Abstract:
An automated multi-speed transmission includes an engine clutch operable to establish a releasable drive connection between the engine and an input shaft, an output shaft adapted to transfer power to the driveline, and a synchromesh geartrain having a plurality of constant-mesh gearsets that can be selectively engaged to establish a plurality of forward and reverse speed ratios. The transmission also includes power-operated dog clutches for selectively engaging the constant-mesh gearsets, and a controller for controlling coordinated actuation of the engine clutch and the power-operated dog clutches. The power-operated dog clutch associated with the low and the top gear are used during downshifts and upshifts, respectively, to actuate a clutch assembly for synchronizing the speed of the input shaft and the selected gear prior to engagement of its corresponding dog clutch.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application claims the benefit of U.S. Provisional Application No. 60/279,088, filed Mar. 27, 2001. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates generally to hydraulic couplings for use in motor vehicle driveline applications for limiting slip and transferring torque between rotary members. More specifically, a drive axle assembly for an all-wheel drive vehicle is disclosed having a pair of hydraulic couplings each having a fluid pump, a multi-plate clutch assembly, and a fluid distribution system operable to control actuation of the clutch assembly. 
     BACKGROUND OF THE INVENTION 
     In all-wheel drive vehicles, it is common to have a secondary drive axle that automatically receives drive torque from the drivetrain in response to lost traction at the primary drive axle. In such secondary drive axles it is known to provide a pair of clutch assemblies connecting each axleshaft to a prop shaft that is driven by the drivetrain. For example, U.S. Pat. No. 4,650,028 discloses a secondary drive axle equipped with a pair of viscous couplings while U.S. Pat. Nos. 5,964,126, 6,095,939 and 6,155,947 each disclose a secondary drive axle with a pair of pump-actuated multi-plate clutch assemblies. In addition to these passively-controlled drive axles, U.S. Pat. No. 5,699,888 teaches of a secondary drive axle having a pair of multi-plate clutches actuated by electromagnetic actuators that are controlled by an electronic control system. 
     SUMMARY OF THE INVENTION 
     An object of the present invention is to provide a drive axle assembly equipped with a pair of hydraulic couplings which are operably arranged for coupling a vehicle drivetrain to a pair of axleshafts. 
     In carrying out the above object, the drive axle assembly of the present invention includes a drive case that is rotatably supported within a housing and driven by the drivetrain, first and second output shafts rotatably supported by the drive case, and first and second hydraulic couplings operably installed between the drive case and the first and second output shafts. Each hydraulic coupling includes a multi-plate clutch assembly and a clutch actuator. The clutch actuator includes a fluid pump and a piston assembly. Each fluid pump is operable for pumping fluid in response to a speed differential between the drive case and the corresponding output shaft. The piston assembly includes a piston retained for sliding movement in a piston chamber and a multi-function control valve. The fluid pump supplies fluid to the piston chamber such that a clutch engagement force exerted by the piston on the multi-plate clutch assembly is proportional to the fluid pressure in the piston chamber. The control valve is mounted to the piston and provides a pressure relief function for setting a maximum fluid pressure within the piston chamber. The control valve also provides a thermal unload function for releasing the fluid pressure within the piston chamber when the fluid temperature exceeds a predetermined temperature value. 
