The invention relates to a differential mechanism particularly but not exclusively for use in motor vehicles.

Differential mechanisms commonly used on vehicles are of the sun and planet gear type and have a well known disadvantage that when one wheel is on a slippery surface such as mud or ice and the other wheel is on a firm surface capable of providing traction, the first wheel will simply spin as it receives all the available power transmitted to the differential.

Limited slip differential mechanisms have been proposed in an attempt to overcome this problem which restrict the extent to which one wheel can spin relative to the other but such differentials are generally more complex and, therefore, most costly to produce.

It is an object of the present invention to provide an improved form of limited slip differential which is less complex and therefore cheaper to produce and whose output torque ratio can be varied.

Thus according to the present invention there is provided a differential mechanism comprising two output cam members rotatable about an axis, each said member having a single frustoconical cam surface thereon of undulating form comprising pairs of mutually inclined surfaces, a plurality of cam followers having end surfaces engaging the cam surfaces of the output cam members, the arrangement being such that relative rotation of said output cam members causes the cam followers to slide axially, an input member which slidably supports the followers and moves the followers circumferentially relative to the output cam members, a first friction surface operatively associated with at least one output cam member and inclined at an acute angle to a plane normal to the axis of rotation of the output cam members, and a second similarly inclined friction surface operatively associated with the input member or the other output cam member, the first and second friction surfaces frictionaily engaging each other and generating a frictional torque which modifies the ratio of the torque output from the output cam members when said relative rotation occurs.

Each output cam member may be provided with an inclined first friction surface which engages a respective inclined second friction surface on the input member or the other output cam member.

The first and second friction surfaces may be detachably

secured to the associated output cam member and input member respectively to enable easy adjustment of the angle of inclination of the surfaces to vary the performance of the mechanism.

The first and second friction surfaces are preferably formed on co-operating friction rings which are secured to the associated output cam member and input member or other output cam member.

The first and/or second friction surfaces may be splined to the associated output cam member and/or input member or other output cam member.

The first and/or second friction surfaces may be pinned or pegged to the associated output cam member and/or input or other output cam member via integral or separate pins or pegs.

The invention also provides a method of providing a differential whose output torque ratio can be varied, said method comprising;-

providing a differential mechanism comprising two output cam members rotatable about an axis, each said member having a single frustoconical cam surface thereon of undulating form comprising pairs of mutually inclined surfaces, a plurality of cam

followers having end surfaces engaging the cam surfaces of the output cam members the arrangement being such that relative contra rotation of said output cam members causes the cam followers to slide axially, and an input member which slidably supports the followers and moves the followers circumferentially relative to the output cam members;

providing a first range of friction rings each having a respective first friction surface thereon which is inclined at a different acute angle to the axis of rotation of the output cam members and which is detachably securable to an associated output cam member;

providing a second range of friction rings each having a respective second friction surface thereon which is inclined at a different acute angle to said axis of rotation, and which is detachably securable to the input member and is designed to cooperate with a respective one of the first range of friction rings to generate a friction torque;

selecting the appropriate first and matching second friction rings which give the required friction torque to modify the ratio of the torque output from the output cam members when said relative contra rotation occurs, and

assembling the selected first and second friction rings into the differential mechanism between the output cam members and the input member to provide a differential mechanism with the required output torque ratio.

The present invention will now be described, by way of example only, with reference to the accompanying drawings in which:-

Figure 1 is a cross-section through a differential mechanism in accordance with the invention taken through output cam members;

Figure 2 is an end view of the differential of Figure 1 shown partly broken away;

Figures 3a - 3d are developments of symmetrical cam surfaces with cam followers shown in positions therebetween at different relative displacements along the cam surfaces;

Figure 4 is a diagrammatic end view of a follower;

Figure 5 is a section on line V-V of Figure l;

Figure 6 is an exploded perspective view of the differential of Figure 1;

Figure 7 shows the forces acting on friction rings located between the output cam members and an input member of the differential mechanism of Figure 1;

Figure 8 shows the forces applied to the output cam members by the cam followers during differential action which give rise to the inherent torque output bias of the differential mechanism;

Figure 9a to 9e show a range of different co-operating friction rings which generate different friction torques;

Figure 10 shows scrolling grooves on a friction ring to promote better lubrication;

Figures 11,12 & 13 show details of alternative friction ring constructions, and

Figure 14 shows an inter axle differential mechanism embodying the present invention.

