse launch. Planetary arrangement  120  includes a compound planetary gearset  124 , two simple planetary gearsets  126  and  128 , two selectively engageable rotating torque transmitting mechanisms  130  and  132 , and four selectively engageable stationary torque transmitting mechanisms  134 ,  136 ,  138  and  140 . 
     Rotating torque transmitting mechanisms  130  and  132  are conventional fluid-operated multi-plate clutch assemblies. Stationary torque transmitting mechanisms  134  and  136  are conventional fluid-operated brake assemblies of either friction plate type or band plate type. In addition, stationary torque transmitting mechanisms  138  and  140  are preferably a band-type brake assembly. As previously noted, the torque transmitting mechanisms are shifted between engaged and disengaged states by a conventional electro-hydraulic mechanism controlled by an ECU that incorporates a conventional programmable digital computer. 
     First planetary  124  is shown to include a sun gear  142 , a ring gear  144 , and a planet carrier assembly  146 . Meshed pairs of pinion gears  150  and  152  are rotatably supported on pinion shafts that extend between laterally-spaced carrier segments of carrier assembly  146 . Pinion gears  150  mesh with sun gear  142  while pinion gears  152  mesh with ring gear  144 . Second planetary gearset  126  includes a sun gear  158 , a ring gear  160 , and a plurality of pinion gears  162  that are meshed with both sun gear  158  and ring gear  160 . Pinion gears  162  are rotatably supported on pinion shafts that extend between laterally-spaced carrier segments of a carrier assembly  164 . Third planetary gearset  128  is shown to include a sun gear  168 , a ring gear  170 , and pinion gears  172  in meshed engagement with both sun gear  168  and ring gear  170 . Pinion gears  172  are rotatably supported on pinion shafts extending between components of a carrier assembly  176 . 
     With continued reference to  FIG. 7 , sun gear  142  is operatively connected with stationary torque transmitting mechanism  138  and rotating torque transmitting mechanism  130 . Ring gear  144  is operatively connected with stationary torque transmitting mechanism  134 . When stationary torque transmitting mechanism  138  is selectively engaged, sun gear  142  is held stationary and when torque transmitting mechanism  130  is selectively engaged, sun gear  142  is connected directly with engine through input shaft  18 , or with a vibration isolator or damper connected between engine  12  and shaft  18 . When torque transmitting mechanism  134  is selectively engaged, ring gear  144  is held stationary. Likewise, ring gear  160  of second planetary gearset  126  is operatively connected with torque transmitting mechanism  132 . In addition, sun gear  158  of planetary gearset  126  is continuously connected with sun gear  168  of planetary gearset  128 . Both sun gears  158  and  168  are operatively connected with torque transmitting mechanism  140  which, when engaged, will hold sun gears  158  and  168  stationary. When torque transmitting mechanism  132  is engaged, ring gear  160  will rotate in unison with engine  12  through input shaft  18 . 
     Ring gear  170  is shown to be continuously connected with ring gear  144  of first planetary gearset  124  and operatively connected with torque transmitting mechanism  134 . Planetary carrier assembly  176  is operatively connected with torque transmitting mechanism  136  and torque transmitting mechanism  132 . When torque transmitting mechanism  134  is selectively engaged, both ring gear members  144  and  170  will be held stationary. When torque transmitting mechanism  132  is selectively engaged, planet carrier assembly  176  will rotate in unison with engine  12  through input shaft  18  and also with ring gear  160 . When torque transmitting mechanism  136  is selectively engaged, planet carrier assembly  176  and ring gear  160  will be held stationary. As seen, torque transmitting mechanism  136  is a combination multi-plate clutch assembly  135  and a controllable overrunning clutch  139 , arranged in parallel, both of which are operable for selectively braking rotation of planet carrier  176 . 
     During operation of the transmission, a neutral condition is established by disengaging all of the torque transmitting mechanisms. To establish the reverse drive ratio, torque transmitting mechanism  136  is fully engaged and torque transmitting mechanism  138  is engaged under controlled conditions dependent on signals received by the ECU. To provide a controlled launch, torque transmitting mechanism  136  is fully engaged prior to initiation of engagement of torque transmitting mechanism  138 . Thus, torque transmitting mechanism  138  is a launch device for the reverse drive ratio. Controlled engagement of torque transmitting mechanism  138  results in controlled launch or acceleration of the vehicle in which the powertrain is installed. In a like manner, to establish the first forward drive ratio from the neutral condition, torque transmitting mechanism  140  is fully engaged and torque transmitting mechanism  138  is engaged under controlled conditions depending on commands from the vehicle operator, as interpreted by the ECU. The ratio thus establishes also designated as the forward launch ratio. To provide a controlled launch, torque transmitting mechanism  140  is fully engaged prior to initiation of engagement of torque transmitting mechanism  138  during the first forward drive ratio. 
