Patent Publication Number: US-6991079-B2

Title: Power transfer device with hydraulically-actuated clutch assembly

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
   This application is a continuation of U.S. patent application Ser. No. 10/769,646 filed Jan. 30, 2004. 

   FIELD OF THE INVENTION 
   The present invention relates generally to power transfer systems for controlling the distribution of drive torque between the front and rear drivelines of a four-wheel drive vehicle and/or the left and right wheels of an axle assembly. More particularly, the present invention is directed to a power transmission device for use in motor vehicle driveline applications having a torque transfer mechanism equipped with a power-operated clutch actuator that is operable for controlling actuation of a multi-plate friction clutch. 
   BACKGROUND OF THE INVENTION 
   In view of increased demand for four-wheel drive vehicles, a plethora of power transfer systems are currently being incorporated into vehicular driveline applications for transferring drive torque to the wheels. In many vehicles, a power transmission device is operably installed between the primary and secondary drivelines. Such power transmission devices are typically equipped with a torque transfer mechanism for selectively and/or automatically transferring drive torque from the primary driveline to the secondary driveline to establish a four-wheel drive mode of operation. For example, the torque transfer mechanism can include a dog-type lock-up clutch that can be selectively engaged for rigidly coupling the secondary driveline to the primary driveline to establish a “part-time” four-wheel drive mode. When the lock-up clutch is released, drive torque is only delivered to the primary driveline for establishing a two-wheel drive mode. 
   A modern trend in four-wheel drive motor vehicles is to equip the power transmission device with an adaptively controlled transfer clutch in place of the lock-up clutch. The transfer clutch is operable for automatically directing drive torque to the secondary wheels, without any input or action on the part of the vehicle operator, when traction is lost at the primary wheels for establishing an “on-demand” four-wheel drive mode. Typically, the transfer clutch includes a multi-plate clutch assembly that is installed between the primary and secondary drivelines and a clutch actuator for generating a clutch engagement force that is applied to the clutch assembly. The clutch actuator can be a power-operated device that is actuated in response to electric control signals sent from an electronic controller unit (ECU). The electric control signals are typically based on changes in current operating characteristics of the vehicle (i.e., vehicle speed, interaxle speed difference, acceleration, steering angle, etc.) as detected by various sensors. Thus, such “on-demand” transfer clutch can utilize adaptive control schemes for automatically controlling torque distribution during all types of driving and road conditions. Such adaptively controlled transfer clutches can also be used in association with a center differential operably installed between the primary and secondary drivelines for automatically controlling interaxle slip and torque biasing in a full-time four-wheel drive application. 
   A large number of adaptively controlled transfer clutches have been developed with an electro-mechanical clutch actuator that can regulate the amount of drive torque transferred to the secondary driveline as a function of the electric control signal applied thereto. In some applications, the transfer clutch employs an electromagnetic clutch as the power-operated clutch actuator. For example, U.S. Pat. No. 5,407,024 discloses an electromagnetic coil that is incrementally activated to control movement of a ball-ramp drive assembly for applying a clutch engagement force to the multi-plate clutch assembly. Likewise, Japanese Laid-open Patent Application No. 62-18117 discloses a transfer clutch equipped with an electromagnetic clutch actuator for directly controlling actuation of the multi-plate clutch pack assembly. Also, U.S. Pat. No. 6,158,561 discloses use of an electromagnetic actuator for engaging a pilot clutch which, in turn, controls energization of a ball ramp unit for engaging the main clutch. 
   As an alternative to such electromagnetic clutch actuation systems, the transfer clutch can employ an electric motor and a mechanical drive assembly as the power-operated clutch actuator. For example, U.S. Pat. No. 5,323,871 discloses a transfer clutch equipped with an electric motor that controls rotation of a sector plate which, in turn, controls pivotal movement of a lever arm that is operable for applying the clutch engagement force to the multi-plate clutch assembly. Likewise, Japanese Laid-open Patent Application No. 63-66927 discloses a transfer clutch which uses an electric motor to rotate one cam plate of a ball-ramp operator for engaging the multi-plate clutch assembly. Finally, U.S. Pat. Nos. 4,895,236 and 5,423,235 respectively disclose a transfer clutch having an electric motor which drives a reduction gearset for controlling movement of a ball screw operator and a ball-ramp operator which, in turn, apply the clutch engagement force to the clutch assembly. 
   In contrast to the electro-mechanical clutch actuators discussed above, it is also well known to equip the transfer clutch with an electro-hydraulic clutch actuator. For example, U.S. Pat. Nos. 4,862,769 and 5,224,906 generally disclose use of an electric motor or solenoid to control the fluid pressure exerted by an apply piston on a multi-plate clutch assembly. In addition, U.S. Pat. No. 6,520,880 discloses a hydraulic actuation system for controlling the fluid pressure supplied to a hydraulic motor arranged which is associated with a differential gear mechanism in a drive axle assembly. 
   While many adaptive clutch actuation systems similar to those described above are currently used in four-wheel drive vehicles, a need exists to advance the technology and address recognized system limitations. For example, the size and weight of the friction clutch components and the electrical power requirements of the clutch actuator needed to provide the large clutch engagement loads make many systems cost prohibitive for use in most four-wheel drive vehicle applications. In an effort to address these concerns, new technologies are being developed for use in power-operated clutch actuator applications. 
   SUMMARY OF THE INVENTION 
   Thus, its is an objective of the present invention to provide a power transmission device for use in a motor vehicle having a torque transfer mechanism equipped with a unique power-operated clutch actuator that is operable to control engagement of a multi-plate clutch assembly. 