In Figures 1 to 3 the differential 10 is mounted in a partially oil filled surrounding casing (not shown) by bearings (also not shown). The differential 10 comprises a housing 11 having a gear 12 on its external surface which receives drive from a pinion (not shown) in known manner.

The gear 12 is drivably connected to end walls 13, 14 which

may be formed integrally of the housing 11 or maybe formed separately and held in the housing 11 by any suitable means such as by bolts 11a.

Two output cam members 16, 17 have splines 15 at the centre thereof to drive output shafts (not shown) extending through bores 18 in the end walls 13, 14. The bores 18 may each have a helical oil feed groove (not shown) on the inner surface thereof in use to feed lubrication into or out of the differential.

The output cam members 16, 17 may be supported for rotation about an axis X within end walls 13 and 14 in bearings (not shown). Alternatively, as shown, no such bearings may be used since the design does not produce any significant radial misalignment forces on members 16 and 17. The output cam members 16, 17 each have a respective undulating cam surface 22, 23 thereon which comprises a frustoconical undulating surface. The cam surface 22 comprises an annular zigzag surface shown in detail in Figure 3 made up from a number of pairs of mutually inclined helical surfaces 24, 25. The cam surface 23 also comprises an annular zigzag surface which, as is apparent from Figure 3, has the same number of pairs of mutually inclined helical surfaces 26, 27 as used on surface 22.

As shown in Figure 1 the undulating cam surfaces 22 and 23 are inclined at an angle P to the axis x-x whereby each cam

surface converges radially inwardly towards the other.

Cam followers 28 are positioned between the cam surfaces 22, 23. Each cam follower is of strut-like elongate form and comprises two sets of mutually inclined end surfaces 29, 30, 32 and 33 which terminate at the side surfaces 34,35 (See Figure 3). The angle of inclination Q (see Figure 3C) between the end surfaces 29, 30 corresponds to the angle of inclination between the mutually inclined surfaces 24, 25.

The angle of inclination between the end surfaces 32, 33 corresponds to the angle of the inclination between the mutually inclined surfaces 26, 27 and is also equal to angle Q. The end surfaces 29,30,32 and 33 are also inclined at angle P as is apparent from Figure l. When viewed from the end, each cam follower is arcuate which enables the followers to be assembled together in an annular array as viewed in Figure 2. Each cam follower has an arcuate embrace of substantially 360/nf degrees where nf is the number of cam followers. Preferably, the arcuate embrace may be less to leave clearance spaces 28' (see Figure 4) between the followers thus preventing drive by abutment between adjacent followers.

Each cam follower includes an elongate drive dog 36 having mutually inclined side surfaces 37, 38 (Figure 4). The drive dogs 36 locate with slight clearance 36a in complementary shaped grooves 39 formed in the inner

periphery of a cylindrical drive input element 40 formed on input housing 11. The clearance 36a is just sufficient to ensure that the arcuate outer periphery (indicated at 28a) of each follower 28 can abut the inner peripheral surface (40a) of the drive input element 40. The grooves 39 provide support for the followers 28 at least adjacent their axial ends and preferably, as shown, for substantially their entire length.

As apparent from Figures 2 and 4 the assembly of the cam followers 28 is preferably such as to place the side surfaces 34, 35 of adjacent followers so that they interengage or lie closely adjacent. In that way maximum use is made of the available circumferential space for the cam followers, the followers together forming a substantially continuous and compact annular array as viewed in Figure 2.

As is best seen in Figure 3, the cam surfaces 22, 23 are identical having symmetrical pairs of inclined surfaces 24, 25 and 26, 27 respectively which are both inclined relative to each other at angle Q. The drive surfaces 24, 26 are a circumferential length LI which is equal to the circumferential length L2 of the overrun surfaces 25, 27.

Thus output cam members 16 and 17 with their equally inclined surfaces 24, 25 and 26, 27 will give an equal bias or torque ratio for turning in either direction, in forward

or reverse drive directions.

In order that the cam followers 28 cannot shuffle through the gap between an opposed peak and an opposed trough on the undulating cam surfaces 22, 23 without providing drive, it is necessary for different cam followers to be provided.