     To establish the second forward ratio, torque transmitting mechanism  138  is disengaged and torque transmitting mechanism  130  is engaged while torque transmitting mechanism  140  remains engaged. The third forward drive ratio is established by releasing torque transmitting mechanism  130  and engaging torque transmitting mechanism  132 . This completes a ratio interchange from the second forward ratio to the third forward ratio. To establish the fourth forward drive ratio, during a ratio interchange from third to fourth, torque transmitting mechanism  140  is disengaged and torque transmitting mechanism  130  is engaged while torque transmitting mechanism  132  remains engaged. To establish the fifth forward drive ratio, with an interchange from the fourth forward drive ratio, torque transmitting mechanism  130  is disengaged and torque transmitting mechanism  138  is engaged while torque transmitting mechanism  132  remains engaged. Finally, the sixth and (highest) forward drive ratio with an interchange from the fifth forward drive ratio is established by disengaging torque transmitting mechanism  138  while engaging torque transmitting mechanism  134  while torque transmitting mechanism  132  remains engaged. 
     Referring now to  FIGS. 9 and 10 , a planetary gear arrangement  220  is disclosed for use in powertrain  10  which uses two separate grounding clutches to launch. That is, one grounding clutch is used for the forward launch and the other clutch is used for the reverse launch. Planetary gear arrangement  220  includes three simple planetary gearsets  224 ,  226  and  228 , two rotating torque transmitting mechanisms  230  and  232 , three stationary grounding-type torque transmitting mechanisms  234 ,  236  and  238 , and a free-wheeling OWC  240 . Planetary gearset  224  is shown to include a sun gear  242 , a ring gear  244 , and pinions  250  meshed therewith. Pinions  250  are rotatably supported on pinion shafts which are connected to laterally-spaced carrier components of a planet carrier assembly  246 . Likewise, second planetary gearset  226  includes a sun gear  258 , a ring gear  260 , and pinion gears  262  meshed therewith. Pinions  262  are rotatably supported on pinions shafts which extend between laterally-spaced components of a planet carrier assembly  264 . As seen, carrier assembly  264  is shown to be directly connected to ring gear  244  of planetary gearset  224 . Third planetary gearset  228  includes a sun gear  268 , a ring gear  270 , and pinions gears  272  meshed therewith. Pinion gears  272  are rotatably supported on pinion shafts which extend between laterally-spaced carrier components of planet carrier assembly  276 . As is also seen, carrier assembly  276  is directly connected for rotation with ring gear  260  of planetary gearset  226  and carrier assembly  246  is directly connected for rotation with ring gear  270 . 
     Torque transmitting mechanism  230  is operably disposed for selectively coupling sun gear  268  for rotation with input shaft  18 . In addition, torque transmitting mechanism  232  is operably arranged for selectively coupling ring gear  260  of planetary gearset  226  and carrier assembly  276  of planetary gearset  228  for rotation with input shaft  18 . Torque transmitting mechanism  234  is operably arranged for selectively coupling sun gear  268  with stationary housing segment  280 , thereby braking rotation of sun gear  268 . Torque transmitting mechanism  236  is operably arranged between stationary housing  280  and planet carrier assembly  276  for selectively braking rotation of planetary carrier  276 . Free-wheeling overrunning clutch  240  is arranged to allow for a 1–2 shift and is disposed between housing  280  and planetary carrier  276  in parallel with torque transmitting mechanism  236 . Torque transmitting mechanism  238  is operably disposed for selectively braking rotation of sun gear  242 . As is similar to all of the previous lever diagrams,  FIG. 10  depicts gear arrangement  220  shown in  FIG. 9  in alternative mode. 
       FIG. 11  is a chart illustrating the shift schedule for the various torque transmitting mechanisms associated with planetary gear assembly  220  required to establish the various forward and reverse gear ratio drive connections. In this embodiment, torque transmitting mechanism  236  is used to launch in reverse and torque transmitting mechanism  238  is used to launch forward. Alternatively, torque transmitting mechanism  236  could be used by itself as the starting clutch for both forward and reverse launch. 
