Abstract:
In a shifting mechanism housed in a case a first relatively rotating member rotates about an axis. A second relatively rotating member is selectively coupled and decoupled with the first member. The coupling has a first spline tooth with a first axial length, and a second spline tooth with a second axial length longer than the first spline tooth. The second spline tooth has an end having a frusto-conical shape. One of the first and second members has a plurality of third spline teeth for engagement with the spline teeth of the coupling. The third spline teeth have a complimentary frusto-conical shape. A selector is moveable for actuating the coupling to mutually connect and disconnect the members. A resilient connection is provided between the coupling and selector.

Description:
BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     This invention relates to the field of power transmission in a driveline for an automotive vehicle. More particularly, it pertains to a shifting mechanism to drivably connect relatively rotating shafts in the driveline of a motor vehicle. 
     2. Description of the Prior Art 
     To drivably connect relatively rotating shafts, a mechanical synchronizer is commonly provided to synchronize the rotational speed of the shafts, an example of which is provided in U.S. Pat. No. 4,375,172. The device of the &#39;172 patent is a relatively effective mechanism, but is produced at high cost and not able to engage over a wide speed variation. 
     The mechanism employed in the &#39;172 patent includes a blocked resilient axial interconnection between a jaw clutch assembly and main shaft  23  to improve the range of engagement. As described with reference to FIGS. 8 and 9 of the &#39;172 patent, a device according to the &#39;172 patent provides a large amount of backlash 208 to ensure adequate initial penetration, or engagement 212 of the teeth. This excessive backlash is not desirable as it provides additional slop in the system. 
     It would be desirable to provide a non-blocked engagement device for engaging relatively rotating shafts. 
     SUMMARY OF THE INVENTION 
     The avoid the difficulties and high cost associated with developing and manufacturing transmissions having a large number of forward speed ratios, and in order to improve the cost and performance of a device for shifting, an improved shift mechanism is provided. 
     In a shifting mechanism housed in a case a first relatively rotating member rotates about an axis. A second relatively rotating member is selectively coupled and decoupled with the first member. The coupling has a first spline tooth with a first axial length, and a second spline tooth with a second axial length longer than the first spline tooth. The second spline tooth has an end having a frusto-conical shape. One of the first and second members has a plurality of third spline teeth for engagement with the spline teeth of the coupling. The third spline teeth have a complimentary frusto-conical shape. A selector is moveable for actuating the coupling to mutually connect and disconnect the members. A resilient connection is provided between the coupling and selector. 
     Such a shift device allows for shifting on-the-go despite the input and output shafts lacking fully synchronized rotational speeds. Such a shift device is useful in many devices, including two-speed axles, subtransmissions (such as secondary transmissions or two-speed gearboxes), 4WD shift mechanisms and power take-off units. The shift mechanism may be coordinated with a computer to synchronize the input and output speeds to improve the shift “feel”. 
     Such a mechanism is further improved using an electronic controller to adjust the input and output rotational speeds closer to synchronous, utilizing engine, transmission and ABS control features in conjunction with adaptive shift motor controls. A shift device according to the present invention thereafter completes the shift at substantially synchronous speeds preferably using a “snap-action” shift device. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a schematic plan view of a powertrain for a motor vehicle that includes a multiple-speed rear axle assembly utilizing an illustrative shift device according to the present invention. 
     FIG. 2 is representation of a cross section taken at plane  2 — 2  of FIG.  1 . 
     FIG. 3 is an enlarged view of a portion of the mechanism shown in FIG.  2 . 
     FIG. 4 is a partial sectional view of the coupling shown in FIG.  2 . 
     FIG. 5 is an end view of the coupling shown in FIG.  4 . 
     FIG. 6 is a partial side view of the spline teeth of the coupling shown in FIG.  4 . 
     FIG. 7A illustrates a secondary transmission using a shift device according to the present invention in a first position rotatably locking the sun and carrier in a direct drive ratio. 
     FIG. 7B illustrates the secondary transmission of FIG. 7A in a second position to engage a gear reduction. 
     FIG. 7C illustrates a secondary transmission using an alternative shift device according to the present invention in a first position rotatably locking the sun and carrier in a direct drive ratio. 
     FIG. 7D illustrates the secondary transmission of FIG. 7C in a second position to engage a gear reduction. 
     FIG. 7E is a schematic illustration of a vehicle using a secondary transmission, for example according to FIGS. 7A-7D. 
     FIG. 8 illustrates a transfer case using a shift mechanism according to the present invention. 
     FIG. 9 illustrates a representative flow chart for synchronizing rotational shaft speeds while using a device according to the present invention. 
