Abstract:
A transfer case for a four-wheel drive vehicle includes a center differential which provides a direct output in two-wheel drive and a slightly reduced speed output in four-wheel drive. In two-wheel drive, the outputs of the center differential are locked together by a clutch collar and drive torque is provided to the primary drive line only. In four-wheel drive, the clutch collar unlocks the differential and couples the output from a sun gear to a chain drive sprocket in the secondary drive line. The center differential is driven through the carrier and operates as an open differential. The ring (outer) gear drives the primary drive line and the sun (center) gear drives the secondary drive line through cooperating chain sprockets and a chain. The chain drive sprockets are of unequal size and effect a speed reduction between the drive and driven sprockets. This speed reduction is nominally the same as the speed reduction from the input to the primary output achieved through the center differential. Thus, in four-wheel drive mode, the drive (speed reduction) ratio through the drive lines is raised, improving the acceleration and pulling power of the vehicle. In two-wheel drive, more favorable fuel economy is provided with the lower drive ratio.

Description:
CROSS REFERENCE TO CO-PENDING APPLICATION 
     This application is a divisional application of Ser. No. 09/191,186, filed Nov. 13, 1998 now U.S. Pat. No. 5,984,821, granted Nov. 16, 1999. 
    
    
     BACKGROUND OF THE INVENTION 
     The invention relates generally to transfer cases for four-wheel drive vehicles and more specifically to a transfer case having a center differential which is configured to selectively provide a four-wheel underdrive operating mode. 
     The obvious benefits of improved traction and vehicle control achieved by four-wheel drive systems in adverse driving conditions such as snow, freezing rain, ice and even water has been known and appreciated by vehicle designers for many years. A not so obvious benefit relates to the use of four-wheel drive when towing a trailer carrying a boat, snowmobiles and the like. Here, too, the improved traction provides improved driver control and stability in adverse driving conditions. When trailer towing is viewed as a specific operational mode, additional features and requirements may be added to the list of vehicle design criteria. For example, when the vehicle weight is augmented by several hundred or several thousand pounds, a slightly higher overall drive ratio provides improved torque and thus improved acceleration. 
     In contrast, when the vehicle is utilized in two-wheel drive the presumption exists that neither the enhanced traction nor improved acceleration would be beneficial or necessary inasmuch as either the vehicle is not towing a trailer, the road conditions are good or both. 
     The present invention is directed a transfer case having a center differential which provides direct two-wheel drive and an underdrive four-wheel drive operating mode which improves vehicle performance. 
     SUMMARY OF THE INVENTION 
     A transfer case for a four-wheel drive vehicle includes a center, planetary gear differential which provides a direct output in two-wheel drive and a slightly reduced speed output in four-wheel drive. In two-wheel drive, the outputs of the center differential are locked together by a clutch collar and drive torque is provided to the primary drive line only. In four-wheel drive, the clutch collar unlocks the differential and couples the output from a sun gear to a chain drive sprocket in the secondary drive line. The center differential is driven through the carrier, operates as an open differential and is preferably configured with a 65% primary (rear), 35% secondary (front) drive line torque split. The ring (outer) gear drives the primary drive line and the sun (center) gear drives the secondary drive line through cooperating chain sprockets and a chain. The chain sprockets are of unequal size and effect a speed reduction between the drive and driven sprockets. This speed reduction is nominally the same as the speed reduction from the input to the primary output achieved through the center differential. Thus, in four-wheel drive mode, the drive ratio through the drive lines is raised and the speed reduced, improving the acceleration and pulling power of the vehicle. In two wheel drive, more favorable fuel economy is provided with the lower (direct) drive ratio. With the open center differential, skid control and torque distribution may be achieved through the vehicle&#39;s antilock or anti-skid braking system. 
     It is thus an object of the present invention to provide a transfer case having both direct drive and underdrive modes of operation. 
     It is a further object of the present invention to provide a transfer case having a center differential which may be locked out to provide direct two-wheel drive and unlocked to provide four-wheel drive at a reduced speed (underdrive) mode. 
     It is a still further object of the present invention to provide a transfer case having an open center differential wherein skid control is provided by, for example, an anti-lock braking system incorporated in the motor vehicle which is independent of the transfer case. 
