Patent Publication Number: US-2006011001-A1

Title: Differential drive actuator

Description:
CROSS-REFERENCE TO RELATED APPLICATION  
      This patent application is a continuation-in-part application of U.S. Ser. No. 10/892,706 filed Jul. 16, 2004, entitled “Ball Ramp Actuator Having Differential Drive.” 
    
    
     BACKGROUND OF THE INVENTION  
      The invention relates generally to a differential drive actuator and more specifically to a linear actuator having differentially driven rotating members which cooperate with ball ramps, cams or threaded members to axially translate an output.  
      Many types of controlled devices utilize linear translation to adjust them between on or off, open or closed or engaged and disengaged positions, as well as modulated or proportional intermediate positions. Motor vehicle clutches, air dampers, and all manner of valves are readily adjusted or controlled by the linear output motion of an actuator.  
      Not surprisingly, there has been extensive development of devices which provide a linear, bi-directional output. Electromagnetic solenoids are perhaps the simplest linear actuators but, of course, provide only two position or on/off operation over a limited distance. Other common bi-directional drive mechanisms include rack and pinion assemblies wherein a bi-directionally translating gear rack is driven by a bi-directionally rotating pinion. Cams and cam followers comprise another group of linear translation generation devices and a third group includes threaded devices such as leadscrews which rotate relative to complementarily threaded members such as nuts and either translate themselves or, if axially restrained, translate the nuts.  
      Frequently, linear translation generation devices are integrated into the controlled device. One such type of device is referred to as a ball ramp clutch. These devices, which include a multiple plate friction clutch pack also include an operator comprising a pair of adjacent plates having a plurality of opposed pairs of arcuate ramped recesses which receive a ball bearing or, alternatively, include a plurality of opposed, complementary oblique cam surfaces. Relative rotation of the plates causes the ball bearings to ride up the ramps of the recesses or the cams to ride up one another and separate the plates, thereby engaging the clutch. An electromagnetic coil may be utilized to create drag which causes the plates to rotate relatively. The electromagnetic coil does not directly engage the clutch but acts upon the ball ramp operator to create drag which, in turn, engages the clutch.  
      The ability to effect clutch engagement independent of a shaft speed difference is seen as a benefit in certain operational conditions. The present invention achieves this goal and has broad application as a bi-directional linear actuator for valves, steering systems, shutters, load leveling, dampers and other similarly controlled devices.  
     SUMMARY OF THE INVENTION  
      A differential drive actuator includes a first gear or circular member having a first plurality of teeth and a second gear or circular member disposed adjacent the first circular member and having a second plurality of teeth distinct in number from the first plurality of teeth. The circular members are commonly driven by a gear or pinion. Because of the distinct or disparate number of teeth on the two circular members, they will rotate at slightly different speeds. The circular members also include complementary cam faces, cam recesses and balls, a cam and cam follower or a threaded member which, upon such different rotational speeds, cause the circular members or the threaded member to axially translate or separate. Such axial translation may be utilized to actuate or move valve stems, dampers, plates, clutches, steering mechanisms, shutters and a wide variety of devices controlled or adjusted by linear motion.  
      It is thus an object of the present invention to provide a linear actuator having two adjacent motor driven differentially rotating members and associated components which generate linear motion.  
      It is a further object of the present invention to provide a linear actuator having two motor driven differentially rotating members and cam devices which generate linear translation of an output member upon relative rotation therebetween.  
      It is a still further object of the present invention to provide a ball ramp actuator for a friction clutch pack having two motor driven, differentially rotating camming members.  
      It is a still further object of the present invention to provide a linear actuator having two motor driven, differentially rotating members and an associated cam and follower.  
      It is a still further object of the present invention to provide a linear actuator having two motor driven, differentially rotating members and an associated nut and leadscrew.  
      Further objects and advantages of the present invention will become apparent by reference to the following description of the preferred embodiments and appended drawings wherein like reference numbers refer to the same component, element, or feature.  
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       FIG. 1  is a diagrammatic view of a drive train of an adaptive four-wheel drive motor vehicle having a transfer case assembly incorporating the present invention;  
       FIG. 2  is a side, elevational view in partial section of a transfer case assembly having a ball ramp actuator and friction clutch assembly according to the present invention;  
       FIG. 3  is an end elevational view in partial section of the ball ramp members of a ball ramp actuator according to the present invention;  
       FIG. 4  is an alternate embodiment cam operator of a ball ramp actuator according to the present invention;  
       FIG. 5  is a diagrammatic view of a motor vehicle drive train including a transaxle and a ball ramp actuator and friction clutch assembly disposed at a rear differential of the drive train;  
       FIG. 6  is an enlarged, full sectional view of a ball ramp actuator and friction clutch assembly according to the present invention disposed at a rear differential;  
       FIG. 7  is a full, sectional view of a twin clutch rear differential incorporating the present invention;  
       FIG. 8  is a fragmentary, full sectional view of a first alternate embodiment of a ball ramp actuator for a friction clutch assembly;  
       FIG. 9  is a fragmentary, full sectional view of a second alternate embodiment of a ball ramp actuator for a friction clutch assembly;  
       FIG. 10  is a full sectional view of a first additional embodiment of a differential drive actuator according to the present invention;  
       FIG. 11  is a full sectional view of the first additional embodiment of a differential drive actuator according to the present invention taken along line  11 - 11  of  FIG. 10 ;  
       FIG. 12  is a full sectional view of a second additional embodiment of a differential drive actuator according to the present invention utilizing a cam and follower;  
       FIG. 13  is a full sectional view of a third additional embodiment of a differential drive actuator according to the present invention utilizing a leadscrew and nut;  
       FIG. 14  is a fragmentary, sectional view of a additional embodiment of a differential drive actuator according to the present invention taken along line  14 - 14  of  FIG. 13 ; and  
       FIG. 15  is a fragmentary, sectional view portion of a third additional embodiment of a differential drive actuator taken along line  15 - 15  of  FIG. 14 . 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS  
      Referring now to  FIG. 1 , a four-wheel vehicle drive train utilizing the present invention is diagramatically illustrated and designated by the reference number  10 . The four-wheel vehicle drive train  10  includes a prime mover  12  which is coupled to and directly drives a transmission  14 . The prime mover  12  may be a gas, Diesel or hybrid power plant. The output of the transmission  14  directly drives a transfer case assembly  16  which provides motive power to a primary or rear driveline  20  comprising a primary or rear prop shaft  22 , a primary or rear differential  24 , a pair of live primary or rear axles  26  and a respective pair of primary or rear tire and wheel assemblies  28 .  
