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Title:
A DIFFERENTIAL GEAR ASSEMBLY AND TRANSMISSION DEVICE USING THE DIFFERENTIAL GEAR ASSEMBLY
Document Type and Number:
WIPO Patent Application WO/1987/003061
Kind Code:
A1
Abstract:
A differential gear assembly comprises first and second shafts (11 and 12) carrying respective first and second driving bevel gears (18 and 19) for rotation therewith. A rotatable carrier (13) has mounted thereon two pairs of differently dimensional idler gears (14, 20 and 15, 21) rotatable with respective lay-shafts (16 and 17) about a fulcrum axis (F). The larger idler gears (14 and 15) engage the first bevel gear (18) and the smaller idler gears (20 and 21) engage the second bevel gear (19), so that the idler gears engage the two bevel gears (18 and 19) at different distances from the fulcrum axis, resulting in the application of different torques to the shafts (11 and 12) upon rotation of the carrier (13).

Inventors:
STIDWORTHY FREDERICK MICHAEL (GB)
Application Number:
PCT/GB1986/000691
Publication Date:
May 21, 1987
Filing Date:
November 07, 1986
Export Citation:
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Assignee:
STIDWORTHY FREDERICK M
International Classes:
F16H3/74; F16H48/08; F16H48/11; F16H48/40; (IPC1-7): F16H1/40; B60K17/34
Foreign References:
FR1434413A1966-04-08
DE3331535A11984-02-23
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Claims:
CLA1MS:
1. A differential gear assembly comprising first and second rotatable shafts carrying respective first and second driving gears for rotation therewith, a rotatable carrier, and idler gear means engaged with the driving gears and mounted on the carrier for rotation about a fulcrum axis such that the distances from the fulcrum axis to the points of engagement of the idler gear means with the first and second driving gears are unequal.
2. An assembly according to claim I , in which the idler gear means comprise first and second idler gears rotatable together about the fulcrum 10 axis.
3. An assembly according to claim 2, in which the first and second idler gears are bevel gears, the first idler gear being of greater diameter than the second idler gear.*& 15.
4. An assembly according to claim 3, in which the first and second driving gears are bevel gears, the first driving gear being of greater diameter than the second driving gear, and the first and second driving gears being engaged with the first and second idler gears respectively.*& 20.
5. An assembly according to claim 4, in which the carrier rotatably engages the second shaft.
6. An assembly according to claim 5, in which the second shaft passes " through the carrier.
7. An assembly according to any one of laims I to 4, in which the carrier rotatably engages the first shaft.
8. 30 8.
9. An assembly according to claim 7, in which the first shaft is a sleeve shaft and the carrier is rotatable with a carrier shaft passing through the first shaft.
10. A transmission device comprising an assembly according to claim 7 or*& 35.
11. 8 and means for applying torque to the first shaft and the carrier from an Input shaft, the second shaft of the differential gear assembly constituting the output shaft.
12. 10 A transmission device according to claim 9, in which the means for applying torque to the first shaft and the carrier comprise a second differential gear assembly driven by the input shaft and driving respective gear trains driving respective transmission gears attached to the first shaft and carrier shaft.
13. A differential gear assembly substantially as hereinbefore described with reference to the accompanying drawings.
14. A transmission device substantially as hereinbefore described with reference to the accompanying drawings.
15. Any novel feature or combination of features described herein.
Description:
Description of Invention

"A differential gear assembly and transmission device using the differential gear assembly"

THIS INVENTION relates to a differential gear assembly which finds particular application in the drive train of a motor vehicle but is also usable in any situation where it is desired to create a torque imbalance.

In order that the invention may be readily understood, an embodiment thereof will now be described, by way of example, with reference to the accompanying drawings, in which:

Figure I is a diagrammatic cross-sectional view of a conventional,

10 four bevel differential gear assembly of the kind commonly found in the rear axle of an automobile; '

Figure 2 is a diagrammatic, partly sectioned view of a simple >r differential gear assembly embodying the present invention;

Figure 3 is a diagrammatic, partly sectioned view of one load- sensitive transmission device embodying the invention; and

2Q Figure is a diagrammatic, partly sectioned view of another load- sensitive transmission device embodying the invention.