   As a related objective of the present invention, the torque transfer mechanism is well-suited for use in motor vehicle driveline applications to control the transfer of drive torque between first and second rotary members. 
   According to each preferred embodiment of the present invention, a torque transfer mechanism and an electrohydraulic control system are disclosed for adaptively controlling the transfer of drive torque between first and second rotary members in a power transmission device of the type used in motor vehicle driveline applications. The torque transfer mechanism includes a main clutch assembly operably disposed between the first and second rotary members, a pilot clutch assembly, and a rotary clutch operator. The rotary clutch operator includes a first and second coaxially aligned components defining a plurality of actuation chambers therebetween. During operation, the magnitude of the fluid pressure delivered by the hydraulic control system to the actuation chambers controls angular movement of the second component relative to the first component. Such relative angular movement controls energization of the pilot clutch assembly which, in turn, controls the magnitude of the compressive clutch engagement force applied to the main clutch assembly, thereby controlling the drive torque transferred from the first rotary member to the second rotary member. 
   According to another feature of the present invention, the electrohydraulic control system includes a fluid pump, an electric motor for driving the pump, vehicle sensors for detecting various operating characteristics of the motor vehicle, and an electronic control unit (ECU) for receiving input signals from the vehicle sensors and controlling energization of the electric motor. The ECU further controls actuation of a control valve for controlling the magnitude of the fluid pressure supplied to the actuation chambers. A pressure sensor provides a pressure signal to the ECU that is indicative of the fluid pressure in the actuation chambers. 
   The torque transfer mechanism of the present invention is adapted for use in a power transmission device for adaptively controlling the drive torque transferred between a primary driveline and a secondary driveline. According to one preferred application, the power transmission device of the present invention is a transfer case with the torque transfer mechanism arranged as a torque transfer coupling for providing on-demand torque transfer from the primary driveline to the secondary driveline. In a related application, the torque transfer mechanism is arranged as a torque bias coupling for varying the torque distribution and limiting interaxle slip between the primary and secondary driveline. According to another preferred application, the power transmission device is a drive axle assembly with the torque transfer mechanism arranged as a torque bias coupling to control speed differentiation and torque distribution across a differential unit. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
     Further objects, features and advantages of the present invention will become apparent to those skilled in the art from analysis of the following written description, the appended claims, and accompanying drawings in which: 
       FIG. 1  illustrates the drivetrain of a four-wheel drive vehicle equipped with a power transmission device according to the present invention; 
       FIG. 2  is a sectional view of a transfer case associated with the drivetrain shown in  FIG. 1  and which is equipped with a torque transfer mechanism according to a preferred embodiment of the present invention; 
       FIGS. 3A and 3B  are enlarged partial views taken from  FIG. 2  showing components of the torque transfer mechanism is greater detail; 
       FIG. 4  is a partial sectional view of a rotary operator mechanism associated with the torque transfer mechanism of the present invention; 
       FIG. 5  is a schematic diagram of a hydraulic control circuit associated with the torque transfer mechanism of the present invention; 
       FIG. 6  is a schematic illustration of an alternative driveline for a four-wheel drive motor vehicle equipped with a power transmission device of the present invention; 
       FIG. 7  is a schematic illustration of a drive axle assembly associated with the drivetrain shown in  FIG. 6  and equipped with a torque transfer mechanism according to the present invention; 
       FIG. 8  is a schematic illustration of an alternative drive axle assembly operable for use with either of the drivetrain shown in  FIGS. 1 and 6 ; 
       FIG. 9  is a schematic illustration of another alternative embodiment of a power transmission device according to the present invention; 
       FIG. 10  illustrates another alternative drivetrain arrangement for a four-wheel drive motor vehicle equipped with another power transmission device embodying the present invention; 
       FIGS. 11 through 14  schematically illustrate different embodiments of the power transmission device shown in  FIG. 10 ; and 
       FIG. 15  is a schematic illustration of an alternative construction for the power transmission device shown in  FIG. 2 . 
   

   DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
   The present invention is directed to a torque transfer mechanism that can be adaptively controlled for modulating the torque transferred from a first rotary member to a second rotary member. The torque transfer mechanism finds particular application in power transmission devices for use in motor vehicle drivelines such as, for example, a torque transfer clutch in a transfer case, a power take-off unit or an in-line torque coupling, a torque biasing clutch associated with a differential unit in full-time transfer cases or power take-off units or in a drive axle assembly, or any other possible torque transfer application. Thus, while the present invention is hereinafter described in association with particular power transmission devices for use in specific driveline applications, it will be understood that the arrangements shown and described are merely intended to illustrate embodiments of the present invention. 
   With particular reference to  FIG. 1  of the drawings, a drivetrain  10  for a four-wheel drive vehicle is shown. Drivetrain  10  includes a primary driveline  12 , a secondary driveline  14 , and a powertrain  16  for delivering rotary tractive power (i.e., drive torque) to the drivelines. In the particular arrangement shown, primary driveline  12  is the rear driveline while secondary driveline  14  is the front driveline. Powertrain  16  includes an engine  18 , a multi-speed transmission  20 , and a power transmission device hereinafter referred to as transfer case  22 . Rear driveline  12  includes a pair of rear wheels  24  connected at opposite ends of a rear axle assembly  26  having a rear differential  28  coupled to one end of a rear prop shaft  30 , the opposite end of which is coupled to a rear output shaft  32  of transfer case  22 . Likewise, front driveline  14  includes a pair of front wheels  34  connected at opposite ends of a front axle assembly  36  having a front differential  38  coupled to one end of a front prop shaft  40 , the opposite end of which is coupled to a front prop shaft  42  of transfer case  22 . 
