GB2160598A - Self regulating transmission - Google Patents

Abstract

A self-regulating transmission comprises an input and an output differential (11, 12) and a lower ratio gearing (13) and a higher ratio gearing (14) providing drive therebetween. The lower ratio gearing (13) drives bevel pinion (32) in the opposite rotational direction to that of the carrier (27) driven by the higher ratio gearing (14) which rotates in the same direction as that of the output shaft (34). When drive commences initially through the lower ratio gearing (13) the worm and wheel assembly (50, 51) in the higher ratio gearing (14) is adapted so as to resist feedback torque experienced in that gearing and to provide to the carrier (27) a driving force in the required rotation direction of the output shaft (34). This produces a driving force causing the output shaft (34) to rotate as well as a feedback torque in the lower ratio gearing which is sensed at the input differential (11) causing a corresponding adjustment of the rotational speeds of the lower ratio gearing (13) and the higher ratio gearing (14) driven thereby. In an alternative embodiment (Fig. 4), planetary gearing with internal ring gears replaces the bevel gear differentials. <IMAGE>

Description

SPECIFICATION Self regulating transmission This invention relates to a self regulating transmission for use as part of, or as a complete, drive transmission system.

The mechanism is suitable for use as a drive transmission in a motor vehicle, e.g. of the type known as a constantly variable transmission (CVT). it could also be used in various other applications such as in cranes, drilling rigs, lifts, trains, ship, aircraft engines, military vehicles and machinery in general.

Various types of constantly variable transmissions have been proposed in the past, such as belt drive systems utilizing variable diameter pulleys and other mechanisms, for example the Van Doorne belt system, the Perbury system and the Ferguson Formula system.

The Van Doorne Variomatic system is based upon a belt coupling between two sets of variable cones, movement of the cones axially making the belt, driven at its edge, couple the power at different diameters, thus changing the effective ratio. This system is reasonably efficient but can only be used for the lower powered car, probably up to 1.5 litre.

The Perbury system relies upon frictional coupling between hardened steel rollers coupling between two concave drive plates: a control system varies the position of the rollers to change the diameter of the drive to change the ratio. The system is quite good at transmitting speed but very poor at transmitting torque. The traction between rollers and surfaces is dependent upon the sheer strength of the oil film. It is unable to handle high shock loads as required in a car, e.g. when the vehicle touches a kerb.

There have been a number of electrical systems based upon the 75 year old Ward Lennard system, wherein a generator driven by the engine is used to supply current to a motor with control circuitry to vary speed and torque. These systems are very expensive, heavy and not really viable for a car.

The Ferguson Formula system uses a hydraulic pump to drive hydraulic motors on all four wheels with control systems to each motor, but is very costly and inefficient.

The object of the present invention is to provide a self regulating transmission which is load-sensitive in operation and is essentially provided by a system of gearing.

The invention provides a transmission mechanism comprising a first differential device having a drive input and a first and a second driven output, a second differential device having first and second drive inputs and a driven output, a first lower ratio gearing for transmitting drive between said first driven output of the first differential device and said first drive input of said second differential device, a second higher ratio gearing for transmitting drive between said second driven output of said first differential device and said second drive input of said second differential device, the arrangement being such that, in use, the rotational directions of the first and second drive inputs to the second differential device, are opposite to one another, and means provided in at least one of said gearings to inhibit reverse drive therethrough from said second differential device to said first differential device and to provide, in operation, a continuous driving force to the drive input of said second differential device with which that gearing is associated in the same rotational direction to that of said driven output thereof thereby to produce a feedback torque in the other gearing which is sensed by said first differential device causing a corresponding adjustment of the rotational speeds of said first and second driven outputs thereof.

The term "differential device" is used herein to include, in addition to a conventional differential mechanism, other gear mechanisms which can operate in a similar fashion, for example an epicyclic gear train.

Said means is preferably in the form of an appropriately constructed worm and wheel arrangement, and may be provided in one or both of said gearings. When provided in one only of said gearings, it is preferred that the worm is provided in the higher ratio gearing because a worm runs more efficiently at higher speeds. By selecting a suitable interface or lead angle of the contacting elements of the worm and worm-wheel, it is possible to produce a device in which resistive torque from a load applied to the driven output of the second differential device and "fed back" into the second gearing, is not only resisted by the worm/worm-wheel assembly but is overcome to produce a net torque in the forward direction through that gearing which acts to drive the aforesaid second drive input of said second differential device in the same direction as the aforesaid driven output thereof.This assembly therefore provides a mechanical advantage in the forward torque transmission direction to enable the feedback torque to be overcome.