Throughout the drawings bearing surfaces are indicated by thick lines.

25

Referring firstly to Figure I , a conventional vehicle differential assembly comprises a carrier or cage 3 fixed to, or part of, a crown-wheel (not shown) driven by a pinion (not shown) attached to a prop-shaft (not shown). Shafts I and 2 are half-shafts each serving a respective road-wheel 0 of the vehicle.

Idler gears 4 and 5 having the same number of teeth are mounted on stub axles 6 and 7 carried by the carrier 3 and are free to rotate about their own axes. The idler gears 4 and 5 mesh with bevel gears 8 and 9 carried respectively by shafts I and 2. 5

In the particular example shown in Figure I , the idler gears 4 and 5 have a similar number of teeth to bevel gears 8 and 9, but this is not essential and in most examples of this type of device bevel gears 8 and 9 would have a greater number of teeth than idler gears 4 and 5. 10

Let us assume that shafts I and 2 are loaded, ie. there is some resistance to rotation present. If a unit of torque (IT) is applied to the carrier 3 (in either direction), then the torque will be split between the two bevel gears 8 and 9, providing each with .SOT in the same direction as that

•5 of the carrier.

If the loading upon shafts I and 2 is equal, and less than the equivalent of .50T, then the two shafts will be rotated. However, if one of the shafts is loaded to a greater extent (say IT) with the other still loaded at 0 .50T or less, then the differential action will allow the idler gears 4 and 5 to rotate, thereby maintaining rotation of the less heavily loaded shaft.

If the carrier is held stationary and bevel gear 8 or bevel gear 9 is given one or more revolutions, then the same number of revolutions will be S experienced by the driven gear. That is, if gear 8 is rotated once, then gear

9 will also be rotated once. However, this equal response is an equal and opposite response. Therefore, if bevel gear 8 is rotated once in a clockwise direction, bevel gear 9 will rotate once in an anti-clockwise direction.

0 |f- for example, bevel gear 8 is held fixed, and the carrier 3 is rotated once, then bevel gear 9 will rotate twice in the same direction as that of the carrier.

This apparent 2 : 1 characteristic is, in some ways misleading, in that, if the carrier 3 were still rotated with a IT capability, and bevel gear 8 were 5 held stationary, the fact that bevel gear 9 would now describe two revolutions makes no difference to the torque spiit: the stationary bevel gear 8 would still have .50T applied to it, and the rotating gear 9 would also

hαve .50T applied to it.

If the idler gears 4 and 5 are regarded as levers, it will be seen that the fulcrum, in each case, coincides precisely with the rotational axis of each idler gear. Therefore, if IT of torque is applied to gear 8, and the carrier 3 is held stationary, there is a IT output torque capability transferred to gear 9 in the opposite direction to that of gear 8, but there is also be a torque of 21 applied to the fulcrum (in this case the stub-axles 6 and 7) in the- same direction as the IT torque applied to gear 8.

As the stub-axles 6 and 7 are fixed to, or part of, the carrier 7, this means tha the carrier is loaded with twice the transferable torque or loading, this being due to the lever lengths L I and L2 on each side of the fulcrum being the same.

Correspondingly, a torque of 2T applied to gear 8 will provide an output of 2T at gear 9 with 4T present at the fulcrum.

Figure 2 shows a split differential gear assembly embodying the present invention employing idler gears in the form of compound gear assemblies. Thus first idler gears 14 and 15 are fixed to, or part of, respective lay-shafts 16 and 17 which also carry respective second idler gears 20 and 21. The lay-shafts 16 and 17 are mounted, freely running on the carrier 13 radially of the axis of rotation of shafts I 1 and 12 which carry respective bevei gears 18 and 19 engaged respectively with the first idler gears 1 and 15 and the second idler gears 20 and 2 1 .

If bevel gear 18 had 40 teeth and idler gears 14, 15, 20, 21 were of similar size, and gear 1 9 had 20 teeth, then an unequal differential could be realised. However, this would rely upon the change in gear ratio to effect an unequal status, and this is not the basis of the present invention.