     Referring to  FIG. 12 , a modified version of the planetary gear assembly shown in  FIG. 9  is described as having a selectively engageable OWC (torque transmitting mechanism  240 ′) in series with the friction launch low/reverse clutch (torque transmitting mechanisms  236 ) instead of being arranged in parallel. This allows use of a one-way clutch for the 1–2 shift and a single clutch for friction launch in both first and reverse gears. In particular, when the reverse gear is established, the grounding torque from torque transmitting mechanisms  236  is transferred through the engaged torque transmitting mechanism  240 ′ as if a direct connect was established in both directions. In first gear launch, the first gear grounding torque is through the selective OWC  240 ′ as if it was a one-way clutch commonly used in conventional automatic transmissions. However, the reverse direction of torque capacity for OWC  240 ′ must be disabled before a 1–2 shift and after reverse or low braking grounding torque is needed. OWC  240 ′ acts as a one-way clutch during the 1–2 and 2–1 shifts, thereby providing improved shifts than could be obtained in a clutch—clutch system. Moreover, in first gear coast, OWC  240 ′ may be enabled in the reverse ground direction to provide ground for coast braking if desired. Selective actuation or disablement of OWC  240 ′ can be controlled via any suitable type of actuation system, with hydraulic controls being preferred. As seen in  FIG. 13 , a slight modification of  FIG. 12  is now shown which locates overrunning clutch  240  in series with the friction launch low/reverse clutch (torque transmitting mechanism  236 ). An additional grounding clutch-type torque transmitting mechanism  245  is located in parallel for use in reverse launch. The arrangement in  FIG. 12  is preferable from a packaging standpoint but further requires the use of a selectively controllable overrunning clutch. 
     One device of particular application to the selective overrunning clutch disclosed for use in  FIG. 12  is a controllable bidirectional one-way clutch. While bi-directional one-way clutches are generally known, a need exists for a one-way clutch which is selectable while both inner and outer members are rotating. Prior speed differential, centrifugal force, electromechanical, electro-magnetic, and friction clutches are not sufficiently controllable within the operating ranges in an automatic transmission. Specifically, electromechanical or electromagnetic devices are not currently present in an automatic transmission and would require separate control systems. Controlling a one-way clutch via a friction brake hub when the power transmission elements are in series requires a non-functional positioning of the brake hub. Conversely, hydraulic power is currently available in an automatic transmission and is a method to discretely control a one-way clutch within the required ranges of operation. 
     With reference to  FIGS. 14 and 15 , a clutch assembly  810  includes a bidirectional, one-way clutch  812  and a friction clutch  814  packaged adjacent one another. A dual stage piston  816  is hydraulically actuated by a friction clutch apply oil. While a dual stage piston  816  is represented in  FIGS. 14 and 15 , one skilled in the art will appreciate that clutch assembly  810  may include a one-way clutch piston  818  and a friction clutch piston  820  operating independently from one another as depicted in  FIG. 16 . 
     Dual stage piston  816  includes a plurality of sub-pistons  822  housed within a perimeter portion of dual stage piston  816 . Sub-pistons  822  may be selectively hydraulically actuated to place one-way clutch  818  in one of a locked mode or an unlocked mode.  FIGS. 14 and 15  depict one-way clutch  812  in the unlocked mode where sub-pistons  822  are not acted upon by hydraulic pressure. One-way clutch  812  also includes an apply piston  824 , a thrust bearing  825 , a clutch plate  826 , a reaction plate  828  and a selection cage  830 . Selection cage  830  includes a pair of dogs  832  in communication with clutch plate  826 . A plurality of torque carrying elements  834  are generally cylindrically shaped members spaced apart from one another. Selection cage  830  includes a serpentine support web  833  having a plurality of pockets formed therein to retain each of torque carrying elements  834  in a desired spaced relationship. An outer race  840  houses clutch plate  826  and selection cage  830 . Outer race  840  is drivingly engaged with a first rotating member. An inner race  844  houses reaction plate  828  and is drivingly coupled to a second rotating member. A cage spring  848  is positioned between each dog  832  and clutch plate  826  to place one-way clutch  812  in the unlocked mode. A plurality of pistons springs  850  interconnect sub-pistons  822  and dual stage piston  816  to return sub-pistons  822  to a position disengaged from apply piston  826 . In the unlocked mode, torque carrying elements  834  are positioned within detents  851  of outer race  840 . Accordingly, outer race  840  and inner race  844  are allowed to freely rotate relative to one another. When one-way clutch  812  is in the locked mode, torque carrying elements  834  are forced to engage outer race  840  and inner race  844  such that outer race  840  and inner race  844  rotate in the same direction at the same speed. 
     To actuate one-way clutch  812  from the unlocked mode to the locked mode, hydraulic pressure is applied to sub-pistons  822 . Each sub-piston  822  axially translates to contact apply piston  824 . Apply piston  824  axially translates to bring clutch plate  826  into contact with reaction plate  828 . Clutch plate  826  includes a friction surface such that upon contact with rotating reaction plate  828 , clutch plate  826  rotates. Rotation of clutch plate  826  drives selection cage  830  to rotate in opposition to cage springs  848 . As selection cage  830  rotates, torque carrying elements  834  are forced to leave detents  851  and follow a contour  852  formed on outer race  840 . As torque carrying elements  834  travel along contour  852 , the torque carrying elements are forced to engage both outer race  840  and inner race  844 . As such, torque is transmitted between the races resulting in both races rotating in the same direction at the same speed. 