     In accordance with an optional construction, the multi-function control valve of the present invention can also provide a flow control function for regulating the fluid pressure in the piston chamber. The flow control function can further include a thermal compensation feature for accommodating viscosity variations in the fluid. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     Further objects, features and advantages of the present invention will become readily apparent from the following detailed specification and the appended claims which, in conjunction with the drawings, set forth the best mode now contemplated for carrying out the invention. Referring to the drawings: 
     FIG. 1 is a schematic view of a motor vehicle drivetrain equipped with a secondary drive axle assembly constructed in accordance with the present invention; 
     FIG. 2 is a sectional view of the secondary drive axle assembly; 
     FIG. 3 is an enlarged partial view taken from FIG. 2 showing components of one of the hydraulic couplings in greater detail; 
     FIG. 4 is a schematic diagram illustrating a hydraulic control circuit associated with the secondary drive axle shown in FIG. 2; and 
     FIG. 5 is a sectional view of a multi-function control valve adapted for use with the hydraulic couplings of the secondary drive axle assembly. 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     In general, the present invention is directed to a hydromechanical limited slip and torque transfer device, hereinafter referred to as a drive axle assembly, for use in connecting the drivetrain to a pair of axleshafts associated with a secondary driveline of an all-wheel drive vehicle. However, the drive axle assembly can also find application in other driveline applications including, but not limited to, limited slip differentials of the type used in full-time transfer cases and front-wheel drive transaxles. Furthermore, this invention advances the technology in the field of hydraulically-actuated couplings of the type requiring pressure relief and thermal unloading to prevent damage to the driveline components. 
     With reference to FIG. 1, a schematic layout for a vehicular drivetrain  10  is shown to include a powertrain  12  driving a first or primary driveline  14  and a second or secondary driveline  16 . Powertrain  12  includes an engine  18  and a transaxle  20  arranged to provide motive power (i.e., drive torque) to a pair of wheels  22  associated with primary driveline  14 . Primary driveline  14  further includes a pair of halfshafts  24  connecting wheels  22  to a differential assembly (not shown) associated with transaxle  20 . Secondary driveline  16  includes a power take-off unit (PTU)  26  driven by transaxle  20 , a prop shaft  28  driven by PTU  26 , a pair of axleshafts  30  connected to a pair of wheels  32 , and a drive axle assembly  34  operable to transfer drive torque from propshaft  28  to one or both axleshafts  30 . 
     Referring to FIG. 2, the components associated with drive axle assembly  34  will be now detailed. Drive axle assembly  34  includes a housing  40 , a pinion shaft  42 , and a differential assembly  44 . Pinion shaft  42  is rotatably supported in housing  40  by bearing assemblies  46  and  48  and is sealed relative to housing  40  via a seal assembly  50 . A yoke  52  is secured to pinion shaft  42  and is adapted for connection to prop shaft  28 . A drive pinion  54  is formed at one end of pinion shaft  42 . Differential assembly  44  includes a multi-piece drive case  56  having a carrier drum  58  rigidly secured to a pair of end caps  60  and  62 . A ring gear  64  is fixed (i.e., bolted) to drum  58  and is meshed with drive pinion  54  such that driven rotation of prop shaft  28  causes rotation of drive case  56 . End caps  60  and  62  are shown supported for rotation relative to housing  40  by bearing assemblies  66 . 
     Differential assembly  44  further includes first and second output shafts  68  and  70  adapted for connection to corresponding axleshafts  30 , and first and second hydraulic couplings  72  and  74 . First coupling  72  is operably connected between first output shaft  68  and drive case  56  while second hydraulic coupling  74  is operably connected between second output shaft  70  and drive case  56 . First and second hydraulic couplings  72  and  74  are substantially identical in structure and function. As such, the remainder of this detailed description will be primarily directed to the construction and operation of first hydraulic coupling  72 . However, common components for second hydraulic coupling  74  are identified in the drawings with primed common reference numerals. 
     Hydraulic coupling  72  includes a transfer clutch  80  and a clutch actuator  82 . Transfer clutch  80  is a multi-plate clutch assembly including a clutch hub  84  fixed (i.e., splined) to first output shaft  68  and a clutch pack  86  of interleaved inner and outer clutch plates that are respectively splined to hub  84  and carrier drum  58 . Clutch actuator  82  includes a fluid pump  88  disposed in a pump chamber formed between first end cap  60  and a piston housing  90 , and a piston assembly  92  retained in an annular piston chamber  94  formed in piston housing  90 . As seen, a bearing assembly  98  supports first end cap  60  for rotation relative to first output shaft  68 . 