     FIGS. 10A and 10B illustrate a partial sectional side view and end view, respectively, of a secondary transmission using a further alternative shift device according to the present invention. 
     FIGS. 11A and 11B illustrate a partial sectional side view and end view, respectively, of a secondary transmission using a further alternative shift device according to the present invention. 
     FIGS. 12A and 12B illustrate a partial sectional side view and end view, respectively, of a secondary transmission using a further alternative shift device according to the present invention. 
     FIG. 12C illustrates an eccentric cam for the device illustrated in FIGS.  12 A-B. 
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     As described in my U.S. Pat. No. 5,888,165, (&#39;165 patent), and my copending application, Ser. No. 09/307,035, filed concurrently with this application having the same inventorship and title, which are incorporated herein by reference in their entirety, as shown in FIG. 1, the powertrain for a rear wheel drive motor vehicle includes an engine  10 ; transmission  12 ; rear drive shaft  14 ; rear axle differential  18 , left-hand and right-hand rear axle shafts  20 ,  22 ; and rear drive wheels  24 ,  26 . The right-hand and left-hand front drive wheels  28 ,  30  are not driven in a rear wheel drive application, as is known to one skilled in the art. The engine  10  is drivably connected to the multiple-speed transmission  12  which is drivably connected to the drive shaft  14 , which is connected to the input shaft of a multiple-speed axle drive mechanism  32  located within a case or housing  34 . 
     As described with reference to FIG. 2 in the &#39;156 patent, the drive shaft  14  is connected to a beveled input pinion  36  drivably connected to a ring gear  38  of a two-speed axle  32  located within housing  34 . The ring gear  38  is rotatably supported by the housing  34  at bearings  40 ,  42 . The ring gear  38  is in continual meshing engagement with a plurality of planetary pinion gears  44  supported for rotation by pinion carrier  46 . The carrier  46  is in continual driving engagement with an interwheel differential, an example of which is disclosed in U.S. Pat. No. 5,316,106. The differential  18  drives the rear drive wheels  24 ,  26  about an axis of rotation  48  via axle shafts  24 ,  26  in a manner known to one skilled in the art. 
     As best shown in FIGS. 2 and 3, a coupling  50  is provided in the axle  32  to mutually drivably connect and disconnect the pinions  44  and the carrier  46 . The coupling  50  comprises an annular sleeve member coaxial with the axis  48 . The coupling  50  carries a sun gear  52  in meshing engagement with pinion gears  44 . The coupling  50  also carries a second gear  54  axially displaced from the sun gear  52 . The coupling  50  is shown in a first position at the right hand side of axis  48 , wherein the coupling  50  provides an underdrive condition by locking pinion gears  44  against rotation with respect to housing  34  when coupling  50  engages housing  34  at the second gear  54 . Furthermore, although the present application is described above with reference to an underdrive ratio across the planetary gearset, in a preferred embodiment, the input gear ratio (for example the beveled pinion ratio) is adjusted so the so-called “underdrive ratio” comprises the equivalent of a direct drive ratio and the so-called “direct drive” ratio comprises an overdrive ratio. 
     In FIG. 2, the coupling  50  is shown at a second position at the bottom of the axis  48 . At this second position, the coupling  50  is axially moved to a second position wherein gear  54  is moved out of engagement with the housing  34 . In this second position, the sun gear  52  remains in meshing engagement with pinion gears  44  while sun gear  52  also engages the carrier  46  to mutually rotate the carrier  46  and pinions gears  44  to produce a direct drive ratio. The coupling  50  is disconnected from the housing  34  prior to the sun gear  52  being drivably connected to the carrier  46 , otherwise the entire planetary gearset would lock up against rotation. 
     As further shown in FIG. 2, sun gear  52  is carried by the coupling  50  and is drivably engaged with pinion gears  44  in the underdrive and direct drive positions. A motor  60  is supported by the housing  34 . The motor  60  moves a shift fork  62  axially to move coupling  50  to a desired position to obtain the proper axle ratio. A preferred embodiment of the motor  60  comprises a rotary electric motor, coaxially rotatably connected to a shift cam  63  through an approximately 58:1 reduction worm gear. Because of the large gearing reduction through the worm gear, only a small electric motor is required. The shift cam  63  includes a spiral groove  67  engaged with the shift fork  62 . Thus, as the motor  60  rotates the shift cam  63 , the spiral groove  67  urges the shift fork  62  axially. The shift fork  62  is supported on a rod  69  which is supported by the housing  34  for axial movement. Alternatively, one skilled in the art recognizes the motor  60  may comprise a linear electric motor or a vacuum motor or any equivalent motor for imparting such linear travel in the shift fork. Alternatively, a mechanical connection may impart the axial movement of the shift fork  62 , such as through a Bowden cable connection as is known to one skilled in the art. 
