Abstract:
A synchronizer for motor vehicle driveline components such as transmissions and transfer cases includes a pair of ball ramp mechanisms which amplify synchronizing force and apply such force to one of a pair of friction clutch packs which achieves synchronization of diversely rotating elements. The synchronizer exhibits relatively low operating force. Each ball ramp mechanism includes a pair of ball ramp members having rotational travel limits which also limit their axial separation and an intermediate compression spring assembly. The spring assembly and the travel limits of the ball ramp members limit the force generated by the ball ramp mechanisms and applied to the friction clutch packs to prevent abrupt and uncontrolled actuation of the friction clutch packs and corresponding abrupt synchronization of the rotating elements.

Description:
BACKGROUND OF THE INVENTION  
         [0001]    The invention relates generally to synchronizers for motor vehicle driveline components such as transmissions and transfer cases and more particularly to a synchronizer for such components having a ball ramp actuator mechanism.  
           [0002]    It is frequently necessary in motor vehicle driveline components to connect one rotating member to another which is rotating at a different speed. In this situation, it is generally desirable to synchronize, or match, the speeds of the two rotating members prior to engagement to achieve a smooth engagement. Synchronizers for transmissions and transfer cases have existed for over seventy years. The most frequently encountered synchronizer is found in manual transmissions utilized in cars and trucks.  
           [0003]    With the increasing sophistication of motor vehicle powertrain products, newer devices utilize sensors which monitor the speed of rotating members to be engaged and adjust, often by braking, the speed of one of the members to synchronize it with the other. In some drivetrains, electronic controllers capable of adjusting the speed of both the engine and transmission achieve synchronism of rotating members without the use of any mechanical synchronizer.  
           [0004]    Notwithstanding such specialized devices, there is still a demand for synchronizers in, for example, transfer cases. A frequent requirement for transfer case synchronizers is rapid operation, that is, synchronization must be achieved typically within less than one second and such synchronizers must operate with relatively low applied force. As is often the case, these two requirements are mutually exclusive: a faster operating synchronizer generally necessitates a larger device with greater delivered force and energy requirements; a smaller device, though requiring less energy, typically generates less force and will require a longer time to achieve synchronization.  
           [0005]    Accordingly, it is apparent that synchronizers exhibiting fast response time and low operating forces and power consumption are desirable.  
         SUMMARY OF THE INVENTION  
         [0006]    A synchronizer for motor vehicle driveline components such as transmissions and transfer cases includes a pair of ball ramp mechanisms which amplify synchronizing force and apply such force to one of a pair of friction clutch packs which achieves synchronization of diversely rotating elements. The synchronizer exhibits relatively low operating force. Each ball ramp mechanism includes a pair of ball ramp members having rotational travel limits which also limit their axial separation and an intermediate compression spring assembly. The spring assembly and the travel limits of the ball ramp members limit the force generated by the ball ramp mechanisms and applied to the friction clutch packs to prevent abrupt and uncontrolled actuation of the friction clutch packs and corresponding abrupt synchronization of the rotating elements.  
           [0007]    Thus, it is an object of the present invention to provide a synchronizer for use with motor vehicle driveline components such as transmissions and transfer cases.  
           [0008]    It is a further object of the present invention to provide a synchronizer having a ball ramp actuator mechanism and friction clutch packs.  
           [0009]    It is a still further object of the present invention to provide a synchronizer having relatively low actuating force.  
           [0010]    It is a still further object of the present invention to provide a synchronizer for motor vehicle driveline components having a pair of ball ramp mechanisms having both rotational and force limiting components.  
           [0011]    It is a still further object of the present invention to provide a synchronizer for a motor vehicle driveline component having a pair of ball ramp mechanisms with stops which limit relative rotation of the ball ramp members.  
           [0012]    Further objects and advantages of the present invention will become apparent by reference to the following description of the preferred embodiment and appended drawings wherein like reference numbers refer to the same component, element or feature. 