A transmission mechanism in accordance with the invention acts to suitably apportion in a self-regulating fashion a power input thereto between a lower ratio gear path and a higher ratio gear path, both of which act continuously to drive a load, in response to variations in the applied and resistive loads.

An embodiment of the invention will now be described by way of example and with reference to the accompanying drawings, in which: Figure 1 is a vertical cross-section through a self-regulating transmission in accordance with the invention; Figure 2 is a diagrammatic representation of a worm/worm-wheel assembly of the transmission illustrating a drive condition from the worm to the worm-wheel; Figure 3 is a diagrammatic representation similar to that of Fig. 2 illustrating the condition of feed-back torque acting in the opposite direction; and Figure 4 is a vertical cross-section through a second embodiment of the invention.

Referring to Fig. 1, there is shown a selfregulating transmission embodying the invention for incorporation in the drive system of a motor vehicle. The gear assembly may be contained in a housing or gearbox of similar size to that of a conventional manual gearbox of a motor car. Generally a clutching device will be provided between the vehicle engine and the input to the transmission to allow engagement or disengagement thereof, or alternatively a fluid flywheel could be used. In order to provide a reverse gear, an additional manually operated gearbox may be provided in the variable transmission system simply allowing selection of a forward drive and a reverse drive and possibly also a neutral position.Alternatively a separate reverse drive transmission path may be selected which is provided in the vehicle transmission system parallel to and therefore bypassing the main self-regulating transmission mechanism.

The self-regulating transmission mechanism is contained in a gearbox housing 10. The transmission comprises two differential units 11 and 1 2 with a main drive input being applied to the differential unit 11, the drive output of the mechanism being taken from differential unit 1 2. There is also a low ratio gearing 13 and a high ratio gearing 14 providing separate drive paths between the two differential units 11 and 1 2.

The differential unit 11 comprises a carrier member 1 5 having an input shaft to the gearbox 1 6 integral therewith and a pair of stub shafts 1 7,1 8 supporting respectively a pair of bevelled pin ions 20 and 23. The differential comprises a further pair of bevelled pinions 21 and 22; the pinion 22 being supported on a stub shaft 24 rotatably mounted on a fixed part of the gearbox housing and the pinion 21 being formed with an integral sleeve 26 which is rotatably mounted on the input shaft 1 6. The carrier 1 5 is also provided with a cylindrical projection 25 which is received within a cylindrical bore within an end face of the stub shaft 24 which is free to rotate with respect to the projection 25.The input differential 11 therefore has a drive input through shaft 1 6 to the carrier 1 5 of the unit and the two driven outputs through the pinion 21 and its integral sleeve 26 and pinion 22 and the stub shaft 24 to which it is fixed.

The output differential 1 2 comprises a carrier 27 formed integrally with a drive input shaft 28 which is rotatably mounted in bearing supports provided within the gearbox housing 10. Carrier 27 has a pair of opposed stub axles 29 and 30 on which are rotatably mounted a pair of bevelled pinions 31 and 32 respectively. The differential comprises a further pair of bevelled pinions 32 and 33; the pinion 33 being fixedly mounted on one end of an output drive shaft 34 of the transmission and the pinion 32 being formed with an integral sleeve 34 which is rotatably mounted on the input drive shaft 28. The carrier 27 is formed with a cylindrical projection 35 which is received in a cylindrical bore in an end face of the output shaft 34 which is able to rotate with respect to the projection 35.The output differential 12, therefore, has a first drive input to the carrier 27 through input drive shaft 28 and a second drive input which can be applied to the bevelled pinion 32. It further has a driven output through shaft 34.

The low gear ratio drive path 1 3 between the input and output differentials comprises a spur gear 36 fixedly mounted on the sleeve 26 of the bevel gear 21 of the input differential unit 11. It further comprises a lay shaft 37 rotatably mounted in bearing supports provided in the gearbox housing 1 0. At one end of lay shaft 37 a spur gear 38 is fixed to mesh with spur gear 36. A bevelled pinion 39 is fixed at the other end of the lay shaft 37.