In Figure 2, idler gears 14 and 1 5 have the same number of teeth as bevel gear 18 and idler gears 20 and 21 have the same number of teeth as bevel gear 1 . The fact that gears 18, 14 and 15 are three times the diameter of gears 1 , 20 and 2 1 causes the fulcrum F (and in this case, the centre of rotation of the compound idler gears) to be displaced. Thus while the rotational gear ratios maintain the i : I characteristics of the device

shown In Figure I , the off -set fulcrum allows for a torque split of unequal proportions between the shafts 1 1 and 12.

This will not, however, cause any difference in the basic principle of operation already described with reference to Figure I , but will result in an unequal loading of bevel gears 18 and 19, while retaining the same I : 1 /2 : I speed differential properties. Thus, it is demonstrated by Figure I that it is the speed which is differentiated and not the torque which remains constant. The assembly of Figure 2 will also behave in the same way.

It will be seen, however, that, in Figure 2, the fulcrum F is not situated half-way between the bevel gears 18 and 19 but is closer to gear 19 than to gear 18. The proportions of this difference are indicated and, in this example, the distance LI from the fulcrum F to the hypothetical pitch- circle of gears 18, 14 and 15 is three times the distance between the fulcrum and the hypothetical pitch-circle of gears 19, 20 and 21.

Therefore, as the leverage characteristics must be . observed, regardless of rotational ratios, the application of a torque of IT to the carrier 13 will cause a torque of 0.25T to be applied to bevel gear 18 and 0.75T to be applied to bevel gear 19. This is an unequal differential which cannot be achieved by an equal differential and supplementary gearing.

In Figure I , the rotational differentiation across the I : I gearing did not effect the torque loadings, for this type of differential does not trade speed for torque, or torque for speed. The same is true for the assembly of Figure 2. The I : I gearing still provides a 2 : l /l : I speed differentiation, but the off-set torque loadings remain constant.

To compare the assemblies of Figures I and 2, it could be useful to consider locking the respective bevel gears 8 in both examples. Then, by applying a torque of 4T to the carrier in each case we would find that for every single carrier revolution in Figure I the bevel gear would rotate twice with a torque capability of .50T multiplied by 4, i.e. 2T. In Figure 2 however, the off -set fulcrum position would cause bevel gear 19 to be torque loaded by 0.75T multiplied by 4, giving a 3T torque output. However, the single revolution of the carrier 13 would still provide two revolutions of

bevel gear 1 , owing to the I : I gear ratios employed.

If alternatively the respective bevel gears 9 and 19 were locked in Figures I and 2, then a torque of 4T applied to the carrier 3 or 13 will still provide, in both cases, two revolutions of bevel gear 8 or 18 for every single revolution of the carrier, but as the fulcrum is not off -set in Figure I , the two revolutions of gear 8 will still retain a 2T capability. In Figure 2, the two revolutions of gear 18 will only have a I T capability owing to the off¬ set input fulcrum.

Ratio changes can be included in these devices, but this is not fundamental to the principle of operation. However, such ratio changes would have an effect upon the rotational and torque characteristics.

A differential gear assembly as illustrated in Figure 2 is capable of allowing a I : I drive coupling through the gears with the carrier 13 held fast, and still allowing for a torque split when drive is applied via the carrier, is a very useful device, and Figures 3 and 4, both illustrating fully variable, constant mesh gear boxes, give some idea of one possible application for such a torque split differential gear assembly.

Figure 3 shows a load-sensitive transmission device having an input shaft 44 and an output shaft 51 the rotational axis of each shaft coinciding with the main centre datum of the device.

Input-shaft 44 is fixed to, or part of, a differential hub or carrier 35, this being provided with two radial stub-axles 34 and 37, the axes of the two stub-axles being at 90° to the axis of input shaft 44.

Held in free-running, bearing location, upon the two stub-axles, by end plates 33 and 36, there are two bevelled idler gears 38 and 32.

Outer case 66 is intended to provide an earth, or stationary reference. Any torque introduced by way of input shaft 44 will, therefore, be subjected to torque splitting. Thus, if one unit of torque I I is applied to input shaft 44, then the idler gears 38 and 32 provide two outputs, each of

0.50T. The two outputs from idler gears 38 and 32 are created across the

engαgement contacts they have with bevelled gears 39 and 31. The tooth ratio is I : I , and the characteristics of operation are similar to those of the differential device shown in Figure I.