       
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0013]    [0013]FIG. 1 is a diagrammatic view of a motor vehicle driveline having a transfer case incorporating the present invention;  
         [0014]    [0014]FIG. 2 is a full, sectional view of a motor vehicle transfer case incorporating a synchronizer according to the present invention;  
         [0015]    [0015]FIG. 3 is an enlarged, fragmentary, sectional view of a synchronizer according to the present invention;  
         [0016]    [0016]FIG. 4 is a greatly enlarged, fragmentary, sectional view of a synchronizer assembly according to the present invention;  
         [0017]    [0017]FIG. 5 is a full, sectional view of a portion of a synchronizer according to the present invention taken along line  5 - 5  of FIG. 4; and  
         [0018]    [0018]FIG. 6 is a flat pattern development of the ramped recesses and load transferring ball of one of the ball ramp actuators of a transfer case electromagnetic clutch taken along line  6 - 6  of FIG. 2. 
     
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT  
       [0019]    Referring now to FIG. 1, a four-wheel vehicle drive train is diagrammatically 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 transmission  14  may either be an automatic or manual type. The output of the transmission  14  directly drives a transfer case assembly  16  which provides motive power to a primary or rear drive line  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 .  
         [0020]    The transfer case assembly  16  also selectively provides motive power to a secondary or front drive line  30  comprising a secondary or front prop shaft  32 , a secondary or front differential assembly  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 pair of front axles  36  or, if desired, a pair of manually or remotely activateable locking hubs  42  may be operably disposed between the pair of front axles  36  and a respective one of the tire and wheel assemblies  38  to selectively connect same. Finally, both the primary drive line  20  and the secondary drive line  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. A control console  46  which is preferably disposed within convenient reach of the vehicle operator includes a switch or a plurality of individual switches or push buttons  48  which facilitate selection of the operating mode of the transfer case assembly  16  as will be further described below.  
         [0021]    The foregoing and following description relates to a vehicle wherein the rear drive line  20  functions as the primary drive line, i.e., it is engaged and operates substantially all the time and, correspondingly, the front drive line  30  functions as the secondary drive line, i.e., it is engaged and operates only part-time or in a secondary or supplemental fashion, such a vehicle commonly being referred to as a rear wheel drive vehicle.  
         [0022]    These designations “primary” and “secondary” are utilized herein rather than “front” and “rear” inasmuch as the invention herein disclosed and claimed may be readily utilized in transmissions and transfer cases wherein the primary drive line  20  is disposed at the front of the vehicle and the secondary drive line  30  is disposed at the rear of the vehicle. Such designations “primary” and “secondary” thus broadly and properly characterize the function of the individual drive lines rather than their specific locations.  
         [0023]    Referring now to FIGS. 1 and 2, the transfer case assembly  16  incorporating the present invention includes a multiple piece, typically cast, metal housing assembly  50  having planar and circular sealing surfaces, openings for shafts and bearings and various recesses, shoulders, flanges, counterbores and the like to receive various components and assemblies of the transfer case assembly  16 . An input shaft  52  includes female or internal splines or gear teeth  54  or other suitable structure which drivingly couple an output of the transmission  14  (illustrated in FIG. 1) to the input shaft  52 . The input shaft  52  is rotatably supported externally by an anti-friction bearing such as a ball bearing assembly  56  and internally by an anti-friction bearing such as a roller bearing assemblies  58 . The roller bearing assemblies  58  are disposed upon a reduced diameter portion of a primary output shaft  60 . An oil seal  62 , positioned between the input shaft  52  and the housing assembly  50 , provides an appropriate fluid tight seal therebetween.  
         [0024]    The opposite end of the output shaft  60  is supported by an anti-friction bearing such as a ball bearing assembly  64 . An end cap or seal  66  closes off the end of an axial passageway  68  in the primary output shaft  60 . A gerotor pump  70  will typically be utilized to provide a flow of lubricating and cooling fluid to the axial passageway  68  which is thence distributed through a plurality of radial ports in the primary output shaft  60  to the components of the transfer case assembly  16 . An oil seal  72  positioned between the housing  50  and an output feature such as a flange  74  achieves a seal between the housing  50  and the primary output shaft  60 .  