      The transfer case assembly  16  also selectively provides motive power to a secondary or front driveline  30  comprising a secondary or front prop shaft  32 , a secondary or front differential  34 , a pair of live secondary or front axles  36  and a respective pair of secondary or front tire and wheel assemblies  38 . The front tire and wheel assemblies  38  may be directly coupled to a respective one of the front axles  36  or, if desired, a pair of manually or remotely activatable locking hubs  42  may be operably disposed between the front axles  36  and a respective one of the tire and wheel assemblies  38  to selectively connect same. Finally, both the primary driveline  20  and the secondary driveline  30  may include suitable and appropriately disposed universal joints  44  which function in conventional fashion to allow static and dynamic offsets and misalignments between the various shafts and components.  
      Disposed in sensing relationship with each of the rear tire and wheel assemblies  28  is a wheel speed sensor  48 . Preferably, the rear wheel speed sensors  48  may be the same sensors utilized with, for example, an antilock brake system (ABS) or other vehicle control or traction enhancing system although they may, of course, be independent of any other system. Alternatively, a single sensor (not illustrated), disposed to sense rotation of the primary or rear prop shaft  22  may be utilized. Signals from the sensors  48  are provided in electrical lines  52  to a microprocessor  56 . Similarly, disposed in sensing relationship with the front tire and wheel assemblies  38  are respective front wheel speed sensors  58  which provide signals to the microprocessor  56  in electrical lines.  62 . Once again, the sensors  58  may be a part of or shared with an antilock brake system or other traction control system or they may be independent thereof.  
      Typically, an operator selectable switch  64  may be utilized and is generally disposed within reach of the vehicle operator in the passenger compartment (not illustrated). The switch  64  may be adjusted to select various operating modes such as two-wheel high gear, automatic, i.e., on-demand or adaptive operation, four-wheel high gear or four-wheel low gear depending upon the particular vehicle, configuration of the transfer case assembly  16  and the driver&#39;s desires. One such system which provides torque delivery to the secondary driveline  30  in increments or decrements in response to a sensed wheel speed difference between the primary driveline  20  and the secondary driveline  30  is disclosed in co-owned U.S. Pat. No. 5,407,024.  
      Referring now to  FIG. 2 , a typical two-speed transfer case assembly  16  includes a cast, multiple piece housing  70  having a plurality of planar sealing surfaces, openings for shafts and bearings and various recesses, shoulders, counterbores and the like which receive, support or mount various assemblies or components of the transfer case assembly  16 . An input shaft  72  includes female or internal splines  74  or other suitable coupling structures which drivingly engage and couple the output of the transmission  14 , illustrated in  FIG. 1 , to the input shaft  72 . In the two-speed transfer case assembly  16 , the input shaft  72  provides motive power to a planetary gear speed reduction assembly  76  which is controlled by a two or three position operator assembly  78  which may be electrically, pneumatically or hydraulically powered and a shift fork and cam assembly  80  to achieve and provide a first, direct drive speed range (high gear), neutral and a second, reduced speed drive range (low gear). The output of the planetary gear speed reduction assembly  76  is provided to a primary output shaft  82  which is coupled to and drives the primary driveline  20 . In a single speed transfer case assembly, the planetary gear assembly  76  and the operator assembly  78  are not present and the input shaft  72  or its equivalent directly drives the primary output shaft  82 . Suitable ball bearing assemblies  84  rotatably support the shafts  72  and  82  and a pair of oil seals  86  provide fluid tight seals between the shafts  72  and  82  and the housing  70 .  
      A modulating clutch assembly  90  is operably disposed between the primary output shaft  82  and a chain drive sprocket  92  freely rotatably disposed about the primary output shaft  82 . The chain drive sprocket  92  is engaged by a drive chain  94  which also engages a driven chain sprocket  96  which is secured to a secondary output shaft  98 . The secondary output shaft  98  is coupled to and drives the secondary driveline  30 .  
      The modulating clutch assembly  90  includes a friction clutch pack assembly  100  having a first plurality of smaller diameter clutch plates  102  which are splined to the primary output shaft  82  by a plurality of interengaging splines  104 . The first plurality of smaller diameter clutch plates  102  are interleaved with a second plurality of larger diameter clutch plates  106  which are coupled by interengaging splines  108  to a bell-shaped clutch housing  110 . The first and second pluralities of interleaved clutch plates  102  and  106  include suitable friction material secured to at least one face of each of the clutch plates  102  and  106 .  
      The bell-shaped clutch housing  110  is rotationally coupled to the chain drive sprocket  92  by interengaging splines, axially extending lugs and apertures, welds, or other permanent or disconnectible rotational coupling means. Between the friction clutch pack  100  and the clutch housing  110  is a circular backup plate  112  which is restrained against axial motion to the left, as illustrated in  FIG. 2 , by a cooperating snap ring and channel  114  in the primary output shaft  82 . On the opposite side or face of the friction clutch pack assembly  100  is a circular apply plate  116 .  
      The modulating clutch assembly  90  also includes a differential gear operator assembly  120 . The differential gear operator assembly  120  includes a fractional horsepower bi-directional electric motor  122  which directly drives a pinion gear  124  through an output shaft  126 . The electric motor  122  is preferably secured to the housing  70  of the transfer case assembly  16  by a plurality of fasteners  128 , one of which is illustrated in  FIG. 2 . A suitable seal such as an O-ring  130  is disposed between the outer surface of the transfer case assembly  16  and a mounting plate of the electric motor  122  and provides a suitable fluid tight seal.  
      Referring now to  FIGS. 2 and 3 , the pinion gear  124  simultaneously engages a first circular cam member  132  and a second, adjacent circular cam member  134 . The circular cam members  132  and  134  have gear teeth  136  and  138  disposed about their peripheries. The number of gear teeth  136  and  138  on each of the circular cam members  132  and  134  is not the same. For example, the first cam member  132  may define or includes 180 gear teeth  138  about its periphery whereas the second cam member  134  defines or has 181 teeth about its periphery. The profiles of the gear teeth  136  and  138 , their pressure angles and overall geometry are chosen to be a compromise with the gear teeth of the pinion  124  such that any errors are split between the gear teeth  136  and  138  of the cam members  132  and  134  and thus are equally but only slightly variant from the nominal or appropriate values. A one tooth difference between the circular cam members  132  and  134  minimizes these errors but a larger difference between the number of teeth, increasing the rotational differential between the circular cam members  132  and  134  is possible, especially if the diameter of the cam members  132  and  134  is increased or the number of gear teeth  136  and  138  is increased and the size of the gear teeth  136  and  138  is decreased.  