This gear connection further includes a transverse shaft 40 rotatably supported in mountings provided within the gearbox casing 10. A bevelled gear 41 is mounted on the shaft 40 to be driven by the pinion 39. A worm 42 is also mounted on the shaft 40 to drive a worm-wheel 43 fixedly mounted on the sleeve 34 of the differential bevel gear 32. It is important that the worm and wheel combinations in the respective gearings 1 3 and 14 are constructed so as to drive the respective input shafts 28 and 34 to the output differential 1 2, in opposite directions to one another, with the input shaft 28 rotating in the same rotational direction to that of the driven output shaft 34, in order to achieve the necessary reaction between the two torque inputs to the output differential so as to effect the required drive of the output drive shaft 34.

The high gear ratio transmission path 1 4 comprises a spur gear 44 fixed to stub shaft 24 to mesh with a spur gear 45 provided on a second lay shaft 46 rotatably supported in bearing mountings provided on the housing 10. A bevelled pinion 47 is mounted at one end of the lay shaft 46 to mesh with a bevelled gear 48 fixedly mounted on a second transverse shaft 49 rotatably supported in the housing 1 0. A worm 50 is also fixedly mounted on the transverse shaft 49 to drive a worm-wheel 51 which is fixedly mounted on the input drive shaft 28 to the second differential 12.

In the embodiment illustrated, the spur gears 36 and 38 have a 1:1 relationship whereas the spur gears 44 and 45 have a 5:2 relationship. In each of the gear trains 1 3 and 14, the bevelled pinions and meshing bevelled gears have a 1:1 relationship and the worm/worm-wheel drives have a 10:1 reduction. In other embodiments, these ratios may be altered as required to achieve the desired input/output variation.

The operation of the gearbox in a motor vehicle which is initially at rest will now be described. Drive from the engine is applied to input shaft 1 6 and therefore to the carrier 1 5 of the input differential 11. The input torque is therefore equally divided between the low ratio transmission path 1 3 and the high ratio path 14. With a high resistive load imposed on output shaft 34 by the inertia of the vehicle, the tendency is for the low ratio path 1 3 to drive the pinion 32 with the carrier 27 of the output differential being rotated in the opposite direction such that there is no net rotational movement of the output bevel gear 33.Therefore in order to achieve a driving output of the shaft 34 it is necessary to prevent rotation of the differential carrier 27 in the opposite direction which would thereby cause a reverse rotational drive through the high ratio gear path 1 4. This reverse rotational drive in path 1 4 is prevented by the construction of the worm and worm-wheel assembly 50,51 which will now be described with reference to Figs. 2 and 3 of the drawings.

In a worm/worm-wheel drive, the interface angle of the contacting elements of the worm and the worm-wheel if selected in a particular range, for example8'-1 1 with respect to the axis of the worm-wheel, can provide a driving assembly which will allow drive to be transmitted in one rotational direction only.By choosing a suitable interface angle, it is possible to provide a sufficient mechanical advantage in respect of drive from the worm 50 to the worm-wheel 51 over a reverse drive in the opposite rotational direction not only to withstand the reaction torque mentioned above which tends to rotate the differential carrier 27 in a direction opposite to the desired rotational direction of output shaft 34 but also to exceed this reaction torque sufficiently to cause the differential carrier 27 to rotate in the same rotational direction as that required of the output shaft 34. Fig. 2 considers drive being applied to the worm-wheel 50 in the direction as applied thereto from the output bevel 22 of the input differential 11.The interface angle of the contacting elements of the worm and worm-wheel is in this case given as 10 . The object is to transmit the drive through an angle of 90 but the direction the worm applies a force (F) to the interface between the engaging elements of the worm and worm-wheel, is deflected by 10 with respect to the desired direction for the application of such a force. The resultant force in the required direction is therefore reduced in accordance with this 10 deflection from the required direction of application of the force thereby correspondingly reducing the effectiveness of the transmission of the force from the worm to the worm-wheel.

Therefore, as the intention is to transmit the force through 90 , the tangential force is reduced in effect by the 10 load angle of the contacting teeth of the worm and wormwheel. In order to obtain a simple realisation as to the effective value, the intended force transmission at 90 is given a value of 100% efficiency. This is reduced by the 10 deflection of the applied from this 90 direction as follows: 90 Force Transmission e 100% 1" 1.111111% 10 11.1111111% Therefore, the ability, or effectiveness, of the worm is reduced, by a 10 angle of contact, by 11.111111 % to 88.888889%.