     The shift cam  63  preferably includes a detent (not shown), preferably comprising a detent position (not shown) in the spiral groove  67 . This detent is positioned to correspond with the sleeve  50  in a “synchronizing” position as described below. The spiral of the groove  67  extends helically around the cam  63 , so as the cam  63  is rotated by the motor  60 , the shift fork  62  is moved axially approximately 4.5 mm past a centered position, which corresponds to “neutral”. The centered “neutral” position is where the second gear  54  is not rotatably engaged with the housing  54  and the sun  52  is not engaged with the carrier  46 . Preferably as illustrated in FIG. 4, the second gear  54  is nearly immediately adjacent the housing  34  at the neutral position, while the sun  52  is approximately 2.0 mm from engagement with the carrier  46 . 
     The coupling  50  preferably moves axially 9 mm in either direction from the centered neutral position, but begins synchronizing with the housing  34  or carrier  46  when the shift fork  62  is moved approximately 4.5 mm axially on either side of centered neutral, the 4.5 mm position being the “synchronizing” position (alternatively called “neutral plus”). At this “synchronizing” position, within the groove  67 , the cam preferably has the detent, comprising a portion of the groove  67  extending circumferentially perpendicular to the axis of rotation of the cam  63  (versus helically), so the shift fork  62  is momentarily not urged further axially by the fork while the shift cam  63  continues to rotate. 
     As shown in FIG. 3, during synchronization, the ball lock mechanism  68  disengages the groove  70  of the sleeve, so the sleeve  50  does not move axially the entire 4.5 mm. While the shift fork  62  is within the detent, the spline teeth of the sleeve  50  are synchronized as described above while the axial spring provides an axial force on the sleeve  50  to urge the sleeve into engagement. As the second gear  54  synchronizes with the housing  34 , the spring  66  urges the sleeve rightwardly and the ball lock  68  will again engage the groove  70 . Once the cam  63  is rotated past the detent, the groove  67  extends further helically, so that the sleeve  50  is urged axially to fully engage the spline teeth as described above for another approximately 4.5 mm axially. Thus, in this preferred embodiment, the spline teeth are engaged approximately 7-9 mm; however one skilled in the art recognizes these distances are application specific and will vary based on the torque being transmitted, as well as the physical characteristics of the splines and gears. 
     One skilled in the art recognizes that the detent could alternatively comprise rotationally stopping the motor  60  at the point where the shift fork  62  is moved axially within groove  67  approximately 4.5 mm, so the synchronization can occur when synchronous speeds are obtained and initial engagement of the spline teeth occur as described above. Speed (RPM) sensors (not shown) preferably detect synchronization, i.e. when the spline teeth are initially engaged, and the motor  60  is started again to rotate until the spline teeth fully engaged. 
     A resilient connection  56 , described in further detail below, is provided between the shift fork  62  and the coupling  50  to ensure proper force is applied during engagement of the various members  52 ,  44 ,  46 ,  54 ,  34  to enable proper synchronization and smooth engagement thereof. This arrangement further provides a “snap-action” engagement of the teeth when the rotational speeds are synchronized. This device further provides shock absorption when the members engage. The resilient connection  56  enables the motor  60  to move the shift fork  62  to an absolute axial position, while the coupling  50  may not necessarily be fully engaged and therefore not properly axially aligned with the shift fork  62 . 
     One skilled in the art recognizes that an equivalent resilient connection  56  may be provided between the motor  60  and shift fork  62 , or any location between the input to move the shift collar and the shaft supporting the member to be engaged (i.e. the gear itself could be axially spring loaded). An example of another preferred resilient connection between the motor and shift fork is shown in U.S. Pat. No. 4,498,350 at  20 ,  20 ′, which is incorporated herein by reference for the relevant teachings provided therein. 
     As shown in FIG. 3, a preferred resilient connection  56  comprises a pair of pre-loaded axial compression springs  64 ,  66  provided between the shift fork  62  and the coupling  50 . The springs  64 ,  66  are axially opposed, each applying an axial force on the coupling  50  when the shift fork  62  is moved in the direction of the particular spring  64 ,  66 . Thus as shown in FIG. 3, the shift fork  62  is moved rightwardly and spring  66  is compressed, thereby imparting an additional axial force on coupling  50  through gear  54  until the gears are engaged and the shift fork  62  and coupling  50  are aligned. The springs  64 ,  66  are selected to provide a proper force on the coupling  50  to ensure proper synchronization and full engagement. The springs bias the coupling in the desired direction, and when synchronous speeds are realized due to the teeth, then the coupling engages rapidly with a “snap action engagement”, where the spring urges the coupling into the final position and the ball lock is reengaged. Further, the springs absorb energy during the initial engagement of the teeth—so as the longer teeth initially engage, the coupling will move axially against the spring force until rotational speeds are synchronous, allowing the coupling to move axially in the desired direction. The springs thus apply an axial force on the coupling  50 . Once the spline teeth described below are aligned on the various members  52 ,  44 ,  46 ,  54 ,  34  to be engaged, the spring force urges the coupling to snap into engagement with the member. Likewise, when the shift fork  62  is moved leftwardly, the second spring  64  imparts a leftward force upon coupling  50  through a stop  72  provided on the coupling  50  to provide proper synchronization and engagement force as described above. 
