Background of the Invention A motor vehicle is essentially a mobile platform capable of transporting people and commerce by virtue of its integrated powertrain, which generates and controls its attributes of torque and speed and delivers them via a mechanism to the drive axles, which in turn drive the wheels that propel the vehicle. At the interface between the output of the transmission and the drive axles is a mechanical assemblage referred to as the final drive, which includes a final drive gear ratio and a differential

The final drive, by way of the gear ratio inherent to it, performs a final manipulation of torque and speed. Another mechanism, known as a differential, performs a critical function that makes possible the drivability of every motor vehicle. Drivability refers to the efficient disposition of torque and speed to the drive wheels in a manner that ensures the safe dynamic and kinematic response of the vehicle while, at the same time, delivering a continuous smooth performance for any tracking or situational interruption that may occur when the two or four drive wheels rotate at different speeds. If not accounted for, the lack of ability internal to the power train to accommodate the need for the two or four drive wheels to negotiate various rotational speed differences, would develop an unstable condition of the vehicle. This lack of ability would render the vehicle unacceptable in terms of driver comfort and ease of operation. In the case of a rear axle drive, the functions performed in the final drive include: 1. A right angle gear set allows the torque to be diverted from the in-line drive shaft to become co-axial with the drive axle centerline. At the same time, the gear set ratio typically amplifies the torque and reduces the speed delivered to the drive wheels. In other cases, the final drive gear ratio acts as a speed increaser, as it decreases driveline torque as a function of ratio and vehicle requirements.

2. The output of the final drive gear set is directed into the differential, which, in turn, is the dispensing element to deliver the torque and speed to the drive wheels. At the same time, the differential must accommodate the varying speed and torque requirements between the drive wheels. In so doing, the differential provides smooth performance and enhances the drivability of the vehicle.

3. Among the many conditions that can force a rotational speed difference between the drive wheels are: a. Difference in the distances covered by the wheels while negotiating a curve. b. Road conditions that present non-uniform tractive effort to the drive wheels (i.e., when traction is different due to ice, snow and sand, etc.) c. Differences in the effective rolling radius between the drive wheels (i.e., inflation or pressure of the tires). d. Condition of the tires for traction coefficient. e. Condition of the roadway. A majority of the present technologies for differentials in motor vehicles involve bevel-geared designs and have been included in virtually every motor vehicle built. The kinematic requirements, outlined above, are not readily replicated in other mechanical devices, which accounts for the longevity of the bevel-gear based differential. Other differential mechanisms are noted in the art and provide similar function as described above. They typically include other features, such as traction aiding or road input damping, but do not address one core issue common to all differential designs, namely that of axial length. Due to the nature and typical configuration of a bevel gear differential, the axial length of the package is fixed by the diameter of the side gear. In turn, the diameter of the side gear is fixed by the torque requirements of the vehicle, typically set as a function of engine power and transmission multiplication.

All devices deemed to provide differential action require a speed ratio of -1 :1. The most common means to providing that ratio is to enmesh three gears in a right-angle gear train. The first and last gears in the train will have equal tooth numbers, but will turn in opposite directions due to the juxtaposition of two right-angle direction changes, thus providing the requisite -1 : 1 ratio for differential action. Other mechanisms can provide a -1 :1 speed ratio as well, but typically require a similar relationship between the input component and the output component.

The bevel gear configuration dictates package geometric size and volume to be roughly square. Thus, the axial length of the differential is dictated by the diameter of the side gear, which in turn is dictated by the torque provided to the differential from the final drive of the vehicle powertrain. This presents a problem in that, especially in front- wheel drive vehicles, axial length further dictates suspension geometry. Components that reduce the room available for the suspension components tend to overly constrain the effect of the suspension on vehicle ride and handling. The longer the transmission housing, particularly in a transaxle application, the less room a designer has for critical suspension components and parameters, such as half-shafts, tire vertical motion, turn radius, etc.

A device that can achieve the required differential ratio of -1 :1 and reduce its axial length, as a function of required diameter, would be a significant improvement to the state of the art in differential design and transmission packaging.