   With continued reference to the drawings, drivetrain  10  is shown to further include an electronically-controlled power transfer system  44  for permitting a vehicle operator to select between a two-wheel drive mode, a locked (“part-time”) four-wheel drive mode, and an adaptive (“on-demand”) four-wheel drive mode. In this regard, transfer case  22  is equipped with a transfer clutch  50  that can be selectively actuated for transferring drive torque from rear output shaft  32  to front output shaft  42  for establishing both of the part-time and on-demand four-wheel drive modes. Power transfer system  44  further includes an electrohydraulic clutch actuator  52  for actuating transfer clutch  50 , vehicle sensors  54  for detecting certain dynamic and operational characteristics of the motor vehicle, a mode select mechanism  56  for permitting the vehicle operator to select one of the available drive modes, and an electronic control unit (ECU)  58  for controlling actuation of clutch actuator  52  in response to input signals from vehicle sensors  54  and mode selector  56 . 
   Transfer case  22  is shown in  FIG. 2  to include a multi-piece housing  60  from which rear output shaft  32  is rotatably supported by a pair of laterally-spaced bearing assemblies  62 . Rear output shaft  32  includes an internally-splined first end segment  64  adapted for connection to the output shaft of transmission  20  and a yoke assembly  66  secured to its second end segment  68  that is adapted for connection to rear propshaft  30 . Front output shaft  42  is likewise rotatably supported from housing  60  by a pair of laterally-spaced bearing assemblies  70  and includes a yoke-type end segment  72  that is adapted for connection to front propshaft  40 . 
   In general, transfer clutch  50  and electrohydraulic clutch actuator  52  define a torque transfer mechanism according to the preferred embodiment of the present invention. Transfer clutch  50  includes a main clutch assembly  74  and a pilot clutch assembly  76 . Main clutch assembly  74  is shown to include a first multi-plate friction clutch  78  and a first ball ramp unit  80 . Likewise, pilot clutch assembly  76  is shown to include a second multi-plate friction clutch  82  and a second ball ramp unit  84 . First friction clutch  78  includes a hub  86  fixed (i.e., splined) for rotation with rear output shaft  32 , a drum  88  and a multi-plate clutch pack  90  that is operably disposed between hub  84  and drum  88 . Clutch pack  90  includes a set of outer clutch plates  92  splined for rotation with drum  88  and which are interleaved with a set of inner clutch plates  94  splined for rotation with hub  84 . As will be detailed, clutch actuator  52  is operable for causing a compressive clutch engagement force to be exerted on clutch pack  90 . Such engagement of clutch pack  90  causes rotary power (“drive torque”) to be transferred from rear output shaft  32  to front output shaft  42  through a transfer assembly  96 . Transfer assembly  96  includes a first sprocket  98  fixed (i.e., splined) for rotation with drum  88 , a second sprocket  100  fixed (i.e., splined) for rotation with front output shaft  42 , and a power chain  102  encircling sprockets  98  and  100 . First sprocket  98  is shown fixed to an end plate segment  104  of drum  88  and is rotatably supported on rear output shaft  32  via a suitable bearing assembly  106 . A thrust bearing  108  is shown disposed between first sprocket  102  and a lock ring  109  fixed to rear output shaft  32 . 
   First ball ramp unit  80  includes a first cam member  110 , a second cam member  112  and rollers  114 . First cam member  110  is splined for common rotation with drum  88  and bi-directional translational movement relative to clutch pack  90 . Specifically, first cam member  110  is axially moveable between a first or “released” position and a second or “locked” position. In its released position, first cam member exerts a minimum clutch engagement force on clutch pack  90  such that virtually no drive torque is transferred from rear output shaft  32  to front output shaft  42 , thereby establishing the two-wheel drive mode. In contrast, movement of first cam member  110  to its locked position causes a maximum clutch engagement force to be exerted on clutch pack  90  such that front output shaft  42  is, in effect, coupled for common rotation with rear output shaft  32 , thereby establishing the part-time four-wheel drive mode. Accordingly, variable control of the movement of first cam member  110  between its released and locked position results in adaptive regulation of the drive torque transferred to front output shaft  42 , thereby establishing the on-demand four-wheel drive mode. 
   Second cam member  112  of first ball ramp unit  80  is operably connected to second friction clutch  82 . In addition, rollers  114  are disposed in a cam channel defined between cam tracks  116  formed in first cam member  110  and cam tracks  118  formed in second cam member  112 . Preferably, a plurality of such cam channels are provided and which are each configured as an oblique section of a helical torus. Balls  114  and cam tracks  116 , 118  may be replaced with alternative components and/or tapered ramp profiles that functions to cause axial movement of first cam member  110  in response to relative angular movement between the cam members. In any arrangement, the load transferring components can not be self-locking or self-engaging so as to permit fine control over the translational movement of first cam member  110  for providing precise control of the clutch engagement force applied to clutch pack  90 . A thrust bearing assembly  120  is disposed between second cam member  112  and a retainer plate  122  that is splined to drum  88 . A lock ring  124  axially locates retainer plate  122  for preventing axial movement of second cam member  112 . 
