EP0258369A1 - Self-adjusting transmissions - Google Patents

Abstract

A transmission device comprises an input shaft (1) rotating around a main axis (X-X) and an output shaft (2) rotating around the same axis. An input gear (6), placed on the input shaft, drives an input bevel gear (13) belonging to a differential unit having a differential support (18) rotating with a first axle (9) supported by rotatably by a support plate (8). The differential carrier comprises two spindles (16, 17) supporting parasitic bevel gears (14, 15) engaging with an input bevel gear and an output gear (21), which engages with a fixed gear (24 ) via a reference gear (23). An intermediate gear (25) rotating with the first axle drives, via a parasitic gear (26), one (28) of two planetary output gears (28, 29), rotating together with a second axle (10) supported by the support plate at a radial interval from the axis (XX) different from the interval from the first axle. The other (29) planetary gear is engaged with an output gear (7) mounted on the output shaft. The torque applied to the input shaft is transmitted to the output shaft with a value inversely proportional to the speed of the output shaft, the rotation of the support plate (8) being in the opposite direction to that of the input and output trees.

Description

Description of Invention

"Self-adjusting transmissions"

THIS INVENTION relates to self-adjusting mechanical transmission devices which find particular, but not exclusive, application in the power train of a motor vehicle.

A gearbox which continuously and steplessly provides the correct coupling ratio between a source of mechanical power, eg a vehicle engine, and a mechanical load, eg road wheels of the vehicle, is an ideal device not only for motor vehicle power trains but for any situation where it is desirable to provide modification of speed or torque between the power source and load.

The present invention aims to provide such a mechanism, thereby opening up a new approach to automatic transmissions for vehicles. However, whilst the following detailed description is slanted towards the motor vehicle applications of the invention, it should be understood that the invention is equally applicable to many other machines including oil drilling rigs and machine tools.

According to the invention, there is provided a transmission device comprising an input shaft rotatable about a main axis of rotation, an output shaft rotatable about the main axis, an input gear rotatable with the input shaft, a carrier body rotatable about the main axis around the input shaft and provided with first and second axles defining respective first and second subsidiary axes of rotation which are parallel to the main axis and at different distances from the main axis, a differential gear assembly mounted on the first axle and having a differential input gear driven by the input gear on the input shaft and first and second differential output gears, reference gear means coupling the first differential output gear to a stationary gear centred on the main axis, and output gear means coupling the second differential output gear to the output shaft, the reference gear means on the output gear means including a gear assembly mounted on the second axle.

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

Figure 1 is an axial cross-section through a first transmission device embodying the invention;

Figure 2 is an axial cross-section through a second transmission device embodying the invention;

Figure 3 is an axial cross-section through a third transmission device embodying the invention;

Figure 4 is an axial cross-section through a fourth transmission device embodying the invention;

Figure 5 is an axial cross-section through a fifth transmission device embodying the invention; and

Figure 6 is an axial cross-section through a modified form of the first embodiment shown in Figure 1 .

Referring to Figure 1 , a first transmission device embodying the present invention comprises input shaft 1 and an output shaft 2 extending into a non-rotatable outer casing 3 and supported by journals 4 and 5 of the casing 3 for rotation about a common main axis X-X. An input sun gear 6 having 42 teeth is rotatable with the input shaft 1 within the casing 3 and an output sun gear 7 having 70 teeth is rotatable with the output shaft 2 within the casing 3.

A main carrier 8 is rotatably supported by the journals 4 and 5 for rotation about the axis X-X and supports a first and second layshafts 9 and 10 for rotation about respective subsidiary axes Y-Y and Z-Z which are parallel to the axis X-X and lie at respective distances of 41.00 mm and 63.50 mm therefrom.

The first layshaft 9 carries an input planet gear 1 1 which has 42 teeth and is engaged at a ratio of 1 : 1 with input sun gear 6. The input planet gear

1 1 is fixed to a sleeve shaft 12 which is freely rotatable on layshaft 9 and is formed with an input bevel gear 13 engaged at a ratio of 1 : 1 with two idler bevel gears 14 and 15 freely rotatably mounted on respective stub axles 16 and 17 extending from a differential carrier 18 which is formed integrally with the layshaft 9. The idler bevel gears are retained on the stub axles 16 and 17 by end caps 19 and 20. A reference bevel gear 21 is formed with a sleeve shaft 22 which is freely rotatably received on the layshaft 9. Each of the bevel gears 13, 14, 15 and 21 has 38 teeth. To the sleeve shaft 22 is fixed a reference gear 23 having 33 teeth. The reference bevel gear 21 is engaged with the idler bevel gears at a ratio of 1 : 1 and the reference gear

23 is engaged at a ratio of 1.5454545 : 1 with a static gear 24 which has 51 teeth and is attached to the outer casing 3 around the journal 4.