Brief Summary of the Invention

A differential capable of, but not limited to, automotive uses, or more generally, a multi-output speed converter made up of, but not limited to, a first rotary speed converter having at least three rotatable components being operably connected together; a first component of the at least three rotatable components of the first rotary speed converter being an input part to the first rotary speed converter, the input part being operably connectable to a driving member; a second and a third component of the at least three rotatable components of the first rotary speed converter being output parts of the first rotary speed converter, the output parts being operably connectable to drivable members; the first component of the at least three rotatable components of the first rotary speed converter being interposed within the second component and the third component to form a radially nested configuration of the at least three rotatable components of the first rotary speed converter; and the at least three rotatable components of the first rotary speed converter comprise a reaction carrier, an inner cam and an outer cam, the inner cam and the outer cam forming a conjugate pair, a second rotary speed converter having at least three rotatable components being operably connected together; a first component of the at least three rotatable components of the second rotary speed converter being an input part to the second rotary speed converter, the input part being operably connectable to a driving member; a second and a third component of the at least three rotatable components of the second rotary speed converter being output parts of the second rotary speed converter, the output parts being operably connected to drivable members; the first component of the at least three rotatable components of the second rotary speed converter being interposed within the second component and the third component to form a radially nested configuration of the at least three rotatable components of the second rotary speed converter; and the at least three rotatable components of the second rotary speed converter comprise a reaction carrier, an inner cam and an outer cam, the inner cam and the outer cam forming a conjugate pair, and the first and the second rotary speed converters being mounted coaxially to form the differential.

More specifically the above embodiment may have the reaction carriers of the first and second rotary speed converters rigidly connected to each other and to a driving member; the inner cams of the first and second rotary speed converters rigidly connected to each other and to a drivable member; the outer cams of the first and second rotary speed converters rigidly connected to each other and to a drivable member; the reaction carriers of the first and second rotary speed converters connected to each other with an angular offset; the inner cams of the first and second rotary speed converters connected to each other with an angular offset; the outer cams of the first and second rotary speed converters connected to each other with an angular offset.

Further, the above embodiments may include a first complement of rolling elements housed within slots in the reaction carrier of the first rotary speed converter, and a second complement of rolling elements housed within slots in the reaction carrier of the second rotary speed converter; wherein the number of slots in the reaction carrier of the first rotary speed converter is equal to the number of slots in the reaction carrier of the second rotary speed converter; the reaction carriers of the first and second rotary speed converters rigidly connected to each other and to a driving member, the inner cams of the first and second rotary speed converters rigidly connected to each other and to a first wheel axle, and the outer cams of the first and second rotary speed converters rigidly connected to each other and to a second wheel axle. In addition, the driving member may be a transmission.

For a better understanding of the present invention, together with other and further needs thereof, reference is made to the accompanying drawings and detailed description and its scope will be pointed out in the appended claims.

Brief Description of the Several View of the Drawings

Fig. 1 A is a cross sectional view of a prior art speed reducer;

Fig. IB is a cross sectional view of the prior art speed reducer of Fig. 1A taken along line IA;

Fig. 2 is an end view of a partial embodiment of the differential/speed

reducer/controller of the present invention;

Fig. 3 A is a schematic illustration of an embodiment of the inner cam utilized with the differential/speed reducer/controller of the present invention; Fig. 3B is a cross sectional view of the cam of Fig. 3A taken along line IIIA;

Fig. 4A is a schematic illustration of an embodiment of the outer cam utilized with the differential/speed reducer/controller of the present invention; Fig. 4B is a cross sectional view of the cams of Fig. 4A taken along line IV A;

Fig. 5 A is a schematic illustration of both the inner and outer cams of the differential/speed reducer/controller of the present invention with the rollers not shown; Fig, 5B is a schematic illustration of the inner cam design utilized with the differential/speed reducer/controller of the present invention with one-half of the rollers shown; Fig. 5C is a chart of the resulting torque output for one lobe cycle of rotation.

Fig. 6A is a schematic illustration of the differential/speed reducer/controller of the present invention; Fig, 6B is a cross sectional view of the differential/speed reducer/controller of the present invention taken along line VI A of Fig. 6A.

Fig. 7 is a prior art schematic representation of a bevel-gear differential. Fig. 8 is a cross section view of a schematic representation of a differential of the present invention.

Description of the Invention The present invention provides various embodiments of a differential or, more generally, but not limited to, a multi-output speed reducer or speed converter. Thus for the purposes of this invention, the term differential will be mainly utilized. The differential may be used for automotive uses, but is not limited thereto, and may find utility in any type of machinery in which a differential may be utilized such as tractors or the like. The differential utilizes a system including uniquely configured cams and rollers, and that is capable of providing the kinematic performance required of differentials in the driveline of a motor vehicle.