If this is then translated into something more representative, e.g. the torque loading as available to the worm, then this would mean that only 88.888889% of the worm's total capability is available to counter the feedback loading caused by the resistive load applied to shaft 34 and transmitted to the worm-wheel 51.

However, the interface angle of 10 , as used in this example, is still a mechanical advantage to the worm over the worm-wheel and while it is accepted as a "locking angle" preventing reverse drive in the gear path 14.

it is necessary to examine the extent of the mechanical disadvantage it bestows upon the worm-wheel.

Fig. 3 illustrates that the rotational direction of the worm-wheel 51 is applied to the interface angle of 10 and, as in the case of the worm, the direction of transmission of force between the worm and the worm-wheel is again deflected. However, the angle of the resultant tangent in this case is not 10 from the intended 90 drive angle, but is in fact 80 away from this intended direction.

This is, in terms of effective percentage transmission of the applied force between the worm-wheel and the worm, as follows: 90 Force Transmission ,"" 100% 1" e 1.1111111% 10 ""88.888888% 100%-88.888888% =11.111112% Therefore, unlike the worm to worm-wheel "advantage" situation, wherein the "loss" was only 11.111111%, in this worm-wheel to worm "disadvantage" situation the "loss" is 88.888888% giving the worm-wheel a capability for transmission to the worm of an applied loading, reduced to only 11.1111112%.

Considering now an input torque of 1T applied to the input drive shaft 16, the torque is equally divided by the input differential 1 5 into 1 /2T to each of the two drive trains 13; 14.

The two trains are of quite different ratio combinations and provide, at the output end, a situation which initially appears to be out of balance. However, if the various facets are examined closely, it will be seen that the feed back and through-drive loadings are compat ible.

If it is considered that bevelled gear 32 of the output differential 1 2 has a total rearward torque loading of 6.6025547:1T (where T is the Torque applied to the input shaft 16) as a result of the gear ratios, then the output differential carrier 27 will have a rearward loading of twice this amount as is normal in such standard differential devices. It will therefore have a rearward loading of 13.205109:1T in a rotational direction opposite to that of the input-shaft 1 6.

The 13.205109:1T is fed-back via the output carrier 27 and its drive shaft 28, to the worm-wheel 50. Worm-wheel 50 is, therefore, given a rearward loading of 13.205109:1T.

At the same time, the input differential 11 is supplying half of its drive capability, viz.

+T, to the low ratio gear path 1 3 train, i.e.

.the 13.205109:1T is derived from the effects of the low gear train upon one side of the output differential 12; and it is also supplying 1 /2T to the high ratio gear path 14. This is transmitted by gears 44,45 so as to provide a force, or loading, capability of 0.2115384T to the worm 50, the face of which is engaged directly with worm-wheel 51.

This state of affairs, at first, may give an impression of imbalance, however, it must be remembered that the ratio from worm to worm-wheel is 1 0:1 in favour of the worm.

Therefore, the ratio from worm-wheel to worm is 1:10, giving the aforesaid reverse or rearwardly loading in the gear path 14 of 13.205109:1T, a ''face'' value of only 1.3205109:1T at the worm 50.

At this point it would appear that the irreversible nature of the chosen interface angle, in this case approximately 10 , of the teeth of the worm and worm-wheel is all that prevents the worm-wheel from driving the worm backwards, as it would seem reasonable to assume that even the diminished 1.3205109:1T is more than a match for the 0.2115384T as manifested by the worm.

However, it is necessary now to take into account the mechanical advantage in force transmission from the worm to the wormwheel as a result of the interface angle described above with reference to Figs. 2 and 3.

Therefore the aforesaid reverse loading of 1.32051 09T is reduced in effectiveness by 88.888888% as a direct result of the chosen interface angle of, in this example, 10 . This means that only 11.11111 12% of the 1.3205109T of the reverse loading is effective across the worm-wheel and worm assembly, i.e. as follows:: 1.3205109T = 100% reverse load transmission efficiency 10% of 1.3205109T= 0.132051T reverse loading available for transmission to worm If therefore, the worm/worm-wheel interface values are examined in the light of this advantage/disadvantage situation, it will be seen that the worm is capable of driving forward with an effective torque value of 0.1880341T while the worm-wheel can only offer a total resistive feed-back loading of 0.1 3205T.