An intermediate gear 25 having 75 teeth is rotatably with the layshaft 9 and is engaged with an idler gear 26 which has 38 teeth and is freely rotatable upon a stub axle 27 fixed to the main carrier 8. The idler gear 26 is retained on stub axle 27 by end cap 30 and is engaged with a first output planet gear 28 having 60 teeth, so that the ratio between the intermediate gear 25 and the output planet gear 28 is 0.8: 1. The output planet 28 is rotatable with the second layshaft 10, to which is also affixed a second output planet gear 29 having 56 teeth and directly engaged with the output sun gear 7 at a ratio of 1.25: 1. Input shaft 1 has a locating portion 31 received in a corresponding locating journals 31 a of carrier 8. A retaining flange 5b is provided on output journal 5.

Means (not shown) are provided on the carrier for rotationally counterbalancing the transmission device.

In order to explain the working cycle of the Figure 1 device, it is necessary to consider the action of the differential unit composed of input bevel gear 13, idler bevel gears 14 and 15, reference bevel gear 21 and differential carrier 18. If a torque of value IT (where T is a unit of torque) were applied to the input bevel gear 13, then there exists a potential for a torque I T to be applied to the reference bevel gear 21 in the opposite rotational direction to that of the input bevel gear and, at the same time, for a torque of 2T to be applied to the differential carrier 18 via the idler bevel gears 14 and 15 in the same rotational direction as the input bevel gear 13. This depends upon there being resistance by both the reference bevel gear 21 and the differential carrier 18.

Given that the main carrier 8, together with its associated gears and shafts does indeed have some specific mass and represents a load, then it wil l be seen that, for the reference bevel gear 21 to rotate, the main carrier 8 must be moved, since the bevel gear 21 is indirectly engaged via reference planet gear 23 with the static gear 24 which is itself unable to rotate. Therefore, if the reference bevel gear 21 is to rotate, it can only do so by the reference planet gear 23 progressing or walking around the static gear 24. In doing so it will cause the main carrier 8 to rotate.

The fact that there is some resistance offered to the combination of reference bevel gear 21 and reference planet gear 23 means that there will be leverage applied to the differential carrier 18, so that two outputs from the differential unit can be shown to exist.

In order for this situation to continue, it must be shown that the action of the reference planet gear walking around the static gear 24 does not simply cause the output planet gear 29 to walk around the loaded output gear 7. This can be determined by considering the contradictory loadings applied to the main carrier 8.

If a torque I T is applied to the input shaft 1 , a torque of IT will be directly applied to the input planet gear 1 1 as the engagement between the input gear 6 and gear 1 1 is 1 : 1. However, the same IT from the input shaft 1 will also forward-load (i.e. in a similar rotational direction to that of the input shaft 1 ) the main carrier 8 with a torque of I T. Therefore, the IT torque directly applied to the input planet gear 1 1 will cause gear 11 (a) to rotate about layshaft 9 in the opposite direction to the rotation of the input shaft 1 and apply, via the differential unit, a torque IT in a rotational direction similar to that of the input shaft gear to the reference bevel gear assembly and (b) to apply a torque of 2T to the differential carrier 18 in a direction opposite the torque on the input shaft 1.

The IT torque capability provided at the reference planet gear 23 will be transmitted to the static gear 24 at a ratio of 1.5454545T, so that a torque of 1.5454545T will be applied to the static gear 24 and, at the same time, 0.5454545T of forward loading will be applied to the main carrier 8 as a result of the layshaft 9 being bearing located within this device.

This means that, for a torque of IT applied via the input shaft 1, a torque of 1.5454545T is applied to the main carrier 8 in the same direction as that of the input shaft 1.

Meanwhile the torque of 2T present upon the differential carrier 18 in a rotational emphasis opposite to that of the input shaft 1 is applied to the intermediate gear 25 via layshaft 9. This torque of 2T is transmitted via the idler gear 26 to the first output planet gear 28 at a ratio of 0.8: 1, so that a torque of 1.6T is applied to the layshaft 10 carrying output planet gear 28. The 1.6T torque is applied to the output gear 7 by way of the second output planet gear 29 at a ratio of 1 .25 : I , the output drive loading upon the output gear 7 being given by multiplication of the two figures, i.e. 1.25 x 1.61 = 2T. Therefore, a driving torque of 2T is applied to the output gear 7.

If the output gear 7 is now considered a variable static gear, the 2T torque output from the output planet gears 28 and 29 will attempt to treat gear 7 as a fulcrum in exactly the same way as the reference gear 23 treats the static gear 24. Thus, by considering the loadings placed upon the static gear 24 and the output gear 7, the forward and rearward loadings can be decided in terms related to the forward and rearward loadings upon the main carrier 8.

If the differential unit provides a IT torque output to the reference gear 23, which is converted into 1.5454545T loading of the static gear 24 acting as a fulcrum, then the output from this lever system will be 0.5454545T applied to the main carrier 8 via the bearings of layshaft 9. If the combination of input gear 6, input planet gear 1 1 , differential unit, reference gear 23 and static gear 24 is considedred as a lever, the reference lever, with the input gear 6 as one end of the lever, the static gear 24 as the fulcrum and the main carrier 8 as the other end of the lever, then a torque of IT applied to one end of the lever, producing 1.5454545T upon the fulcrum, must apply the difference to the other end of the lever. Therefore, the result of the input gear 6 applying a torque of IT to the system will be:-

(i) forward loading of the main carrier 8 with a torque of IT resulting from torque introduced.