The differential of the present invention is capable of more efficiently performing the functions of conventional differentials in that:

1) the differential of this invention disseminates equal torque and speed to the drive wheels as the vehicle travels in a straight track on the roadway. In this condition, the two drive wheels are rotating at the same speed, which is also the same speed of their output gears in the differential. When this is the case, the differential will provide no relative internal rotation, and all internal components will rotate at the output speed of the final drive and so, in turn, will the two drive wheels. This will cause the combination of the final drive and differential mechanisms to continue to propel the vehicle in a straight-line direction along the highway.

2) the differential of the present invention accommodates those situations presented above where the drive wheels on a drive axle rotate at different speeds. This is demonstrated, in the extreme, by a situation when a drive wheel is lifted off the roadway and the other remains in contact with the roadway, and power is delivered to the differential. The grounded wheel remains stationary and the lifted wheel rotates at twice the speed of the input to the differential. This doubling of the input speed by the differential is an inherent property of a differential for this condition. In the case of straight tracking, the ratio of wheel speeds is 1 to 1, which together adds up to 2. In the lifted- wheel example above, the ratio of wheel speeds is 0 to 2, which again add up to 2. In general, any situation reflecting unequal wheel speeds must have a ratio of wheel speeds whose sum equals 2. By way of a further example, when a vehicle makes a 90° turn such that the turn radius equals the center distance between the drive wheels, the slower wheel will rotate at 0.67x and the faster gear will rotate 1.33x, where x equals the input speed of the final drive to the differential. The sum of the wheel speeds is once again equal to 2.00x.

It is possible to regard the deviations between the wheel speeds and the speed of the drive train as the actual differential action. In the case of straight-line motion, the differential action is zero, as there is no relative speed difference between the two wheels. In the extreme case of the lifted wheel, the differential action is maximized, and equals the speed of the drive train itself: one wheel speed is reduced to zero by the differential action, while the other is doubled relative to the drive train speed. These examples pertain to differences between the left and right wheel speeds in a two-wheel drive vehicle. In four-wheel drive vehicles, a further differential action is used to divide motion between the front and rear wheels. In order to better understand the present invention, reference is initially made to a prior art cam and roller speed converter 10 shown in Figs. 1 A and IB, which is presented in greater detail in the present assignee's U.S. Patent No. 5,988,145, incorporated herein in its entirety by reference, in order to describe the main components of speed converter 10. Speed converter 10 of Figs. 1A and IB is capable of producing 6 speed ratios depending on the function assigned to the three elements, namely inner cam 15, outer cam 20, and reaction carrier 25, the latter containing a complement of rollers 26 in their slots 27. Depending on which element is driven, which element is driving and which element is grounded, the three main elements can rotate at different speeds within each of the 6 possible speed ratios. A unique variation of speed converter 10 is incorporated within the present invention and is described below with respect to Fig.2. Essentially, the assembly 30 of the present embodiment is a configuration that includes, but is not limited to, six lobes 31 on the inner cam 35, six lobes 32 on the outer cam 40 and 12 slots 46 and rollers 47 in the reaction carrier 45.

Such a design of the present invention presents a unique configuration wherein, for example, but limited thereto, the same speed ratio exists between the input reaction carrier 45, with its 12 rollers 47, interacting with six lobes on the inner cam 35, as does between the input reaction carrier 45, with its 12 rollers 47, interacting with six lobes in the outer cam 40 and in which the cam parts 35 and 40 are conjugate. This equality of ratios enables the present invention to perform as a differential, with all of the special characteristic of the prior art bevel gear differential as shown in Fig.7, wherein the two- axle bevel gear drives 77 and 78 have identical speed ratios with the side gears 76, a fact that results in a -1 :1 speed ratio between the two bevel gears 77 and 78.