If we substract 0.1 3205T from 0.1880341T we find that the worm is capable of not only balancing the feed-back from the worm-wheel but is able to drive it forward by 0.0559841T which is, of course, then amplified by at a ratio of 10:1 as the worm/worm-wheel is a 10:1 coupling in favour of the worm. Therefore the output carrier differential is fed with 0.559841T over and above the feed-back balance requirement, and as the 0.559841T is divided equally across the output differential, it means that 0.2799205T is used as a "Brake" or "False" load upon the low gear train while compensating for this by contributing a similar amount (0.2799205T) directly upon the output bevelled gear.Therefore, the low gear train senses a greater load than that of the true output loading, while the high gear train offers a by-pass contribution of 0.2799205T towards the driving torque applied to output shaft 34.

This reasoning does not, of course, take into account the frictional content of the argument, which in fact also favours the worm and not the worm-wheel.

From Figs. 2 and 3, it will be seen that 904 angle of intention is not obtain in either case.

However, in Fig. 2 it is clear that the 104 interface angle only deflects the tangential force by 10 , losing only 11.1111111% of the mechanical ability of the worm. On the other hand, in the case of Fig. 3, it will be seen that only 10 of mechanical effectiveness is available from the worm-wheel when this is the "driving" element. This fact is emphasised by the almost linear matching of the rotational datum (R) to the resultant tangent (R2), both are very closely aligned with the rotational axis of the worm 'X'-'X'.

The available 10 interface angle equivalent to 11.1111 12% of the feed-back torque is therefore lost within the frictional margins, and while this will also effect the torque output from the worm it does so to a much lesser extent. This leaves the worm with a high ability factor over the worm-wheel's feedback loadings.

This indicates that, providing the initial reduction in feed-back torque is able to bring the remaining value down to within the mechanical conversion limits of the interface angle, a seemingly smaller torque force from the worm will always be able to drive, what appears to be, a greater opposing feed-back loading. In this example, the 1:10 wormwheel/worm engagement coupled with the hypothetical frictional effects, ensure that 0.2115384T can, and does, drive forward, a reverse or feed-back loading of 13.205109T, once it has been reduced by the intrinsic 1:10 ratio of the worm-wheel to worm combination.

Therefore, as a principle, it can be established that providing the feed-back path includes a ratio of torque reduction (or speed multiplication) which is capable of bringing the feedback loading down within acceptable limits, i.e. within the parameters of the mechanical advantage/disadvantage factors inherent in the particular interface angle, the worm will always be able to drive the worm-wheel forward, i.e. against the established direction of feed-back rotation.

This means that the interface angle can be chosen in order to be compatible with the feed-back demands and, if the feed-back loading is light, then the angle can be greater, e.g. 10"-20". If, however, there is a considerable amount of feed-back torque (force) to be handled, then the angle will be smaller, e.g. 10"-5". The exact angle is dictated by the loading requirements.

As mentioned above, when a load is applied to the input shaft 1 6 with an inertia resistive loading applied to the output shaft 34 of the transmission, it is necessary in order to produce drive at the output differential 1 2 for the differential carrier 27 to be held against rotation in the direction opposite to that intended for the output drive shaft 34. As described above in relation to Figs. 2 and 3, the reverse or rearward "feed-back" loading through the high gear path 14 is not only balanced by the forward loading in that gear path as a result of the geometry of the worm/worm-wheel combination 50,51 but is indeed exceeded.