The different types of cam followers are arranged such that the number of followers is a multiple of a single wave length. For example, in the construction shown in Figures 1 to 6 there are two cam followers per wave length (one wave length being the distance spanned by each pair of mutually inclined surfaces 24, 25 and 26, 27).

Preferably the cam followers 28 are provided in groups of four followers 28A, 28B, 28C, 28D. The followers 28 have peaks which are offset from a reference line, in this case the centre line of each follower.

The followers 28A and 28D are identical and the followers 28B and 28C are identical, the follower 28B being a mirror image (in plan view only) of the adjacent follower 28A and the follower 28C being a mirror image of the adjacent follower 28D. Followers 28A' and 28B' are in the next group. There are therefore two types of follower and to assist recognition of the different types, one type of follower has its drive dogs 36 grooved at 36a (See Figures 5 and 6) .

In Figure 3a the followers 28A and 28C provide drive, whilst the followers 28B and 28D take the load in reverse.

When drive input is applied through drive input housing 11, and assuming that a vehicle having the differential is being driven in a straight line, the cam followers apply a load to cam surfaces 22, 23 so as to rotate the output cam members 16, 17 at equal speeds. As is apparent from Figure 3, with a driving load applied in direction Y the cam follower 28A on the extreme left has its end surfaces 29, 32 in driving engagement with surfaces 24, 26 and alternate followers are similarly in driving engagement with the cam surfaces 22, 23. However, intermediate cam followers have their surfaces in non-driving engagement with the cam surfaces as discussed above.

The driving force applied by the followers 28 to the inclined surfaces 24, 26 produces a reaction force F as illustrated in Figure 4. The inclination of the end surfaces of the cam followers at angle P causes the application of forces, which are shown only for cam 22 having the angle P. The application of force F creates an outward force G thereby producing a resultant force R which passes radially outboard of edge E preferably approximately through or adjacent a corner Cl between the drive dog 36 and an adjacent outer peripheral part of the follower 28. In that way the loading on the cam follower tends to wedge it firmly against a corner C2 of the drive input element 40 in

such a way that tipping of the follower about its edge E is avoided.

The operation of the differential can be appreciated by reference to Figures 3a - 3d which illustrate progressive movement of the cam surface 23 in the direction Y relative to the cam surface 22.

The relative movement of the cam surfaces 22, 23 causes the cam followers 28 to move axially and it can be seen in 3c that the followers 28B and 28C are sitting on a cam peak on cam surfaces 23 and 22 respectively and provide no drive.

The follower 28A is providing drive whilst the follower 28D is available to take overrun or reverse loading.

Because of the off-set follower design when the peaks and troughs are opposite each other, as in Figure 3d, the followers 28 cannot shuttle through. The followers 28A, 28B in each set provide drive, whilst the followers 28C and 28D can take reverse loads. The diagrams 3a-3d show progressive relative movement over half a wave. The other half of the movement would be similar.

In all situations the cam followers take drive load, although the sum of the areas available for load bearing is not constant, with the minimum load bearing area available being dependant upon the offset of the cam follower tips

from the centre line.

The length (and hence area) of the follower drive faces 29 and 32 (which engage the inclined surfaces 24 and 26 respectively) are in a ratio a/b and c/d in relation to the length of the overrun follower faces 29 and 33, which engage the inclined cam surfaces 25 and 27 respectively. Typically the ratio of a/b is about 2:1 and the ratio c/d is about 1:2, where a=d and b=c.

There are provided four cam followers for two cam waves and for a balanced design from two upto ten cam waves.

Designs using either eight or twelve cam followers are preferred.

As there is a considerable amount of friction between the followers 28 and the cams; torque will be transmitted to one cam when the other is drivably connected to a wheel spinning on a slippery surface, which is highly advantageous over conventional differential systems. The moving of one wheel faster than the other will result in a reduction in net torque applied to that wheel through the associated cam due to the load applied by the axially moving cam followers to which input torque is applied. There will be, in that case, an increase in the net torque applied at the other cam and the ratio between the net torques will be dependent upon the value of the portion QF of angle Q (i.e. the torque

bias ratio when driving depends on the inclination of the driving surfaces 24,26 (angle QF) and the torque bias ratio when on the overrun depends on the inclination of the overrun surfaces 25,27 (angle Q-QF). The greater the angle QF, the greater will be the friction at the cam surfaces during driving due to axial loading applied thereto by the followers.