         [0025]    Referring now to FIGS. 2 and 3, the transfer case assembly  16  includes a planetary gear speed reduction assembly  80 . The planetary gear speed reduction assembly  80  includes a sun gear  82  which may be a collar coupled by interengaging splines  84  to the input shaft  52  or may be integrally formed therewith. The sun gear  82  includes gear teeth  86  which are in constant mesh with a plurality of pinion or planet gears  88 . The planet gears  88  may be rotatably disposed upon roller bearings  92  which in turn are supported by fixed stub shafts  94  or the pinion gears  88  may be rotatably supported directly upon the stub shafts  84 , if desired. The stub shafts  94  are retained and secured within a planet carrier  96  which includes a bell shaped extension  98  and male or external splines or gear teeth  100 . The planet carrier  96  is also supported by a circular disc  102  which engages a shoulder  104  on the input shaft  52  on one side and is axially positioned by a spacer  106  on its opposite side.  
         [0026]    The plurality of pinion or planet gears  88  are in constant mesh with gear teeth  112  of a stationary ring gear  114  which is secured within the housing  50  by, for example, a snap ring  116 .  
         [0027]    The input shaft  52  includes an elongate sleeve or quill  122 . The sleeve or quill  122  is rotatably supported by the a pair of roller bearing assemblies  58 . The sleeve or quill  122  of the input shaft  52  includes male splines or gear teeth  126  which are spaced from and axially aligned with the male splines or gear teeth  100  on the planet gear carrier  96 .  
         [0028]    Referring now to FIGS. 3 and 4, the transfer case assembly  16  also includes a synchronizer assembly  130 . The synchronizer assembly  130  includes an outer annular shift collar  132  which may be bi-directionally translated from its center position by corresponding bidirectional motion of a shift fork  134  which is received within a circumferential channel or groove  136 . The annular shift collar  132  includes internal or female splines or gear teeth  138  which are complementary to and in constant engagement with male or external splines or gear teeth  142  formed on an annular member  144  of the primary output shaft  60 . The annular shift collar  132  also includes internal or female splines or gear teeth  148  which are complementary to and axially aligned with the male splines  100  on the planetary gear carrier  96  and the male splines  126  on the input shaft sleeve or quill  122 . An inner detented collar  150  having external or male splines or gear teeth  152  is received within and rotates with the annular shift collar  132 . The detented collar  150  includes a circumferential channel  154  which is capable of receiving a contractable circumferential spring  156  which, in its relaxed state, resides within a shallow, oblique-walled circumferential recess  158  in the outer annular shift collar  132 . The circumferential groove  154 , the contractable spring  156  and the oblique-walled circumferential recess  158  cooperate to provide a detenting action between the detented collar  150  and the annular shift collar  132  which provides some resistance to motion of the detented collar  150 , when it is moved out of the position illustrated in FIG. 3 and assists return of the detented collar  150  to the center position illustrated in FIG. 3.  
         [0029]    Referring now to FIGS. 3, 4 and  5 , centrally disposed on the inner surface of the detented collar  150  is a region of axially extending internal or female splines or gear teeth  162 . The splines or gear teeth  162  engage first or inner left and right circular ball ramp members  164 A and  164 B which include splines  166 A and  166 B complementary to and engaged with the female splines or gear teeth  162  on the detented collar  150 . The circular members  164 A and  164 B include a plurality of oblique walled, ramped recesses  168 A and  168 B which receive a like plurality of load transferring balls  172 A and  172 B. Preferably, the ramped recesses  172 A and  172 B extend angularly over approximately 90° to 100°. The circular members  164 A and  164 B also include internal or female splines or gear teeth  174 A and  174 B which engage complementarily configured male or external splines or gear teeth  178  on a collar or hub  180  which is freely rotatably disposed upon the input shaft sleeve or quill  122 . The splines or gear teeth  178  are non-standard in that only three splines or gear teeth  178  disposed at 120° intervals reside on the collar or hub  180 . It will be appreciated that the primary output shaft  60 , the annular shift collar  132 , the detented collar  150 , the first or inner clutch members  164 A and  164 B and the collar or hub  180  all rotate together.  