The fulcrum provided by the differential carrier 27 is not only maintained in a fixed position to achieve drive in the output shaft 34 but is itself subjected to a torque in the forward direction which is generated in the worm/worm-wheel assembly as described above. This means that the carrier 27 of the output differential which is driven by the high ratio gear train 14, is continuously driven in forward rotational direction thereby producing a torque which is divided equally between the output bevel gear 33 and the input bevel gear 32. In this way it contributes to the output torque produced in the output shaft 34 as well as providing a "feed-back" torque in the low ratio gear train 1 3 which acts to diminish the torque in this train in the forward torque transmitting direction from the input differential to the output differential.This feed-back torque is sensed at the input differential 11 and acts to reduce the speed of the bevel gear 21 producing a corresponding increase in the speed of the bevel gear 22 and thereby increases the rotational speed throughout the higher ratio gear train 1 4. Therefore with this arrangement there is achieved a feed-back torque generated by the rotational movement of the output differential carrrier 27 in the same direction as the output drive shaft 34, which acts against the driving input torque applied to the low ratio gear train 1 3.This feed-back torque is sensed by the input differential and results in a corresponding adjustment of the relative speeds of the bevel gears 36 and 22 of the input differential 11 thereby controlling the rotational speeds of the two gear paths 1 3 and 14 in dependence upon the loading applied to the input shaft 1 6 and the output shaft 34 of the transmission providing a self-regulating transmission which is load sensitive.

Initially when the vehicle is stationary and a driving load is applied to the input shaft 16, a driving torque is applied through each of the gear paths 1 3 and 1 4 to the output differential 1 2. As described above, the output differential carrier 27 has a force applied to it urging it to rotate in the same direction as that required of the output drive shaft.However, the resistive loading on the output drive shaft 34 derived from the static inertia of the vehicle and the unfavourable gear ratio between the carrier and the output bevel gear 33 is such that the carrier 27 and hence bevel gear 22 will not move so that the initial drive will be through the low gear train 1 3 with its associated bevel gear 32 in the output differential 1 2. However, as soon as the vehicle starts to move and the output drive shaft 34 rotates, then the output differential carrier 27 turns in the same rotational direction of the output shaft 34 but at a slower rotational speed.The aforesaid self-regulating load sensitive action with respect to apportionment of the drive between the high gear path 1 3 and the lower gear path 14 thereby takes place whereby the output torque and rotational speed imposed on output drive shaft 34 is derived from the drive transmission through both the gear paths 1 3 and 14, the rotational speed and torque transmitted by each such path being controlled in dependence on the loading associated with input drive shaft 1 6 and output drive shaft 34.

In certain arrangements and under conditions of relatively high vehicle speed, the feedback torque transmitted through the low gear path 1 3 may be sufficiently high to stop completely the rotation of this gearing.

The ability to quantify the value of the worm-wheel's disadvantage and also to establish the value of the worm's mechanical advantage for a given interface angle, will allow any set of torque loadings, both drive and feed-back, to be matched precisely with its respective interface angle. This means that the working cycle of a particular design can be predicted and the "overlap" created across the worm/worm-wheel coupling used to a lesser, or greater, effect, i.e. if the feed-back is light and the angle of a low order, say 8"-10", then the ability of the worm over the worm-wheel would be considerable, and the output derived, over and above that required to nullify the feed-back loading, would also be considerable (within the terms of reference).

As it has been established that the torque drive feed from the output differential carrier is of a "driving" (to the output bevelled gear) and "braking" (in its effect upon the low gear train) nature, then this higher residue throughdrive available to the output differential carrier will have a more direct and profound effect upon the rate of change within the automatic cycle as it is this "residue" through-torque from the high gear train that provides the basic control element.

Therefore it is clear to see that the fulcrum.

as represented by the output differential carrier, centre-line lay-shaft and worm-wheel assembly is not a "locked" fulcrum device but a driven unit and one which contributes at all times to the overall output capability of the transmission thereby constituting a true and complete Automatic Power Management System.

In constructing practical versions of a gearbox according to the invention, it is desirable to design the high and low gearing assemblies such that they have similar mechanical efficiencies. By balancing the inertial and frictional characteristics of the two gearing assemblies, there is less likelihood of a reduction in the required torque transmitting difference between the two for effective operation of the transmission. Moreover, in order to improve the mechanical efficiency, it is preferable to use a worm drive of the type known as a "Squirm Drive", as described for example in U.S. Patent No. 3,820,413, which utilizes a worm wheel having a multiplicity of roller members to provide a rolling engagement with the worm screw threads.