     Preferably, the resilient connection  56  further includes a ball lock mechanism  68  provided on the shift fork  62 . The ball lock mechanism  68  is radially displaceable from engagement in a groove  70  provided on coupling  50 . Thus, when the motor  60  rotates and moves the shift fork  62  axially, which then urges the coupling  50  rightwardly to engage the gear  54  with the housing  34 , if the spline teeth on gear  54  and housing  34  are not synchronized, the spline teeth axially oppose each other at the conical portion of the spline teeth described below with reference to FIG.  4 . Because the motor  60  forces the shift fork  62  rightwardly beyond the centered “neutral” position before the rotational speeds are synchronized, the unsynchronized opposing spline teeth resist axial movement of the coupling  50 . This resistance causes the ball lock  68  to come out of engagement from the groove  70 , but the axial spring  66  continues to impart an axial force upon the coupling  50  to engage the second gear  54  with housing  34 . Once the rotational speeds are synchronized, the spline teeth on the gear  54  engages the housing  34  and the axial spring  66  causes the coupling  50  to move rightwardly into engagement with the housing  34  and the ball lock mechanism  68  is aligned with the groove  70  and is engaged therein. Likewise, when the sun gear  52  engages the carrier  46 , the shift fork  62  is moved leftwardly. The ball lock mechanism  68  disengages the groove  70  leftwardly and the second spring  64  urges the coupling  50  leftwardly until the coupling  50  is synchronized with the carrier  46  and engaged therewith, allowing the coupling  50  to align the groove  70  with the ball lock mechanism  68  of the shift fork  62 . 
     As shown in FIG. 3, the coupling  50  is illustrated in a position where the motor  60  has moved the shift fork  62  rightwardly and disengaged the ball lock mechanism  68 . Because the second gear  54  is not synchronized with the housing  34 , the second gear  54  occupies the leftward position abutting the housing  34  as shown in FIG.  3 . As the second gear  54  synchronizes rotation with the housing  34 , the second gear  54  moves rightwardly as illustrated in phantom. During this rightward movement, the sun gear  52  also moves rightwardly, away from the carrier  46 . As shown the FIG. 3, during synchronization of the coupling  50  with the housing  34 , the sun gear  52  occupies the center position shown in the right hand portion of FIG.  3 . In this position, sun gear  52  is spaced axially approximately 2 mm from the carrier  46 , and is therefore not engaged with carrier  46  and the drive is in a “neutral” state. As the coupling  50  moves rightwardly into the underdrive position as described above, or leftwardly, into the direct drive position as described above, the sun occupies the respective position as shown in phantom. 
     The engagement of the members  52 ,  44 ,  46 ,  54 ,  34  is provided through a plurality of circumferentially spaced spline teeth. As shown in FIG. 4, the sun gear  52  is preferably formed integrally on the sleeve  50 . As shown in end view FIG. 5, the sun gear  52  comprises a plurality of circumferentially spaced spline teeth  51 ,  53 . The sun gear  52  teeth  51 ,  53  have flat contact surfaces for engagement with complimentary teeth provided on the planetary pinion gears  44  and the carrier  46 . As is known to one skilled in the art, the flat contact surfaces of the teeth  51 ,  53  may include small spiral shaped grooves (not shown) for carrying lubrication. 
     In a preferred embodiment, the teeth are synchronized mechanically. As shown in FIG. 6, every other tooth  51  is preferably recessed axially from adjacent teeth  53 , so lockup is more easily obtained at synchronizing speeds. If the rotational speeds are synchronized electronically as explained below, the recessed teeth are less necessary. As is shown in FIG. 4, the teeth of sun gear  52  include a cone angle  57  optimized for synchronization with a complimentary cone angle provided on the teeth of the carrier  46 . The teeth of the sun gear  52  preferably further include a tapered surface  59  at the leading edge of the teeth  51 ,  53  to facilitate engagement of the sun gear  52  and carrier  46 . The spline tooth spacing is optimized to minimize backlash. The second pair of teeth  54  on the coupling  50  are similarly formed to synchronize the rotational speed of the coupling  50  when engaging the teeth on the housing  34 . 
