Abstract:
Provided are various vehicular transmission combinations utilizing selectively engageable spiral-type couplings or slipper ring clutches, operable to provide smooth shifting between gear sets, in place of traditional shifting mechanisms, such as hydraulic plate-type clutches, or synchronizers.

Description:
CROSS REFERENCE TO RELATED APPLICATION 
     This application claims the benefit of U.S. Provisional Application No. 60/642,390, filed Jan. 7, 2005, which is hereby incorporated by reference in its entirety. 
    
    
     TECHNICAL FIELD 
     The present invention relates to vehicular transmissions and, more particularly, to the use of slipper ring clutch mechanisms to control various types of vehicular transmissions. 
     BACKGROUND OF THE INVENTION 
     Automobiles require shiftable transmissions as a result of the physics of the internal combustion engine. All internal combustion engines have a maximum engine rotational speed above which the internal combustion engine can not effectively operate due to limitations of various internal combustion engine components, such as the valvetrain or reciprocating components. Additionally, internal combustion engines have an engine speed range where horsepower and torque are at their maximum. The transmission, through changes in gear ratio, allows the ratio between the rotational speed of the engine and the drive wheels to change as the automobile accelerates or decelerates. By varying the gear ratios within the transmission, the internal combustion engine may operate below the maximum rotational speed and preferably near the engine speed range for best performance and/or fuel economy. The transmission provides a gear set operable to provide a reverse gear for backing the automobile. The transmission also provides a low gear, effective to allow adequate acceleration of the automobile while not sacrificing the top speed of the automobile. 
     Vehicular transmissions typically employ a clutch mechanism such as a hydraulically actuated plate-type clutch or a set of synchronizers in order to effect a gear selection. Engineers strive to make this gear selection as smooth and imperceptible to the driver as possible. Modern vehicular transmissions are designed with a very small footprint to allow tighter packaging within the engine bay resulting in an increased passenger cabin volume. Therefore, it is important to ensure that the usable space within the transmission housing is effectively utilized in the design of the transmission. 
     Automatically and manually shiftable transmissions constitute the two main varieties of vehicular transmissions. Automatically shiftable transmissions select the gear ratio with no input from the driver, whereas the manually shiftable transmissions require a gear selection input from the driver. Different gear selection techniques exist for manual transmissions such as, dual clutch shifting and synchronized shifting. The synchronized transmission designs utilize synchronizers to ensure that the drive gear rotational speed closely matches that of the input shaft during a shift to effect a smooth transition through the gear sets. The dual clutch transmission, or DCT, employs two separate input shafts, each with its own corresponding clutch, to engage and disengage the input shafts. In a 5-speed DCT, for instance, one input shaft would account for forward ranges 1, 3, and 5, and the second shaft would account for forward ranges 2 and 4 as well as a reverse range. Automatically shiftable transmissions typically employ hydraulically actuated plate-type clutches. Other variations of the automatic transmission may employ both hydraulically actuated plate-type clutches as well as one-way clutches. In this variation, the one-way clutches are used to improve shift feel and shift timing by enabling the smooth release of the off-going clutching element as the on-coming element gains torque capacity. 
     SUMMARY OF THE INVENTION 
     Provided is a shiftable vehicular transmission having a rotatable first input shaft and an output shaft. At least one drive gear is rotatably mounted with respect to the first input shaft, and at least one driven gear is fixedly mounted with respect to the output shaft and in constant mesh with the drive gear. Additionally, at least one selectively engageable slipper ring clutch is disposed between the drive gear and the first input shaft. The selectively engageable slipper ring clutch operates to lock the drive gear with the first input shaft when engaged so that the drive gear may rotate with substantially the same rotational speed as the first input shaft. Subsequently, the selectively engageable slipper ring clutch operates to unlock the drive gear from the first input shaft when disengaged so that the drive gear may rotate independently with respect to the first input shaft. 