(ii) forward loading of the main carrier 8 by a torque of 0.5454545T resulting from the lever action of the lever comprising the input gear, static gear and main carrier.

This produces a total forward loading of 1 .5454545T on the main carrier 8.

If similar logic is applied to the output gear side of the device, then the IT torque introduced via the input gear 6 will be at one end of a lever, the output lever, the output gear 7 being the fulcrum and the main carrier 8 again representing the second end of the lever. Therefore a torque of IT applied to one end of this lever, producing a torque of 2T upon the fulcrum (the output gear 7) produces IT of rearward loading upon the main carrier 8 via the bearings of the second layshaft 10.

The main carrier 8 is itself a lever with 0.5454545T being applied to one end, and IT being applied to the other. The fulcrum loading is not important as the fulcrum coincides with the main axis X-X and is a fixed fulcrum. Therefore, the IT more than balances the 0.5454545T and, if the lever were of equal length each side of the fulcrum axis, the rearward emphasis would be 0.4545455T in favour of rearward rotation of the main carrier. However, it must not be overlooked that there is a further IT of forward loading being applied to the main carrier by the input gear 6. Therefore instead of 0.4545455T trying to rotate the main carrier rearwards there is, in fact, 0.5454545T trying to rotate the main carrier forwards. It will be seen, however, that the points of loading application to the main carrier are not at equal distances from the fulcrum axis X-X and this off-set must be taken into account. As the IT is applied to the main carrier in a rearward direction at 63.50mm from the axis X-X and the 0.5454545T is applied at only 41.00mm from the axis X-X, the off-set ratio of 1.5487805 :

1 must be included in the rotational equation.

This means that the IT applied to the main carrier via the bearings of layshaft 10 will be 1.5487805 times more effective than the 0.5454545T applied to the main carrier via the bearings of layshaft 9, so that a rearward loading of IT x 1.5487805 = 1.5487805T is present and this loading is greater than the total forward loading of the main carrier by 0.0033600T (approximately).

Therefore, there is an overlap torque in favour of rearward rotation of the main carrier of 0.0033600T simply as a result of the leverage created by the axes Y-Y and Z-Z of the layshafts 9 and 10 being at unequal distances from the axis of rotation X-X of the main carrier.

|f any frictional losses make either lever train more or less effective than the other or if these figures are incorrect in any way, then simply by increasing or decreasing the radial length (distance between axis X-X and axis X-Y or Z-Z) of either lever arm the forward or rearward torque loadings as 'applied to the main carrier can be balanced or, as in this device, over-balanced in order to create a change-up emphasis to the mechanism.

This is achieved without any ratio change being made to the gears themselves and the same ratios applied at greater disparity will provide greater rearward emphasis. If the output gear 7 can be accelerated to reach the same rotational forward speed as the input gear 6, then for every single forward revolution of the two gears 6 and 7 the main carrier 8 will rotate rearwards 1.8333352 revolutions. This means that the differential unit is rotating en masse with ail the bevel gears held non-rotating relative to one another. Therefore, during 1 .8333352 rearward rotations of the main carrier, the differential unit will rotate rearwards 2.8333362 revolutions, with 2.8333362 revolutions of the input planet gear 1 1 producing 1 .000001 forward revolutions of the input gear 6. Also, 2.8333362 revolutions of the differential unit wil produce 2.8333362 revolutions of the intermediate gear 25, resulting in 3.5416703 revolutions of the output planet gears 28 and 29. This will result in 1.000001 forward rotations of the output gear 7 in unison with the 1.000001 forward revolutions of the input gear 6 representing a 1 : 1 input/output or top gear situation.

The total regulation from a 2 : 1 bottom gear ratio to a 1 : 1 top gear ratio results from the speed of the output gear 7 as it is accelerated from rest. The range of ratios achieved with the Figure 1 device is of course merely an example.

Outer casing 3 is shown as being lubrication tight. However, the framework necessary to support the various bearing assemblies etc., can be to any enclosed or open design providing sufficient lubrication can be maintained at the various bearings and contact surfaces.

Static gear 24 is indicated as being fixed to, or part of, outer casing 3. However, the gear 24 could be provided with a means of becoming non-rotatable. For example, gear 24 could be attached to the outer casing 3 by way of a clutch, or brake mechanism, thereby allowing gradual locking of what could be a normally free-running gear. This would allow for disengagement of the whole transmission, for if the static gear 24 were allowed to rotate freely upon its own bearings, there would be no output felt by the output gear 7, and the whole transmission would remain idle. As braking resistance is applied to gear 24 drive would be gradually applied to gear 7 and the total locking of gear 24 to the casing would create full effectiveness, with the 2 : 1 / 1 : 1 envelope established. This gradual locking capability would be ideal in motor vehicle applications as it would allow the engine to be started without there being any output drive established, and the gradual braking of the gear 24 to a non-rotating relationship with casing 3 would be similar to letting in the clutch of a conventional vehicle.