     In another preferred embodiment, the rotational speeds of the members  52 ,  44 ,  46 ,  54 ,  34  are synchronized electronically using the engine controller and/or the antilock braking system of the motor vehicle. As shown in FIG. 1, sensors  73 ,  74  are provided to measure the rotational speed of the input and output of the differential  18 . The input speed is preferably measured by obtaining the output speed of the transmission  12  using sensors  73 ,  74  as is known in the art. As shown in FIG. 2, based on the reduction of the input pinion  36 , the rotational speed of the ring gear  38  is known. The rotational speed of the planetary pinion gears  44 , sun gear  52 , and carrier  46  is calculated based on the position of the coupling which mutually connects and disconnects several of the members  52 ,  44 ,  46 ,  54 ,  34  as described above. 
     The output speed of the differential  18  is preferably inferred by measuring the rotational speed of the wheel  24  using an antilock braking system (ABS), which is known to one skilled in the art and not described here in detail. In a preferred embodiment, the ABS system includes an ABS sensor illustrated as sensor  74 , such an ABS sensor being known to one skilled in the art. The speed of the wheel  24  may be used to estimate the rotational speed of the carrier  46  when differential action is not occurring. Thus, to electronically control the synchronization of the members  52 ,  44 ,  46 ,  54 ,  34 , the input speed of the input gear  36  or output speed of the differential  18  may be controlled. As will be appreciated by the description provided herein, the sensors  73 ,  74  may be located in various positions to provide the signal indicating the input and output rotational speeds. 
     Preferably, the sensors  73 ,  74  send a signal to a computer  76 , such as an engine control unit (ECU). The computer  76  then determines whether it is proper to have the axle in an underdrive or direct drive position based on the rotational speeds of the driveline. Once this determination is made, the computer  76  provides a signal to control the rotational speeds of the input or output shaft to synchronize the rotation of the members  52 ,  44 ,  46 ,  54 ,  34  by controlling the engine speed, antilock brakes or transmission. The speeds are thus synchronized by using the ECU to increase or decrease the rotational speed of the engine  10  or transmission  12  in a manner known to one skilled in the art, or by decreasing the output rotational speed of the differential  16  by using the antilock brake system (ABS) to apply a brake at one or more of the rear wheels  48 ,  26  as is also known to one skilled in the art. As the rotational speeds are thus synchronized, the motor  60  is commanded by the computer  76  to move the shift fork  62  to the desired position to create the proper ratio. 
     In a preferred embodiment, a further sensor  75  is provided to sense the position of the shift fork  62  and to determine if the shift fork is in the proper position and preferably within the proper “synchronization timing window” to engage smoothly and to obtain the desired ratio. This “timing window” is provided in the period at which the rotational speeds are substantially synchronous. In FIG. 2, the sensor  75  is illustrated schematically as an encoder provided on the motor  60 , but could be incorporated in the case to sense the fork or coupling, or any other part of the mechanism. The rotational speed sensors  73 ,  74  then measure the rotational speeds and the computer  76  calculates whether the proper ratio is actually engaged. Such a sensor  75  may be of any known form, such as an encoder, a linear position sensor, a Hall Effect sensor, a limit switch, or any other known positional sensing devices. 
     The positional signal provided by the sensor  75  is preferably further used to enable the controller to adjust the axial shifting speed provided by the motor  60  and thereby position the mechanism in the proper axial position when the rotational speeds are synchronized—i.e. the shaft speeds are synchronized within a short “time window” through which the device preferably axially moves the shift fork to soften the shift harshness; the motor  60  is controlled to shift through this “time window” at which the rotational speeds are substantially synchronous. 
     Selection of the underdrive ratio may be performed automatically by the computer  76  commanding a shift when appropriate as described above. Otherwise, such a shift may be commanded manually by the operator moving a lever or a switch  78  to a desired position, such as commanding an underdrive position. Preferably the switch  78  includes a digital display to indicate the presently engaged ratio or mode (such as underdrive or performance). For example, a light may be illuminated when underdrive is engaged. Alternatively, an indicator may be provided on the instrument panel cluster to indicate the ratio. 
     An axle according to the present invention may thus be used to multiply the number of gear ratios in an existing transmission. In such an arrangement, a shift of the axle may be commanded simultaneously during a shift of a gear in the transmission to multiply the transmission ratio across the axle to obtain a wider range transmission. For example, third gear may be reduced using the axle to produce a final drive ratio between first and second gears in the transmission. In such an example, movement of a manual shift lever to what was previously second gear position would cause third gear to be engaged and the axle simultaneously shifted to underdrive. Upon movement of the shift lever to what was previously third gear, the second gear would be engaged and the axle simultaneously shifted to the direct drive position. 