     It is a still further object of the present invention to provide a motor vehicle transfer case wherein an open center differential and unequal chain sprockets provide drive torque to primary and secondary drive lines in a four-wheel underdrive mode to improve vehicle performance. 
     Further objects and advantages of the present invention will become apparent by reference to the following description and appended drawing wherein like reference numbers refer to the same component, element or feature. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a diagrammatic, plan view of a four-wheel drive motor vehicle powertrain having a transfer case incorporating the present invention; 
     FIG. 2 is a full, sectional view of a motor vehicle transfer case incorporating a center differential according to the present invention; and 
     FIG. 3 is an enlarged, fragmentary sectional view of an open differential according to the present invention. 
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     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 . 
     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  are preferably directly coupled to a respective one of the pair of front axles  36 . Alternately, 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. 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. 
     An operator control console or assembly  46  is preferably disposed within easy reach of the vehicle operator and includes a switch or a plurality of pushbuttons  48  which select one of four operating modes of the transfer case assembly  16 : H2—two-wheel drive by the primary drive line  20  in high gear (direct drive); N—neutral; H4—four-wheel drive through the primary and secondary drive lines  20  and  30  in reduced high gear (underdrive) and L4-four-wheel drive through the primary and secondary drive lines  20  and  30  in low gear. 
     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. 
     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. 
     Referring now to FIGS. 1 and 2, the transfer case assembly  16  incorporating the present invention includes a multiple piece, typically cast, 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 or assemblies of the transfer case  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 the ball bearing assembly  56  and internally by an anti-friction bearing such as the roller bearing assembly  58 . The roller bearing assembly  58  is disposed upon a portion of a stepped, intermediate shaft  60 . A suitable oil seal  62 , positioned between the input shaft  52  and the housing assembly  50 , provides an appropriate fluid tight seal therebetween. The opposite end of the intermediate shaft  60  is supported by an anti-friction bearing such as a roller bearing assembly  64 . An end cap or seal  66  closes off the end of an axial passageway  68  in the intermediate shaft  60 . A gerotor pump P illustrated in phantom lines 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 intermediate shaft  60  to the components of the transfer case assembly  16 . 
     Referring now to FIG. 2, the transfer case assembly  16  includes a two-speed planetary (epicyclic) gear assembly  70  disposed generally about the input shaft  52 . The planetary gear assembly  70  includes a sun gear  72  having a plurality of external gear teeth  74  and a plurality of internal splines or gear teeth  76  which are both formed on an axial extension  78  of the input shaft  52 . Radially aligned with the sun gear  72  and its teeth  74  is a ring gear  80  having internal gear teeth  82 . The ring gear  80  is fixedly retained within the housing assembly  50  by any suitable retaining structure such as a projection or lip  84  formed in portions of the housing assembly  50  and a cooperating snap ring  86 . A plurality of pinion gears  88  are rotatably received upon a like plurality of anti-friction bearings such as roller bearings  90  which, in turn, are supported and located by a like plurality of stub shafts  92 . The plurality of stub shafts  92  are mounted within and secured to a planet carrier  94 . The planet carrier  94  includes a plurality of internal splines or gear teeth  96  disposed generally adjacent the internal splines or gear teeth  76  on the extension  78  of the input shaft  52 . The planetary gear assembly  70  is more fully described in co-owned U.S. Pat. No. 4,440,042 which is herein incorporated by reference. 
     The planetary gear assembly  70  also includes a first dog clutch or clutch collar  100  defining elongate internal splines or gear teeth  102 . The internal splines or gear teeth  102  of the first clutch collar  100  are slidably received upon a complementary plurality of external splines or gear teeth  104  on the intermediate shaft  60 . The first clutch collar  100  thus rotates with the intermediate shaft  60  but may translate bi-directionally therealong. The first clutch collar  100  also includes external splines or gear teeth  106  on one end which are in all respects complementary to the internal splines or gear teeth  76  on the axial extension  78  of the input shaft  52  and the internal splines or gear teeth  96  on the planet carrier  94 . The opposite end of the first clutch collar  100  defines a circumferentially and radially extending flange  108 . 