     Also provided is a hybrid vehicular transmission having a plurality of electric motors and a plurality of planetary gear sets. A plurality of selectively engageable slipper ring clutches is also provided, each being operable to selectively couple at least one of the plurality of planetary gear sets with at least one of the plurality of electric motors when the slipper ring clutch is engaged. Additionally, the selectively engageable slipper ring clutch operates to de-couple the at least one the plurality of gear sets from the at least one of the plurality of electric motors when the slipper ring clutch is disengaged. The selectively engageable slipper ring clutch may be selectively engaged by a hydraulic actuator. Additionally, an input shaft mounted damper assembly operable to attenuate the transmission of torsional vibrations to the hybrid vehicular transmission may be provided. One of the plurality of selectively engageable slipper ring clutches may operate to bypass the damper assembly when engaged. Alternately, the selectively engageable slipper ring clutches are operable to allow the damper assembly to function when disengaged. 
     Also provided is an automatically shiftable transmission having a low forward range planetary gear system having a low gear and a selectively engageable slipper ring clutch operable to control the low forward range planetary gear system, thereby providing a measure of engine braking in the low gear. 
     The selectively engageable slipper ring clutch may include an inner slipper ring having an outer periphery and an outer slipper ring also having an inner periphery and an outer periphery each being circumferentially disposed about the inner slipper ring. The outer slipper ring has a slit operable to allow the radial expansion of the outer slipper ring. A plurality of arcs are formed on the outer periphery of the inner slipper ring, and the outer slipper ring has complementary number of arcs formed on the inner periphery. The plurality of arcs and complementary arcs form a plurality of cavities between the inner slipper ring and the outer slipper ring. A plurality of roller elements are disposed respectively within the plurality of cavities. An outer race is rotatably mounted with respect to the outer periphery of the outer slipper ring. At least one positionable selector sleeve may be provided to selectively engage and disengage the selectively engageable slipper ring clutch. Additionally, a bimetallic spring mechanism may be disposed between the inner periphery of the outer race and the outer periphery of the outer slipper ring. The bimetallic spring operates to transmit drag forces between the outer race and the outer slipper ring when the fluid within the selectively engageable slipper ring clutch is cold and to subsequently reduce the drag forces when the fluid within the selectively engageable slipper ring clutch is warmer. 
     By employing selectively engageable slipper ring clutch mechanisms within vehicular transmission combinations, many benefits may be garnered. These benefits may include, reduced part cost, smaller packaging footprint, lower torque loss, reduced control scheme complexity, increased power density, and the ability to shift gears in both a synchronous and non-synchronous state. 
     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 modes for carrying out the invention when taken in connection with the accompanying drawings. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIGS. 1   a ,  1   b ,  1   c , and  1   d  taken together comprise a series of four schematic cross-sectional diagrams of a slipper ring clutch illustrating the four states of operation; 
         FIG. 2  is a schematic sectional fragmentary view of an exemplary manual transmission input shaft mounted slipper ring clutch mechanism, showing the portion of the slipper ring clutch mechanism above the centerline of the input shaft; 
         FIG. 2   a  is a schematic cross-sectional view, taken along line A-A of  FIG. 2 , of an exemplary slipper ring clutch design; 
         FIG. 3  is a schematic fragmentary sectional view of a Dual Clutch Transmission, or DCT, architecture employing input shaft mounted, selectively engageable slipper ring clutches; 
         FIG. 4  is a schematic diagram illustrating the various components of a hybrid powertrain system employing a plurality of slipper ring clutches; 
         FIG. 5  is a schematic sectional fragmentary view of an alternate embodiment of a slipper ring clutch mechanism as used in an vehicular transmission showing the portion of the slipper ring clutch mechanism above the centerline of the input shaft and illustrating a simplified construction and an improved locking mechanism; 
         FIG. 5   a  is a schematic fragmentary view illustrating a mechanism for locking the slipper ring clutch of  FIG. 5 ; 
         FIG. 6   a  is a schematic sectional view of a slipper ring clutch mechanism illustrating a bi-metallic drag spring device in a cold state of operation; 
         FIG. 6   b  is a schematic sectional view of a slipper ring clutch mechanism illustrating the bimetallic drag spring device shown in  FIG. 6   a  in a warm state of operation; 
         FIG. 7   a  is a schematic lever diagram representing an automatically shiftable transmission operating in a first forward range and illustrating the combination of a slipper ring clutch within the automatically shiftable transmission operable to enable engine braking in first gear; 
         FIG. 7   b  is a schematic lever diagram representing the automatically shiftable transmission of  FIG. 7   a  operating in a second forward range; and 
         FIG. 7   c  is a schematic lever diagram representing the automatically shiftable transmission of  FIG. 7   a  operating in the first forward range illustrating an engine braking condition. 