   Second friction clutch  82  includes a multi-plate clutch pack  128  that is operably disposed between second cam member  112  of first ball ramp unit  80  and hub  86  of first friction clutch  78 . Clutch pack  128  includes a set of outer plates  130  splined for rotation with second cam member  112  and which are interleaved with a set of inner clutch plates  132  splined for rotation with hub  86 . Second ball ramp unit  84  includes a first cam ring  134 , a second cam ring  136 , and rollers  138 . First cam ring  134  is fixed to a pressure plate  140  which, in turn, is splined for rotation with hub  86  of first friction clutch  78 . Rollers  138  are disposed in cam channels defined between cam tracks  142  formed in first cam ring  136  and cam tracks  144  formed in second cam ring  136 . Again, it is preferred that a plurality of such cam channels be provided between the cam rings with each being configured as an oblique section of a torus. Furthermore, second ball ramp unit  84  is also not self-locking or self-engaging to permit accurate control of bi-directional translational movement of pressure plate  140  relative to clutch pack  128  that is caused in response to relative rotation between cam rings  134  and  136 . A thrust bearing  146  is disposed between second cam ring  136  and a retainer ring  148  which, in turn, is rigidly secured for rotation with hub  86  via bolts  150 . Such translational movement of pressure plate  140  is operable for controlling the magnitude of pilot actuation force exerted on clutch pack  128  which, in turn, controls energization of first ball ramp unit  80 . With pressure plate  140  in a first or “retracted” position, a minimum pilot actuation force is exerted on clutch pack  128  such that first and second cam members of first ball ramp unit  80  are permitted to rotate together, thereby maintaining first cam member  110  in its released position. In contrast, movement of pressure plate  140  to a second or “extended” position causes a maximum pilot actuation force to be exerted on clutch pack  128  which, in turn, causes second cam member  112  to rotate relative to first cam member  110 . Such relative rotation results in axial movement of first cam member  110  to its locked position. 
   To provide means for moving pressure plate  140  between its retracted and extended positions, clutch actuator  52  generally includes a rotary operator  152  and an electrohydraulic power unit  154 . Power unit  154  is secured to housing  60  and includes an electric motor  156  and a fluid pump  158 . Rotary operator  152  is shown to include a first or “reaction” ring  160  that is concentrically aligned with a second or “actuator” ring  162 . The rings are retained between clutch hub  86  and retainer ring  148 . As seen, bolts  150  also pass through mounting bores  164  in reaction ring  160  such that reaction ring  160  is fixed to hub  86  for common rotation with rear output shaft  32 . 
   As best seen from  FIG. 4 , reaction ring  160  includes a cylindrical body segment  166  and a plurality of radially outwardly projecting lugs  168 . Lugs  168  define a complementary number of longitudinally extending channels  170  therebetween with a like number of radial inlet ports  172  extending through body segment  166  and communicating with channels  170 . Actuator ring  162  also has a cylindrical body segment  174  that is fixed via a spline connection  176  to second cam ring  136  of second ball ramp unit  84 . In addition, a plurality of radially projecting lugs  180  extend inwardly from body segment  174 . Each lug  180  extends into a corresponding one of channels  170  so as to define a series of actuation chambers  182  delimited by a face surface  184  of lugs  168  and a face surface  186  of lugs  180 . Likewise, a distal end surface  188  on each lug  168  is in sliding engagement with an inner wall surface  190  of body segment  174  while a distal end surface  192  on each lug  180  is in sliding engagement with outer wall surface  194  of body segment  166  so as to further delimit each actuation chamber  182 . 
   As noted, reaction ring  160  includes a series of inlet ports  172  that are in communication with actuation chambers  182 . As will be described, increasing the fluid pressure delivered through inlet ports  172  to actuation chambers  182  causes actuator ring  162  to move (i.e., index) in a first rotary direction (i.e., clockwise) relative to reaction ring  160  for energizing pilot ball ramp unit  84  which, in turn, causes corresponding movement of pressure plate  140  toward its extended position, In contrast, decreasing the fluid pressure in actuation chambers  182  causes actuator ring  162  to move in a second rotary direction (i.e., counterclockwise) relative to reaction ring  160  for de-energizing pilot ball ramp unit  84  which, in turn, causes corresponding movement of pressure plate  140  toward its retracted position. 
   Main ball ramp unit  80  further includes a torsional return spring  196  that is operably connected between first cam member  110  and second cam member  112 . Return spring  196  functions to angularly bias the cam members for moving first cam member  110  toward its released position so as to de-energize main ball ramp unit  80 . Such angular movement between the cam members due to the biasing of return spring  196  also results in rearward translation of first cam ring  134  toward its retracted position for de-energizing pilot ball ramp unit  84 . The resulting angular movement of second cam ring  136  relative to first cam ring  134  acts to index actuator ring  162  in the second direction relative to reaction ring  160  toward a first or “low pressure” position, as is shown in  FIG. 4 . Rotary operator  152  is designed to provide fluid leakage paths which permit fluid in actuation chambers  182  to leak out at a predetermined rate so as to permit the biasing force of return spring  196  to angularly bias actuator ring  162  to move toward its low pressure position. 
   In operation, the delivery of fluid to actuation chambers  182  causes actuator ring  162  to rotate relative to reaction ring  160  in the first direction from its low pressure position toward a second or “high pressure” position which, in turn, results in corresponding relative rotation between cam rings  134  and  136  for moving first cam ring  134  from its retracted position toward its extended position. In essence, such angular movement of actuator ring  162  acts to initiate energization of pilot ball ramp unit  84 . Accordingly, the profile of cam tracks  142  and  144  establishes the resultant amount of forward axial movement of first cam ring  134  which causes pressure plate  140  to exert a corresponding pilot actuation force on clutch pack  128 . Engagement of clutch pack  128  effectively couples second cam member  112  of main ball ramp unit  80  for rotation with hub  86  and rear output shaft  32 . This action results in relative rotation between the cam members  110  and  112  and translational movement of first cam member  110  toward its locked position. 