     The first clutch collar  100  is capable of three positions and operational modes. In the lower portion of FIG. 2, the first clutch collar  100  is illustrated in its leftmost or direct drive position. Direct drive is achieved when the external splines or gear teeth  106  of the first clutch collar  100  engage the internal splines or gear teeth  76  on the axial extension  78  of the input shaft  52  thereby directly coupling the input shaft  52  to the intermediate shaft  60  and providing direct or high gear drive therebetween. When the first clutch collar  100  is moved to the right from the position illustrated in the lower portion of FIG. 2, to the position illustrated in the upper portion of FIG. 2, the speed reduction achieved by the planetary gear assembly  70  is engaged through engagement of the external splines or gear teeth  106  on the first clutch collar  100  with the internal splines or gear teeth  96  on the planet carrier  94 . So engaged, the planetary gear assembly  70  is active and provides a speed reduction, typically in the range of from 3:1 to 4:1 between the input shaft  52  and the intermediate shaft  60 . Between these two positions is a neutral position. In the center, neutral position both the input shaft  52  and the planet carrier  94  are disconnected from the intermediate shaft  60  and no power is transmitted therebetween. 
     The position of the first clutch collar  100  is commanded by an electric shift control motor  110 . The shift control motor  110  rotates a drive shaft or shift rail  112  which is supported for rotation in the housing assembly  50  by a pair of bushings or bearings  114 . The shift rail  112  includes a first pair of radially extending, spaced apart pins or cam followers  116 . Disposed between the first pair of cam followers  116  is a first shift fork assembly  120 . The first shift fork assembly  120  includes a first shift fork body  122  defining a through passageway  124  which receives the shift rail  112 . At opposite ends of the first shift fork body  122  are complementarily configured, spaced apart helical cam surfaces  126  and axially extending discontinuities or flats  128 . The helical cam surfaces  126  are configured and spaced a distance slightly less than the separation of the cam followers  116  such that rotation of the shift rail  112  axially translates the shift fork assembly  120  a limited distance along the shift rail  112 . The discontinuities or flats  128  function as stops which limit rotation of the shift rail  112  in one direction. The shift fork assembly  120  also includes a yoke  132  which defines a semicircular channel or groove  134 . The semicircular groove  134  is complementary to and engages a portion of the flange  108  on the first clutch collar  100  thus axially and bidirectionally repositioning the first clutch collar  100  in response to rotation of the shift rail  112 , as will be readily appreciated. 
     It should be understood that the planetary gear assembly  70  including the first clutch collar  100  which provides dual range, i.e., high and low speed, capability of the transfer case assembly  16  is optional, that the four-wheel vehicle drive train  10  and the transfer case assembly  16  are fully functional as a single speed, direct drive unit and that the present invention may be utilized without the planetary gear assembly  70  and the dual speed range capability provided thereby. 
     Turning now to FIGS. 2 and 3, the transfer case assembly  16  also includes a planetary gear, center differential assembly  140 . The center differential assembly  140  is preferably also an epicyclic gear train device and includes a planet carrier  142  having internal splines or gear teeth  144  which engage and are driven by complementarily configured external splines or gear teeth  146  formed on a stepped, terminal portion  148  of the intermediate shaft  60 . The planet carrier  142  receives a plurality, preferably three, axially extending stub shafts  150  which support needle bearing assemblies  152  which rotatably receive a plurality of pinion or planet gears  154 . The plurality of pinion or planet gears  154  each include gear teeth  156  which are in constant mesh with internal gear teeth  158  formed on the interior of a ring gear annulus  162 . The ring gear annulus  162  is engaged by, coupled to and supported by a plurality of external gear teeth  164  formed on the periphery of an obliquely extending circular member  166  which constitutes a portion of a primary output shaft  168 . A snap ring  170  seated in a circumferential groove in the gear teeth  158  retains the circular member  166  of the output shaft  168  within the ring gear annulus  162 . 
     As noted above, the roller bearing assembly  64  rotatably supports the stepped, terminal portion  148  of the intermediate shaft  60  in a counterbore in the primary output shaft  168 . The primary output shaft  168  is preferably rotatably-supported upon an anti-friction bearing such as a ball bearing assembly  172 . The primary output shaft  168  may include external splines or gear teeth  174  which mate with internal splines or gear teeth  176  in an output flange  178  or other output or driveline structure. The output flange  178  is preferably secured to the primary output shaft  168  by a suitable retainer such as a nut  182 . An oil seal  184  provides a fluid tight seal between the flange  178  and the housing  50  of the transfer case assembly  16 . 