      On the opposed, adjacent faces of the cam members  132  and  134  are camming features. On the cam member  132  are curved (arcuate), teardrop shaped ramped recesses  142 . Complementarily configured curved (arcuate), teardrop shaped ramped recesses  144  are formed in the second cam member  134 . Received within the curved, ramped recesses  142  and  144  are load transferring members such as ball bearings  146 . It will be appreciated that as the circular cam members  132  and  134  rotate relative to one another, the ball bearings  146  and the ramped recesses  142  and  144  axially separate the cam members  132  and  134 . The ramps and ball bearings may be readily replaced with other, analogous mechanical assemblies such as ramps and roller bearings or oblique, opposed camming surfaces to name but two.  
      Referring briefly to  FIG. 4 , an alternate embodiment utilizing cam surfaces, noted above, is illustrated. In a first circular member  132 A, a plurality of projections  148 A and recesses  150 A form oblique ramps. On a second circular member  134 B, a complementary plurality of projections  148 B and recesses  150 B form opposed complementary, oblique ramps. As the circular members  132 A and  134 B rotate differently relative to one another, they are driven axially apart.  
      It should be appreciated that both the factor of differential rotation, i.e., how quickly differential rotation of the circular cam members  132  and  134  occurs and how quickly such rotation causes axial, clutch engaging, translation may be characterized as the amplification of the operator assembly  120  and may be adjusted to satisfy various design criteria. Slower differential rotation and shallower cam angles requires significant rotation such that even a small electric motor  122  has the capability to apply significant compressive force to the associated friction clutch pack assembly  100 , which may be characterized as high (force) amplification. On the other hand, a greater numerical gear tooth difference and steep cams (low amplification) will achieve quicker clutch engagement and generally require a more powerful electric motor  122 .  
      Between the first cam member  132  and the apply plate  116  is a first ball or roller thrust bearing  152  which transmits axial force but allows the adjacent apply plate  116  and first cam member  132  to rotate fully independently of one another. Adjacent the second cam member  134  is a second ball or roller thrust bearing  154 . Adjacent the second ball or roller thrust bearing  154  is a backup washer  156  which has axial travel limited by a complementary snap ring and groove  158  formed in the primary output shaft  82 . The second thrust bearing  154  allows the second cam member  134  to rotate fully independently of the primary output shaft  82  and the backup washer  156 . The circular backup plate  112  and the backup washer  156  as well as the adjacent snap rings and grooves  114  and  158  function as stops and the termini of a reaction force circuit against which the differential gear clutch operator  120  functions and contains all forces and reaction forces within the length of the primary output shaft  82  between the snap rings and grooves  114  and  158 .  
      Referring now to  FIG. 5 , an adaptive four-wheel vehicle drive train is diagrammatically illustrated and designated by the reference number  200 . The four-wheel vehicle drive train  200  includes a prime mover  202  such as a gasoline, Diesel or natural gas fueled internal combustion engine or hybrid power plant which is coupled to and directly drives a transaxle  204 . The output of the transaxle  204  drives a primary or front driveline  210  and a secondary or rear driveline  220 . The primary driveline  210  comprises a front or primary propshaft  212 , a front or primary differential  214 , a pair of live front axles  216  and a respective pair of front tire and wheel assemblies  218 . It should be appreciated that the front or primary differential  214  is conventional.  
      The transaxle  204 , through a power takeoff  206 , also provides drive torque to the secondary or rear drive line  220  comprising a secondary propshaft  222  having appropriate universal joints  224 , a rear or secondary axle assembly  226 , a pair of live secondary or rear axles  228  and a respective pair of secondary or rear tire and wheel assemblies  230 .  
      As utilized herein with regard to the secondary axle assembly  226 , the term “axle assembly” is used to identify a device for receiving drive line torque, distributing it to two generally aligned, transversely disposed drive axles and accommodating rotational speed differences resulting from, inter alia, vehicle cornering.  
      Furthermore, the foregoing and following description relates to a vehicle wherein the primary drive line  210  is disposed at the front of the vehicle and, correspondingly, the secondary drive line  220  is disposed at the rear of the vehicle, such a vehicle commonly being referred to as a (primary) front wheel drive vehicle or adaptive four-wheel drive vehicle. Nonetheless, it should be appreciated that this invention is equally suited for use in a (primary) rear wheel drive vehicle having the drive component location reversed.  
      Associated with the vehicle drive train  200  is a controller or microprocessor  240  which receives signals from a plurality of sensors and provides a control, i.e., actuation, signals to the rear or secondary axle assembly  226 .  
      The vehicle drive train  200  also includes a first variable reluctance or Hall effect sensor  246  which senses the rotational speed of the left primary (front) tire and wheel assembly  218  and provides a signal to the microprocessor  240 . A second variable reluctance or Hall effect sensor  248  senses the rotational speed of the right primary (front) tire and wheel assembly  218  and provides a signal to the microprocessor  240 . A third variable reluctance or Hall effect sensor  250  associated with the left secondary (rear) tire and wheel assembly  230  senses its speed and provides a signal to the microprocessor  240 . Finally, a fourth variable reluctance or Hall effect sensor  252  associated with the right secondary (rear) tire and wheel assembly  230  senses its speed and provides a signal to the microprocessor  240 . It should be understood that the speed sensors  246 ,  248 ,  250  and  252  may be those sensors mounted in the vehicle to provide signals for anti-lock brake systems (ABS) or other speed sensing and traction control systems or may be independent, i.e., dedicated, sensors. It is also to be understood that an appropriate and conventional counting or tone wheel (not illustrated) is associated with each of the respective tire and wheel assemblies  218  and  230  in proximate sensing relationship with each of the speed sensors  246 ,  248 ,  250  and  252 .  
      Referring now to  FIGS. 5 and 6 , a modulating rear axle clutch assembly  260  is operably disposed between the output of the power takeoff  206  and the secondary axle assembly  226 . The rear axle clutch assembly  260 , incorporating the present invention, includes a generally bell shaped housing  262  having an annular flange  264  defining a plurality of through openings  266  which receive complementarily sized threaded fasteners (not illustrated) utilized to secure the clutch assembly  260  and specifically the housing  262  to the housing of the secondary axle assembly  226  illustrated in  FIG. 5 . The clutch assembly  260  includes an input shaft  270  which is supported within the housing upon an antifriction bearing such as a ball bearing assembly  272 . A suitable oil seal  274  provides a seal between the housing  262  and the rotating input shaft  270  to prevent the ingress of contaminants and egress of clutch lubricating fluid. The input shaft  270  may include a collar or hub  276  or other component which may be a portion of a universal joint or other driveline feature. A threaded lock nut  278  may be utilized to secure the collar  276  to the input shaft  270 .  