The torque developed by an unmodified assembly 30, shown in Fig. 2 would be of a pulsating nature, due to the equal number of lobes on cams 35 and 40. The disposition of rollers 47 relative to cams 35 and 40 at their extreme (that is, at minimum or maximum radial displacement) positions illustrates a condition of zero torque being transferred, as no reactive forces are possible between cam and roller. The snapshot of the assembly of Fig. 2 will produce zero torque throughput, and as input cam 35 rotates clockwise in the illustration, rollers 47 at positions B will start increasing their torque output towards a maximum, which occurs at position 15° at the midpoint of cam lobe rise. The torque output of those same rollers will then decrease to zero again, when they achieve position A. Meanwhile, the rollers 47 at position A generate zero torque while they proceed from position A to position B, and then proceed to generate torque as rollers in position B did initially. It can be seen that an interchange of torque production between position A and position B rollers will continuously occur as input cam 35 rotates, and a pulsating torque output will be produced. To eliminate this condition and thus achieve the desired substantially constant torque output of the differential of this invention, a modified inner cam assembly 37, as shown Figs. 3A, 3B, and modified output cam assembly 41, as shown in Figs. 4A, 4B are presented. Cam 35, Fig. 2, is replaced by modified inner cam assembly 37 as shown in Figs. 3A, 3B. Modified inner cam assembly 37 is made up of a pair of identical cams 37A and 37B, Figs. 3 A and 3B that are joined together side by side, by way of illustration but not limited thereto, by welding or the like, or by cutting or molding a single cam structure. The two cams 37A, 37 B are coaxial, but angularly offset with respect to one another by a phase angle Θ as dictated by the application and number of cam lobes, being equal to 15° in the particular case illustrated but not limited thereto. The geometry of each of the cam lobes 37A and 37B, Figs. 3A, 3B, is identical to the cam lobes of inner cam 35 shown in Fig. 2.

Modification of outer cam 40 shown in Fig. 2 is shown in Figs. 4A, 4B, as modified outer cam assembly 41, and includes two identical and coaxial cams 41 A and 41B that are joined side by side at the identical offset angle Θ shown in Fig. 4A in the manner as described above with respect to the inner cam assembly 37. Again it should be pointed out that the inner cams of the cam assembly 37 and the cams of the outer cam assembly 41 are conjugate cams with respect to each other, respectively. The geometry of each of the cam lobes 41 A and 41B, Figs. 4A, 4B is the same as cam lobes 40, Fig. 2.

In Fig. 5A a schematic overlay is shown of the modified inner cam assembly 37, Figs. 3 A, 3B with its offset configuration, and modified outer cam assembly 41, with its offset configuration, together with reaction carrier 50, without rollers 51 and 52 being shown for clarity. Fig. 5B is a simplified schematic illustration of Fig. 5A showing the modified inner cam assembly 37 and reaction carrier 50, with cam lobes or tracks 37A and 37B, and with rollers 51 and 52, at positions corresponding to A, B, C, D, J, K, and L. These new dual rollers 51 and 52 are axially separated and replace the single roller 47, Fig. 2, and are further illustrated in Fig. 6B in side view. It should be further noted for clarity only seven (7) roller sets are shown in Fig. 5B. In reality, in this embodiment, but not limited thereto, twelve (12) roller sets are utilized.

Utilizing two separate rollers 51 and 52 allows one set of rollers 51 to be in continuous contact with cam lobes 37A, while the other set of rollers 52 is in continuous contact with cam lobes 37B. As shown in Fig. 5B rollers 51 and 52 could be both active in the same slots or different slots (not shown) of the reaction carrier 50 as the original single roller 47, not shown for clarity, with rollers 51 interacting between inner cam 37A and outer cam 41 A shown in Fig. 5 A and rollers 52 interacting between inner cam 37B and outer cam 4 IB also shown in Fig. 5 A, not shown for clarity. Positions E, F, G, H, and I are not necessary to describe, as they will duplicate the reaction of the rollers that they are 180° apart from: position A is duplicated at position G, B is duplicated at H, C is duplicated at I, L is duplicated at F, and K is duplicated at E. For further clarification, a chart is shown in Fig. 5C that describes the interaction of the modified cam 37 and rollers as the reaction carrier 50, rotates one full lobe cycle starting at zero degree position A, and rotating to 60° position C. It can be seen from the chart in Fig. 5C that the differential of the present invention possesses a continuous and symmetric torque- transmission capability, independent of the location of the rollers on the various cams. As one set of rollers unloads, another set of rollers loads in unison, such that any potential pulsation in the transmission of torque is eliminated. This is described in greater detail below.