   With pressure plate  140  in its retracted position, first cam member  110  is located in its released position such that virtually no drive torque is transferred from rear output shaft  32  to front output shaft  42  through transfer clutch  50 , thereby effectively establishing the two-wheel drive mode. In contrast, movement of pressure plate  140  to its extended position causes corresponding movement of member  110  to its locked position. As such, a maximum amount of drive torque is transferred to front output shaft  42  for, in effect, coupling front output shaft  42  for common rotation with rear output shaft  32 , thereby establishing the part-time four-wheel drive mode. Accordingly, controlling the position of pressure plate  140  between its retracted and extended positions permits variable control of the amount of drive torque transferred from rear output shaft  32  to front output shaft  42 , thereby establishing the on-demand four-wheel drive mode. Thus, the magnitude of the fluid pressure supplied to actuation chambers  182  controls the angular position of actuator ring  162  relative to reaction ring  160  for controlling actuation of pilot ball ramp unit  84  and, in turn, axial movement of pressure plate  120  between its retracted and extended positions. 
   A hydraulic flow circuit is provided within transfer case  22  for supplying fluid from pump  158  to actuation chambers  182 . Referring initially to  FIG. 5 , a schematic of the hydraulic flow circuit will be described. Specifically, hydraulic fluid from a source of fluid or sump  200  maintained with transfer case housing  60  is drawn through a first flow path  202  to an inlet of pump  158 . Actuation of motor  156  controls the magnitude of the line pressure delivered through a second flow path  204  from an outlet of pump  158  to an inlet of an electrically-actuated control valve  206 . Control valve  206  includes a moveable valve element  208  (see  FIG. 3B ) that regulates the delivery of fluid from its inlet to at least one of a pair of outlets. It will be understood that any type of electrically-actuated control valve capable of regulating the fluid pressure supplied to actuation chambers  182  can be used. The first outlet supplies fluid to actuation chambers  182  of rotary operator  152  through a third flow path  210  while the second outlet supplies fluid to cool and lubricate clutch pack  90  through a fourth flow path  212 . 
   ECU  58  sends electrical control signals to both electric motor  156  and control valve  206  for accurately controlling the fluid pressure supplied through third flow path  210  to actuation chambers  182  by utilizing a predefined control strategy that is based on the mode signal from mode selector  56  and the sensor input signals from vehicle sensors  54 . A pressure sensor  214  sends a signal to ECU  58  that is indicative of the fluid pressure in actuation chambers  182 . In addition, a temperature sensor  216  sends a signal to ECU  58  that is indicative of the fluid temperature in fourth flow path  212  for permitting improved control over the cooling of clutch pack  90 . Finally, a pressure relief valve  218  is provided for selectively venting fluid from actuation chambers  182  into fourth flow path  212  so as to limit the fluid pressure within actuation chambers  182  to a predetermined maximum pressure value. 
   Referring primarily now to  FIGS. 3A and 3B , the structure associated with transfer case  22  for providing the flow paths schematically shown in  FIG. 5  will now be described in greater detail. As seen, a central chamber  220  is formed in rear output shaft  32  and is sealed via a seal cup  222 . A separator  224  is retained within chamber  220  and includes a cylindrical hub segment  226  and an elongated tube segment  228 . Hub segment  226  has a series of radial flow ports  230  which communicate with a large diameter longitudinal flow port  232  formed in tube segment  228 . In addition, an end portion of tube segment  228  is retained in a small diameter portion of central chamber  220  and has a flange  234  delimiting an intermediate diameter portion of central chamber from its large diameter portion. A ring seal  236  provides a fluid-tight interface between the intermediate and large diameter portions of central chamber  220 . In addition, one or more by-pass ports  238  extend through hub segment  226  of separator  224 . Suitable seal rings  240  provide a fluid-tight seal between radial flow ports  230  and large diameter portion of chamber  220 . 
   First flow path  202  includes an inlet hose  242  which provides fluid communication between the internal fluid sump  200  provided within housing  60  and the inlet to pump  158 . Second flow path  204  includes a flow port  244  which delivers fluid at line pressure to a valve chamber  246  within which valve element  208  is retained. Flow port  244  and valve chamber  246  are formed in a valvebody segment  60 A of housing. Third flow path  210  includes a flow passage  250  formed in valvebody housing segment  60 A which communicates with the first outlet of valve chamber  246 , an annular chamber  252  which communicates with passage  250 , and a series of radial passages  254  formed in rear output shaft  32  which provide fluid communication between chamber  252  and the intermediate diameter portion of central chamber  220 . Radial ports  230  and longitudinal port  232  in separator  224  are also part of third flow path  210  and are in fluid communication with intermediate diameter portion of central chamber  220  via one or more throughbores  256  in tube segment  228 . Third flow path  210  also includes a plurality of radial flow passages  258  formed through rear output shaft  32  which connect radial ports  230  in separator  224  with radial inlet ports  172  extending through body segment  166  of reaction ring  160 . As such, the fluid supplied from valve chamber  246  to the inlet of flow passage  250  will flow through third flow path  210  into actuation chambers  182 . 