Bevelled gear 39 is fixed to, or part of, sleeve-shaft 40 and bevelled gear 31 is fixed to, or part of, shaft 30.

Shafts 40 and 30 have the same rotational axis datum as shafts 44 and 51 , and shaft 40 is concentric with shaft 44 but bearing located in such a way as to allow totally indpendent rotational capability.

Shaft 30 is provided with a spur-gear 29, located at the opposite end ot the bevelled gear 31. Gear 29 is engaged with gear 25 at a suggested ratio of 2 : I i.e. if gear 29 has 20 teeth, then gear 25 will have 40 teeth.

Sleeve shaft 40 is also provided with a spur-gear 43 and this is engaged at a suggested ratio of I : I with spur -gear 42.

It will be seen, therefore, that the input torque applied to the input shaft 44, is split, or divided equally by the differential action across gears

38, 39, 31 , 32, and provides two torque-paths. Path 31 , 30, 29, 25, and path

39, 40, 43, 42. Note that, if a torque of I T is introduced to input shaft 44, then 0.50T will be applied to each of the gears 31 and 39. This will be transmitted to gears 43 and 29, and the 2 : I ratio across gears 29 and 25 will result in I T being present at gear 25, and the I : I coupling of gears 43 and 42 will result in there being 0.50T present at gear 42.

Spur-gear 25 is fixed to, or part of lay-sdtøft 24 and spur-gear 42 is fixed to, or part of, lay-shaft 41 . Lay-shaft 24 is provided with compound spur-gear 23, and lay-shaft 41 is provided with compound spur-gear 27. Both assemblies 25, 24, 23 and 42, 41 , 27 are case-held, free-running items.

Spur-gear 23 is engaged with sun-gear 22 at a suggested ratio of 4 : I and spur-gear 27 is engaged with spur-gear 28 at a suggested ratio of I : I . This means that gears 28 and 22 will rotate in the same direction as the input shaft 44.

Gear 28 is fixed to, or part of, lay-shaft 26, and compound spur-gear 19 is also fixed to, or part of, lay-shaft 26. Gear 19 is engaged with sun- gear 21 at a ratio of 2 : I , so that gear 21 will rotate in a direction opposite to that of the input shaft 44 at a torque value similar to that of the input shaft: i.e. if I T is introduced via shaft 44, then IT will be present at gear 21.

As gear 23 is engaged at a ratio of 4 ; 1 with sun-gear 22 and gear 22 is fixed to, or part of, shaft 43, it will mean that the assembly 22, 63 will rotate in the same direction as the input-shaft 44 and with a torque value four times as great: i.e., if IT is introduced by shaft 44 then shaft 63 will rotate ai 4T.

Gear 71 is fixed to, or part of, sieeve-shaft 70 which is concentric with shaft 63 but bearing separated. Therefore, sleeve-shaft 70 rotates in an opposite direction to that of shaft 63 without any direct reference to shaft 63.

Shaft 70 is also fixed to, or part of, bevelled gear 68 and shaft 63 is fixed to, or part of, differential carrier 58, the whole of the resultant output section being similar to the differential assembly described with reference to Figure 2.

It will be seen, that the output differential is a torque splitting or dividing assembly and the rotational inputs from the two torque paths from input shaft 44 enable a balanced drive to be established.

The out-of-balance torque split incorporated in this particular example, is one using a lever system across the fulcrum F in which one side of the lever is three times longer than the other. Therefore, if the fulcrum (as represented by stub-axles 60 and 56) is driven forward (i.e. in the same direction as input shaft 44), the torque present upon the fulcrum will be divided unequally: i.e. if IT were introduced via the fulcrum, then, as in Figure 2, the torque split will provide the shorter arm with three quarters of the available torque, and the longer arm with only one quarter of the available torque. Assuming the fulcrum is provided with 4T, then the short lever L2 will have an output of 3T and the longer L I will have an output of I T.