      Concentrically disposed about the input shaft  270  is a friction clutch pack assembly  280 . The friction clutch pack assembly  280  includes a first plurality of smaller diameter clutch plates  282  having internal or female splines  284  which are complementary to and engage external or male splines  286  formed on a portion of the input shaft  270 . Interleaved with the first plurality of smaller diameter clutch plates  282  is a second plurality of larger diameter clutch plates  288 . The second plurality of larger diameter clutch plates  288  includes external or male splines or gear teeth  292  which are complementary to and engage internal or female splines  294  formed on the inner surface of a bell shaped output housing  300 . The bell shaped output housing  300  defines a concentric aperture having a plurality of female or internal splines or gear teeth  302  which are complementary to and engage male splines or gear teeth  304  on a stub output shaft  306 . The stub output shaft  306  is received within a counterbore  308  in the input shaft  270 . The inner surface of the stub output shaft  306  preferably includes internal or female splines or gear teeth  312  which may be engaged by a driven member (not illustrated) disposed within the rear axle assembly  226 .  
      Between the friction clutch pack assembly  280  and the bell housing  300  is a backup or stop plate  316  which is maintained in position on the input shaft  270  by a cooperating snap ring and groove  318  formed in the input shaft  270 . On the opposite face of the friction clutch pack assembly  280  is an apply plate  320 .  
      The friction clutch pack assembly  280  is actuated by a differential cam actuator assembly  330  which includes a fractional horsepower, bi-directional electric motor  332 . The electric motor  332  drives a pinion gear  334  through an output shaft  336 . The pinion gear  334  includes uniform axially extending gear teeth  338  about its periphery. The gear teeth  338  of the pinion  334  simultaneously engage a first circular cam plate  340  and a second circular cam plate  342 . The first circular cam plate  340  includes gear teeth  344  disposed about its periphery. The second circular cam plate  342  also includes gear teeth  346  disposed about its periphery. The numbers of gear teeth  344  and  346  are not equal. Preferably, for example, there are 180 gear teeth  344  on the first circular cam plate  340  and  181  gear teeth  346  on the second circular cam plate  342 . Thus, as the circular cam plates  340  and  342  are driven by the pinion  344 , they differentially rotate, i.e., the second cam plate  342  rotates slightly slower than the first cam plate  340 . The numbers of gear teeth  344  and  346  recited are given by way of example only and it should be understood that the numbers can vary widely depending upon the size of the gears  340  and  342 , the size of the pinion  334  and the operating speed desired.  
      The first circular cam plate  340  includes a plurality of arcuate, teardrop shaped ramped recesses  348  and the second circular cam plate  342  includes a like plurality of arcuate, teardrop shaped ramped recesses  352 . Within these ramped recesses are captured load transferring members such as ball bearings  354 . As the circular cam plates  340  and  342  differentially rotate, the load transferring members  354  drive them apart.  
      Between the friction clutch pack assembly  280  and the second circular cam plate  342  is disposed a ball or roller thrust bearing  356  which allows free relative rotation between the apply plate  320  and the second circular cam plate  342 . To the left of the first circular cam plate  340  is disposed a second ball or roller thrust bearing assembly  358 . Adjacent the thrust bearing assembly  358  is a stop washer  360  which is held in a fixed axial position by a snap ring and groove  364  formed in the input shaft  270 . The thrust bearing assembly  358  allows free relative rotation between the first cam member  340  and the stop washer  360 . The snap rings and grooves  318  and  364  provide reaction stops and contain the forces and reaction forces of the clutch operator within the input shaft  270 .  
      As illustrated in  FIG. 5 , the secondary axle modulating clutch  260  functionally precedes the secondary differential or rear axle  226  and controls delivery of torque thereto. Another application for the differential cam actuator assembly  330  encompasses a rear axle assembly without the single modulating clutch  260  and conventional caged differential. These components are replaced by a rear axle assembly  226  having twin independently operable modulating clutches which independently deliver torque to the left and right rear axles  228  and tire and wheel assemblies.  
      Referring now to  FIGS. 5 and 7 , the rear or secondary axle assembly  226  includes an input shaft  370  which receives drive torque directly from the secondary propshaft  222 . The input shaft  370  may include a flange or cup  372  or similar component which forms a portion of, for example, a universal joint  224  or other connection to the secondary propshaft  222 . The flange  372  may be retained on the input shaft  370  by a lock nut  374  or similar device. The input shaft  370  is received within a centrally disposed, axially extending center housing  376  and is surrounded by a suitable oil seal  378  which provides a fluid impervious seal between the housing  376  and the input shaft  370  or an associated portion of the flange  372 . The input shaft  370  is preferably rotatably supported by a pair of anti-friction bearings such as the tapered roller bearing assemblies  380 . The input shaft  370  terminates in a hypoid or bevel gear  382  having gear teeth  384  which mate with complementarily configured gear teeth  386  on a ring gear  388  secured to a flange  392  on a centrally disposed tubular drive member  394  by suitable threaded fasteners  396 .  
      The tubular drive member  394  is rotatably supported by a pair of anti-friction bearings such as ball bearing assemblies  402 . The tubular drive member  394  is hollow and defines an interior volume  404 . A pair of scavengers or scoops  406  extend radially through the wall of the tubular drive member  394  and collect a lubricating and cooling fluid  408  driving it into the interior volume  404 . The lubricating and cooling fluid  408  is then provided to components in the rear axle assembly  226  through passageways  410  in communication with the interior volume  404  of the tubular drive member  394 .  
      The rear or secondary axle assembly  226  also includes a pair of bell housings  412 A and  412 B which are attached to the center housing  376  by threaded fasteners  414 . The housings  412 A and  412 B are mirror-images, i.e., left and right, components which each receive a respective one of a pair of modulating clutch assemblies  420 A and  420 B. But for the opposed, mirror-image arrangement of the two modulating clutch assemblies  420 A and  420 B, the components of the two clutch assemblies  420 A and  420 B described below are identical. Accordingly, and for purposes of clarity in  FIG. 7 , numerical component cal louts may appear in either or both of the left and right clutch assemblies  420 A and  420 B, it being understood that such components reside in and such callouts refer to both assemblies.  