         [0030]    Operably disposed between the adjacent faces of the first or inner left and right circular ball ramp members  164 A and  164 B is a compression spring  182 . The compression spring  182  may be a Belleville washer, a wave washer or a circular disc having a plurality of small compression springs disposed along axes parallel to and equidistant from the center line of the primary output shaft  60 .  
         [0031]    The synchronizer assembly  130  also includes second or outer left and right circular ball ramp members  184 A and  184 B each having a corresponding plurality of oblique walled, ramped recesses  188 A and  188 B. Preferably, the ramped recesses  188 A and  188 B extend angularly over approximately 90° to 100°. As illustrated in FIG. 5, the three male or external splines  178  spaced at 120° intervals engage with a corresponding number, i.e., three, spaced apart splines  192  on each of the pair of second circular members  184 A and  184 B. Accordingly, the second circular members  184 A and  184 B are free to rotate through a limited range of travel relative to the collar or hub  180 . Such range of travel is on the order of 80 to 90 angular degrees and thus the relative rotation and the axial displacement of the circular members  164 A and  164 B relative to the corresponding circular member  184 A and  184 B are limited.  
         [0032]    It will be appreciated that the ramped recesses  168 A,  168 B,  188 A and  188 B and the load transferring balls  172 A and  172 B may be replaced with other analogous mechanical elements which cause axial displacement of the circular members  164 A,  164 B,  184 A and  184 B in response to relative rotation therebetween. For example, tapered rollers disposed in complementarily configured conical helices may be utilized.  
         [0033]    Each of the second or circular outer members  184 A and  184 B includes a respective shoulder  192 A and  192 B which traps and engages a corresponding flat washer  194 A and  194 B. The opposite faces of each of the flat washers  194 A and  194 B engage the internal splines or gear teeth  162  on the inner detented collar  150 . Thus, as the detented collar  150  moves to the left or right from the position illustrated in FIG. 3, the female or internal splines or gear teeth  162  engage and translate one of the flat washers  194 A or  194 B in a direction corresponding to the direction of travel of the detented collar  150  and correspondingly translate one of the second or outer circular members  184 A or  184 B into engagement with a corresponding left and right friction clutch pack  200 A or  200 B.  
         [0034]    The left and right friction clutch packs  200 A and  200 B include a first plurality of larger clutch plates or discs  202 A and  202 B. The larger friction plates or discs  202 A on the left engage complementarily configured splines or gear teeth  204 A on the bell shaped portion  98  of the planet carrier  96 . A second plurality of smaller diameter friction clutch plates or discs  206 A on the left engage the splines  178  on the collar or hub  180 . Correspondingly, a first set of larger friction clutch plates or discs  202 B on the right engage a complementary plurality of internal or female splines or gear teeth  204 B on the elongate quill or sleeve  122  of the input shaft  52 . A second, interleaved plurality of smaller diameter friction clutch plates or discs  206 B on the right also engage the splines  178  on the collar or hub  180 .  
         [0035]    Referring again to FIG. 2, the shift fork  134  is part of a shift operator assembly  210 . The shift fork  134  extends radially from a cylindrical body  212  having a pair of identical cams  212 A at each end. The cams  212 A are engaged by a pair of spaced apart cam followers  214  which are secured to a bi-directionally rotatable shift shaft or rail  216 . The shift rail  216  is bi-directionally rotated by an electric motor drive mechanism  218  which selectively, bi-directionally rotates the shift rail  216  and axially translates the shift fork  134  to axially, bi-directionally, move the outer elongate shift collar  132 .  