     To switch one-way clutch  812  from the locked mode to the unlocked mode, hydraulic pressure is released from sub-pistons  822 . Piston springs  850  disengage sub-pistons  822  from apply piston  824 . The torque generating frictional interface between clutch plate  826  and  828  is also released. Cage springs  848  rotate selection cage  830  to reposition torque carrying elements  834  within detents  851 . Once torque carrying elements  834  are positioned within the detents, torque is no longer transferred between outer race  840  and inner race  844 . 
     With reference to  FIGS. 17 through 19 , an alternative clutch assembly  854  is depicted. Clutch assembly  854  includes a one-way clutch  856  and a friction clutch  858 . A dual stage piston  860  is actuated by a friction clutch apply oil which provides the desired force for positioning one-way clutch  856  in one of a locked mode or unlocked mode. In similar fashion to clutch assembly  810  previously described, independent pistons (not shown) may be used in lieu of dual stage piston  860 . A plurality of sub-pistons  862  are positioned within a plurality of recesses  864  extending into dual stage piston  860 . A plurality of ports  866  are in communication with recesses  864  to allow pressurized hydraulic fluid to act on sub-pistons  862 . One-way clutch  856  includes an apply piston  868 , a cage  870 , an outer race  872 , an inner race  874  and torque carrying elements  876 . As opposed to using frictional plate elements to actuate one-way clutch  812 , one-way clutch  856  is actuated using a camming action. A set of cams described hereinafter convert axial motion of apply piston  868  to rotational movement of cage  870 . 
     A spring  878  is positioned to act upon outer race  872  and return cage  870  to an initial position depicted in  FIG. 17 . Cage  870  includes a flange  880  having an outer surface  882  in contact with a retainer  884 . As best shown in  FIGS. 18 and 21 , an inner surface  886  of cage  870  includes a driven cam face  888  positioned thereon. Outer race  872  includes a driving cam face  890  positioned proximate driven cam face  888 .  FIGS. 18 through 20  depict one-way clutch  856  in the locked mode.  FIGS. 21 through 23  depict one-way clutch  856  in the unlocked mode. As mentioned earlier, as long as hydraulic pressure is not supplied to sub-pistons  862 , spring  878  maintains clearance between driven cam face  888  and driving cam face  890  as shown in  FIG. 21 . A spring  891  returns apply piston  868  and sub-pistons  862  to their initial positions. At this time, cage  870  is positioned to maintain the location of torque carrying element  876  within detents  892 . Accordingly, outer race  872  may rotate independently of inner race  874 . 
     To operate one-way clutch  856  in the locked mode, hydraulic fluid axially displaces sub-pistons  862  to further axially displace apply piston  868 . Apply piston  868  axially moves a thrust washer  894  into contact with flange  880  of cage  870 . The axial displacement force from hydraulic acting on sub-pistons  862  is sufficient to overcome spring  878  thereby engaging driven cam face  888  with driving cam face  890 . Due to the angle of the respective cam faces, axial movement is converted to rotational movement of cage  870 . Rotation of cage  870  causes torque carrying elements  876  to be moved from detents  892  into engagement with both outer race  872  and inner race  874 . One-way clutch  856  is now in the locked mode drivingly interconnecting outer race  872  with inner race  874 . 
     It should be appreciated that the bi-directional one-way clutch may be a roller, a diode, a sprag, or a pawl-type one-way clutch. For example, a diode type one-way clutch  896  is illustrated in  FIGS. 24 through 27 . One-way clutch  896  includes an apply plate  898 , a reaction plate  900 , a selector plate  902 , a diode  903 , a first sprag  904  and second sprag  906 . To change the mode of operation of diode clutch  896  from a locked to an unlocked mode or vice versa, hydraulic pressure is supplied to a plurality of sub-pistons  908  to axially displace apply plate  898  toward reaction plate  900 . In similar fashion to that earlier described, each of apply plate  898  and reaction plate  900  include selectively engageable cam faces. Selector plate  902  rotates upon engagement and disengagement of the cam faces. Selector plate  902  includes a tab  910 . Tab  910  is selectively rotated into contact with first sprag  904  to place diode clutch  896  in the desired locked or unlocked mode. Diode clutch  896  is switched to the opposite mode by simply releasing pressure from sub-pistons  908 . A spring  912  disengages apply plate  898  from reaction plate  900  thereby disengaging the respective cam faces. At this time, selector plate  902  rotates to its initial position.