      Both of the modulating clutch assemblies  420 A and  420 B are driven by the input shaft  370  through the bevel gears  382  and  388  and the tubular drive member  394 . Specifically, the ring gear  388 , as noted above, is secured to the tubular drive member  394 . A tubular extension  422  of the ring gear  388  includes external or male splines  424 , which mate with internal or female splines or gear teeth  428 A, formed on a left drive collar  430 A. The left drive collar  430 A also includes external or male splines or gear teeth  432 A which mate with complementarily configured internal or female splines or gear teeth  434 A on a left clutch end bell  440 A. With regard to the drive to the right modulating clutch assembly  420 B, the tubular drive member  394  includes external or male splines or gear teeth  436 , which engage complementarily configured female splines or gear teeth  428 B on a right drive collar  430 B. Correspondingly, the right drive collar  430 B includes male or external splines or gear teeth  432 B which are complementary to and engage internal or female splines or gear teeth  434 B formed on a right clutch end bell  440 B.  
      The clutch end bells  440 A and  440 B are identical but disposed in mirror image relationship. Each of the clutch end bells  440 A and  440 B includes internal splines  442  which drivingly engage complementarily configured external splines  444  on a first plurality of larger diameter friction clutch plates or discs  446 . Interleaved with the first plurality of larger diameter friction clutch plates or discs  446  is a second plurality of smaller diameter friction clutch plates or discs  448 . At least one face of each of the friction clutch plates or discs  446  and  448  includes suitable friction clutch material. Each of the smaller diameter friction clutch plates or discs  448  includes internal or female splines  450  which engage complementarily configured male or external splines  452  on a circular collar or hub  454 . The hub  454  is, in turn, coupled by internal or female splines or gear teeth  456  to male splines or gear teeth  458  on respective left and right output shafts  460 A and  460 B for rotation therewith. The output shafts  460 A and  460 B may include male splines  462 A and  462 B.  
      The modulating clutch assemblies  420 A and  420 B also include ball ramp actuator assemblies  470 A and  470 B. The ball ramp actuator assemblies  470 A and  470 B each include a first circular camming member  472  having gear teeth  474  disposed about its periphery and a plurality of arcuate, tear drop shaped ramped recesses  476  on one face. Adjacent the first circular cam member  472  is a second circular cam member  482  having gear teeth  484  disposed about its periphery and a plurality of arcuate, tear drop shaped recesses  486  facing the similarly configured ramped recesses  476  on the first circular cam member  472 . Disposed and retained within the ramped recesses  476  and  486  are a plurality of load transferring members such as ball bearings  490 .  
      As noted above, as the camming members  472  and  482  rotate relative to one another, the ball bearings  490  move along the ramped recesses  476  and  486  and separate the first and second camming members  472  and  482 . Once again, it should be understood that analogous devices such as tapered roller bearings in complementarily configured ramped recesses or opposed, oblique cam surfaces will affect similar axial motion upon relative rotation of the camming members  472  and  482  and thus are also suitable.  
      Adjacent the first camming member  472  is a first thrust bearing  492  which is maintained in its axial position by adjacent snap rings. Adjacent the second camming member  482  is a second thrust bearing  494 . Between the friction clutch plates  446  and  448  and the second thrust bearing  496  is a circular apply plate  496 . The second thrust bearing  496  allows free relative rotation between the apply plate  496  and the second camming member  482 .  
      The ball ramp actuator assembly  470 A also includes a fractional horsepower, bi-directional electric motor  500  which is secured to the housing  412 A by a plurality of threaded fasteners  502 . A suitable fluid tight seal (not illustrated) between the housing of the electric motor  500  and the housing  412 A may be included. The electric motor  500  includes an output shaft  504  which drives a pinion  506  having gear teeth  508 . The gear teeth  508  of the pinion  506  engage the gear teeth  474  and  484  on both the first camming member  472  and the second camming member  482 . As noted, there are different numbers of gear teeth  474  and  484  on the respective camming members  472  and  482  such that they rotate differentially, i.e., at different speeds, thereby generating relative rotation which causes separation of the camming members  472  and  482 . Such separation compresses the adjacent, associated friction clutch pack assembly and transmits torque from the drive tube  394  to the output shaft  460 A.  
      Turning now to  FIG. 8 , a first alternate embodiment of the ball ramp actuator is illustrated and designated by the reference number  520 . Here, a shaft  522  which may be either an input shaft such as within a transfer case assembly  16  as illustrated in  FIG. 2  or an output shaft as in a rear differential as illustrated in  FIG. 5  and a second shaft  524  which may be an output shaft as illustrated in the transfer case assembly  16  or an input shaft as illustrated in the rear differential are coupled to and drive a respective first plurality of smaller clutch plates or disks  526  which are splined to the first shaft  522  by a plurality of interengaging male and female splines  528 . The first plurality of clutch plates or disks  532  are interleaved with a second plurality of friction clutch plates or disks  532  which include a plurality of interengaging splines or gear teeth  534  which couple the second plurality of friction clutch plates or disks  532  to a bell shaped housing  536  which in turn is coupled through splines or other interengaging means such as lugs to the second shaft  524 . A backing or stop plate  532  is maintained in position on the first shaft  522  by a cooperating snap ring and groove  544 .  
      An actuator assembly  550  includes a fractional horsepower, bi-directional electric motor  552  which drives an output shaft  554 . The output shaft  554  is secured to or integrally formed with a pinion  556  having a first region of gear teeth  504  having a number of gear teeth distinct from a second region of gear teeth  562 . For example, the first region of gear teeth  558  may include twenty teeth whereas the second region of gear teeth  562  may include twenty-one teeth. It will be appreciated that these numbers may be varied to achieve a desired differential rotation of the driven members. Aligned with the first region of gear teeth  558  is a first camming member  566  which includes gear teeth  568  disposed about its periphery which engage the first region of gear teeth  558  on the pinion  556 . Disposed adjacent the first camming member  556  is a second camming member  572  having a plurality of gear teeth  574  about its periphery which engage the second region of gear teeth  562  on the pinion  556 . The number of gear teeth  568  on the first camming member  566  may be equal to or unequal to the number of gear teeth  574  on the second camming member  572 . In order to achieve differential rotation of the camming members  566  and  572 , however, the numbers of teeth on the first and second camming members  566  and  572  must not be such as to achieve the same drive ratio given the number of teeth on the two regions of the pinion  556 .  