     
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Referring to the figures wherein like reference numbers refer to like or similar components,  FIGS. 1   a ,  1   b ,  1   c , and  1   d  are schematic diagrams of a selectively engageable slipper ring clutch  10  each illustrating one of four states of engagement. The outer slipper ring  12  and the inner slipper ring  14  of the slipper ring clutch  10  define a cavity  16  with a roller element  18  disposed therein. The outer slipper ring  12  is of a split ring design, described hereinafter, allowing the outer slipper ring  12  to expand radially outward.  FIG. 1   a  represents a state in which the slipper ring clutch  10  is locked or engaged in both rotational directions, clockwise and counter-clockwise. With clockwise rotation of the inner slipper ring  14 , the roller element  18  will engage ramp  20  of the inner slipper ring  14 . Subsequently, the roller element  18  will engage ramp  22  of the outer slipper ring  12 . This engagement of ramp  22  will force the outer slipper ring  12  to expand outward and frictionally engage a generally cylindrical inner surface  24 , also shown in  FIG. 2 , defined by an outer race  26 , thereby locking the outer race  26  and the slipper ring clutch  10  together for unitary rotation. If the direction of rotation of the slipper ring clutch  10  is reversed, i.e. moving the inner slipper ring  14  in the counter-clockwise direction, the roller element  18  will engage ramp  20 ′ of the inner slipper ring  14  and ramp  22 ′ of the outer slipper ring  12 . The roller element  18  will force the outer slipper ring  12  to expand radially outwardly to frictionally engage the cylindrical inner surface  24  of outer race  26  in the same fashion as previously described, thereby locking the outer race  26  and the slipper ring clutch  10  for unitary rotation. 
     In  FIG. 1   b , a first key  28  is inserted between tab  30  of the outer slipper ring  12  and tab  32  of the inner slipper ring  14 . The presence of the first key  28  will allow relative motion between the outer slipper ring  12  and the inner slipper ring  14  in the clockwise direction, thereby allowing the slipper ring clutch  10  to engage the outer race  26  in the way previously described. However, the presence of first key  28  will disallow relative motion between the inner slipper ring  14  and the outer slipper ring  12  in the counter-clockwise direction disallowing the expansion of the outer slipper ring  12 . In effect, the slipper ring clutch  10  will “freewheel” in the counter-clockwise direction. 
     In  FIG. 1   c , a second key  34  is inserted between tab  30 ′ of the outer slipper ring  12  and tab  32  of the inner slipper ring  14 . The slipper ring clutch  10  will operate in the opposite fashion to that of  FIG. 1   b , in that it will “freewheel” in the clockwise direction and lock in the counter-clockwise direction. 
       FIG. 1   d  illustrates a condition where the relative motion of the inner slipper ring  14  to that of the outer slipper ring  12  is restricted by the inclusion of both the first key  28  and the second key  34  within the slipper ring clutches  10 . This will force the slipper ring clutch to “freewheel” in both the clockwise and counter-clockwise direction. 
       FIG. 2  is a schematic sectional fragmentary view of a transmission  38  illustrating an input shaft  36  mounted slipper ring clutches  10  and  10 ′. The slipper ring clutches  10  and  10 ′ operate to substantially lock gears  40  and  40 ′, respectively, to the input shaft  36  of the transmission, thereby providing the requested gear ratio. The gears  40  and  40 ′ are rotatably mounted with respect to the input shaft  36  by bearings  41  and  41 ′, respectively. The slipper ring clutch  10  has the ability to dissipate energy upon frictional engagement of the outer slipper ring  12  with the outer race  26 . This attribute allows both synchronous and non-synchronous gear selections. Therefore, the relative rotational speed between the input shaft  36  to that of the gear  40  is of little consequence to provide smooth gear engagements.  FIG. 2   a  is a schematic sectional view, taken along line A-A of  FIG. 2 , of a typical slipper ring clutch  10  illustrating the structure of the slipper ring clutch  10  whose operation was described with reference to  FIG. 1 . An inner race  42  is fixedly mounted, and coaxially oriented with respect to the input shaft  36 . The inner slipper ring  14  is fixedly attached to, and coaxially oriented with the inner race  42 . The inner race  42  and the inner slipper ring  14  may be separate elements or may be a single element. The outer slipper ring  12  is circumferentially disposed about the inner slipper ring  14 . The outer slipper ring  12  has a split  44  to allow the radial expansion of the outer slipper ring  12 . A plurality of arcs  46  are provided on the outside periphery of the inner slipper ring  14  and a plurality of arcs  48  are provided on the inside periphery of the outer slipper ring  12  to form a plurality of cavities  16  within each of which the generally cylindrical roller element  18  is disposed. 