     Although described here with reference to a differential on a rear wheel drive vehicle, the present concepts may readily be applied by one skilled in the art to another drive configuration. For example, the present invention may be added before or after the transmission in either a front wheel drive or rear wheel drive vehicle to provide additional gear reduction or increase the number of gear ratios provided thereby. An example of such an application in a front wheel drive application is described in U.S. Pat. No. 5,474,503, assigned to the assignee of the present invention, which is incorporated herein by reference. In such an instance, the input to the planetary gearset comprises a direct rotational input instead of a beveled pinion gear as illustrated in FIG.  1 . In this case, the secondary transmission (or two-speed gearbox) provides an additional reduction to increase the number of gear ratios available. A clutch according to the present invention may be provided in a device according to the &#39;503 patent to engage the ring with the one way clutch, or such a device may be used in place of the transfer clutch. As would be appreciated by one skilled in the art, the present invention is capable of doubling the number of gear ratios produced by such a transmission. For example, a four speed transmission may be used in an application to provide up to eight forward speed ratios using a secondary transmission or an axle according to the present invention. 
     A rear wheel drive secondary transmission (alternatively called a subtransmission or two-speed gearbox) is illustrated in FIGS. 7A-D. FIGS. 7A and 7B illustrate a first embodiment, while  7 C- 7 D illustrate a second embodiment. The reference numbers remain the same in each  7 A- 7 D except where the design differs. 
     In a secondary transmission according to the present invention, a shift mechanism  710  is provided to shift a secondary transmission  712  for a rear wheel drive vehicle. The secondary transmission  712  is located behind the primary transmission  12  illustrated in FIG.  1 . Preferably, the transmission  12  includes a flange at the rear end thereof and the secondary transmission  712  may be selectively mounted at  713  thereto on an optional basis to provide additional gear ranges, or an optional overdrive system. The shift device  710  is similar in many manners to the device previously described in FIGS. 2-6, but the shift fork of that device is replaced by a lever attached to a ball screw drive  716 . As motor  720  rotates, ball screw drive  716  is forced axially. This translates the end of lever  714  attached thereto. 
     The lever  714  rotates about a pivot  718  to translate the opposite end of the lever  714  a proportional distance (of course the lever  714  travels in an arc, the linear vector is presently of interest). The lever  714  includes a bifurcated end  722  (for the sake of clarity, one end is shown in phantom in this partial sectional view) which engages an annular groove  724  provided in a sleeve  726  engaged with a coupling  750 . Preferably the motor  720  includes a known encoder  721 , illustrated schematically, for determining the rotational position thereof. The controller preferably interprets a signal from the encoder  721 , and after interpreting the position of the motor  720 , the controller commands the motor  720  to shift the coupling within the “time window” during which the input and output speeds are substantially synchronous. 
     The coupling  750  has a splined connection  727  to the sun gear  752  and one skilled in the art appreciates this device operates in a manner similar to that described above with reference to the axle above and therefore the operation will not be described in great detail here. As shown in FIG. 7B, the coupling  750  is slid from the position shown in FIG. 7A where the sun  752  and carrier  746  were locked to a position where the coupling  750  is moved rightwardly as viewed in FIG. 7B to a position where the coupling  750  is drivably disengaged from the carrier  747 . Preferably this produces a reduction to develop an underdrive ratio across the planetary gearset. One skilled in the art could develop a variety of reductions and rotational reversals in a known manner and therefore these will not be discussed here in detail. 
     A compression spring  730  is provided between the sleeve  726  and coupling  750  and functions in a manner similar to the springs  64 ,  66  described above with reference to FIGS. 2 and 3, by providing a resilient connection at either end  732 ,  734  between the input force provided by the shift mechanism  710  and the coupling  750 . Further, a shift position detent, or ball lock mechanism  736 , is provided to retain the coupling  750  in a manner similar to that described above, thereby retaining the desired gear engagement. A screw  738  is provided to install the ball lock mechanism  736  on the coupling  750 , and in one embodiment is used to adjust the force of the ball lock mechanism. As shown in FIG. 7A, the ball lock mechanism engages one of a pair of grooves provided in the sleeve  750 , each groove corresponding to an “end detent position”, such that the ball lock mechanism  736  in this embodiment operates to engage a pair of terminal grooves, versus the central groove  70  shown in FIG.  2 . The planetary gear engagement, as illustrated in FIGS. 7A-B, includes a helical engagement between the sun gear  752  (part of the splined  727  sleeve  750 ) and planets  744 . As appreciated by one skilled in the art, this design provides axial thrust bearings adjacent the gears  752 ,  754  to accommodate the resultant thrust loads. 