Axial thrust applied to the output cam members 16, 17 by the followers 28 is transmitted to the end walls 13 and 14 through first friction rings 16a and 17a which are detachably secured to cam members 16 and 17 by pins 16b and 17b respectively. Friction rings 16a and 17a each have first friction surfaces 16c and 17c respectively which are inclined to a plane normal to axis X - X at an acute angle W and which contact second friction surfaces 13c and 14c on second friction rings 13a and 14a which are splined to end walls 13 and 14 at 13b and 14b respectively. Axial float of the output cam members 16 and 17 within end walls 13 and 14 is controlled by a shim ring 60 and a belleville washer 61 located between friction ring 14a and end wall 14 and between friction ring 13a and end wall 13 respectively.

As can be seen from Figure 7, the contact between friction rings 13a and 16a (and also rings 14a and 17a) generates a friction force (perpendicular to the plane of Figure 7) this results in a friction torque FT which acts about the axis X-X of the differential which is equal to .N.R. [where LK

is the coefficient of friction between the friction rings, N is the normal reaction between the contacting surfaces, and R is the effective radius of the contact area between the contacting surfaces] i.e. FT = f*. H .R

The normal reaction force N is balanced by an axial thrust component FA and a radial thrust component FR (see Figure 7) applied at the contact area between rings 13a and 16a. The radial component FR of thrust is internally balanced within ring 13a leaving axial component FA to contribute to the friction torque generated between rings 13a and 16a.

Thus since FT = U.N.R.

FT = . FA/COSW.R

Thus the friction torque FT increases as the thrust angle W increases.

As indicated above, the basic geometry of the output cam members 16 and 17 and the cam followers 28 is such that if one output cam member rotates slower than the other a higher torque output is passed to the slower rotating output cam member. This can be seen from Figure 8 which shows a view generally in a plane perpendicular to Figure 1 indicating the forces applied to output cam member 17 which is rotating slower than output cam member 16 by a follower 28a driving

in direction Y as per Figure 3.

In figure 8, the normal reactions N16 and N17 are substantially equal due to the equal inclination of cam surfaces 24 and 26 and the frictional forces F16 and F17 applied to the output cam members via faces 24 and 26 are in opposite directions. Thus a higher resultant driving force D17 is applied to slower rotating output cam member 17 than the resultant driving force D16 applied to faster rotating output cam member 16. Resultant driving forces D16 and D17 operate in a tangential direction about axis X - X to provide drive torque to the two output cam members.

These drive torques are modified by the effect of the friction torques FT developed by the contacting friction rings 13a, 16a and 17a associated with the respective output cam members 16 and 17. The friction torque FT 17 (see Figure 8) acting on the slower rotating output cam member 17 acts in the same sense as the drive force D17 thus increasing the net torque applied to the output cam member 17. Conversely the friction torque FT 16 acting on the faster rotating output cam member 16 acts in the opposite sense as the drive force D16 thus decreasing the net torque applied to the output cam member 16.

Thus the provision of the friction torque generated by friction rings 13a, 16a and 14a, 17a significantly increases the torque bias ratio of the differential. This means that

when differential action occurs a higher proportion of the available torque will be transmitted to the slower rotating output cam member.

So that by the provision of the appropriate friction torque, from contacting friction rings 13a, 16a and 14a, and 17a, the torque output bias ratio of the differential can be adjusted as required by a particular vehicle application or particular course on which the vehicle is to be run. For example, on particular tracks with more severe corners it may be advantageous to have a higher torque bias.

Figures 9a to 9e show a range of co-operating friction rings 13a, 16a with different angles of inclination W which generate different levels of friction torque FT. Rings 13a and 16a in Figure 9a generate the highest friction torque and the rings 13a and 16a in Figure 9e generate the lowest friction torque.

Figure 1 shows lubrication apertures 70 formed in end walls 13 and 14 through which oil contained within the differential casing is forced as the input member 11 and end walls 13, 14 rotate. This oil flows between the contacting friction ring surfaces 13c 16c and 14c, 17c via one or two way scrolling grooves formed in the friction surfaces (see grooves 71 of Figure 10). Additionally or alternatively the friction ring surfaces can be provided with grooves 72 through which the oil can flow to lubricate these surfaces.