     Piston assembly  92  is supported for axial sliding movement in piston chamber  94  for applying a compressive clutch engagement force on clutch pack  86 , thereby transferring drive torque and limiting relative rotation between drive case  56  and first output shaft  68 . Similarly, piston assembly  921  of hydraulic coupling  74  functions to exert a clutch engagement force on clutch pack  861  for transferring torque and limiting slip between drive case  56  and second output shaft  70 . The amount of torque transferred is progressive and is proportional to the magnitude of the clutch engagement force exerted by piston assembly  92  on clutch pack  86  which, in turn, is a function of the fluid pressure within piston chamber  94 . Moreover, the fluid pressure generated by pump  88  and delivered to piston chamber  94  is largely a function of the speed differential between drive case  56  and first output shaft  68 . 
     With particular reference to FIGS. 3 and 4, a fluid distribution and valving arrangement is shown for controlling the delivery of fluid to piston chamber  94 . The fluid distribution system includes a first flow path  100  for supplying hydraulic fluid from a sump  102  to an inlet reservoir  104  located at the inlet or suction side of fluid pump  88 , and a second flow path  106  for supplying fluid from the discharge or outlet side of pump  88  to piston chamber  94 . A third flow path  108  extends through piston assembly  92  for venting fluid from piston chamber  94  into a clutch chamber  110  in close proximity to clutch pack  86 . A multi-function control valve  112  forms part of piston assembly  92  and provides at least two functional modes of operation. The first mode, hereinafter referred to as its pressure relief function, is schematically illustrated by a pressure relief valve  114 . The second mode of operation, hereinafter referred to as its thermal unload function, is schematically indicated by a thermal unload valve  116 . With each function, fluid discharged from piston chamber  94  is delivered to clutch chamber  110  for cooling clutch pack  86  and is then returned to sump  102  via a fourth flow path  118 . According to the structure shown, exhaust ports  122  formed in drum  58  define fourth flow path  118 . 
     First flow path  100  is defined by a pair of inlet ports  130  formed through first end cap  60 . A One-way check valve  132  is provided for selectively opening and closing each of inlet ports  130 . Specifically, one-way check valves  132  move between “open” and “closed” positions in response to the direction of pumping action generated by fluid pump  88 . Rotation of the pump components in a first direction acts to open one of check valves  132  and to close the other for permitting fluid to be drawn from sump  102  into inlet reservoir  104 . The opposite occurs in the case of pumping in the reverse rotary direction, thereby assuring bi-directional operation of pump  88 . Check valves  132  are preferably reed-type valves mounted on rivets secured to first end cap  60 . Check valves  132  are of the normally-closed type to maintain fluid within inlet reservoir  104 . 
     A valving arrangement associated with second flow path  106  includes a second pair of one-way check valves  134  that are located in a pair of flow passages  136  formed in piston housing  90  between the outlet of pump  88  and piston chamber  94 . As before, the direction of pumping action establish which of check valves  134  is in its “open” position and which is in its “closed” position to deliver pump pressure to piston chamber  94 . Upon cessation of pumping action, both check valves  134  return to their closed position to maintain fluid pressure in piston chamber  94 . Thus, check valves  134  are also of the normally-closed variety. 
     As noted, fluid pump  88  is operable for pumping hydraulic fluid into piston chamber  94  to actuate transfer clutch  80 . Fluid pump  88  is bi-directional and is capable of pumping fluid at a rate proportional to speed differential between its pump components. In this regard, pump  88  is shown to include a gerotor pump assembly having a pump ring  152  that is fixed (i.e., keyed or splined) to first output shaft  68 , an eccentric ring  154  that is mounted on first end cap  60 , and a stator ring  156  that is operably disposed therebetween. Pump ring  152  has a plurality of external teeth that rotate concentrically relative to first output shaft  68  about a common rotational axis. Stator ring  156  includes a plurality of internal lobes and has an outer circumferential edge surface that is journally supported within a circular internal bore formed in eccentric ring  154 . The internal bore is offset from the rotational axis such that, due to meshing of internal lobes of stator ring  156  with external teeth of pump ring  152 , relative rotation between pump ring  152  and eccentric ring  154  causes eccentric rotation of stator ring  156 . However, fluid pump  88  can be any type of mechanical pump capable of generating pumping action due to a speed differential. 