      Differential rotation may be readily achieved by utilizing the same number of gear teeth  568  and  574  on the first and second camming members  566  and  572 , respectively, with the pinion  556  described above. Both of the camming members  566  and  572  include a plurality of arcuate, teardrop shaped ramped recesses  576  and  578  which receive and retain a like plurality of load transferring members such as ball bearings  580 . A first thrust bearing  582  is disposed adjacent the first camming member  566  and between a flat washer  584  and a cooperating snap ring and groove  586 . The first thrust bearing  582  allows free rotation of the first camming member  566  relative to the first shaft  522 . The washer  584  and the snap ring and groove  586  provide a reaction force stop. Adjacent the second camming member  572  is a second thrust bearing assembly  592  which likewise allows free rotation of the second camming member  572  and transmits clutch actuating force therethrough. An apply plate  594  is disposed adjacent the second thrust bearing  592  and applies compressive, clutch actuating force to the clutch plates  526  and  532  upon relative rotation and separation of the first and second camming members  566  and  572 .  
      It will be appreciated that the recesses  576  and  578  and the load transferring balls  580  may be replaced with other analogous mechanical elements which cause axial displacement of the camming members  566  and  572  in response to relative rotation therebetween. For example, tapered rollers disposed in complementarily configured conical helices or opposed, oblique cam surfaces may be utilized.  
      Turning now to  FIG. 9 , a second alternate embodiment of the ball ramp actuator is illustrated and designated by the reference number  600 . A shaft  602  which may be either an input shaft  72  of a transfer case assembly  16  as illustrated in  FIG. 2  or an output shaft of a secondary axle assembly  226  as illustrated in  FIG. 7  and a second shaft  604  which may be an output shaft  82  as illustrated in the transfer case assembly  16  or an input shaft as illustrated in the secondary differential  226  are coupled to and drive a respective first plurality of smaller clutch plates or disks  606  which are splined to the first shaft  602  by a plurality of interengaging male and female splines  608 . The first plurality of clutch plates or disks  606  are interleaved with a second plurality of friction clutch plates or disks  612  which include a plurality of interengaging splines or gear teeth  614  which couple the second plurality of friction clutch plates or disks  612  to a bell shaped housing  616 . The bell shaped housing  616  is, in turn, coupled through splines or other interengaging means such as lugs to the second shaft  604 . A backing or stop plate  622  is maintained in position on the first shaft  602  by a cooperating snap ring and groove  624 .  
      An actuator assembly  630  includes a fractional horsepower, bi-directional electric motor  632  which drives an output shaft  634 . The output shaft  634  is secured to or integrally formed with a first pinion gear  636  having gear teeth  638 . The first pinion gear  636  engages and bi-directionally drives a larger spur gear  640  which is disposed upon a stub shaft  642  rotatably supported within a housing  644 . The larger spur gear  640  is integrally formed with or coupled to a second pinion gear  646 . The gear reduction achieved by the first pinion gear  636 , the larger spur gear  640  and the second pinion gear  646  reduces the speed of the bi-directional electric motor  632  and increases its torque. The speed reduction i.e., gear reduction ratio, suitable for a particular application will be a function of the speed and torque of the electric motor  632 , the angles of the ramped recesses  658  and  662 , the desired engagement time of the clutch and other factors.  
      The second pinion gear  646  engages and drives a first camming member  650  having a first number of gear teeth  652  disposed about its periphery. Adjacent the first camming member  650  is a second camming member  654  having a second number of teeth  656  disposed about its periphery which are distinct in number from the teeth  652  on the first camming member  650 . Preferably, the number of gear teeth  652  will be a hundred or more and the number of gear teeth  656  will be different only by one or two. It is apparent that a larger number of teeth will permit a larger difference in the number of teeth and that the number of teeth is loosely related to the overall size of the device. A number of teeth between 75 and 250 will likely encompass most applications.  
      As in the other embodiments, the first and second camming members  650  and  654  include defined cam or ramp components which drive the members  650  and  654  axially apart from a proximate, rest position. As illustrated in  FIG. 9 , the first camming member includes a plurality of arcuate, teardrop shaped ramped recesses  658  and the second camming member  654  includes a like plurality of opposed arcuate, teardrop shaped ramped recesses  662 . A like plurality of load transferring members such as ball bearings  664  are received and retained within the opposed arcuate, ramped recesses  658  and  662 . It will be appreciated that the ramps and balls may be replaced by oblique camming surfaces and other configurations causing separation of the camming members  650  and  654  upon relative rotation.  
      A first thrust bearing  672  is disposed between the first camming member  650  and a flat washer  674  and a cooperating snap ring and groove  676 . The first thrust bearing  672  allows free rotation of the first camming member  650  relative to the shaft  602  and the flat washer  674 . The snap ring and groove  676  provide a reaction force stop. Adjacent the second camming member  654  is a second thrust bearing assembly  678  which likewise allows free rotation of the second camming member  654  and transmits clutch actuating force therethrough. An apply plate  682  is disposed adjacent the second thrust bearing  678  and applies compressive, clutch actuating force to the first plurality of clutch plates  606  and  612  upon relative rotation and separation of the first and second camming members  650  and  654 .  
      Upon energization of the electric motor  632 , the first pinion gear  636  rotates, driving the second pinion gear  646  at a reduced speed. Since the camming members  650  and  654  have different numbers of gear teeth  652  and  656 , they rotate at different (differential) speeds causing relative rotation therebetween which results in axial separation of the members  650  and  654  by action of the ramped recesses  658  and  662  and the load transferring ball bearings  664 , compression of the pluralities of friction clutch plates  606  and  612  and energy transfer through the clutch assembly  600 .  
      Referring now to  FIG. 10 , a first additional embodiment of a differential drive actuator is illustrated and designated by the reference number  700 . The differential drive actuator  700  includes an irregularly shaped multiple piece, preferably die cast housing  702  having various apertures, grooves, shoulders, counterbores and features which accept, position, support and mount various components and assemblies of the differential drive actuator  700 . A bi-directional, fractional horsepower electric motor  704  is secured to the housing  702  by any suitable fastening arrangement and is provided with electrical energy through a multiple conductor cable  706 . The bi-directional electric motor  704  includes an output shaft  708  which includes a positively secured, radially extending pin or vane  710  which engages and drives a drive pinion  712 . The drive pinion  712  is preferably fabricated of a strong, long wearing plastic such as nylon. The drive pinion  712  includes peripheral gear teeth  714  which are complementary to and engage gear teeth  716  on an intermediate or idler gear  718 . The intermediate or idler gear  718  is a composite gear and includes a peripheral metal annulus  720  which includes the gear teeth  716  and an inner hub  722  having a capital “I or H” cross-section which receives a stub shaft  724 . The peripheral metal annulus  720  and the inner hub  722  may be secured together by any suitable means including a snap ring  726  which seats within complementarily configured semi-circular recesses on the inner surface of the metal annulus  720  and the outer surface of the inner hub  722 . This composite design reduces the mass and thus the inertia of the idler gear  718  while still providing rugged, long wearing metal gear teeth  716 . The stub shaft  724  may be maintained in position by a cover plate  732  which is secured to the housing  702  by suitable fasteners  734 .  