   Fourth flow path  212  includes a flow passage  270  in valvebody housing segment  60 A which communicates with the second outlet of valve chamber  246 , an annular chamber  272  which communicates with passage  270 , and a series of radial passages  274  formed in rear output shaft  32  which interconnect chamber  272  to a first chamber  276  formed in large diameter portion of central chamber  220 . First chamber  276  surrounds tube segment  288  of separator  224  and is in fluid communication with a second chamber  278  via by-pass ports  238 . Fourth flow path  212  further includes a series of radial passages  280  formed through rear output shaft  32  which communicate with throughbores  282  formed in clutch hub  86 . As such, low pressure fluid supplied from valve chamber  246  to the inlet of flow passage  270  will flow through this flow path and through the interleaved clutch plates of clutch pack  90  before returning to sump  200 . In this manner, the heat generated within clutch pack  90  can be transferred to the fluid prior to its return to sump  200 . 
   In operation, if the two-wheel drive mode is selected, control valve  206  is de-actuated such that valve element  208  moves to a position where the inlet to flow passage  250  is blocked. As such, the biasing of return spring  196  and the leakage paths within rotary operator  152  cause actuator ring  162  to index relative to reaction ring  160  in the second direction toward its low pressure position, whereby pilot ball ramp unit  84  is de-energized and pressure plate  140  is permitted to return to its retracted position. With pilot clutch  82  released, main ball ramp unit  80  is de-energized such that main clutch  78  is also released. In contrast, upon selection of the part-time four-wheel drive mode, control valve  206  is fully activated to move valve element  208  to a position where flow passage  250  receives the full line pressure from pump  158  through port  244 . This high pressure fluid is delivered through third flow path  210  to actuation chambers  182  for causing actuator ring  162  to index relative to reaction ring  160  in the first direction until located in its high pressure position, whereby pilot ball ramp unit  84  is fully energized and pressure plate  140  is moved to its extended position for fully engaging pilot clutch  82 . As such, main ball ramp unit  80  is energized to move first cam member  110  to its locked position for fully engaging main friction clutch  78 . As stated, the pressure signal sent from pressure sensor  214  to ECU  58  in conjunction with the use of pressure relief valve  218  function to limit the maximum fluid pressure that can be delivered to actuation chambers  182 , thereby preventing damage to clutch pack  90 . 
   When mode selector  52  indicates selection of the on-demand four-wheel drive mode, ECU  58  energizes motor  156  for initiating a fluid pumping action in pump  158  and also controls energization of control valve  206  for supplying a predetermined initial fluid pressure to actuation chambers  182  that results in a slight indexing of actuator ring  162  relative to reaction ring  160  in the first direction. This angular movement causes actuator ring  162  to move from its low pressure position to an intermediate or “ready” position which, in turn, results in main ball ramp unit  80  moving first cam member  110  from its released position to a “stand-by” position. Accordingly, a predetermined minimum amount of drive torque is delivered to front output shaft  42  through transfer clutch  50  in this adapt-ready condition. Thereafter, ECU  58  determines when and how much drive torque needs to be transferred to front output shaft  42  based on the current tractive conditions and/or operating characteristics of the motor vehicle, as detected by sensors  54 . Sensors  54  detect such parameters as, for example, the rotary speed of the output shafts, the vehicle speed and/or acceleration, the transmission gear, the on/off status of the brakes, the steering angle, the road conditions, etc. Such sensor signals are used by ECU  58  to determine a desired output torque value utilizing a control scheme that is incorporated into ECU  58 . This desired torque value is used to actively control actuation of electric motor  156  and control valve  206  to generate a corresponding pressure level in actuation chamber  182 . One non-limiting example of a clutch control scheme and the various sensors used therewith is disclosed in commonly-owned U.S. Pat. No. 5,323,871 which is incorporated by reference herein. 
   In addition to adaptive torque control, the present invention permits automatic release of transfer clutch  50  in the event of an ABS braking condition or during the occurrence of an over-temperature condition. Furthermore, while the control scheme was described based on an on-demand strategy, it is contemplated that a differential or “mimic” control strategy could likewise be used. Specifically, the torque distribution between rear output shaft  32  and front output shaft  42  can be controlled to maintain a predetermined rear/front ratio (i.e., 70:30, 50:50, etc.) so as to simulate the inter-axle torque splitting feature typically provided by a mechanical differential unit. Regardless of the control strategy used, accurate control of the fluid pressure delivered from pump  156  to actuation chambers  182  of rotary operator  152  will result in the desired torque transfer characteristics across transfer clutch  50 . Furthermore, it should be understood that mode select mechanism  56  could also be arranged to permit selection of only two different drive modes, namely the on-demand 4WD mode and the part-time 4WD mode. Alternatively, mode select mechanism  56  could be eliminated such that the on-demand 4WD mode is always operating in a manner that is transparent to the vehicle operator. 