     Referring back to  FIG. 2 , the outer race  26  is rigidly mounted with respect to the gear  40  and rotates coaxially about the slipper ring clutch  10 . The cylindrical inner surface  24  of outer race  26  defines the surface upon which the outer slipper ring  12  will frictionally engage upon engagement of slipper ring clutch  10 . This frictional engagement substantially locks the input shaft  36  to the gear  40 , thereby providing the requested gear ratio. The selective engagement of the gears  40  and  40 ′ may be performed by the translational movement of a shift fork  50  slidably mounted to a shaft  52 . The shift fork  50  will engage a selector sleeve  54  having three axial positions relative to the slipper ring clutches  10  and  10 ′. These three axial positions are maintained by a detent mechanism  56 . The detent mechanism  56  includes a spring  58  operable to bias a ball  60 , which operates to engage recesses  64 ,  66 , and  68  formed on the selector sleeve  54 . First keys  28 ,  28 ′ and second keys  34 ,  34 ′ are fixedly mounted with respect to the selector sleeve  54  and are operable to selectively engage the split  44 , shown in  FIG. 2   a , formed by the outer slipper ring  12 . 
     By sliding the shift fork  50  to the right, the selector sleeve  54  will move to the left-most detent position  64  and will extract the first key  28 ′ and the second key  34 ′ from the split  44 ′ allowing the outer slipper ring  12 ′ to expand radially and frictionally engage the cylindrical inner surface  24 ′ of the outer race  26 ′. The frictional engagement between the slipper ring clutch  10 ′ and the outer race  26 ′ will dissipate the energy generated by a non-synchronous shift and cause the gear  40 ′ to become substantially locked to the input shaft  36  for unitary rotation therewith. The first key  28  and second key  34  will remain engaged with slipper ring clutch  10 , forcing the slipper ring clutch  10  to “freewheel”. Upon movement of the selector sleeve  54  to the middle detent position  66 , as shown in  FIG. 2 , both first keys  28  and  28 ′ and second keys  34  and  34 ′ will remain engaged with both slipper ring clutches  10  and  10 ′. This condition will force both slipper ring clutches  10  and  10 ′ to “freewheel” creating a neutral gear selection state. Moving the selector sleeve  54  to the right-most detent position  68  will extract first key  28  and second key  34  allowing the outer slipper ring  12  to expand and frictionally engage the cylindrical inner surface  24  of the outer race  26 . This frictional engagement between the slipper ring clutch  10  and the outer race  26  will dissipate the energy generated by a non-synchronous shift and cause the gear set  40  to become substantially locked to the input shaft  36  for unitary rotation therewith. The first key  28 ′ and second key  34 ′ will remain engaged with slipper ring clutch  10 ′, forcing the slipper ring clutch  10 ′ to “freewheel”. 
       FIG. 3  is a schematic sectional fragmentary view of a Dual Clutch Transmission or DCT  70  employing a plurality selectively engageable slipper ring clutches  10 ,  10 ″ mounted with respect to the input shaft  36 . Additionally, a plurality of selectively engageable slipper ring clutches  10 ′,  10 ′″ are mounted with respect to the input shaft  36 ′. The use of slipper ring clutches  10 ,  10 ′,  10 ″, and  10 ′″ within the DCT  70  may simplify the control scheme of input clutches  72  and  72 ′. Slipper ring clutches  10  and  10 ″ are fixedly mounted with respect to the input shaft  36  containing odd gear sets  76  and  76 ″, while the slipper ring clutches  10 ′ and  10 ′″ are fixedly mounted with respect to the input shaft  36 ′, which contains the even gear sets  76 ′ and  76 ′″. 