     Referring now to FIGS. 3 and 4, piston assembly  92  is shown to include a piston  158  and control valve  112 . Piston  158  includes a radial web segment  160  sealed by seal rings  162  and  164  for movement relative to piston housing  90 . Piston  158  further includes an axial rim segment  166  extending from web segment  160  and which engages clutch pack  86 . Piston  158  further defines a cup segment  168  within which control valve  112  is retained. Seal rings  170  are provided to seal control valve  112  relative to cup segment  168  and a circlip  172  is provided to retain control valve  112  in cup segment  168 . Control valve  112  includes a tubular housing  174  defining a series of inlet ports  176  communicating with a pressure chamber  177 , and a valve chamber  178  having a series of outlet ports  180 . Pressure chamber  177  and valve chamber  178  are delineated by a rim section  182  having a central valve aperture formed therethrough. A thermal actuator  186  is retained in pressure chamber  177  of housing  174  and includes a post segment  190  which extends through the valve aperture into valve chamber  178 . A head segment of a valve member  192  is seated against the valve aperture. A spring  196  mounted between an end cap  198  and valve member  192  is operable to bias valve member  192  against the seat surface defined by the valve aperture for normally preventing fluid flow from inlet ports  176  to outlet ports  180 . Control valve  112  is arranged such that inlet ports  176  communicate with piston chamber  94  with valve member  192  directly exposed to the fluid pressure in piston chamber  94 . 
     Hydraulic coupling  72  includes a flow regulator  200  which is operable for setting the predetermined minimum pressure level within piston chamber  94  at which transfer clutch  80  is initially actuated and which is further operable to compensate for temperature gradients caused during heating of the hydraulic fluid. Preferably, flow regulator  200  is a reed-type valve member secured to piston assembly  92  such that its terminal end is normally maintained in an “open” position displaced from a by-pass port  202  formed through piston  158  for permitting by-pass flow from piston chamber  94  to clutch chamber  110 . During low-speed relative rotation, the pumping action of fluid pump  88  causes fluid to be discharged from piston chamber  94  through by-pass port  202  into clutch chamber  110 . Flow regulator  200  is preferably a bimetallic valve element made of a laminated pair of dissimilar metallic strips having different thermal coefficients of expansion. As such, the terminal end of the valve element moves relative to its corresponding by-pass port  202  regardless of changes in the viscosity of the hydraulic fluid caused by temperature changes. This thermal compensation feature can be provided by one or more bimetallic valves. However, once the fluid in piston chamber  94  reaching its predetermined pressure level, the terminal end of the bimetallic valve element will move to a “closed” position for inhibiting fluid flow through by-pass port  202 . This flow restriction causes a substantial increase in the fluid pressure within piston chamber  94  which, in turn, causes piston  158  to move and exert a large engagement force on clutch pack  86 . A bleed slot (not shown) is formed in one of by-pass port  202  or bimetallic valve element and permits a small amount of bleed flow even when the flow regulator is in its closed position for gradually disengaging transfer clutch  80  when fluid pump  88  is inactive. 