      The intermediate or idler gear  718  engages a first gear, circular cam plate or member  740  having a plurality of peripheral gear teeth  742  disposed about its periphery. One face of the circular cam plate  740  includes a plurality of arcuate camming recesses  744  which each receive a load transferring component such as a ball bearing  746 . Immediately adjacent the first circular cam plate  740  and commonly driven by the idler gear  718  is a second gear, circular cam plate or member  750  having a plurality of gear teeth  752  disposed about its periphery and the same number of arcuate camming recesses  754  as reside in the face of the opposing and adjacent first circular cam plate  740 .  
      Just as in the other embodiments, the number of teeth on the peripheries of the first gear or circular cam plate  740  and the second gear or circular cam plate  750  are distinct, the number of teeth on the two cam plates  740  and  750  preferably differing by one when there are approximately 100 to 200 gear teeth on each of the cam plates  740  and  750 . This number may be increased to two, three, four or more if the teeth are relatively small, there are a large number of teeth, or both, such that the engagement error will be negligible. As noted above, since the differential drive assembly operates upon and presumes a certain small mismatch between the gear teeth of the circular cam plates  740  and  750  with the teeth  716  of the intermediate or idler gear  718 , it has been found preferable, though not necessary, to divide this error between the two sets of gear teeth  742  and  752  in order to optimize operation. Increasing the gear tooth disparity by increasing the difference in the number of teeth, of course, increases this small error and decreases the amplification achieved by the assembly but increases the axial output speed and thus actuation speed.  
      The circular cam plates  740  and  750  are each journalled and ride upon a friction reducing metal or plastic bushing  758  which are freely rotatably received and supported upon an input shaft  760 . The input shaft  760  extends beyond the housing  702  and includes male splines  762  which are complementary to and engaged by female splines  764  on an input flange  766 . The input flange  766  is secured and maintained upon the input shaft  760  by a washer  768  and a threaded nut  772  which is received upon a complementarily threaded portion  774  of the input shaft  760 . A ball bearing assembly  776  rotatably supports the input shaft  760  and the input flange  766  within the housing  702 . An oil seal  778  disposed between the input flange  766  and the housing  702  provides a suitable fluid tight seal therebetween and inhibits the ingress of any foreign material into the differential drive actuator  700 . A flat washer  782  axially restrained by the input flange  766  and a thrust bearing  784  maintain the axial position of the first circular cam plate  740 . A second thrust bearing  786  is disposed on the opposite side of the second circular cam plate  750  and transfers axial thrust and motion from the second circular cam plate  750  to an annular actuator member  790 . The annular actuator member  790  is plate  750  to an annular actuator member  790 . The annular actuator member  790  is biased to the left as illustrated in  FIG. 10  by a circular flat spring or Belleville washer  792 . The annular actuator member  790  engages a circular apply plate  794  which applies compression to a friction clutch pack assembly  800 .  
      The friction clutch pack assembly  800  includes a first plurality of smaller diameter clutch plates  802  which include internal or female splines which engage a plurality of male splines  804  on the input shaft  760  and rotate therewith. Interleaved with the first plurality of friction clutch plates  802  is a second plurality of larger diameter friction clutch plates  806 . Both the friction clutch plates  802  and  806  include suitable friction clutch material on at least one face. The second plurality of larger clutch plates  806  include exterior or male splines which mate with internal or female splines  808  on the inner surface of a bell-shaped output member  810 . The output member  810  includes a circular opening  812  having internal, female splines  814  which receive and drivingly engage a driven (output) member (not illustrated). A roller bearing assembly  816  received within a blind aperture  818  of the input shaft  760  receives and freely rotatably supports a suitably sized terminal portion of the driven (output) member. An O-ring  822  may be utilized to provide a fluid tight seal between the housing  702  and a housing (not illustrated) of an associated component.  
      Referring now to  FIG. 12 , a second additional embodiment of a differential drive actuator is illustrated and designated by the reference number  830 . The second additional embodiment differential drive actuator  830  includes a multiple piece, preferably cast housing  832  which may be assembled and secured together by threaded fasteners  834  disposed in suitable opening  836 . The housing  832  preferably includes various apertures, grooves, shoulders, counterbores, flanges and the like which receive, rotatably support and provide mounting for various components of the differential drive actuator  830 . A fractional horsepower bi-directional electric motor  840  is suitably secured to the housing  832  and is provided with electrical energy through a multiple conductor cable  842 . The bi-directional electric motor  840  includes an output shaft  844  having a positively secured vane or pin  845  which engages a drive pinion  846  having a plurality of gear teeth  848  disposed about its periphery. The drive pinion  846  is preferably fabricated of a strong and durable plastic such as nylon. The drive pinion  846  engages and bi-directionally drives an intermediate or idler gear  850 . The idler gear  850  includes a peripheral metallic annular gear  852  having gear teeth  854  which are complementary to, engaged by and driven by the gear teeth  848  on the drive pinion  846 . The idler gear  850  also includes an inner annular member  856  having a capital “I or H” shaped cross-section. The inner annular member  856  defines a central through passageway  858  which freely rotatably receives a stub shaft  860  which is supported within the housing  832 . The two components of the idler gear  850  are secured together by a snap ring  862  which is received within aligned, semi-circular channels formed in the inner surface of the annular gear  852  and outer surface of the inner member  856 .  
      The gear teeth  854  of the idler gear  850  engage both a first circular member or gear  866  having a plurality of peripheral gear teeth  868  and an adjacent, second circular member or gear  872  having a distinct number of peripheral gear teeth  874 . Just as with the other embodiments, the difference between the number of teeth on the first gear  866  and the second gear  872  is preferably one when there are on the order of a hundred to two hundred gear teeth. In circumstances wherein there are significantly more gear teeth either because the teeth are very small, because the gear is relatively large, or both, the difference in the number of gear teeth between the two gears  866  and  872  may be two, three, four or more, as described above.  