The differential hub or carrier 58 Is provided with two, or more, radial stub-axles 60 and 58, these being provided with end plates 61 and 64 in order that the two free-running compound bevelled gear assemblies 7, 9,

17 and 5, 4, 3 are retained in constant, but bearing located communication with the hub 58.

The two conical assemblies can be of any suitable configuration which provides for an off-set rotational centre: i.e. a centre of rotation that does not coincide with the centre datum situated between the two engaged pitch circles: i.e. gears 67 and 53 are engaged with output gear 52 and compound idler gears 57 and 55 are engaged with the first bevelled gear 68.

As the two compound elements of the idler gear assemblies are of a I

: I nature (other ratios are possible), then the conical construction lends itself to easy realisation. In fact, the tooth face could have been extended right across the surface of each cone, however, the separating into two gear tooth face areas helps to define the exact location intentions.

This driven fulcrum arrangement only splits the torque derived from the off-set axle datum of the compound idler units. However, the unequal lever arm lengths L I and L2 between the pitch-circle engagement of gears 68, 57, 55 and the fulcrum, and the fulcrum and the pitch-circle engagement of gears 67, 53, 52 would indicate that the IT input torque introduced by way of gear 68 would cause the fulcrum to be loaded with 4T in a direction opposite to that of the input shaft 44, and the output gear 52 to be loaded with 3T in the same direction as the input shaft 44. As the fulcrum is, itself, forward loaded with 4T in the same (forward) direction similar as the input shaft 44, then the IT introduced by gear 68 cannot reverse the fulcrum with its 4T loading in the opposite (rearward) direction to that of shaft 44, and the fulcrum is balanced. If, in this situation, the fulcrum is unable to move one way or the other, then gear 31 is also unable to rotate, however, this would cause gear 39 to make two revolutions for every single revolution of shaft 44 and gear to make one revolution in a direction opposite to that of shaft 44 for every single revolution of sdj ft 44: i.e; with the fulcrum locked, gear 68 will be at I : I with the input shaft 44.

If, therefore, the forward 4T torque of the fulcrum is negated by the

reαrwαrd 4T loading derived from the IT introduced by gear 68, it must be remembered that the 4T rearwards loading of the fulcrum was only established as a resuit of 3T forwards having been established upon gear 52 by the lever L I /L2 action across the fulcrum. Therefore, in order for there to be a 4 torque rearwards loading of the fulcrum, there must have been established a 3T forwards loading of the output gear 52.

If the other possibility is examined, i.e. the fulcrum input torque is dominant, then the off-set, or biased input situation is again effected by the lever lengths. Thus distance L I being three times greater than distance L2 will cause the 4T introduced by the fulcrum to be split into 3T output to gear 52 and IT output to gear 68. As gear 68 is loaded with IT in a direction opposite to that of the driven fulcrum, the IT is balanced out, and gear 68 is unable to rotate. However, the 3T provided by the fulcrum is still loading gear 52 in the same direction as shaft 44.

Either way, gear 52 is forward loaded with 3T. However, if gear 68 should rotate, with the fulcrum locked, then the output gear 52 will rotate at I : I with the input shaft 44, with gear 31 stationary and gear 39 rotating twice for every single revolution of shaft 44. If gear 68 is stationary, together with gear 39, then the fulcrum will move forwards by 0.25 of a revolution.

The torque loading of gear 52 remains constant at 3T.

In order to establish the output capability of the overall device of Figure 3, it will be seen that, with the fulcrum iocked, a IR /revolution) via gear 68 will remain a IR directionally opposite output on gear 52 providing the idlers can rotate, and gear 52 is capable of being driven by IR., i.e. I : I with shaft 44, and the torques remaining constant, with only the speeds differentiating. Like any differential device, only the speeds are differentiated, the torques remain constant i.e. speed can be traded for speed but torque wiil remain constant regardless of the speed of gear 52.

The output gear 52 is fixed to, or part of, output shaft 5 1 and, assuming this is the loaded termination, the establishment of 3T at gear 2 suggests that a 3T load connected to shaft I can be driven.