The embodiment of Figure 1 does not include provision for driving the output in the reverse direction. This could however be added as a separate system or, if required, means of reversing the whole transmission can be included.

A driving relationship between the input and output shafts 1 and 2 can be established in marry ways. However, adjustable locking of the static gear 24 to the outer casing 3 or to the input shaft 1 or main carrier 8 will enable solid drive to be established without any gear movement other than en masse operation.

The stepless performance of the device without any torque-converter or clutch intervention, coupled with en masse operation of the differential gearing at 1 : 1 , suggests very high overall efficiency. Furthermore, applying one such transmission device to each driving wheel of a vehicle would negate the need for a separate differential and allow for perfect 4 wheel drive.

Figure 2 shows a second transmission device embodying the invention in which the single counter-balanced differential unit and output planet gear arrangement of Figure 1 is replaced by a balanced arrangement having a double differential unit and double output planet gear with an offset ratio of 1.976 1 905: 1. In Figure 2, components which are the same as in Figure 1 have the same reference numerals and duplicated components have the same reference numerals with the addition of the reference letter "a" or "b".

The Figure 2 embodiment differs from Figure 1 in that the idler gear 26 is rotatable with a sleeve shaft 32 rotatably mounted on input shaft 1 and carrying a second idler gear 33 engaged with the first output planet gears 28. The input gear 6 has 42 teeth, the input planet gears 1 1 a and 1 1 b have 42 teeth, the reference gears 23a and 23b have 31 teeth, the intermediate gears 25a and 25b have 53 teeth, the first idler gear 26 has 30 teeth, the first output planet gears 28a and 28b have 83 teeth, the second idler gear 33 has 83 teeth, the second output planet gears 29a and 29b have 60 teeth, the output gear 7 has 106 teeth and the bevelled gears 1 3a, 13b, 14a, 14b, 1 5a, 15b, 21 a and 21 b have 38 teeth.

In Figure 2, the axes Z are each at 83.00mm from the axis X-X and the axes Y-Y are at 42.00mm from axis X-X.

The single differential/counterbalanced layout of Figure 1 is replaced by a double differential and double output planet layout with an off -set ratio of 1.976 1 905 : 1. If reference gear 23a, 23b were of 30 tooth construction rather than 3 1 tooth construction as illustrated, then the off-set ratio would be 83 - 41.50 = 2 : 1.

If the main carrier 8 rotates rearwards 1.4090909 revolutions for every single revolution forward of the input gear 6 then the output gear 7 will also rotate one revolution forwards, establishing a 1 : 1 input/output ratio.

The reference lever formed by input gear 6, static gear 24 (fulcrum) and main carrier 8 drives the main carrier 8 forwards with 0.7096774T and loads the fulcrum (gear 24) with 1.7096774T. To this is added IT from the input gear to main carrier, giving a total forward loading of the main carrier with 1.7096774T.

The output lever formed by input gear 6, output gear 7 (fulcrum) and main carrier 8 drives the main carrier rearwards with IT, and loads the fulcrum (gear 7) with 2T. Nothing is to be added to or subtracted from this as the IT input to the main carrier loading has already been accounted for giving total rearward loading of the main carrier of IT.

The offset ratio of 1.9761905 : 1 must now be taken into account.

The fact that the leverage disparity favours the Z-X lever-arm, rather than the Y-X lever-arm, means that the IT applied to the main carrier 8 via the bearings of layshafts 10a and 10b will be increased by a multiplication of the

IT by the offset ratio, IT x 1.9761905 = 1.9761905T.

Therefore, the total forward loading of the main carrier 8 of 1.7096774T is adequately balanced by the total rearwards loading of the main carrier by the 1.976 1 905T as applied to the main carrier via the bearings of layshafts 10a and 10b.

The 0.2665131 T rearwards emphasis overlap provides definite change-up drive torque. However, the main carrier 8 cannot rotate rearwards unless the output-gear 7 is accelerated. Therefore, the change from 2 : 1 up to 1 : 1 is a driven situation which is totally load sensitive, and clearly demonstrates the torque/speed/speed/torque trading capabilities of the invention.

Figure 3 il lustrates a third embodiment of the invention in which a stepped differential is used in order to widen the torque envelope beyond the range of from 2 : 1 to 1 : 1 provided by the embodiments of Figures 1 and 2.

Components in Figure 3 which correspond to the components in Figures 1 and 2 have the same reference numerals.