     FIGS. 7C-7D illustrate a variation to the embodiment shown in FIGS. 7A-7B. In this embodiment, the coupling  750 ′ carries the sun gear  752 ′ and the splined connection  727  of FIGS. 7A and 7B is eliminated. A feature of this embodiment is that the gear engagement between the sun  752 ′ and planetary gear  744  comprises a simple spur gear profile, thus enabling translation of the coupling  750  directly and minimizing any axial loading. One skilled in the art appreciates the straight spur gear engagements, such as the sun  752  to planets  744  in FIG. 7C-7D, contrasted to the embodiment of FIGS. 7A-B, provide for minimal axial gear reactions. Further alternative shifting devices are provided in FIGS. 10-12. These embodiments are similar to the devices described above, in that they utilizes many of the same components but these embodiments have a generally more simple shift device. These devices are illustrated in use as a secondary transmission, but one skilled in the art appreciates the applicability to other devices as described above. In the embodiment of FIGS. 10A and 10B, an electric motor  720 ′; is connected through a shaft to a link  714 ′. The link  714 ′ is in the form of a shift fork and engages a slot in the coupling  750 ″ through a snap-action device  724 ′. The snap-action device  724 ′ provides a resilient connection between the link  714 ′ and coupling  750 ″ in a manner similar to the embodiments described above and is therefore not described in greater detail here. 
     In the embodiment of FIGS. 11A and 11B, an electric motor  720 ′; is connected to a reduction gearbox  716 ′, which is subsequently connected to a link  714 ′. The link  714 ′ is in the form of a shift fork and engages a slot in the coupling  750 ″ through a snap-action device  724 ′. An encoder illustrated schematically at  721 ′, senses the position of the gearbox  716 ′, or alternatively the motor  720 ′. The snap-action device  724 ′ provides a resilient connection between the link  714 ′ and coupling  750 ″ in a manner similar to the embodiments described above and is therefore not described in greater detail here. 
     In the embodiment of FIGS. 12A, B and C, an electric motor  720 ′; is connected to an eccentric pivot  723 , which is subsequently connected to a link  714 ′. The link  714 ′ is in the form of a shift fork and engages a slot in the coupling  750 ″ through a snap-action device  724 ′. The snap-action device  724 ′ provides a resilient connection between the link  714 ′ and coupling  750 ″ in a manner similar to the embodiments described above and is therefore not described in greater detail here. The link  714 ′ rotates about a pivot  725  to effect a translation of the coupling  750 ″. The eccentric device is illustrated in FIG. C from right to left in an end view of a mid position, then a side view of the same position. As the motor  720 ′ rotates, the eccentric pivot device  723  rotates in a bifurcated end  729  of the link  714 ′, thereby causing rotation of the link  714 ′ about the pivot  725 . 
     Preferably, at the time the shift fork is in its “detented end positions”, the eccentric cam effect of this embodiment generates the additional shift force required to overcome the ball lock mechanism  738  so the ball is forced out of the detent, thereby reducing the shift torque requirement for the electric motor  720 . Thus, a smaller motor  720  may be used and/or the gear reduction  716  (ref. FIG. 11B) may be reduced or eliminated. 
     Furthermore, the present invention may use an adapter to bolt onto an existing transmission case and thereby require no additional modifications to the transmission, particularly when this device is used on an optional basis in production. 
     As illustrated schematically in FIG. 7E in a preferred embodiment, the secondary transmission  712  of FIGS. 7 through 7D are utilized in combination with an automatic transmission  12 ′ attached to an engine  10 ′. In this arrangement, the electronic control logic of the transmission  12 ′ is preferably adapted to change the gear shift sequence and clutch slippage in a known mariner to further improve the synchronization of the input and output shaft speeds during a shift of the secondary transmission  712 , and thereby improve the shift smoothness of the secondary transmission  712 , bringing about the shifts in a coordinated manner. The transmission controls may be used in conjunction with the engine and antilock brake controls as described above. 