Theref ore, if gears 39 and 31 rotate in unison with shaft 44 (i.e. one revolution of shaft 44 producing one revolution of both gears in the same direction), then gear 68 will rotate 0.50 revolutions in the opposite direction to sahft 44, and carrier 58 will rotate only 0.125 revolutions in the same rotational direction as shaft 44.

This will cause the conical assemblies to rotate 0.50 revolutions (as a result of the I : I engagement across 68, 55, 57) plus the distance they must orbit or walk around gear 68 as a result of the hub 58 rotating in the opposite direction to 68 by 0.1 25 revolutions. This means that gears 67 and

53 rotate 0.625 revolutions.

To this rotational factor there must be added the +0.1 25 forward revolutions of the differential carrier 58 to obtain a rotational value for gear 52, i.e. 0.625 + 0.125 = 0.75 revolutions. Therefore, with the input differential assembly rotating en masse, the output gear 52 can be rotated at 0.75 revolutions with 3T capability. This gives an input to output speed ratio of I : 0.75.

If. for example, gear 39 is stationary, and input shaft 44 is rotated by one revolution, then gear 22 will rotate only 0.25 revolutions (providing 29/25 is 2 : I is 4 : I ) and gear 68 will be stationary. This results in carrier 58 rotating forward 0.25 revolutions, thereby causing gear 52 to rotate forward by 0.50 revolutions.

These two speed/speed examples indicate the limited-slip type of operation possible, and other speed variations can be demonstrated. For example, with gear 39 rotating forward at 0.75 revolutions shaft 44 rotating forward at I revolution, and gear 31 rotating forward at 1.25 revolutions, 9 ear 2 will rotate forwards 0.6875 revolutions, i.e. input to output ratio between shafts 44 and 51 will be 1 : 06875 in respect of speed at a constant 3T.

The other interesting example to contemplate is the one In which gear 3 1 is stationary, shaft 44 is rotated forward by one revolution, and gear 39 is rotated forward by 2 revolutions.

This produce α stationary carrier 58 with gear 68 rotating one revolution rearwards: i.e. this produces a single forward revolution of gear 52, with the input to output ratio being I : I at 3T.

Therefore, the speed differentiation between input 44 and output 51 gives an operating range of 2 : I to I : I with a 3T output capability available at ail times-.

The: spreed change is governed by the speed of the output shaft (the load)aπd assuming that the load can be accelerated, then the differentiation across the input differential will take place, enabling equalisation to take place.

The input/output capability can be extended, by making the basic reduction path 31 , 19, 25, 23, 22 more capable. For example, assuming gear 29 is still provided with 0.50T, then by making the 29, 25 engaged 4 : I instead of the present 2 : 1 , the overall ratio between gear 31 and hub 58 would be 8 : I . However, this would require the off-set fulcrum " to be corrected, i.e. repositioned so as to aliow L I to be seven times longer than L2 if a top-speed ratio of I : I is to be maintained. This would also increase the torque capability to a 7 : I output.

The speed difference would, however, give the following bottom speed capability a lower differential factor: i.e. if gear 39 were stationary, and shaft 44 were to rotate forward one revolution at IT, then the 8 : I reduction between gears 29 and 21 wouid cause the carrier 58 to rotate forward by 0. 125 revolutions. This muitipled by the 2 : 1 differential factor (i.e.. rotational ratio across the output differential) will give the speed of gear 52, i.e. 0.25 revolutions, a speed difference ratio between input 44 and output 51 of 4 : I , with a maintained torque of 7T.

Any change of rotational status by gear 52 from stationary, will allow the input differential to equalise. However, if a prejudice is included in either path, i.e. a ratio difference which does not quite balance across the output differential, then, aside from having a small proportion of unreacted torque i.e. if, say, gear 18 was capable of providing IT and one revolution, with gear 3 1 stationary, and gear 39 rotating twice for every single

SUBSTITUTE SHEET

revolution of shaft 44, and the fulcrum datum was moved slightly, so as to allow less than a IT output from the lever action, or as in Figure 3, say the 4: I train was only capable of providing, say, 3.75 : I , then the IT capability of 18 would become dominant and it coudl reverse the fulcrum, but only 5 after applying 3.75T to the fulcrum and 2.75T to gear 2; in other words, the fulcrum could not be reversed if the load upon shaft I was anything less than 2.75T.