In other embodiments of the invention a worm may be provided in only one of the aforesaid gearings, preferably the higher-ratio gearing. Moreover an epicyclic gear train may be used in place of one or both conventional differentials used in the above described embodiment. Fig. 4 shows an embodiment in which a single worm and wheel assembly 60 is used in the higher ratio gearing and two differential devices in the form of epicyclic gear trains 61,62 (each having a reduction ratio of 2:1) are used in place of the differentials 11,1 2 of the first embodiment. The lower ratio gearing is constituted by an input drive shaft 63 to the sun gear 64 of the output epicyclic gear train 62, the drive shaft 63 being driven by the ring gear 65 of the input epicyclic gear train 61.The higher ratio gearing, which includes the worm and wheel assembly 60, is driven by the planet carrier 66 of the input gear train 61 and drives the planet carrier 67 of the output gear train 62.

The input shaft 68 drives the sun gear 69 of the input gear train 61 and the output shaft 70 is driven by the ring gear 71 of the output gear train 62.

As stated above, it is necessary to have opposite directions of rotation of the two input members of the output differential in order to produce the required drive to the output shaft thereof. When conventional differential units are used as in the first embodiment, the two driven output members (36,44) of the input differential rotate in the same direction so that the low ratio gearing 1 3 must be adapted to provide a reversal of the rotational direction of the drive applied thereto so that the input member 43 of the output differential 14 is rotated in the opposite sense to that of the input member 28 associated with the higher ratio gearing 14, which -itself rotates in the same rotation sense as that of the output drive shaft 34.However when epicyclic gear trains are used, as in the Fig. 4 embodiment, no such reversal of the rotational direction applied to the lower gear ratio gearing is required, because the driven output members 65,66 of the first differential device 61 are rotated in the required opposite directions.

It will be appreciated that many modifications of the above described embodiments are possible within the scope of the invention as defined in the accompanying claims, for example the axes of rotation of the components of the output differential may be at right angles to those of the input differential and the low and high intermediate gearing being adapted accordingly to provide the required drive therebetween. Such an arrangement would provide an arrangement suitable for a motor vehicle with an output shaft having a gear located to provide an input to a main output differential of the vehicle which drives the half shafts thereof.

Claims (9)

1. A transmission mechanism comprising a first differential device having a drive input and a first and a second driven output, a second differential device having first and second drive inputs and a driven output, a first lower ratio gearing for transmitting drive between said first driven output of the first differential device and said first drive input of said second differential device, a second higher ratio gearing for transmitting drive between said second driven output of said first differential device and said second drive input of said second differential device, the arrangement being such that, in use, the rotational directions of the first and second drive inputs to the second differential device, are opposite to one another, and means provided in at least one of said gearings to inhibit reverse drive therethrough from said second differential device to said first differential device and to provide, in operation, a continuous driving force to the drive input of said second differential device with which that gearing is associated, in the same rotational direction to that of said driven output thereof thereby to produce a feedback torque in the other gearing which is sensed by said first differential device causing a corresponding adjustment of the rotational speeds of said first and second driven outputs thereof.

2. A transmission mechanism according to Claim 1, wherein the aforesaid means is constituted by an appropriately formed worm and wheel combination.

3. A transmission mechanism according to Claim 2, wherein said worm and wheel combination is provided in at least the higher ratio gearing.

4. A transmission mechanism according to any preceding Claim, wherein one or each, differential device comprises a rotary carrier with a pair of idler bevel gears mounted on coaxial stub shafts on the carrier and a pair of further bevel gears meshing with said idler gears.

5. A transmission mechanism according to Claim 4, wherein the drive input of the first differential device is the carrier thereof; said first gearing is driven by one of said further bevel gears of the first differential device and drives one of said further bevel gears of the second differential device; said second gearing is driven by the other of said further bevel gears of said first differential device and drives the carrier of the second differential device; and said driven output of the second differential device is the other of said further bevel gears thereof.

6. A transmission mechanism according to any of Claims 1 to 3, wherein one, or each, differential device comprises an epicyclic gear train.

7. A transmission mechanism according to Claim 6, wherein said drive input of the first differential device is associated with the sun gear thereof; said first gearing is driven by the ring gear of the first differential device and drives the sun gear of the second differential device; said second gearing is driven by the planet carrier of the first differential device and drives the planet carrier of the second differential device; and said driven output of the second differential device is associated with the ring gear thereof.

8. A transmission mechanism substantially as hereinbefore described with reference to, and as illustrated in, Figs. 1 to 3 or Fig. 4 of the accompanying drawings.

9. A motor vehicle having a transmission mechanism as claimed in any preceding Claim.