   To illustrate an alternative power transmission device to which the present invention is applicable,  FIG. 6  schematically depicts a front-wheel based four-wheel drivetrain layout  10 ′ for a motor vehicle. In particular, engine  18  drives a multi-speed transmission  20 ′ having an integrated front differential unit  38 ′ for driving front wheels  34  via axle shafts  33 . A transfer or power take-off unit (PTU)  300  is also driven by transmission  20 ′ for delivering drive torque to the input member of a torque transfer mechanism, such as an in-line torque transfer coupling  302 , via a drive shaft  30 ′. Torque transfer coupling  302  is preferably integrated with the components of conventional axle assembly  26  to define a drive axle assembly  26 ′. In particular, the input member of torque coupling  302  is coupled to drive shaft  30 ′ while its output member is coupled to a drive component of rear differential  28  which, in turn, drives rear wheels  24  via axleshafts  25 . Accordingly, when sensors  54  indicate the occurrence of a front wheel slip condition, ECU  58  adaptively controls actuation of torque coupling  302  such that drive torque is delivered “on-demand” to rear wheels  24 . It is contemplated that torque transfer coupling  302  includes a transfer clutch and an electrohydraulic clutch actuator that are similar in both structure and function to the torque transfer mechanism previously described herein. Accordingly, common reference numerals will be used hereinafter to identify components previously described. 
   Referring to  FIG. 7 , torque coupling  302  is schematically illustrated to be operably disposed between drive shaft  30 ′ and rear differential  28 . Rear differential  28  includes a pair of side gears  304  that are connected to rear wheels  24  via rear axle shafts  25 . Differential  28  also includes pinions  306  that are rotatably supported on pinion shafts fixed to a carrier  308  and which mesh with side gears  304 . A right-angled drive mechanism is associated with differential  28  and includes a ring gear  310  that is fixed for rotation with carrier  308  and meshed with a pinion gear  312  that is fixed for rotation with a pinion shaft  314 . Torque coupling  302  is schematically shown to include various components of transfer clutch  50  that are operably disposed between driveshaft  30 ′ and pinion shaft  314 . In particular, transfer clutch  50  is schematically shown to include main friction clutch  78  and main ball ramp unit  80  as well as pilot friction clutch  82  and pilot ball ramp unit  84 . Torque coupling  302  also is shown schematically to include clutch actuator  52  that can be adaptively actuated for controlling the magnitude of the clutch engagement force applied to transfer clutch  50  and thus the amount of drive torque transferred from drive shaft  30 ′ to rear differential  28 . Actuator  52  includes rotary operator  152  and the electrohydraulic control system previously disclosed in  FIG. 5  for adaptively controlling actuation of rotary operator  152 . In this regard, power transfer system  44  is illustrated in block format and is contemplated to include all electrical and hydraulic components and sub-systems required to adaptively control actuation of rotary operator  152 . 
   Torque coupling  302  permits operation in any of the drive modes previously disclosed. For example, if the on-demand 4WD mode is selected, ECU  58  regulates activation of clutch actuator  52  in response to the operating conditions detected by sensors  54  by controllably varying the electric control signal sent motor  128  and control valve  206 . Selection of the part-time 4WD mode results in complete engagement of main clutch pack  90  such that pinion shaft  314  is, in effect, rigidly coupled to driveshaft  30 ′. Finally, in the two-wheel drive mode, main clutch pack  90  is released such that pinion shaft  312  is free to rotate relative to driveshaft  30 ′. Alternatively, elimination of mode select mechanism  56  would provide automatic adaptive operation of torque coupling  302 . 
   The arrangement shown for drive axle assembly  26 ′ of  FIG. 7  is operable to provide on-demand four-wheel drive by adaptively controlling the transfer of drive torque from the primary driveline to the secondary driveline. In contrast, a drive axle assembly  320  is shown in  FIG. 8  wherein torque coupling  302  is now operably installed between differential case  308  and one of axleshafts  25  to provide an adaptive “side-to-side” torque biasing and slip limiting feature. As before, torque coupling  302  is schematically shown to include a transfer clutch  50  and an electrohydraulic clutch actuator  52 , the construction and function of which are understood to be similar to the detailed description previously provided herein for each sub-assembly. 
   Referring now to  FIG. 9 , a drive axle assembly  322  is schematically shown to include a pair of torque couplings  302 L and  302 R that are operably installed between a driven pinion shaft  314  or  30 ′ and axleshafts  25 . The driven pinion shaft drives a right-angled gearset including pinion  312  and ring gear  310  which, in turn, drives a transfer shaft  324 . First torque coupling  302 L is shown disposed between transfer shaft  324  and the left one of axleshafts  25  while second torque coupling  302 R is disposed between transfer shaft  324  and the right axle shaft  25 . Each torque coupling includes a corresponding transfer clutch  50 L,  50 R and electrohydraulic clutch actuator  52 L,  52 R. Accordingly, independent torque transfer and slip control is provided between the driven pinion shaft and each rear wheel  24  pursuant to this arrangement. 
   To illustrate additional alternative power transmission devices to which the present invention is applicable,  FIG. 10  schematically depicts a front-wheel based four-wheel drive drivetrain layout  10 ″ for a motor vehicle. In particular, engine  18  drives multi-speed transaxle  20 ′ which has an integrated front differential unit  38 ′ for driving front wheels  34  via axle shafts  33 . As before, PTU  300  is also driven by transaxle  20 ′ for delivering drive torque to the input member of a torque transfer coupling  330 . The output member of torque transfer coupling  330  is coupled to propshaft  30 ′ which, in turn, drives rear wheels  24  via axleshafts  25 . Rear axle assembly  26  can be a traditional driven axle with a differential or, in the alternative, be similar to the drive axle arrangements described in regard to  FIG. 8  or  9 . Accordingly, in response to detection of certain vehicle characteristics by sensors  54  (i.e., the occurrence of a front wheel slip condition), power transfer system  44  causes torque coupling  330  to deliver drive torque “on-demand” to rear wheels  24 . It is contemplated that torque coupling  330  would be generally similar in structure and function to that of torque transfer coupling  302  previously described herein. As such, its primary components of transfer clutch  50  and electrohydraulic clutch actuator  52  are again schematically shown. 