     At the opposite end of the ring gear annulus  162  from the obliquely extending circular member  166  is a circular end plate  192 . The circular end plate  192  includes external gear teeth  194  which are complementary to and engage the internal gear teeth  158  on the ring gear annulus  162 . The circular end plate  192  is disposed adjacent the planet carrier  142  and is axially restrained by a snap ring  196  received within a suitable circumferential groove in the gear teeth  158 . The circular end plate  192  also includes a plurality of internal splines or gear teeth  198 . It will be appreciated that various friction reducing flat washers  200  may be disposed between components of the center differential assembly  140  which rotate at different speeds. 
     A drive sleeve  202  is freely rotatably disposed about the intermediate shaft  60 . A friction reducing bushing  204  may be disposed between the drive sleeve  202  and the intermediate shaft  60 , if desired. The drive sleeve  202  includes a plurality of external gear teeth  206  which are in constant mesh with the gear teeth  156  of each of the plurality of pinion or planet gears  154 . The drive sleeve  202  and its external gear teeth  206  thus function as a sun gear, cooperating with the pinion or planet gears  154  which, in turn, cooperate with the gear teeth  158  of the ring gear annulus  162  to form a planetary or epicyclic gear train which, as arranged, functions as an open center differential. 
     The drive sleeve  202  also includes a region of external splines or gear teeth  208  which are disposed generally opposite the external gear teeth  206  which form the sun gear. The drive sleeve  202  is retained and axially positioned by a snap ring  212  which is received within a suitable circumferential groove  214  formed in the intermediate shaft  60 . Axially, bi-directionally slidable upon the external splines or gear teeth  208  of the drive sleeve  202  is a second dog clutch or clutch collar  220 . The second clutch collar  220  includes internal splines or gear teeth  222  which are complementary to and drivingly engaged by the external splines or gear teeth  208  on the drive sleeve  202 . The end of the second clutch collar  220  more proximate the center differential assembly  140  includes a first set of external splines or gear teeth  224  which are in all respects complementary to and engageable with the internal splines or gear teeth  198  on the circular end plate  192 . 
     The second clutch collar  220  includes a medially disposed, radially and circumferentially extending flange  226 . The second clutch collar  220  also includes a second set of external splines or gear teeth  228  disposed at the end of the second clutch collar  220  opposite the first set of external splines or gear teeth  224 . 
     As illustrated in FIG. 2, a second shift fork assembly  230  is also disposed upon the shift rail  112 . A second pair of pins or cam followers  232  engage spaced apart helical cams  234  on opposite ends of a second shift fork body  236 . The helical cams  234  may include axially extending discontinuities or flats  238 . The second shift fork body  236  also includes a through passageway  240  adapted to receive the shift rail  112 . The second shift fork assembly  230  also includes a yoke  242  defining a semi circular groove  244  which receives a portion of the flange  226  of the second clutch collar  220 . 
     Returning on FIGS. 2 and 3, a smaller chain drive sprocket  250  is freely rotatably disposed upon the intermediate shaft  60 . The smaller chain drive sprocket  250  is located and axially restrained between a shoulder  252  formed on the intermediate shaft  60  and a snap ring  254  which is seated within a complementarily configured circumferential groove  256 . The smaller chain drive sprocket  250  include internal splines or gear teeth  258  which are complementary to and may be selectively engaged by the second plurality of external splines or gear teeth  228  on the second shift collar  220 . The smaller chain drive sprocket  250  also includes a plurality of chain drive teeth  262  which drivingly engage a drive chain  264 . The drive chain  264 , in turn, drivingly engages chain teeth  266  formed about the periphery of a larger driven chain sprocket  268 . The larger driven chain sprocket  268  preferably includes internal splines or gear teeth  272  which are complementary to and engage external splines or gear teeth  274  on a secondary output shaft  276 . The secondary output shaft  276  is preferably supported upon a pair of spaced apart anti-friction bearings such as the ball bearing assemblies  282 . The secondary output shaft  276  includes a flange  278  or other structure compatible with the related, driven components of the secondary drive line  30 . An oil seal  284  provides a suitable fluid tight seal between the secondary output shaft  276  and the housing  50  of the transfer case assembly  16 . 