Also as shown in Figure 11 one of the friction surfaces of each co-operating pair (e.g. 13c, 16c or 14c, 17c) can be slightly crowned (i.e convex) to ensure that contact between the surfaces takes place in the generally central regions of these contacting surfaces. The oil which enters via apertures 70 fills-up the radially outer portion 73 of the volume encased by input member 11 and end walls 13 and 14.

Figure 12 shows an alternative construction in which friction ring 16a is connected with output cam member 16 via cast in pegs 16b.

Figure 13 shows a construction in which separate pins 16b engage slots 16d in friction ring 16a rather than simple peg bores.

The concept of using friction rings 13a, 16a and 14a 17a to generate a friction torque to vary the torque output bias of the differential is also applicable to differentials in which the output cam members do not have equal numbers of inclined helical surfaces 24,25 and 26, 27. For example differential of the form described in the Applicants European Patent Application EP-A-0326289.

The invention is also applicable to interaxle differentials of the forum described in the Applicants European Patent Application EP-A-0619863, one example of which is shown in

Figure 14. with components of a similar function to components shown in Figure 1 multiplied by 10. Input to this differential is via an input hub 160 which is connected via splines 161 to an input shaft 162. Output cam member 170 is connected via splines 171 with output shaft 172 and the other output cam member is formed on end wall 130. Cam followers 280 act between cam surfaces 220 and 230 which face inwards towards the axis of rotation X-X. Followers 280 are connected with input hub 160 via drive dogs 360 which engage groove 390 in hub 160.

A pair of cooperating frictions rings 170a. 140a act between the outputs of the differential and again increase the torque bias ratio of the differential.

If a differential with identical output cam members is used in a circuit racing application, it may be that on a particular circuit, there are one or more very tight corners in one direction, while the corners in the opposite direction are mainly of a more sweeping nature. In such cases it may improve the handling of the vehicle, and hence the overall performance, if the torque bias ratio for cornering in one direction is set higher than that for the opposite direction.

When a vehicle is cornering near the limit of adhesion, the inner wheel tends to lose adhesion a speed up. This causes a transfer of torque to the outer wheel

at or near the limit. Using friction rings with a larger thrust angle on the ouput cam member driving the inner wheel will cause a larger change in torque bias (by subtracting a given amount of torque from the lower level) . Thus where it is desirous to create more torque transfer in one direction than the other, this can be achieved by a suitable choice of different friction rings on each side.

This effect may also be advantageous in centre differential applications (i.e inter axle differentials) in four wheel drive vehicles.

At low speeds the rear wheel of a vehicle rotate slower than the front wheels in a corner. This difference in speed will cause a transfer of torque towards the rear.

As vehicle cornering speeds increase, there will be a tendencyfor the rear to reach its adhesion limit first, causing speeding up of those wheels. This will cause a transfer of torque towards the front.

Depending on the weight distribution and suspension 'set-up' of the vehicle, it may be advantageous to provide a higher torque bias towards the front at the cornering limit, than was being transferred towards the rear at the start of cornering. This can be achieved by choosing different

friction ring angles for the two output cam members. (In the example given, a higher thrust angle on the friction ring attached to the output cam member driving the rear will provide a higher ratio to the front at the adhesion limit) .

In differential designs where the output cam members have different numbers of pairs of inclined surfaces 24, 25 and 26,27, these normally give a different torque bias ratio for left and right hand cornering. By choosing different angled friction rings at each end, this unequal torque bias effect can be reduced or eliminated.

In all cases where friction rings with different angles can be used, one option is to have a needle thrust bearing at one end and an angled friction ring at the other end. This arrangement can enable an unequal torque bias set-up to be equalised, as well as providing an extension to the options where an unequal performance is desired (lowering the bottom end of the range of possible torque bias ratios) .

It will also be appreciated that the separate friction rings 16a, 13a, 17a, 14a described could be formed integrally with components 16,13,17 and 14 respectively, if desired, so that friction surfaces 16c, 13c, 17c and 14c are formed on components 16,13,17 and 14 respectively. This gives a non-adjustable but possibly cheaper construction.