The stepped differential unit incorporated in the Figure 3 embodiment is a 4 : 1 ratio device comprising input bevel gear 1 3 engaged with first idler bevel gears 14a, 15a of compound idler gears 14, 15 and reference bevel gear 21 engaged with second idler bevel gears 14b, 15b of the gears 14, 15. The reference bevel gear 21 is compounded with reference gear 23 which engages an idler gear 27 engaged with a first reference planet 34a gear of a compound reference planet gear 34 having a second gear 34b engaged with static gear 24. Compound gear 34 comprises a layshaft 35 journalled in the main carrier 8 which also incorporates a counter-balancing weight 36. An end cap 37 is provided on layshaft 9. Output planet gear 29 is rotatable with layshaft 9 and directly engages the output gear 7, the intermediate gear 25, idler gear 33 and first output planet gear 28 of Figure 1 being omitted.

Input gear 6 has 30 teeth, input planet gear 1 1 has 30 teeth, reference gear 23 has 36 teeth, static gear 24 has 30 teeth, output planet gear 28 has 30 teeth, idler gear 27 has 24 teeth, first reference planet gear 34a has 18 teeth, second reference planet gear 34b has 12 teeth, output gear 7 has 30 teeth, bevel gear 1 3 has 25 teeth, bevel gears 14a, 15a have 50 teeth, bevel gears 14b, 15b have 40 teeth and bevel gear 21 has 80 teeth.

If a torque of IT is introduced via the input shaft 1 and the output shaft 2 is loaded, then the IT forward rotation of input gear 6 will cause input planet gear 1 1 to rotate in the opposite rotational direction at IT, thereby rotating the input bevel gear 13 at IT in a direction opposite to that of the input gear.

The input bevel gear 1 3 and the first idler bevel gears 14a, 15a are at a ratio of 2 : 1 so that these cause the second idler bevel gears 14b, 15b to rotate in opposite directions at 2T. The engagement between the second idler bevel gears 14b, 15b and the reference bevel gear 21 is also at a ratio of 2 : 1. Therefore, the 2T capability bevel gears 14b, 15b is transmitted to the reference bevel gear 21 , causing this to be driven in the opposite rotational direction to that of input bevel gear 13 at 4T.

If 4T is applied to reference bevel gear 21 then 5T must be applied to the differential carrier 18, in the same direction as input bevel gear 13.

As layshaft 9 is rotatable with the differential carrier 18, the 5T is transmitted to the output planet-gear 28 causing this to apply 5T to the output gear 7. Meanwhile, the 4T present at reference gear 21 , which is responsible for the 5T reaction in the differential carrier and layshaft combination 18, 9 is directly applied to the reference gear 23 which is engaged with the idler gear 27. The 36 tooth to 24 tooth engagement between gears 23 and 27 gives a ratio of 0.6666666 : 1. This ratio multiplied by 4T gives the output capability of the idler gear 27, which is 2.6666667T. An idler gear 27 is engaged with reference planet gear 24a at a ratio of ( 18 - 24) 0.75 : 1, the multiplication of 2.66666611 by 0.75 results in 2T being applied to layshaft 35.

The second reference planet gear 34b has 12 teeth and is engaged with static gear 24 which has 30 teeth. Therefore, the ratio is 2.5000000 : 1 and the 2T must now be multiplied by 2.2000000 in order to give the torque loading as applied to static gear 24 by reference planet gear 34b, giving

5.000000T.

If 5.000000T is applied to gear 24 (the fulcrum) by the introduction of IT via input shaft I then the output to the main carrier 8 via the layshaft 35 bearings will be 4.0000000T. However, the IT subtracted from the original figure must be added back as it represents the IT as introduced in forward fashion to the main carrier 8 by the input gear 6. The main carrier 8 is thus forward loaded by 5.0000000T, Against this, there is only 4T rearward loading of the main carrier 8 at the bearings of layshaft 9. However, the offset ratio of 1.4285714 : 1 must now be applied to the advantage of the 4T at the axis Z-Z of layshaft 9. The 4T multiplied by 1.42857 14 gives 5.7142856T and this is greater than 5.0000000T by 0.7142856T.

The 0.7142856T is an overlap torque capable of driving the main carrier 8 in a rearwards direction and thereby, as in all these devices, changing the ratio across the differential and, in so doing, changing the whole input/output ratio. However, for the input shaft 1 and output shaft 2 to rotate once ( 1 revolution forward in unison), the main carrier 8 will rotate rearwardly 4 times.

The number of rearward revolutions of the main carrier 8 can be reduced by increasing the offset ratio and then, as a result, increasing the torque applied to the static gear. This will then enable the differential unit to rotate quicker, thereby reducing the number of main carrier rotations required to ensure that the 1 : 1 situation can be achieved.

Furthermore, an overdrive situation can be included as there is a considerable amount of overlap torque in this device, namely 0.7 142856T. There is room to alter the output planet to output gear ratio for example, instead of 30/30 it is possible to make the output gear 7 into a 29 tooth gear and the output planet gear 28 into a 31 tooth gear. This would result in 4 rearward revolutions of the main carrier 8 producing 4 rearward revolutions of the differential unit (en masse), which would then produce 1 forward revolution of the input gear 6 and 1.3448276 forward revolutions of the output gear 7. This gives the envelope of the device a spread of 4.67741 94 : 1 to 1 : 1.3448275, the overlap torque being reduced to 0.253456T rearward emphasis.