     The pressure relief function of control valve  112  occurs when the fluid pressure in piston chamber  94  is greater than that required to close bimetallic flow regulator  200  but less than a predetermined maximum value. In this pressure range, the bias of spring  196  is adequate to maintain valve member  192  seated against the aperture such that fluid is prevented from flowing from piston chamber  94  through outlet ports  180 . However, when the fluid pressure in piston chamber  94  exceeds this maximum value, valve member  192  is forced to move in opposition to the biasing of spring  196 . As such, fluid in piston chamber  94  is permitted to flow through the aperture into valve chamber  178  from where it is discharged from outlet ports  180 . The fluid discharged from outlet ports  180  circulates in clutch chamber  110  to cool clutch pack  86  and is directed to flow across actuator section  210  of thermal actuator  186  prior to discharge to pump through exhaust ports  122  in drum  58 . Use of this pressure relief function torque limits hydraulic coupling  72  and prevents damage thereto. The thermal unload function is actuated when the fluid temperature detected by actuator section  210  of thermal actuator  186  exceeds a predetermined maximum value. In such an instance, post segment  190  moves from its retracted position shown to an extended position for causing valve member  192  to move away from seated engagement against aperture (or maintain valve member  192  in its displaced position during pressure relief) and permit fluid in pressure chamber  92  to vent into clutch chamber  110 , thereby releasing transfer clutch  80 . Once piston chamber  94  has been unloaded, the fluid and thermal actuator  186  will eventually cool to a temperature below the predetermined value, whereby post segment  190  will return to its retracted position for resetting the thermal unload function. Thermal actuator  186  is of a type manufactured by Therm-Omega Tech of Warminster, Pa. or Standard-Thomson of Waltham, Mass. 
     Referring now to FIG. 5, an alternative construction for a control valve assembly  212  is shown. In general, control valve assembly  212  is substantially identical to control valve assembly  112  with the exception that it now provides a third functional mode of operation. Namely, control valve assembly now also provides a flow control function (shown schematically in FIG. 4 as flow control valve  214 ) in addition to its pressure relief and thermal unload functions. In particular, control valve assembly  212  includes one or more bimetallic valves  218  having one end fixed to housing  174  and a second end overlying a flow port  216  formed in rim section  184  of housing  174 . Flow port  216  provides a flow path between inlet ports  176  and outlet ports  180 . Bimetallic valve  218  is adapted to move from an open position displaced from flow port  216  based on the fluid pressure acting thereon. In operation, when the speed differential between drum  58  and first output shaft  68  is less than a certain actuation value, the pressure in piston chamber  94  permits the second end of bimetallic valve  218  to remain in its open position. However, when the speed differential exceeds the actuation valve, the fluid pressure causes the second end of bimetallic valve  218  to move to its closed position. Thus, flow through flow port  216  is inhibited and causes a substantial increase in fluid pressure within piston chamber  94  which, in turn, cases piston assembly  92  to exert a larger engagement force on clutch pack  86 . Once relative rotation ceases, the bleed flow to clutch chamber  110  through the bleed slot functions to reduce the pressure in piston chamber  94  for disengaging transfer clutch  80  and permitting subsequent movement of bimetallic valve  218  to its open position. 
     As a further feather, bimetallic valve  218  is thermally compensating for accommodating temperature gradients caused during cyclical heating and cooling of the hydraulic fluid. More specifically, due to its construction of two metallic strips laminated together having different thermal expansion coefficients, the second end of bimetallic valve  218  is caused to move relative to flow port  216  as its temperature is varied for controlling flow through flow port  216  independent of changes in the viscosity of the hydraulic fluid caused by such temperature variations. In addition to being thermally-compensating, bimetallic valve  218  is also speed dependent for delaying engagement of transfer clutch  80  at higher vehicle speeds. Specifically, the spring function of bimetallic valve  218  provides a centrifugal effect for delaying movement of the second end of bimetallic valve  218  to its closed position as a function of increasing rotary speed. In this regard, the centrifugal effect caused by increasing rotary speed requires a higher flow rate to close bimetallic valve  218 , whereby a greater speed differential is required to overcome the centrifugal resistance and move the second end of bimetallic valve  218  to its closed position. 
     Twin clutch drive axle  34  is operable to control interaxle slip between front driveline  14  and rear driveline  16  and is further adapted to control intra-axle slip between rear wheels  32 . The application of a multi-purpose control valve  112  or  212  provides a significant advantage in that the valving required to perform the pressure relief, temperature unload and flow control functions is significantly simplified. Those skilled in the art will appreciate that variations can be made to the disclosed structure without extending beyond the scope of the proprietary drive axle described herein.