      Axially extending from the first gear  866  toward the second gear  872  is a cam follower arm  876  which terminates in a cam follower  878 . The second gear  872  includes a cam  880  having a cam profile or track  882  which may be of a uniform rise or fall, may include plural dwell and rise and fall regions or may have a highly irregular profile as illustrated. Whatever the profile of the cam track  882 , it will be appreciated that as the first and second gears  866  and  872  rotate, relative rotation therebetween will cause the cam follower  878  to proceed along the cam track  882  and axially translate the cam follower arm  876  and the first gear  866 .  
      The second gear  872  is preferably secured to a shaft  886  one end of which is rotatably received within a suitable tapered roller bearing  888 . The first gear  866  which both rotates and axially translates, is secured to a stub shaft  892  which is supported within a roller bearing assembly  894 , which in turn is supported within a linear output shaft  900 . The other end of the stub shaft  892  may telescope and nest within a suitable counterbore (not illustrated) in the shaft  886 . Between the output shaft  900  and the first gear  866  is a thrust bearing assembly  902  and between the output shaft  900  and the stationary housing  832  is a ball bearing assembly  904 . If desired, a compression or return spring  906  may be utilized to positively withdraw or return the linear output shaft  900  into the actuator housing  832  as the cam follower  878  descends the cam and the first gear  866  moves to the right as illustrated in  FIG. 12 .  
      Referring now to  FIGS. 13, 14  and  15 , a third additional embodiment of a differential drive actuator is illustrated and designated by the reference number  920 . The third additional embodiment differential drive actuator  920  includes a multiple piece, preferably cast housing  832 ′ which may be assembled and secured together by threaded fasteners  834 ′ disposed in suitable opening  836 ′. The housing  832 ′ preferably includes various apertures, shoulders, counterbores, flanges and the like which receive, rotatably support and provide mounting for various components of the actuator  920 . A fractional horsepower bi-directional electric motor  840 ′ is suitably secured to the housing  832 ′ and is provided with electrical energy through a multiple conductor cable  842 ′. The bi-directional electric motor  840 ′ includes an output shaft  844 ′ having a positively secured pin or vane  845 ′ which engages and drives a drive pinion  846 ′ having a plurality of gear teeth  848 ′ disposed about its periphery. The drive pinion  846 ′ is preferably fabricated of a strong and durable plastic such as nylon. The drive pinion  846 ′ engages and bi-directionally drives an intermediate or idler gear  850 ′. The idler gear  850 ′ includes a peripheral metallic annular gear  852 ′ having gear teeth  854 ′ which are complementary to, engaged by and driven by the gear teeth  848 ′ on the drive pinion  846 ′. The idler gear  850 ′ also includes an inner annular member  856  having a capital “I or H” shaped cross-section. The interior annular member  856 ′ defines a central through passageway  858 ′ which freely rotatably receives a stub shaft  860 ′ which is supported within the housing  832 ′. The two components of the idler gear  850 ′ are secured together by a snap ring  862 ′ which is received within aligned, semi-circular channels formed in the inner surface of the annular gear  852 ′ and outer surface of the inner member  856 ′.  
      The third additional embodiment differential drive actuator  920  also includes a bi-directional linear translation actuator assembly  930 . The linear translation actuator assembly  930  includes a first circular member or spur gear  932  having a plurality of peripheral teeth  934 .  
      The first gear or circular member  932  defines a central, circular, smooth walled aperture  936  which includes a plurality, preferably four, axially and radially inwardly extending, equally spaced apart keys  938 . The smooth walled aperture  936  receives a threaded shaft or lead screw  940  having a plurality, preferably four, keyways  942  equally spaced around its periphery. The keyways  942  are complementary to and receive the keys  938  formed in the first circular member  932 . Accordingly, the lead screw  940  rotates with the first gear or circular member  932 . Additionally, the lead screw  940  is free to translate axially within and relative to the first circular member  932 .  
      Adjacent the first gear or circular member  932  is the second gear or circular member or gear  946  having a plurality of teeth  948  disposed about its periphery. The gear teeth  854 ′ of the idler gear  850 ′ engage both the first gear or circular member  932  having a plurality of external gear teeth  934  and the adjacent, second gear  946  having a distinct plurality of peripheral gear teeth  948 . The number of teeth on the first circular member  932  is distinct from the number of teeth on the second circular member  946  such that driving both circular members  932  and  946  with a common drive will result in a speed disparity and relative rotation between the circular members  932  and  946 . Once again, a one tooth difference has been found suitable for gear or circular members  932  or  946  having approximately one hundred to two hundred teeth although a greater disparity between the number of teeth will increase actuator speed, reduce amplification and may be suitable when the gear or circular members  932  and  946  have a relatively large diameter or have a relatively large number of teeth. In such circumstances the difference in the number of gear teeth between the two gears  932  and  946  may be two, three, four or more, as described above.  
      The second circular member  946  includes a through opening  952  which includes female threads  954  which are complementary to the threads on the threaded shaft  940 . Three thrust bearings  956  are preferably disposed between each of the circular members  932 ,  946  and the housing  832 ′. As the first and second circular members  932  and  946  rotate differentially, the threaded shaft or lead screw  940  will be subjected to this differential rotation and this differential rotation will cause axial translation thereof. Inasmuch as the leadscrew  940  will rotate relatively rapidly while it undergoes axial translation, it should be understood that it may be desirable to dispose one or both ends of the lead screw within an assembly which transmits the axial motion but does not transmit the rotational motion such as the bearing and return spring assembly illustrated in the lower portion of  FIG. 12  discussed above.  
      It should be appreciated that the intermediate or idler gears  716 ,  850 , and  850 ′ may be eliminated if desired but the drive pinion  710 ,  846  and  846 ′ should be replaced with a metal gear if this is done.  
      It should also be appreciated that either with or without the idler gears  716 ,  850  and  850 ′, the drive pinions  710 ;  846  and  846 ′, the driven cam plates  740 ,  750  and the gears  732 ,  736 ,  866  and  872  may be sized to effect a rotational speed reduction (or increase), thereby increasing (or decreasing) output torque and force but reducing (or increasing) operating speed (response time).  
      It should further be appreciated that while described above in connection with an electric motor, the invention is equally usable with pneumatic, e.g., vane, motors and hydraulic motors which may be utilized with and are within the purview of this invention  
      The foregoing disclosure is the best mode devised by the inventor for practicing this invention. It is apparent, however, that devices incorporating modifications and variations will be obvious to one skilled in the art of motor driven ball ramp clutches. Inasmuch as the foregoing disclosure is intended to enable one skilled in the pertinent art to practice the instant invention, it should not be construed to be limited thereby but should be construed to include such aforementioned obvious variations and be limited only by the scope and spirit of the following claims.