Figure 4 shows a transmission device which features two differential units, these being bearing-located upon axles defining rotational axes Z-Z parallel to the axis of rotation X-X of the input and output shafts 51 and 52.

Two pairs of compound reference gears run on similar axles defining axes Y-Y parallel with the axis X-X but at a smaller radial distance from the axis X-X than are the axes Z-Z. For example, the axis Z-Z may be at a 50.00mm radial distance out from the axis X-X while the axis Y-Y is only

35.00mm radially distant from the axis X-X. The off-set ratio thus 1.4285714 : 1 in favour of axis Z-Z. The Figure 4 embodiment comprises an input shaft 51 , output shaft 52, fixed axle 53, fixed axle 54, differential layshaft 55, differential layshaft 56, sleeve shaft 57, sleeve shaft 58, sleeve shaft 59, sleeve shaft 71 , input location shaft 60, main carrier 6 1 , carrier gear 62, external/sun carrier 63, external/sun carrier 64, input main journal 65, output main journal 66 and outer casing 67.

The axles 53, 54 define respective axes of rotation Z-Z and Y-Y which are spaced at 50.00mm and 35.00mm respectively from the axis X-X.

Input gear 70 has 50 teeth, carrier gear 71 has 50 teeth, reference planet gear 72 has 23 teeth, static gear 73 has 47 teeth, output external planet gear 74 has 50 teeth, differential reference external planet gear 75 has 60 teeth, reference sun 76 has 40 teeth, compound reference sun 77 has 35 teeth, compound reference planet 78 has 35 teeth, differential sun gear 79 has 15 teeth, differential sun gear 80 has 17 teeth, differential planet gears 81 and 82 have 1 1 teeth each and differential planet gears 83 and 84 have 13 teeth each. The external gears and sun gears shown in the differential unit design are themselves orbited by their own planets and could be described as moon gears.

The embodiment of Figure 4 includes double differentia! units and double reference planetary pairs and, like the Figure 3 embodiment, it uses a stepping of the reference section of the device rather than the stepping of the output section as used in Figures 1 and 2. These variations indicate either a diametric reduction or increase in respect of the off-set.

In Figure 4, the off-set ratio is 50 - 35 = 1.4285714 : 1 and the operating range or envelope is from 3.9464289 : 1 to 1 : 1 , and the change-up (overlap) torque is 0.1952083 T at IT input torque.

The differential units revolve 3.76 times in order to produce 1 : 1 ratio from input to output with the main carrier 61 rotating rearward 2.76 times.

The total forward loading of the main carrier is 4.0139757T and the total rearward loading of the main carrier (including the off-set ratio multiplication) is 4.2091841 T (assuming IT forward from the input shaft).

An example of alternative off-setting (all other ratios remaining the same) would be to increase the 50.00mm lever arm to 55.00mm whilst leaving the 35.00mm lever arm the same. This would produce an off-set ratio of 55 - 35 = 1.57 14286 : 1 , thereby increasing the rearward torque loading of the main carrier to 4.6301026T (an overlap factor of 0.6161269T instead of 0.1952083T). This is of coruse, simply one example and any offset ratio can be used, in conjunction with any desired gear ratio set, in order to effect any desired set of performance parameters.

The device of Figure 4 is fundamentally a 4 : 1 to 1 : 1 device with the ability to change gear according to load status.

Figure 5 illustrates a further variation of the fundamental principle of the present invention and includes an overdrive facility in the basic layout as drawn. However, any of the previously described embodiments can be modified to include this capability as has been explained.

In top-gear, i.e. with the main carrier rotating 2.4 revolutions in a rearward direction, opposite to that of the input shaft, the differential units (in this case, two epicyclic units of 3 : 1 capability), revolve rearwards 3.4 times. This produces a situation in which the input shaft is rotating forwards 1 'revolution and the output shaft is rotating forward 1 revolution and the output shaft is rotating forward 1.3578947 revolutions. In this case, an input to output ratio of 1 : 1.3578947 is achieved.

This overdrive capability does, of course, reduce the overall torque capability of the bottom-gear slightly, this being 3.6 190476 : 1. This bottom-gear ratio would have been 4T, and not 3.6190476T, had the ratio between the output planet and the output gear been 1 : 1 and not, as drawn,

0.9047619 : 1.

The Figure 5 embodiment comprises input shaft 101 , output shaft 102, layshafts 103 and 104, sleeve shaft 105, sleeve shaft 106, epicyclic stub axles 107 and 108, centre-line concentric sleeve shaft 109, input shaft location axle 1 10, main carrier 1 1 1 , annulus carrier I 12, outer casing 1 13, planet gear carrier 1 14, input main journals 1 15 and output main journals 1 16. As this is a dif ferentially balanced device, rather than a counterbalance weight balanced device, there is a duplication of mechanisms as in Figure 4.