     The shifting operation of the DCT  70  is best understood by way of example. In first gear, slipper ring clutch  10  will be engaged; thereby locking the gear set  76  to the input shaft  36  to provide the requisite gear ratio to an output shaft  78 . Additionally, to transfer torque from an engine  77  to the DCT  70 , the input clutch  72  must also be engaged. The second slipper ring clutch  10 ′ is operable to frictionally lock the gear set  76 ′, corresponding to the second forward range, to the input shaft  36 ′. The slipper ring clutch  10 ′ may engage the gear set  76 ′ while the gear set  76  is engaged. However, the input clutch  72 ′ must be de-clutched or disengaged during this pre-selection operation to avoid lockup of the DCT  70 . Prior to the shift to the gear set  76 ′, the slipper ring clutch  10  will remain engaged with the gear set  76 , but will be set to a “freewheel” ready state. When the shift is desired, the input clutch  72 ′ will engage and slipper ring clutch  10  will be set to “freewheel”. The pre-selected gear set  76 ′ will then provide the requisite gear ratio to the output shaft  78 . With the slipper ring clutch  10  “freewheeling”, the de-clutching of the input clutch  72  during the shift will not be required. This characteristic may reduce the control system complexity for the operation of the DCT  70 . The shifting of the DCT  70  will cause little torque interruption and will therefore provide a smooth shift feel. Shifts from gear set  76 ′ to gear set  76 ″ to gear set  76 ′″ as well as downshifts will follow a similar sequence. 
     The slipper ring clutch  10  may also be used within hybrid powertrain applications.  FIG. 4  is a schematic diagram illustrating the various components of a hybrid powertrain system  80  having an engine  79 , hybrid transmission  81 , and a final drive  83 . The hybrid transmission  81  employs multiple slipper ring clutches  10  and  10 ′ and favors the use of a plurality of motors  82  and  82 ′ combined with a plurality of planetary gear sets  84  and  84 ′ selectively engageable by slipper ring clutches  10  and  10 ′. The hybrid powertrain system  80  will enable multiple modes of operation with the possibility of higher system efficiency. The shifting from one mode of operation to another is synchronous due to the minimal slip speed experienced by the slipper ring clutches  10  and  10 ′. This configuration may also result in improved packaging and lower spin losses than traditional clutch mechanisms. 
     The slipper ring clutch  10  may be integrated within the hybrid transmission  81  in place of a traditional hydraulically actuated plate-type clutch. The slipper ring clutches  10  and  10 ′ may use a hydraulic actuators  88  and  88 ′, respectively to effect engagement. The actuators  88  and  88 ′ are similar to the actuators employed for hydraulic plate-type clutch engagement. However, the hydraulic actuators  88  and  88 ′ are operable to selectively engage the outer slipper ring with first keys  28 ,  28 ′ and second keys  34 ,  34 ′. Therefore, for operating modes requiring the engagement of the slipper ring clutch  10 , the hydraulic actuator  88  will be de-stroked. When the hydraulic actuator  88  is stroked, the slipper ring clutch  10  will “freewheel”. The same principle of operation applies to the slipper ring clutch  10 ′ and actuator  88 ′. 
     An additional slipper ring clutch  10 ″ placement within the hybrid transmission  81  is shown in  FIG. 4 . The slipper ring clutch  10 ″ may be used to bypass torsional springs  90  located within a damper  91  when starting or shutting down the internal combustion engine  79 . In this embodiment, a hydraulic actuator  94  will engage and disengage the slipper ring clutch  10 ″. The damper  91  operates to isolate the transmission from the torsional vibrations caused by the firing pulses of the engine  79 . However, when shutting down or restarting the engine  79 , such as when transitioning into an out of electric mode, the compliance of the damper  91  may impart a resonance to the hybrid transmission  81 . Therefore, it is beneficial to bypass the springs  90  when shutting off and restarting the engine  79 . The slipper ring clutch  10 ″ located within the damper  91  may result in improved packaging and lower torque losses than traditional clutching mechanisms. 