         [0036]    Referring now to FIGS. 2 and 6, the transfer case assembly  16  also includes an electromagnetically actuated disc pack type clutch assembly  220  which effects selective torque transfer from the primary output shaft  60  to the secondary drive line  30 . The disc pack clutch type assembly  220  is disposed about the primary output shaft  60  and includes a circular drive member  222  coupled to the primary output shaft  60  through, for example, a splined interconnection. The circular drive member  222  includes a plurality of circumferentially spaced-apart recesses  226  in the shape of an oblique section of a helical torus. Each of the recesses  226  receives one of a like plurality of load transferring balls  228 .  
         [0037]    A circular driven member  232  is disposed adjacent the circular drive member  222  and includes a like plurality of opposed recesses  234  defining the same shape as the recesses  226 . The oblique side walls of the recesses  226  and  234  function as ramps or cams and cooperate with the balls  228  to drive the circular members  222  and  232  apart in response to relative rotation therebetween. It will be appreciated that the recesses  226  and  234  and the load transferring balls  228  may be replaced with other analogous mechanical elements which cause axial displacement of the circular members  222  and  232  in response to relative rotation therebetween. For example, tapered rollers disposed in complementarily configured conical helices may be utilized.  
         [0038]    The circular driven member  232  extends radially outwardly and is secured to a soft iron rotor  236 . An armature  242  is disposed adjacent the face of the rotor  236 . The rotor  236  surrounds an electromagnetic coil  244  on three sides.  
         [0039]    The electromagnetic coil  244  is provided with electrical energy preferably from a pulse width modulation (PWM) controller through an electrical conductor  246 . The pulse width modulation scheme increases or decreases the average current to the electromagnetic coil  244  of the electromagnetic clutch assembly  220  and thus the torque throughput of the disc pack type clutch assembly  220 , as will be more fully described below, by increasing or decreasing the on time (duty cycle) of a drive signal. It will be appreciated that other modulating control techniques may be utilized to achieve engagement and disengagement of the electromagnetic disc pack type clutch assembly  220 .  
         [0040]    Providing electrical energy to the electromagnetic coil  244  causes magnetic attraction of the armature  242  with the rotor  236 . This magnetic attraction results in frictional contact of the armature  242  to the rotor  236 . When the primary output shaft  60  is turning at a different speed than the armature  242  this frictional contact results in a frictional torque being transferred from the primary output shaft  60 , through the circular drive member  222 , through the load transferring balls  228  and to the circular driven member  232 . The resulting frictional torque causes the balls  228  to ride up the ramps of the recesses  226  and  234 , causing axial displacement of the circular drive member  222 . Axial displacement of the circular drive member  222  translates an apply plate  248  axially toward a disc pack clutch assembly  250 . A compression spring  252  which may comprise a stack of Belleville washers provides a restoring force which biases the circular drive member  222  toward the circular driven member  232  and returns the load transferring balls  228  to center positions in the circular recesses  226  and  234  to provide maximum clearance and minimum friction between the components of the electromagnetic clutch assembly  220  when it is deactivated. An important design consideration of the recesses  226  and  234  and the balls  228  is that the geometry of their design and the design of the compression spring  252  and the clearances in the disc pack assembly  250  ensure that the electromagnetic clutch assembly  220  is not self-locking. The electromagnetic clutch assembly  220  must not self-engage but rather must be capable of controlled, proportional engagement and torque transfer in direct response to the modulating control input.  
         [0041]    The disc pack clutch assembly  250  includes a first plurality of smaller friction plates or discs  254 . The first plurality of discs  254  are coupled by interengaging splines to a clutch hub  256  which is coupled to the primary output shaft  60  for rotation therewith. A second plurality of larger friction plates or discs  258  are coupled to an annular housing  260  by interengaging splines for rotation therewith and are interleaved with the first plurality of friction discs  254 .  