   Referring now to  FIG. 11 , torque coupling  330  is schematically illustrated in association with an on-demand four-wheel drive system based on a front-wheel drive vehicle similar to that shown in  FIG. 10 . In particular, an output shaft  332  of transaxle  20 ′ is shown to drive an output gear  334  which, in turn, drives an input gear  336  that is fixed to a carrier  338  associated with front differential unit  38 ′. To provide drive torque to front wheels  34 , front differential unit  38 ′ includes a pair of side gears  340  that are connected to front wheels  34  via axleshafts  33 . Differential unit  38 ′ also includes pinions  342  that are rotatably supported on pinion shafts fixed to carrier  338  and which are meshed with side gears  340 . A transfer shaft  344  is provided for transferring drive torque from carrier  338  to a clutch hub  84  associated with transfer clutch  50 . PTU  300  is a right-angled drive mechanism including a ring gear  346  fixed for rotation with drum  88  of transfer clutch  50  and which is meshed with a pinion gear  348  fixed for rotation with propshaft  30 ′. According to the present invention, the components schematically shown for torque transfer coupling  330  are understood to be similar to those previously described. In operation, drive torque is adaptively transferred on-demand from the primary (i.e., front) driveline to the secondary (i.e., rear) driveline. 
   Referring to  FIG. 12 , a modified version of the power transmission device shown in  FIG. 11  is now shown to include a second torque coupling  330 A that is arranged to provide a limited slip feature in association with primary differential  38 ′. As before, adaptive control of torque coupling  330  provides on-demand transfer of drive torque from the primary driveline to the secondary driveline. In addition, adaptive control of second torque coupling  330 A provides on-demand torque biasing (side-to-side) between axleshafts  33  of primary driveline  14 . 
     FIG. 13  illustrates another modified version of  FIG. 9  wherein an on-demand four-wheel drive system is shown based on a rear-wheel drive motor vehicle that is arranged to normally deliver drive torque to rear wheels  24  while selectively transmitting drive torque to front wheels  34  through a torque coupling  330 . In this arrangement, drive torque is transmitted directly from transmission output shaft  332  to power transfer unit  300  via a drive shaft  350  which interconnects input gear  336  to ring gear  346 . To provide drive torque to front wheels  34 , torque coupling  330  is shown operably disposed between drive shaft  350  and transfer shaft  344 . In particular, transfer clutch  50  is arranged such that drum  88  is driven with ring gear  346  by drive shaft  350 . As such, clutch actuator  52  functions to transfer drive torque from drum  88  through clutch pack  90  to hub  84  which, in turn, drives carrier  338  of differential unit  38 ′ via transfer shaft  344 . 
   In addition to the on-demand four-wheel drive systems shown previously, the power transmission technology of the present invention can likewise be used in full-time four-wheel drive systems to adaptively bias the torque distribution transmitted by a center or “interaxle” differential unit to the front and rear drivelines. For example,  FIG. 14  schematically illustrates a full-time four-wheel drive system which is generally similar to the on-demand four-wheel drive system shown in  FIG. 13  with the exception that an interaxle differential unit  360  is now operably installed between carrier  338  of front differential unit  38 ′ and transfer shaft  344 . In particular, output gear  336  is fixed for rotation with a carrier  362  of interaxle differential  360  from which pinion gears  364  are rotatably supported. A first side gear  366  is meshed with pinion gears  364  and is fixed for rotation with drive shaft  350  so as to be drivingly interconnected to the rear driveline through power transfer unit  300 . Likewise, a second side gear  368  is meshed with pinion gears  364  and is fixed for rotation with carrier  338  of front differential unit  38 ′ so as to be drivingly interconnected to the front driveline. Torque coupling  330  is now shown to be operably disposed between side gears  366  and  368 . Torque coupling  330  is operably arranged between the driven outputs of interaxle differential  360  for providing an adaptive torque biasing and slip limiting function between the front and rear drivelines. 
   Referring now to  FIG. 15 , a full-time 4WD system is shown to include a transfer case  22 ′ which is generally similar to transfer case  22  of  FIG. 2  except that an interaxle differential  380  is provided between an input shaft  382  and output shafts  32  and  42 . As is conventional, input shaft  382  is driven by the output of transmission  20 . Differential  380  includes an input defined as a planet carrier  384 , a first output defined as a first sun gear  386 , a second output defined as a second sun gear  388 , and a gearset for permitting speed differentiation between first and second sun gears  386  and  388 . The gearset includes a plurality of meshed pairs of first planet gears  390  and second pinions  392  which are rotatably supported by carrier  384 . First planet gears  390  are shown to mesh with first sun gear  386  while second planet gears  392  are meshed with second sun gear  388 . First sun gear  386  is fixed for rotation with rear output shaft  32  so as to transmit drive torque to the rear driveline. To transmit drive torque to the front driveline, second sun gear  388  is coupled to transfer assembly  100  which again includes first sprocket  102  rotatably supported on rear output shaft  32 , second sprocket  106  fixed to front output shaft  42 , and power chain  110 . 
   A number of preferred embodiments have been disclosed to provide those skilled in the art an understanding of the best mode currently contemplated for the operation and construction of the present invention. The invention being thus described, it will be obvious that various modifications can be made without departing from the true spirit and scope of the invention, and all such modifications as would be considered by those skilled in the art are intended to be included within the scope of the following claims.