The fact that a 0.25T bias has been added to gear 68 means an - . overlap will have been created, and one which will favour, at ail times, the 1 : 1 speed input/output situation: i.e. the 0.25 will drive the device towards I : I for the sake of having sacrificed 0.25T in overall torque output capability.

1 These figures are based upon the premise that both torque paths have similar frictional losses and similar inertial values.

The input differential described is a I : I device. However, this could be a split differential, or a ratio stepped differential, if required, and 0 imbalances can be included and almost any speed ratios can be contemplated.

Figure 4 illustrates a device similar to that of Figure 3, but in which input differential is not driven by the carrier. In Figure 4, the input 5 differential is driven by a input applied to the first bevelled gear 136.

Therefore, IT applied to input shaft 138 will be applied to input bevel 136 which is fixed to, or part of, shaft 138. Gear 136 is engaged with idler gears 130 and 129 which are free-running upon stub axies 131 and 132, these 0 being provided with end plates 133 and 134 in order to ensure that gears 130 and 129 are retained upon the stub axles.

Stub axles 131 and 1 32 are fixed to, or part of, differential carrier 135 which is, itself, fixed to or part of datum shaft 127 which is also 5 provided with a spur-gear 122 and a bearing location shaft 137.

Idlers 130 and 129 are also engaged with bevelled gear 1 28 which is fixed to, or part of, concentric sleeve shaft 126. Sleeve shaft 126 is bearing

locαted around shaft 127 sharing the same axial datum, and is provided with a spur-gear 123. Lay-shaft 1 17 is a free-running item provided with compound planet gears 124 and 1 16, these being engaged with sun-gears 123 and 1 15 respectively. The through-ratio from gear 1 23 to gear 1 15 is I : I .

As gear 1 15 is fixed to or part of sleeve shaft 1 14, which is itself fixed to, or part of gear 1 13, gear 1 13 is also driven at a ratio of I : I relative to gear 28 and in the same directions i.e. opposite to that of input shaft 138. The direction of gear 128 is rotationaliy reversed by idler gears 130 and 129-.

Sleeve lay -shaft 120 is concentric with lay-shaft 1 17 and is provided with two compound planet gears 121 and 1 19, these being engaged with sun- gears 122 and 1 18 respectively. The through-ratio from carrier 135 to gear I 18 is 5 : 1 , so that shaft I 12 which is fixed to, or part of, gear I 18 is driven at 5 : I down 5 revolutions of shaft 138 resulting in one revolution of shaft 1 12 in the same rotational direction.

Output differential carrier I 10 is fixed to, or part of, shaft 1 12 and is therefore driven forward at a rate of 5 : I . A torque of IT applied to shaft 138 will thus result in 5T being applied to carrier I 10.

The output differential shown in Figure 4 is similar to that shown in Figure 3. However, the split, or off-set fulcrum F provides at 4 : 1 imbalance.

Gear 1 13 is engaged with 109 and I I I . Compound bevelled idlers 103 and 105 are engaged with output bevelled gear 102 and all enagements are indicated as being 1 : 1.

The fact that L I is four times longer than L2, means that the 5T present at the indicated fulcrum (i.e. the axial datum of axles 107 and 106) will be split into IT and 4T proportions IT out from gears 109 and I I 1 to gear I 1 3 and 4T out from gears 103 and 1 05 to gear 102.

Torque split differential assemblies embodying the present invention can be incorporated into transmission trains in many different ways, and

Ffgures 3 and 4 are included merely as examples. Furthermore, the I : I ratios Indicated in Figures 2, 3 and 4 can be modified to include other variations, maintaining the basic off-set fulcrum principle, however.

In Figure 4, the concentric lay-shaft arrangement is purely hypothetical, and outer case 108 is assumed to be earthed.

Intermediate gearing can be introduced between gears 103 and 109 and 105 and I I I etc.

The off-set torque split differential assemblies shown in this specification are based upon bevelled differential arrangements. However, it is envisaged that the principle could equally be applied to other forms of gear, such as epicylic and compound spur-gears, for example.