         [0042]    The annular housing  260  is disposed concentrically about the primary output shaft  60  and is coupled to a chain drive sprocket  262  by a plurality of interengaging splines or lugs and recesses  264 . The chain drive sprocket  262  is freely rotatably disposed on the primary output shaft  60  and is supported by a journal or needle bearing assembly  266 . When the clutch assembly  220  is engaged, it transfers torque from the primary output shaft  60  to the chain drive sprocket  262 . A drive chain  268  is received upon the chain drive sprocket  262  and engages and transfers energy to a driven chain sprocket  270 . The driven chain sprocket  270  is coupled to a front or secondary output shaft  272  of the transfer case assembly  16  by interengaging splines  274 . The secondary output shaft  272  is rotatably supported by a pair of roller bearing assemblies  276  and an oil seal  278  provides a fluid tight seal between the secondary output shaft  292  and the housing  50 .  
         [0043]    In operation, the synchronizer assembly  130  according to the present invention provides rapid synchronism while utilizing relatively low engagement force. Thus, the associated shift operator may be lighter, smaller and exhibit lower power consumption than many conventional designs. When a shift is commanded, the shift fork  134  begins to move the outer annular shift collar  132  to the right or to the left from the position illustrated in FIGS. 2, 3 and  4 . In the following explanation, it will be assumed that the outer annular shift collar  132  is being moved to the left as illustrated in FIGS. 2, 3 and  4  to engage the reduced speed output from the carrier  96  of the planetary gear speed reduction assembly  80 . Translation of the outer annular shift collar  132  to the right engages direct drive from the input shaft  52  but the action of the synchronizer assembly  130  is essentially the same.  
         [0044]    As the outer annular shift collar  132  moves to the left, the contractable spring  156  is driven by the oblique sidewalls  158  into the circumferential channel  154  of the inner detented collar  150 . The detented collar  150  likewise begins to move to the left and the female or internal splines or gear teeth  162  translate the flat washer  194 A which in turn, translates the second or outer left circular ball ramp member  184 A into increased frictional engagement with the left friction clutch pack  200 A. The drag so created causes relative rotation between the outer circular ball ramp member  184 A and the inner circular ball ramp member  164 A causing the load transferring balls  172 A to axially separate the circular members  164 A and  184 A.  
         [0045]    Both the relative rotation of the inner and outer circular members  164 A and  184 A and thus their axial separation is limited by the cooperative action of the splines  178  and  192 . The axial separation of the inner and outer circular members  164 A and  184 A compresses the friction clutch pack  200 A and begins to drive the planetary gear carrier  96  into synchronism with the primary output shaft  60 . The compressive force applied to the friction clutch pack assembly  200 A is controlled and limited by the compressive force generated by the compression spring  182  and, in fact, can be no greater than that force generated by the compression spring  182 . It must be appreciated that the adjacent first or inner circular members  164 A and  164 B must not be permitted to touch or contact one another as this would allow force in excess of that controlled or limited by the compression spring  182  to be applied to the friction clutch packs  200 A and  200 B and provide abrupt and unacceptable synchronizer operation.  
         [0046]    In this regard, it should also be appreciated that selection of the spring rate of the compression spring  182  will control the force applied to the friction clutch packs  200 A and  200 B and thus the relative speed of synchronization achieved by the synchronizer assembly  130 . That is, a higher or stiffer spring rate will allow more force to be applied to the friction clutch packs  200 A and  200 B resulting in faster synchronization and a lower or softer spring rate will achieve a slower rate of synchronization.  
         [0047]    When the speed of the planet carrier  96  matches that of a primary output shaft  60 , the outer annular shift collar  132  may be further advanced to the left such that the female or internal splines or gear teeth  148  may be engaged with the male splines or gear teeth  100  on the planetary gear carrier  96 . In this condition, drive torque is transferred directly from the planetary gear carrier  96  through the outer annular shift collar  132 , through the inter-engaging splines  138  and  142  and to the primary output shaft  60 .  
         [0048]    The foregoing disclosure is the best mode devised by the inventors for practicing this invention. It is apparent however, that devices incorporating modifications and variations will be obvious to one skilled in the art of mechanical synchronizers. Inasmuch as the foregoing disclosure presents the best mode contemplated by the inventors for carrying out the invention and is intended to enable any person skilled in the pertinent art to practice this 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 spirit and scope of the following claims.