     The operation of the transfer case assembly  16  and specifically the center differential assembly  140  in the two-wheel drive, high gear (direct drive) mode and the four-wheel drive, reduced speed (underdrive) mode according to the present invention will now be described. In FIG. 3, the second clutch collar  220  is illustrated in the two-wheel, direct drive operating mode. In this operating mode, the second clutch collar  220  is in its right-most position with the external splines or gear teeth  224  of the second clutch collar  220  engaged with the internal splines or gear teeth  198  on the circular end plate  192 . Since the second clutch collar  220 , through the drive sleeve  202  is rotationally connected to the sun gear teeth  206  and thence to the pinions or planet gears  154 , the center differential assembly  140  is effectively locked and drive torque entering the center differential assembly  140  through the carrier  142  is provided to the primary output shaft  168  without differentiation, speed increase or speed reduction. 
     In the four-wheel, underdrive operating mode, the second shift collar  220  is translated by the shift control motor  110  to the left-most position illustrated in phantom lines in FIG.  3 . Drive torque is introduced into the center differential assembly  140  from the intermediate shaft  60 , through the carrier  142  and the associated pinion or planet gears  154 . The second shift collar  220  is engaged with and drives the smaller chain drive sprocket  250 . The center differential assembly  140  is thus unlocked and may achieve differentiation between the primary output shaft  168  and the smaller chain drive sprocket  250 . Preferably, the center differential assembly  140  is configured to provide a 65/35 torque split between the primary drive line  20  and the secondary drive line  30 . 
     In this operating mode, speed reduction is achieved through the center differential assembly  140  to the primary output shaft  168  relative to the direct drive operating mode discussed above. A corresponding speed reduction is achieved through the ratio of the gear teeth on the smaller chain drive sprocket  250  and the larger driven chain sprocket  268 . For example, the smaller chain drive sprocket  250  may include thirty-one teeth whereas the larger driven chain sprocket  268  may include forty-one teeth thereby achieving a speed reduction ratio of 1.1323 to 1 which, when combined with the ratio between the pinions or planet gears  154  and the sun gear teeth  206  on the drive sleeve  202  achieves the same nominal reduction as that through the pinions or planet gears  154 , the ring gear annulus  162  and the primary output shaft  168 . Reduction ratios between the smaller chain drive sprocket  250  and the larger driven chain sprocket  268  in the range of 1.15 to 1.50 to 1.00 are suitable. 
     While this ratio may be varied to accommodate various vehicle sizes, weights and horsepowers, the overall operating reduction ratio in the underdrive mode of 1.13 to 1 has been found satisfactory. Ratios in the range of about 1.05 to 1 to 1.25 to 1 are also considered suitable. This reduction (underdrive) ratio effectively increases the axle ratios of the front and rear differentials from, for example, 3.31 to 1 to 3.73 to 1, thereby improving the performance and towing capability of the vehicle. Of course, when the vehicle is shifted back to two-wheel, direct drive, it performs in accordance with its standard, for example, 3.31 to 1 axle ratio as will be readily appreciated. 
     As noted above, the center differential assembly  140  providing an underdrive speed range in high gear may be utilized with or without the two speed planetary gear assembly  70  and the high (direct) and low (reduced) speed ranges it provides. Furthermore, the center differential assembly  140  according to the present invention in conjunction with the primary differential  24  and the secondary differential  34  provide complete differentiation, i.e., permit independent variable speeds among the tire and wheel assemblies  28  and  38 . Accordingly, the antilock or antiskid brake control system of the associated motor vehicle may be utilized to control and limit torque distribution to the various tire and wheel assemblies  28  and  38  in order to reduce wheel spin and improve vehicle performance and handling, particularly in adverse weather and driving conditions. 
     The foregoing disclosure is the best mode devised by the inventor for practicing this invention. It is apparent, however, that apparatus incorporating modifications and variations will be obvious to one skilled in the art of transfer cases and motor vehicle drive lines. Inasmuch as the foregoing disclosure presents the best mode contemplated by the inventor 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.