Input gear 1 17 has 40 teeth, input planet gear 1 18 has 40 teeth, epicyclic sun gear 1 19 has 16 teeth, epicyclic idler planet gears 120 and 12 1 have 16 teeth each, epicyclic annular gear 122 has 48 teeth, intermediate gear 123 has 48 teeth, reference sun gear 124 has 32 teeth, compound reference sun gear 125 has 25 teeth, reference planet gear 126 has 25 teeth, compound reference planet gear 127 has 16 teeth and static gear 128 has 36 teeth.

The basic operating principle of all of the transmission devices described herein is similar and the various examples of Figures 1 to 5 are included in order to demonstrate the flexibility of the application of the principle.

The main carrier of the device is, in each case, prevented from rotating in the forward direction by reason of the off-set characteristics of the outputs from the main carrier at the layshaft axes, and the rearward rotational behaviour is dictated by the overlap torque created by said offset ratio being prejudiced in favour of rearward rotation.

The main carrier cannot simply rotate rearward without the output gear being simultaneously driven forward, so that the status of the output gear will always dictate the rotational status of the main carrier.

The description has been based upon the introduction of IT (one unit of torque). However, any torque loading can be applied to the main input shaft.

The overlap torque diminishes in direct proportion with the increase in the velocity of the output shaft.

At the 1 : 1 ratio between input and output many of the gears of the device are stationary relative to one another. The greater the required overlap torque, the greater the change-up drive.

The greater the forward torque compensation required, the greater the disparity between layshaft centres so that any off-set ratio can be included for any given set of gear ratios: the fact that they are relatively larger or smaller does not affect the rotational ratios of the gear components.

Figure 6 illustrates a variation of the embodiment of Figure 1 , in that the output mechanism is changed to include a larger diameter idler gear 26, thereby providing a 1 : 1 ratio between the intermediate and output planet gears 25 and 28, 29 and also between the output planet gears and the output gear 7.

The idler gear is eccentrically located with regard to the axis X-X, but this is acceptable. Indeed, the idler gear could be replaced, for example, by a drive belt or chain with the intermediate gear and output being replaced by similar ratio sprockets or pul leys.

A transverse right-angle drive could also be used to extend the offset ratio centres.

Crank/ connecting-road and flexible drives can also be incorporated if desired.

The embodiment of Figure 6 has a total forward loading on the main carrier in the same direction as that of the input shaft of 1 .5454545T (assuming IT applied ai the input) and a total rearward loading of the main carrier (off-set included) of 1.6 190476T on the assumption that IT is introduced by the input and the inertial output loading does not exceed 2T.

This indicates an overlap torque (change-up drive) of 0.073593 1 T. This can of course be increased or decreased according to requirement simply by changing the off-set ratio.

The static gears in the varius embodiments described can be made variable by way of braking, locking or clutching mechanisms in order to effect engagement of the transmission device. However, no such clutching or braking mechanisms are actually required in order for the automatic, load-sensitive gear changing action of the device of the present invention. The transmission devices are self-regulating without need for external stimulation. However, should external control mechanism be considered desirable, then they may be added, if only for cosmetic reasons.

Throughout the foregoing description, the resultant "balances" have been calculated by multiplying the off-set ratio with a given torque, it is also understood that a similar result (different in subsequent figures) may be obtained by dividing a given torque by the off-set ratio, or by both methods.

Either way, the off-set ratio can be seen as effective.

No reference to frictϊonal losses has been included in the above description, as the transmission does not rely upon friction to make it work. However, frictional considerations are understood and the input to output figures will be modified if such losses are included in the calculations.

It is assumed throughout this specification, that in the examples shown, the main carrier etc., is balanced, and that frictional losses are also balanced. However, as can be demonstrated, the ability to change the level length and still retain the same gear ratios will enable rearward rotational emphasis tp be determined at will. If, for example, in Figure 1, the two main carrier lever centres were changed from 41.00mm and 63.50mm, to say, 41.00mm and 73.50mm, then the IT rearward loading of the main carrier would change from 1.5487805T (effective) to 1.7926829T (effective). In terms of overlap torque, i.e. change-up drive this would be an increase from 0.0033600T to 0.2472284T, an increase of 0.2438684T without any gear ratio change (all gear diameters (not tooth count) being increased in concert).

This Indicates that by simply changing the ratio of the relative radial positions of the layshaft centres (bearing centres) relative to the main axis X, both torque and/or frictional balancing can be accomplished, and the overlap torque quantified according to reqirement. The basic designed through ratios (2 : 1 / 1 : 1 envelope) remaining constant.

Claims

CLAIMS:

1. A transmission device comprising an input shaft rotatable about a main axis of rotation, an output shaft rotatable about the main axis, an input gear rotatable with the input shaft, a carrier body rotatable about the main axis around the input shaft and provided with first and second axles defining respective first and second subsidiary axes of rotation which are parallel to the main axis and at different distances from the main axis, a differential gear assembly mounted on the first axle and having a differential input gear driven by the input gear on the input shaft and first and second differential output gears, reference gear means coupling the first differential output gear to a stationary gear centred on the main axis, and output gear means coupling the second differential output gear to the output shaft, the reference gear means on the output gear means including a gear assembly mounted on the second axle.