       FIG. 5  is a schematic sectional fragmentary view of a design for a slipper ring clutch  100  as used in a transmission  102  and illustrating a simplified design and improved selective engagement mechanism  103 . An inner slipper ring  104  is coaxially located and fixedly attached to the rotatable input shaft  106 . By mounting the inner slipper ring  104  directly to the input shaft  106 , the need for the inner race  42 , as shown in  FIG. 2 , is obviated. The inner slipper ring  104  has at least one tab  108 , shown in  FIG. 5   a , that extends upwardly into the space defined by slit  110 , shown in  FIG. 5   a , formed in an outer slipper ring  112 . The outer slipper ring  112  is coaxially disposed about the outer periphery of the inner slipper ring  104 . A plurality of arcs, similar to reference number  20  of  FIG. 2   a , are formed on the outside periphery of the inner slipper ring  104  and a complimentary number of arcs, similar to reference number  22  of  FIG. 2   a , are formed on the inside diameter of the outer slipper ring  112 . The plurality of arcs form a like number of cavities, similar to reference number  16  of  FIG. 2   a , within which a roller element  114  is disposed. Two spring-biased pins  116  and  118  are disposed within gear  120  and are operable to selectively engage and disengage the slipper ring clutch  100 . The selective engagement is accomplished by the movement of fork  122  forcing the selection sleeve  124  to insert pin  116  into the space formed in the slit  110  by the tab  108  and the edge of the outer slipper ring  112 , as shown in  FIG. 5   a . In this state of engagement, the slipper ring clutch  100  will be engaged in only one rotational direction, while “freewheeling” in the other. The slipper ring clutch  100  may be made to “freewheel” in both directions by the engagement of both selector sleeves  124  and  124 ′. This condition will force pins  116  and  118  into the space formed from slit  110  by the tabs  108  and the opposite faces of the outer slipper ring  112 . Upon disengagement of both selector sleeves  124  and  124 ′ and subsequent extraction of pins  116  and  118 , the slipper ring clutch  100  will be engaged in both rotational directions. 
     The selector sleeves  124  and  124 ′ may be made from stamped metal due to the light forces that act upon them. Additionally, a detent feature  126  may be provided to the selector sleeve  124  and the gear  120  to retain the selector sleeve  124  in the desired state. The selector sleeve  124  may be formed from one or two parts, and may be manipulated by one or two shift forks  122  and  122 ′. The various states of operation described with reference to  FIG. 1  are possible with the present embodiment. This alternate embodiment may be utilized in any of the previously discussed transmission architectures. The detent mechanisms  126  and  126 ′ may also include a spring mounted ball operable to engage recesses formed within the selector sleeves  124  and  124 ′, similar to that shown in  FIG. 2 . This engagement will ensure that the desired mode of engagement of slipper ring clutch  100  is maintained. 
       FIGS. 6   a  and  6   b  are schematic diagrams of a slipper ring clutch  10  employing a bi-metallic spring  136  and illustrating two states of operation. Modern transmissions must operate in extreme conditions, such as temperature. Transmission fluid has a high viscosity, or resistance to flow, at low temperatures, which may cause a hydrodynamic film to develop at the interface between the outer slipper ring  12  and outer race  26 . With the development of this film, there may be insufficient friction available to enable the roller element  18  to effectively expand the outer slipper ring  12  and frictionally engage the inner surface  24  of the outer race  26 . An additional force may be required to shear the cold fluid form this interface. A solution may be to employ a drag producing device such as the bi-metallic spring  136 . The bimetallic spring  136  will impart a drag force between the outer slipper ring  12  and the outer race  26  at low temperatures. The variance in thermal expansion coefficients between the two metals used in the bimetallic spring  136  will enable a change in shape at higher operating temperatures thereby reducing the drag as the fluid temperature increases and the viscosity of the transmission fluid decreases. 
       FIGS. 6   a  and  6   b  illustrate one possible design for a drag inducing bi-metallic spring  136 . The bi-metallic spring  136  is disposed between the cylindrical inside surface  24  of the outer race  26  and the outside diameter of the outer slipper ring  12  and has at least one extension  138  projecting radially inward and engaging a groove  140  formed on the outside diameter of the outer slipper ring  12 . At low temperatures, the differences in thermal expansion between the two metal layers constituting the bi-metallic spring  136  will force the bi-metallic spring  136  into an oblong or oval shape, as shown in  FIG. 6   a . This oblong shape will cause the bi-metallic spring  136  to exert a drag force by frictionally engaging the cylindrical inner surface  24  of the outer race  26 . At higher temperatures, the drag forces within the slipper ring clutch  10  will decrease as the bi-metallic spring  136  returns to a nearly circular shape and disengages the outer race  26 , as shown in  FIG. 6   b.    