At the reference point of 0° at position A, Fig. 5B, cam lobes 37A are at maximum radius (or distance from the rotation axis), and the cam lobe slope is zero, whereas cam lobes 37B are halfway between maximum and minimum radius, and the cam lobe slope is negative. Positions C and K are equivalent to position A. Therefore, the interaction of cam lobes 37A and 37B with rollers 51 and 52, will lead the same torque throughput profile at positions A, C and K during the one full rotation, as described at each quarter cycle of rotation of the modified inner cam assembly 37 in Chart 5C. Likewise, rollers 51 and 52 are in equivalent positions with respect to cam lobes 37A and 37B at positions, B, D, J, and L, and their interaction with the cam lobes will produce identical torque throughput profiles, again as described in Chart 5C.

Accordingly, in the instantaneous position illustrated in Fig. 5B at 0°, rollers 52 at positions A, C, and K and their complementary rollers at E, G and I (not shown) are all at the point of highest loading and in the process of transferring all of the torque applied to the differential, while rollers 51 at the same position A are not under any load, and not transferring any torque. At the same time, rollers 51 and 52 at positions B, D, J, and L, and their complementary rollers F, and H, are not transferring any torque, as they are at either minimum radial displacement, or on the inactive half of the cam lobe, as described in Chart 5C.

As shown in Figs. 6A and 6B as inner cam assembly 37 rotates a quarter of a cycle with respect to the aforementioned reference position, the actions of the rollers 51 and 52 evolve in terms of their participation in transferring torque to the output cam assembly 41. It is shown that at 0° that rollers 52 on 37B at positions A, C, E, G, I, and K are transferring maximum torque. At the quarter-cycle position, it can be seen that rollers 51 on 37A at positions B, D, F, H, J, and L are now transferring maximum torque. These latter rollers, at the 0° reference position, are just coming into contact, but still in a state of zero load. Thus, during the quarter-cycle rotation of inner cam assembly 37 rollers 52 on 37B at positions A, C, E, G, I, and K, are reducing their contribution to the overall torque transfer from maximum to zero, while rollers 51 on 37A at positions B, D, F, H, J, and L are increasing their share of torque transfer from zero to maximum. As a result, the instantaneous value of torque transfer is essentially constant and receives maximum contribution from rollers 52 on 37B at positions A, C, E, G, I, and K, and rollers 51 on 37A at positions B, D, F, H, J, and L.

Accordingly, it can be seen that continuity of maximum torque transfer is maintained for the completion of one cycle: • From quarter to half of one full cycle of rotation of inner cam assembly 37, the torque contribution of rollers 51 on 37A at positions B, D, F, H, J, and L is decreasing to zero, while that of rollers 52 on 37B at positions B, D, F, H, J, and L is increasing to maximum.

• From half to three-quarters of one full cycle of rotation, the torque contribution of rollers 52 on 37B at positions B, D, F, H, J, and L is decreasing to zero, while that of rollers 51 on 37A at positions A, C, E, G, I, and is increasing to maximum.

• From three-quarters to one full cycle of rotation, the torque contribution of rollers 51 on 37 A at positions A, C, E, G, J, and K is decreasing to zero, while that of rollers 52 on 37B at positions A, C, E, G, I, and K is increasing to maximum.

With the complete cycle of rotation, it has been shown that the modified inner cams 37A and 37B, interacting with rollers 51 and 52 and modified outer cams 41A and 4 IB, Fig. 6B, can transfer continuous maximum torque.

Figs. 6A and 6B illustrate the final assembly of the differential 34 with modified inner cam assembly 37, modified outer cam assembly 41, and reaction carrier 50 with twelve slots 53, and two rollers 51 and 52 in each slot 53. In operation, power is applied to the differential assembly through reaction carrier 50 and rollers 51 and 52, and outputted through modified inner cam assembly 37, which transfers power to one axle 60, and modified outer cam assembly 41, which transfers power to the other axle 65 of the vehicle (not shown). The wheels of the vehicle would be connected to the axles 60 and 65, respectively, and power from the engine would be applied to the reaction carrier 50.

In the case of straight ahead tracking of a vehicle, the input speed and torque from the final drive must be equally distributed to the two axles such that equal torque and equal speed are realized by the two drive wheels on axles 60 and 65 of Fig. 8, which, in turn, are driven by modified cam assemblies 37 and 41. The input speed and torque to the differential from the final drive 80 and 81/50, Fig. 8, are reacted by the drive wheels in contact with the roadway. Assuming no slippage occurs, the wheels will rotate at the same speed. The two wheel speeds (and hence the rotational speeds of the axles and modified cam assemblies 37 and 41, Fig. 8 will equal the speed input into the reaction carrier 81/50, Fig.8 by the final drive in the vehicle transmission. This will give rise to zero differential speed, such that all elements of the differential rotate at the same speed as the output speed of the final drive. Also, because both drive wheels are reacting to the roadway, the input torque is equally distributed to the two drive wheels. It should be noted that, although in straight line motion there is no speed differential between the wheels, a -1 :1 speed ratio still exists between the two drive wheels relative to the reaction carrier.