     As described above, the shifting of the device in FIGS. 7-7D are most smoothly accommodated by nearly synchronizing the rotational speeds of the input and output prior to engaging the shift mechanism  710 . This is best accomplished by monitoring the input/output speeds using sensors as is known to one skilled in the art, for example using a transmission sensor  761  and driveshaft sensor  762 . Examples of such sensors include ABS sensors, turbine speed sensors, or any other such known sensor used to measure the rotational speed of the vehicle driveline. A controller  763  receives signals from the sensors and adjusts the input/output speeds by controlling the rotational speed of the engine  10 ′ and/or the wheels  24 ′- 30 ′. Such a controller  763  comprises one or more known controllers, such as an engine controller, an antilock brake controller, a traction control controller (utilizing ABS and/or engine controls), and/or an automatic transmission controller, preferably while simultaneously adjusting the shift motor speed by monitoring the shift motor position sensor  721 ′ to allow adequate time for input/output shaft rotational speed changes in order to substantially synchronize the speeds thereof. Simultaneously, the controller adjusts the shift motor speed to allow adequate time for input/output shaft speed changes before the coupling is urged into position, thereby smoothing the engagement thereof. Of course the device  710  acts to provide the snap-action shift as described above, so the speeds need not be synchronized for engagement. However, by controlling the motor  720 , the engagement is timed to enable smooth shifting. 
     An exemplary logic is provided in the flow chart of FIG.  9 . As illustrated in FIG. 9, an input is provided at  771  to indicate a shift is desired, either manually or using a controller as described above. Preferably the shift motor position is known  772  and the motor is commanded to shift at  773  in the direction to actuate a shift. However, the motor is shifted to the “neutral plus” position described above and not completely to the shifted position. At this time, the controller compares the rotational speeds  774  and determines at  775  whether the shaft speeds are adequately synchronized. If so, the shift motor commands the shift to be completed  779 . If not, one or more controllers, such as the engine and/or transmission controller  776  or ABS and/or traction controller  777  control one or more rotational speeds  778  as described above to synchronize the rotational speeds, which is compared again at  774 - 775  and the shift is completed  779  once acceptable. 
     As illustrated in FIG. 8, a shift device  810  according to the present invention may be applied in an application including a four wheel drive transfer gearbox. The planetary gearset  812  would provide a gearing reduction in a transfer gearbox to provide a reduction from a four wheel high ratio to a four wheel low ratio in a manner known to one skilled in the art. Such a device is described in U.S. Pat. No. 4,718,303, which is incorporated herein by reference. However, the coupling mechanism, embodied as clutch plates in the  303  patent, are replaced by the shift mechanism  810  to replace the clutch plates as the coupling mechanism. 
     A device according to the present invention enables a shift to produce either a transfer to 4WD or a 4WD Low reduction to occur while the vehicle is moving, because the synchronization device and techniques taught herein provide for such reduction in a transfer gearbox in a smooth manner. The function of this device is similar to the other devices described above, and is therefore not described in great detail. An electric motor  820  acts through a reduction gearbox  821  having an internal sensor (not shown) to detect position to move a rotating cam device  816 , similar to that described above. The rotating cam device  816  includes a cam follower sleeve provided at the end of the shift fork  862  to actuate a shift fork  862  to translate a spring-loaded coupling  850  as described above. The coupling  850  is splined  827  to the output shaft for axial movement while remaining rotatably engaged thereto. 
     The coupling  850  engages the planetary carrier  846  for a reduction across the planetary gearset for 4WD Low range, or alternatively, the sun gear  852  for 4WD high or 2WD ranges (not shown in the alternate position). One skilled in the art appreciates that this device can be equally applied to a secondary transmission as described above for a gear reduction in 2WD mode, or for a 2WD system (versus the 4WD system illustrated in FIG.  8 ). A separate device  870  is provided in FIG. 8 to engage the 4WD feature. This device  870  could be a similar snap-lock device as described above or a conventional 4WD engagement as known to ins skilled in the art. 
     Although not illustrated, one skilled in the art also appreciates the present invention may be used in a layshaft transmission to engage a journalled gear with a relatively rotating shaft and thereby replace a blocked synchronizer as is typically used. 
     One skilled in the art will appreciate the disclosed mechanism is capable of reliably engaging the relatively rotating members at relatively high differential rotational speeds, but such engagement may be perceived by the driver or passengers of the motor vehicle as being too harsh. Therefore, a preferred embodiment further includes some synchronization of rotational speeds prior to engagement. These methods, as described above and appreciated by one skilled in the art, include the use of engine speed control through the powertrain control module, ABS systems or traction control systems. Using these techniques, one is readily able to improve the smoothness of engagement, and therefore improve the feel of the shift to the passengers of the vehicle. Preferably the shift smoothing capabilities of an automatic transmission controller and mechanisms are also used to synchronize a device and provide smooth engagement thereof. 
     The forms of the invention shown and described herein constitute the preferred embodiments of the invention; they are not intended to illustrate all possible forms thereof. The words used are words of description rather than of limitation, and various changes may be made from that which is described here without departing from the spirit and scope of the invention.