       FIGS. 7   a ,  7   b , and  7   c  are schematic diagrams illustrating the implementation of a slipper ring clutch  10  in place of the hydraulic plate-type clutch and a one-way clutch first forward range/reverse system of an automatically shiftable transmission. The slipper ring clutch  10  can provide a mode of engagement in which the system can be engaged in both the clockwise and counter-clockwise rotational directions. This attribute makes the slipper ring clutch  10  particularly useful to provide engine braking in the first forward range. Engine braking is a method of decelerating a vehicle using the work expended by the internal combustion engine during the compression cycles. Implementation of engine braking in combination with a conventional friction based brake system increases the lifespan of the conventional brake system by reducing wear and heat generation. The slipper ring clutch  10  in this embodiment may require less packaging space as well as reduced control complexity. 
       FIGS. 7   a ,  7   b , and  7   c  show an automatically shiftable transmission  150  in lever diagram form as will be readily understood by those skilled in the art. More precisely,  FIGS. 7   a ,  7   b , and  7   c  show a series of lever diagrams representing the automatically shiftable transmission  150  during the first forward range of operation, the second forward range of operation, and the first forward range of operation with engine braking enabled, respectively. The lever diagrams of  FIGS. 7   a ,  7   b , and  7   c  each include a vertical line  152  having a first second, third, and fourth node A, B, C, and D, respectively, and a diagonal line  154 . The vertical line  152  generally represents the automatically shiftable transmission  150  at rest, while the diagonal line  154  represents the dynamic state of the automatically shiftable transmission  150 . The nodes A, B, C, and D, represent planetary gear members, which are known in the art. 
     Referring to  FIG. 7   a , a lever diagram representing the automatically shiftable transmission  150  operating in the first forward range is shown. The slipper ring clutch  10  engages such that the relative speed across the slipper ring clutch is zero, therefore the rotational speed at node C is zero. The engine, not shown, provides an input torque T 1  such that the rotation speeds of nodes A and B are V 1  and V 2 , respectively in the clockwise direction. Additionally, the rotational speed of node D is V 3  in the counterclockwise direction. A torque T 2  is provided by the slipper ring clutch to ensure that the rotational speed of node C remains zero. A torque T 3  is provided to node B by the final drive, not shown, of the vehicle. The torque T 3  is a reaction torque to the torques T 1  and T 2 . The key  34  is inserted in the slipper ring clutch  10  to allow the slipper ring clutch  10  to freewheel in the direction opposite the torque T 2 . 
     Referring to  FIG. 7   b , a lever diagram representing the automatically shiftable transmission  150  operating in the second forward range is shown. In this mode of operation, the slipper ring clutch  10  is disengaged, while a reaction clutch  156  is engaged such that the rotational velocity of node D is zero. The engine, not shown, provides an input torque T 4  such that the rotation speeds of nodes A, B, and C are V 4 , V 5 , and V 6 , respectively in the clockwise direction. A torque T 5  is provided by the reaction clutch  156  to ensure that the rotational speed of node D remains zero. A torque T 6  is provided to node B by the final drive, not shown, of the vehicle. The torque T 6  is a reaction torque to the torques T 4  and T 5 . 
     Referring now to  FIG. 7   c , a lever diagram representing the automatically shiftable transmission  150  operating in the first forward range engine braking condition is shown. The slipper ring clutch  10  engages such that the relative speed across the slipper ring clutch is zero, therefore the rotational speed at node C is zero. The final drive, not shown, in slowing the vehicle provides an input torque T 1  such that the rotation speeds of nodes A and B are V 7  and V 8 , respectively in the clockwise direction. Additionally, the reaction clutch  156  disengages to allow the node D to rotate with a rotational speed of V 9  in the counterclockwise direction. A torque T 9  is provided by the slipper ring clutch to ensure that the rotational speed of node C remains zero. A torque T 8  is provided to node A by the engine, not shown, of the vehicle. The torques T 8  and T 9  are the reaction torques to the torque T 7 . The key  34  is removed from the slipper ring clutch  10  to allow the slipper ring clutch  10  to lock or engage in both directions. By operating the slipper ring clutch engaged in both directions the torque T 8  may be transferred to the engine, and the corresponding resistance to the transfer of torque T 8  by the engines compressive forces, a measure of engine braking is provided. 
     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.