2. A transmission device according to claim 1 , wherein the differential gear assembly comprises a differential carrier rotatable with the first axle which is rotatable supported by the carrier body for rotation about the first subsidiary axis, the differential input gear is an input bevel gear rotatably mounted on the first axle and engaged with idler bevel gears rotatably mounted on respective stub axles presented by the differential carrier, the first differential output gear is an output bevel gear rotatably mounted on the first axle and engaged with the idler bevel gears, and the second differential output gear is a gear mounted on the first axle for rotation with the first axle.

3. A transmission device according to claim 2, wherein the reference gear means comprises a reference gear rotatable with the output gear and engaging the stationary gear.

4. A transmission device according to claim 3, wherein the second differential output gear engages an output gear rotatable with the output shaft via output idler means and a pair of output planet gears rotatable together with the second axle which is rotatably supported by the carrier body, one of the output planet gears engaging the output idler gear and the other engoging the output gear on the output shaft.

5. A transmission device according to claim 4, wherein the output idler means is a gear rotatable about an axis parallel to the main axis on a stub axle presented by the carrier body.

6. A transmission device according to claim 4, wherein the output idler gear is mounted for rotation about the main axis.

7. A transmission device according to claim 2, wherein the reference gear means comprises a reference gear rotatable with the output bevel gear and engaging the stationary gear via a reference idler gear and a pair of reference planet gears rotatable together with the second axle which is rotatably supported by the carrier body, one of the reference planet gears engaging the reference idler gear and the other engaging the stationary gear.

8. A transmission device according to claim 7, wherein the reference idler gear is mounted for rotation about the main axis.

9. A transmission device according to claim 7 or 8, whereia the second differential output gear directly engages an output gear rotatable with the output shaft.

10. A transmission device according to any one of claims 2 to 9, wherein the idler bevel gears have a first bevel gear face which engages the input bevel gear and a second bevel gear face which engages the output bevel gear and is further from the first subsidiary axis than the first bevel gear face.

11. A transmission device according to claim 1 , wherein the first axle is non-rotatably supported by the carrier body, the differential input gear comprises a planet carrier rotatably mounted on the first axle and carrying a planet layshaft which is rotatable about an axis parallel to the first subsidiary axis and which has first and second differential planet gears rotatable therewith, and the first and second differential output gears are rotatably mounted on the first axle with the planet carrier therebetween and each comprise an internal differential sun gear face engaged with the respective differential planel gear and an external gear face.

12. A transmission device according to claim 1 1 , wherein the reference gear means comprises a first reference sun gear rotatable about the main axis and engaged with the external gear face of the first differential output gear, a second reference sun gear rotatable with the first reference sun gear, and a pair of reference planet gears rotatable together around the second axle which is non-rotatabiy supported by the carrier body, one of the reference planet gears engaging the second reference sun gear and the other engaging the stationary gear.

13. A transmission device according to claim 12, wherein the external gear face of the second differential output gear is directly engaged with an output gear rotatable with the output shaft.

14. A transmission device according to claim 1 , wherein the differential gear assembly comprises a planet carrier which is rotatable with the first axle about the first subsidiary axis and presents a fixed stub axle on which a differential planet gear is rotatable, the differential input gear comprises a differential sun gear rotatable on the first axle and engaged with the differential planet gear, the first differential output gear is a differential annular gear carrier rotatable on the first axle and engaged with the differential planet gear, and the second differential output gear is a gear mounted on the first axle for rotation with the first axle.

15. A transmission device according to claim 14, wherein the reference gear means comprises an intermediate gear rotatable with the first differential output gear, a first reference sun gear rotatable about the main axis and engaged with the intermediate gear, a second reference sun gear rotatable with the first reference sun gear, and a pair of reference planet gears rotatable together with the second axle which is rotatably supported by the carrier body, one of the reference planet gears engaging the second reference sun gear and the other engaging the stationary gear.

16. A transmission device according to claim 1 5, wherein the second differential output gear directly engages an output gear rotatable with the output shaft.

17. A transmission device according to claim 4, wherein the idler gear is rotatable about an axis parallel to but eccentrically located relative to the main axis.

18. A transmission device according to any preceding claim, wherein the carrier body comprises counterbalancing means for rotataionally balancing the differential gear assembly mounted on the first axle.

19. A transmission device according to claim 18, wherein the counterbalancing means comprises a counterbalance weight carried by the carrier body.

20. A transmission device according to claim 18, wherein the counterbalancing means comprises a duplicate differential gear assembly.

21. A transmission device substantially as hereinbefore described with reference to any one of Figures 1 to 6 of the accompanying drawings.

22. Any novel feature or combination of features described herein.