In those situations where the drive wheels are experiencing relative rotation, such as in making turns, these relative speeds are reflected via the wheel axles 60 and 65 to their respective drive cams 37 and 41, Fig. 8, which, with their unique -1 :1 speed ratio relative to the reaction carrier, can adjust to these relative wheel speeds. With a vehicle traveling in a straight ahead tracking at a constant speed, the differential of the present invention rotates as a rigid unit and the rollers are not oscillating. When the vehicle negotiates a turn and a turn radius is established in accordance with the speed of the vehicle, a differential speed ratio is established by the reflected wheel speeds. In that case, the rollers in the differential of the present invention will oscillate in accordance with the differential speed ratio and the speed of the vehicle. The differential speed ratio between the two wheels, in turn, is established by the magnitude of the radius of the turn and the distance between the wheels. The following illustration is for the sake of example only and is not to be construed as limiting the scope of the invention. For a vehicle making a turn at 100-rpm speed input, with an inner turn radius of 10 feet and an outer turn radius of 16 feet and assuming true arcs of curvature:

Ratio of wheel speeds is 16/10 = 1.6:

At this ratio, the inner wheel will rotate at 0.769 times the reaction carrier speed and the outer wheel will rotate at 1.231 times the reaction carrier speed, where 1.231/0.769 = 1.6. Because the reaction carrier speed is 100 rpm, this will result in an inner wheel speed of 76.9 (rpm) and an outer wheel speed of 123.1 (rpm). In this particular case, the turn leads to a differential speed of 23.1 (rpm), which is

accommodated by the oscillations of the rollers 51 and 52. The vehicle velocity will be the same during the turn as it was before the turn. The wheel-to-reaction carrier speed ratios of 0.769: 1.0 and 1.231 : 1.0 reflect a sum of 2.000, just as with the 1 : 1 for the straight tracking.

For a further example, again not to be construed in a limiting sense, consider one wheel raised off the ground and the other in contact with the ground. In this case the wheel on the ground is effectively arrested, and the differential of the present invention performs as a normal speed reducer, as the following description makes clear (see Fig. 8): as shown in Fig. 8, the grounded wheel (not shown) restrains modified inner cam assembly 37 via the wheel axle 60, an input speed is provided to the reaction carrier 81/50, and the output is taken through the modified outer cam assembly 41 , which rotates at twice the speed as the reaction carrier. The doubled speed of the output cam 41 will rotate wheel axle 65, and so its wheel will rotate at twice the speed of the input into the reaction carrier 50. This situation normally does not happen, but illustrates how the -1:1 ratio between the modified cams interacting through rollers 51 and 52 accommodates different speeds of the two drive wheels, up to an including the case where one wheel is effectively stopped.

Fig. 8 with axles 60 and 65 illustrates the present invention differential as a substitute for the prior art bevel gear differential in Fig. 7. The new design has as one of numerous advantages over conventional differentials, a decided advantage of compact axial length when compared with such prior art differentials. This is because the differential of the present invention uses no right-angle gear meshes to achieve a -1 :1 ratio between the drive wheels. Therefore, the new design requires no side gears that add axial length to the differential and reduce the space available to other vehicle

components, such as suspension elements. Although not illustrated herein, the present invention is also capable of the differential requirements in a front wheel drive. In that case, the engine is usually mounted transversely such that its output is parallel to front wheel axle, and is coupled to an in-line final drive that directly drives the differential.

The present invention described herein and shown in Figs. 6A and 6B, with its unique -1 :1 speed ratio between the modified output cams 37 and 41 that drive the axles 60 and 65 of the two drive wheels, can provide continuous maximum torque throughput to the drive wheels in an axially compact package, and thus is capable of performance superior to that of automotive differentials of prior art.

Although the invention has been described with respect to various embodiments, it should be realized this invention is also capable of a wide variety of further and other embodiments within the spirit and scope of this invention and the scope of the various embodiments of the invention are set forth in the claims.

What is claimed is: