WO1987005085A1 - Self-adjusting transmissions - Google Patents

Abstract

A transmission device comprises an input shaft (1) and an output shaft (2) rotatable about a main axis (X-X). An intermediate shaft (6) is interposed between the input and output shafts and a main carrier (9) is rotatable about the main axis in a stationary casing (2) of the device. A differential gear assembly (10-12, 20-23) provides a first output to reference planet gears (25, 26) mounted on the carrier (9) for rotation about a first subsidiary axis (Y-Y), planet gear (26) being engaged with a stationary gear (27). The differential assembly provides a second output via a sun gear (28) rotatable with differential carrier (10) to output planet gears (29-31) which drive the output shaft and are carried by a carrier gear (33) which drives further planet gears (38, 39) carried by the carrier (9) for rotation about a second subsidiary axis (Z-Z), gear (39) being engaged with a second stationary gear (40).

Description

Des ription of Invention

"Self-adjusting transmission*: "

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

It is an object of the invention to provide a steplessly variable, load- sensitive transmission device and, to this end, there is provided a transmission device comprising an input shaft rotatable about a main axis of rotation, an output shaft, an input gear rotatable with the input shaft, an intermediate rotary assembly mounted for rotation about the main axis and engaged at respective first and second points of engagement at different radial distances from the main axis by first and second gear assemblies driven by the input gear, the first gear assembly transmitting the input torque applied to the input gear to the intermediate assembly in one rotational direction and the second gear assembly transmitting the input torque to the rotatable assembly in the other rotational direction and being coupled to transmit torque to the output shaft.

A transmission device embodying the present invention is suitable for use in any situation where the load is of a variable nature or where a load is required to be accelerated to a constant speed from an initial stationary condition. The device is also of use where the power input of a power source during the load is variable or in situations where both the power input and load are variable, either independently of one another or in concert.

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 I is an axial cross-section through a first transmission device embodying the present invention; Figure 2 is art oxiol cross-section through σ 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 oxial cross-section through a fourth transmission device embodying the invention;

Figure 5 is an oxial cross-section through a fifth transmission device

10 embodying the invention;

Figure 6 is an axial cross-section through a sixth transmission device embodying the invention; and

■ 5 Figure 7 is a diagram for use in explaining in more detail the principle of operation of the sixth embodiment.

Referring to Figure I , a first transmission device embodying the invention comprises on input shaft I extending into a casing 2 through an o input journal 3 of the casing and an output shaft 4 extending from casing 2 through an output journal 5. A free-running intermediate shaft 6 is located between the input and output shafts I and 4 by means of locating shaft portions 7 and 8 received in corresponding locating bores in the shafts I and 4. A main carrier 9 is borne by the journals 3 and 5 and the intermediate 5 shaft 6 so as to be independently rotatable about the same main axis X-X as the shafts 1 , 4 and 6.

A differential unit of the device comprises a differential carrier 10 which is fixed to or port of the intermediate shaft 6 and presents stub axles 0 1 1 and 12 which are fixed to or part of the carrier 10. First and second layshaft axles 13 and 14 are fixed to or part of the main carrier 9 and support free-running sleeve layshafts 15 and 16 for rotation about respective first and second subsidiary axes Y-Y and Z-Z.

5 A differential input bevel gear 20 having 44 teeth is fixed to or part of input shaft I and engages idler bevel gears 21 and 22 respectively rotatable on stub axles 1 1 and 12 and having 44 teeth each. A differential output bevel gear 23 having 44 teeth engages gears 21 and 22 and is rotαtαble with α sleeve shaft bearing located on the intermediate shaft 6 and carrying a dif ferential output gear 24 having 47 teeth. Engaged with gear 2k is a planet gf-ar 25 having 47 teeth and rotatable with sleeve shaft 15 which also carries compound planet gear 26 having 36 teeth. Gear 26 is engaged with a static gear 27 having 58 teeth non-rotatably supported by input journal 3 concentrically of the input shaft I .

Sun gear 28 having 28 teeth is fixed to or part of intermediate shaft 6 and engages planet gears 29 and 30 having 28 teeth each and rotatable with respective planet layshσfts 31 and 32 rotatably carried by a carrier gear 33 having an external gear face 34 provided with 123 teeth. The gear 33 could be a single component as shown or two separate components coupled together for rotation together. Compound planet gears 35 and 36 having 33 teeth each are also rotatable with the layshafts 31 and 32 respectively and engage an output sun gear 37 which is fixed to or part of output shaft 2 and has 23 teeth. The gear face 34 of carrier gear 33 is engaged with a planet gear 38 having 1 23 teeth and rotatable with sleeve shaft 16 which also carries a compound, planet gear 39 having 54 teeth and engaged with a second static gear 40 having 87 teeth and non-rotatably carried by the output journal 5 concentrically of output shaft 2.

The principles of operation of the transmission device of Figure I are as follows.

If one unit of torque (IT) is applied to the device via the input shaft I and input gear 20 in a forward direction and the output shaft 2 is considered to be loaded, then the differential unit will cause bevel gear 23 to be driven in a backward direction (a direction opposite to that of gear 20) with IT capability. The differential carrier 10 will also be driven in the forward direction with a capability of 2T. This is standard differential behaviour.

As gear 23 is compounded with gear 24, the latter gear is also provided with a IT rearward torque loading and this is transmitted to the gear 25 with which gear 24 is engaged. Gear 25 transmits this torque to gear 26 which is connected to gear 25 by sleeve shaft 15. However, as gear

26 is engaged with static gear 27 at a ratio of 58 - 36 = 1.61 1 1 1 1 I : I , there is 1.6 1 I I 1 1 I T applied to the static gear. This means that the IT input -μ.

torque must be subtracted from the 1.61 I I I N T in order to establish the forward loading applied to the main carrier 9 at axis Y-Y, i.e. 1.61 1 1 1 I I - I = 0.61 1 1 1 1 I T. This means that for every IT of torque introduced by the input shaft I , 0.61 I I I N T is applied to the main carrier 9 in the same rotational direction at axis Y-Y.

Meanwhile, and at the same time, the 2T torque output from the differential carrier 10 is transmitted along the intermediate shaft 6 to the sun gear 28, thereby providing the sun gear 28 with 2T of torque in the forward direction.

The torque of 2T introduced to the output unit is in the forward direction and therefore, as sun gear 28 is engaged with planet gears 29 and 30 at a ratio of I : I , the 2T is applied to gears 29 and 30. However, the planets 29 and 30 are of course caused to rotate in an opposite direction to gear 28 at 2T and transmit this rotation to compound planet gears 35 and 36 via layshaf ts 31 and 32. As gears 35 and 36 have 33 teeth each and output gear 37 has 23 teeth, the ratio at which the 2T is transferred to gear 37 is 23 - 33, i.e. 0.6969697 : I . Therefore, the 2T from provides only I .393994T in the forward direction for the output gear 37, the difference between this and 2T, namely 0.6060606T, being applied to the carrier gear 33 in the forward direction.

If carrier gear 33 were unrestrained, then gears 35 and 36 would simply progress or walk forwardly around the loaded output gear 37.

However, gear 33 is engaged at a ratio of I : I with planet gear 38, so that the 0.60606 06T is transferred to gear 38 causing it to rotate in the rearward direction.

Planet gear 38 is fixed to or part of planet gear 39 which is engaged with the second static gear 40 at a ratio of 1.61 1 1 1 1 1 : 1 (similar ratio to the engagement of the planet gear 26 and static gear 27). Therefore, the 0.6060606T is multiplied by 1.61 1 1 1 1 1 to give a loading of 0.9764309T upon the second static gear 40. From this figure must be subtracted the initial 0.60606606T in order to establish the rearward loading upon the main carrier at the axis Z-Z (0.9764309T - 0.6060606T = 0.3703703T). This is then multiplied by the off-set ratio of 1.7446809 between the axis Y-Y and Z-Z in ordcr to see if it is capable of counteracting the 0.61 1 1 M I T of forward drive applied to the main carrier via the layshaft axle 13, giving 1.7446809 x O.3703703T = 0.64 178T, which is greater than 0.6 M I M I T by 0.0350669T). Therefore, not only does the torque applied to carrier 9 at axle \ k counterbalance the forward loading of the carrier 9 but it applies to the carrier a torque biasing the carrier towards secured rotation, thereby establishing a change-up drive factor capable of ensuring a main carrier rotational status responsive to the change in speed and acceleration (and/or deceleration) of the output shaft 2.

The main carrier 9 will, therefore, rotate 1.6363436 revolutions rearward in order for gears 23 and 24, together with carrier gear 33 to be driven forward one revolution. This represents a I : I input to output ratio.

If gears 23, 2k and 33 are rotating in unison in the forward direction, then there will be no relative rotation of the differential bevelled gears, or any relative motion between gears 28, 29, 30, 35, 36, 37. The only gears running in this situation will be the compound planets. Therefore, the output shaft and input shaft I must be rotating in unison. This gives to operating envelope of device Figure I as being 1 .3939394 : I / I : I .

The static gears 27 and 40 may be variable between the non- rotatable, fixed condition and a rotatable condition by the interposition of a clutch mechanism or the like between the gear and the stationary casing 3.

The device of Figure I is therefore, a simple version of a fully variable, constant-mesh transmission made possible by way of the off-set torque compensation applied to the main carrier 9. The explanation indicates something of the possibilities of such a device, and the question as to what enables the transmission to know when to change gear is shown to be a function of the overlap torque working in favour of the rearward rotation of the main carrier 9. This rearward rotational bias cannot simply allow the main carrier to rotate backwards producing negative output drive, as the ratios dictate that for every 1.6363636 revolutions of the main carrier in the rearward direction, the output gear 37 must rotate forwards one revolution.

The overlap torque of 0.0350669T always tries to cause rearward rotαtion of the main carrier but is unable to move the main carrier 9 without forward rotation of the output gear 37. A soon σs gear 37 starts to rotate, the main carrier 9 will immediately begin to creep slowly rearwards and, σs it does so, the overall ratio from input to output will start to change as the rearward rotation of the main carrier will cause gears 25 and 26 to ensure that gear 2k is given forward rotational movement. This will change the speed differences across the differential unit and, as a result, change the speed of the through-gearing. This will be accomplished in direct relationship with the change in status of the output-gear 37 from stationary to rotating and the faster the output gear moves, the more the rearward rotations of the main carrier 9 will match the forward speed of the output gear until the^re are 1 .6363636 rearwards revolutions of the main carrier 9, for one forward revolution of both the input and output shafts is reached and the I : I input/output ratio is achieved.

The overlap torque on the main carrier 9 will diminish in direct relationship with the increase in rotational speed of the main carrier 9 and output gear 37. Indeed, at the I : I input/output ratio, there will be no overlap torque present. However, any slowing of the output will immediately cause overlap torque to be established and rearward emphasis to be applied to the main carrier. However, this will not be eroded if the output gear 37 is unable to accelerate, and if a disparity between input and output revolutions exists, then until the output can accelerate, the status quo will be maintained. A degree of torque increase will be present until the 1.6363636 rearward revolutions of the main carrier matches the I forward revolution of the input and output shafts.

The rearward overlap torque on the carrier 9 can be increased without a change in gear ratio being involved by retaining similar ratios but by increasing the length of the relevant lever arm, i.e. the distance between axes X-X and Z-Z. If this distance is increased to, say, 90mm, the off-set ratio will increase from 1.7446809 : I so that the 0.6060606T output from the planet gear 39 will again be multiplied by the 1.61 1 1 1 I I ratio between planet gear 39 and static gear 40 to give the resultant 0.9764309T, from which is subtracted the input value of 0.6060606T to give 0.3703703T as the rearward torque bias on the applied to the carrier 9 at axis Z-Z. This will, however, now be multiplied by the increased off-set value working in favour of axis Z-Z, i.e. 0.3703703T will be multiplied by the increased off-set ratio of 1.9148936 and a totσl rearward capability of 0.70921 8T will emerge. Therefore, an overlap torque can be selected to conform with any desired specification, regardless of the ratios of the gears used in the device.

Figure 2 illustrates a second embodiment of the invention which is a variation of the layout depicted in Figure I . However, in Figure 2, the differential unit is placed within the output section of the transmission, and the input section includes an epicyclic/annular gear assembly capable of generating a k : I ratio.

Referring to Figure 2, a second transmission device embodying the invention comprises an input shaft 101 extending into a casing 102 through an input journal 103 of the casing and an output shaft 104 extending from the casing 102 through an output signal 105. A free-running intermediate shaft 106 is located between the input and output shafts 101 and 104 by means of locating shaft portions 107 and 108 received in corresponding locating bores in the shafts 101 and 104.

A main carrier 109 is borne by the journals 103 and 105 and the intermediate shaft 106 so as to be independently rotatable about the same main axis X-X σs the shafts 101 , 104 and 106.

A differential unit of the device comprises a differential carrier 1 10 which is fixed to or part of the intermediate shaft 106 and presents stub axles I 1 1 and 1 12 which are fixed to or part of the carrier 1 10 first and second layshaft axles 1 13 and I 14 are fixed to or part of the main carrier 109 and support free-running sleeve layshafts I 15 and 1 16 for rotation about first and second subsidiary axes Y-Y and Z-Z.

An input gear 1 17 having 28 teeth is fixed to or part of input shaft 101 and engages epicyclic planet gears 1 18 and 1 19 having 42 teeth each and mounted on respective stub axles 120 and 121 which are fixed to or part of an epicyclic planet gear carrier 122 which is rotatable with intermediate shaft 106. The planet gears 1 18 and 1 1 are engaged with an annular gear

1 23 having 1 12 teeth and rotatably mounted on intermediate shaft 106 by annular gear carrier 123a. Epicyclic output sun gear 124 has 47 teeth and is rotatable with annular gear 123 to which it is connected by sleeve shaft 125. Sun gear 124 is engaged with a reference planet gear 126 having 47 teeth and rotatable with sleeve shaft 1 15 on the first axle 1 13. Compound reference planet gear 127 has 36 teeth is rotatable with sahft 1 15 and engages a static gear 128 having 58 teeth and non-rotatably supported by casing 102 concentrically of main axis X-X.

A sun gear 129 having 123 teeth is mounted concentrically and rotatably on intermediate shaft 106 and is coupled for rotation with differential bevel gear 130 having 44 teeth and engaged with idler bevel gears 131 and 132 having 44 teeth and rotatably mounted on shaft axles I I I and ! 12. The idler bevel gears 131 and 1 32 are also engaged with an output bevel gear 133 having 44 teeth and fixed to or part of output shaft 104.

Gear 129 is engaged with planet gear 134 having 123 teeth and rotatable with sleeve shaft I 16 on the second axle 1 14. A compound planet gear 135 having 54 teeth is also rotatable with sahft 1 16 and engages a second static gear 136 having 87 teeth and non-rotatably carried by output journal 105 on a mounting flange 137.

Subsidiary axes Y-Y and Z-Z are radially spaced by 47.00mm and 82.00mm respectively from the main axis X-X.

The transmission device of Figure 2 can produce a bottom gear ratio of 2.5 : I and a top gear ratio of I : I . The input epicyclic assembly produces (for each IT torque introduced) a torque of 4T upon the annular gear 123 in the opposite rotational direction to the torque applied to the input shaft 101 and a torque of 5T upon the epicyclic planet carrier 122 in the same direction as that on the input shaft 101.

The 5T torque is fed along the intermediate shaft 106 to the differential carrier 1 10 and the free-running idler bevel gears 131 and 132 by way of the stub-axles I I 1 and I 12, applying 2.50T to gear 133 and 2.50T to gear 130. In both cases the application of torque is in the same rotational direction as that of input shaft 101. The 2. SOT applied to gear 133 is, of course, the 2.50T output drive, as gear 133 is rotatable with output shaft 104. However, the 2.50T applied to gear 130 is the torque factor responsible for ultimately driving the main carrier 109 rearwards and it is applied to gear 129 from gear 130.

The 2.50T present upon gear 129 is applied to gear 134 causing it to rotate in a direction opposite to the input shaft 101. However, as gear 134 is compounded to gear 135, which is engaged with the static gear 136, the resultant engagements cause the main carrier 109 to be loaded in a direction opposite to the input shaft 101. The 2.50T is multiplied by the engagement ratio between gear 135 and gear 1 36 i.e. 87 — 54 = 1.61 1 1 1 1 1 , producing 4.0277778T upon the static gear 136. Therefore, 4.0277778T minus the 2.50T introduced gives I .5277778T applied to the main carrier 109 at the axis Z-Z by the bearings of the sleeve shaft 1 1 .

Meanwhile, the 4T applied to the annular gear 123 is transferred along the sleeve shaft 125 to sun gear 124. This in turn, transfers the 4T to planet gear 126. Gear 126 being coupled to gear 127, causes the 4T to be multiplied by I .6 I I 1 I I 1 , as a result of the engagement ratio (58 -■} 36) between gears 127 and 128. Therefore, we have a forward torque loading upon the main carrier 109 of 1.6 1 I I I I I x 4T - 4T = 2.4444444T.

The 2.4444444T is applied to the main carrier 109 at the subsidiary axis Y-Y which is the centre of rotation of the sleeve shaft I 15.

The I .52777778T present at subsidiary axis Z-Z of axle 1 16 is multiplied by the off-set ratio of 1 ,7446809 between the two axes Z-Z and Y-Y, giving a total rearward torque of 2,6654847T, this being 0.22 I 0403T greater than the total forward loading applied to the main carrier 109 by way of axle I 15. The required overlap torque and change-up drive is therefore established.

If the input epicyclic gearing were a 6 : 1 arrangement instead of a 4 :

I arrangement, then 6T would have been provided at gear 124, resulting in _a forward loading upon the main carrier 109 at the axis Y-Y of 3.6666666T and, at the same time, 7T output from the planet gear 122 would be transmitted to the axis Z-Z responsible for the rearward loading of the main -10-

carrier 109. This would give 7Tf2 = 3.5T x 1.61 1 M 1 1 = 5.6388889 - 3.5 = 2. I 388889T x 1.7446809 = 3.73 I 6785T of rearward torque and 3.50T output drive torque to the output shaft 104.

The rearward drive to the main carrier 109 is thus 0.06501 1 T greater than the forward drive (3.73 I 6785T - 3.6666666T). However, if this rate of change drive factor is required to be greater, then by simply increasing the off-set ratio of axes Z-Z and Y-Y to say 47/90, instead of the 47/82 indicated, the overlap torque would be increased to 90 - 47 = 1. 148936 : I x 2.1388889T = 4 095744T - 3.6666666T = 0.4290781T.

This example indicates the flexibility of the invention.

Figure 3 illustrates a third transmission device which employs an output section similar to that of the Figure 1 embodiment and an input section similar to that of Figure 2.

In Figure 3, an input shaft 201 extends into casing 202 through input journal 203 of the casing and an output shaft 204 extends from casing 202 through output journal 205. A free-running intermediate shaft 206 is located between the input and output shafts 201 and 204 by means of locating shaft portions 207 and 208 received in corresponding locating bores in the shafts 201 and 204. A main carrier 209 is borne by the journals 203 and 205 and the intermediate shaft 206 so as to be rotatable about the same axis X-X as the shafts 201 , 204 and 206. First and second Iayshaft axles 213 and 214 are fixed to or part of the main carrier 209 and support free-running sleeve layshafts 215 and 216 for rotation about respective first and second subsidiary axes Y-Y and Z-Z.

Aninput gear 217 having 18 teeth is fixed to or part of input shaft 201 and engages epicyclic planet gears 218 and 219 having 27 teeth each and mounted on respective stub axles 220 and 221 which are fixed to or part of an epicyclic planet gear carrier 222 which is rotatable with intermediate shaft 206. The planet gears 218 and 219 are engaged with annular gear 123 having 72 teeth and rotatably mounted on intermediate shaft 106 by annular gear carrier 223a. Epicyclic output sun gear 224 has 47 teeth and is rotatable with annular gear 123 to which it is connected by sleeve shaft 225. Sun gear 224 is engaged with reference planet gear 226 having 47 teeth and rotatable with sleeve shaft 215 on the first axle 213. Compound reference planet gear 227 has 36 teeth, is rotatable with shaft 215 and engages static gear 228 having 58 teeth and non-rotatably supported by casing 202 concentrically of main axis X-X.

A sun gear 229 having 32 teeth is fixed to or part of input shaft 201 and engages planet gears 230 and 231 having 24 teeth each and rotatable with respective planet layshafts 232 and 233 rotatably carried by a carrier gear 234 having an external gear face 235 provided with 123 teeth.

Compound planet gears 236 and 237 having 30 teeth each are also rotatable with layshafts 232 and 233 respectively and engage an output sun gear 238 which is fixed to or part of output shaft 204 and has 26 teeth. The gear face

235 of carrier gear 234 is engaged with a planet gear 239 having 123 teeth and rotatable with sleeve shaft 21 6 which also carries a compound planet gear 240 having 54 teeth and engaged with a second static gear 24 1 having

87 teeth.

Subsidiary axes Y-Y and Z-Z are radially spaced by 32.25mm and 82.00mm respectively from main axis X-X, giving an off-set ratio of 2.326241 1 : I .

The operating principles of the Figure 3 embodiment are exactly similar to those previously described in conjunction with Figures I and 2. However, the ratios included have been changed in order to demonstrate further variation.

If an input of IT in a forward direction is applied, then gear 229 wil l be provided with 5T also in the forward direction, with gear 224 being provided with 4T in the reverse direction.

Gear 224, via planet gears 226 and 227, applies 4T (58 36) = 6.4444444T to static gear 228. However, 6.4444444T - 4T = 2.4444444T forward drive is applied to the main carrier 209 at axis Y-Y via the bearings between sleeve shaft 215 and axle 21 3. Meαnwhile, the 5T forword loading of gear 229 is applied to the engaged planets 230 and 231 at a ratio of 24 » 32 providing layshaft 232 and 233 with 3.75T. This is then applied to gears 236 and 237, which gears drive gear 238 in a forward direction with 26 - ' 30 x 3.75T = 3.25T.

5

The original 5T forward torque is now subjected to a subtraction of the output torque generated via the planet gears, so that 5T - 3.75T is the forward output from carrier gear 234.

I Q Carrier gear 234 is engaged with planet gear 239 at a ratio of I : I thereby driving planet gear 240 with I .75T. This results in I .75T x (87 • 54) = 2.8 I 44444T being applied to static gear 241. This means that 2.8 I 4444T - 1.75 = I .0694444T is applied to the main carrier 209 in a rearward direction.

The 2.4444444T forward loading of the main carrier 209 is now met with only 1.0694444T of rearward loading. However, in order that the main carrier should not simply run forward, producing negative output on gear 238, the off set ratio of 2.326241 I : I between axes Y-Y and Z-Z has been included. The layshaft combination 21 3, 215 is only 35.25mm radially out 0 from the centre rotation of the main carrier 209, while the combination 2 14, 216 is 82.00mm radially out from the rotating centre of the main carrier 209. Therefore, multiplying the I .0694444T by the off -set ratio 2.326241 I gives 2.4877855T effective rearward drive to the main carrier 209. By subtraction the 2.4444444T of forward drive from the 2.4877855T rearward 5 drive, a 0.04334 1 I T overlap of rearward change-up drive torque becomes apparent. Therefore, the main carrier is not only torque-balanced, but has a rearward rotational prejudice.

Throughout the specification, the torque applied to the main carrier 0 at axis Z-Z is always multiplied by the off-set ratio in order to discover the overlap. However, this is for convenience only and even if the torque present at axis Y-Y is divided by the off-set ratio, the overlap always comes out in favour of the effective torque being in favour of axis Z-Z. 5

Figure 4 is an all-spur, fully variable constant mesh transmission device which includes a single rotating carrier mechanism. -13-

In Figure 4, input shaft 301 extends into casing 302 through input journal 303 and output shaft 304 extends from the casing 302 through output journal 305, the shafts being coaxially located by locating shaft portion 306 of output shaft 304 rotatably engaging in a corresponding locating bore of

⁵ input shaft 301.

A main carrier 309 is borne by the casing 302 and shaft 304 for rotation about the same main axis X-X as the shafts 301 and 304. A pair of first layshafts 310, 31 1 are rotatably mounted on main carrier 309 for

' 0 rotation about respective first subsidiary axes Z-Z equally spaced from the main axis X-X and a pair of second layshafts 312, 313 are rotatably supported by the carrier 309 for rotation about respective subsidiary axes Y-

Y equally spaced from the main axis X-X but at a smaller distance than the axes Z-Z. A pair of third layshafts 314, 31 5 are rotatably supported by the

' -> casing 302 for rotation about respective third subsidiary axes equal ly sapced from the main axis X-X and lying intermediate the axes Z-Z and Y-Y. The output shaft 304 carries first and second adjacent, independently rotatable inner sleeve shafts 316 and 317 and an outer sleeve shaft 318 rotatably mounted on the first inner sleeve shaft 31 . 0

An input gear 320 having 75 teeth is fixed to or part of input shaft 301 and engages planet gears 321 , 322 having 75 teeth each respectively rotatable with layshafts 310, 31 1 which also carry respective compound planet gears 323, 324 having 60 teeth. Gears 323, 324 engage sun gear 325 5 which is fixed to or part of sleeve shaft 318 and has 90 teeth. Compound sun gear 326 is connected to sun gear 325 by sleeve shaft 318, has 60 teeth and is engaged with planet gears 327, 328 having 60 teeth and rotatable with layshafts 314, 31 5 respectively whcih also carry respective compound planets gears 329, 330 having 30 teeth. Sun gear 331 is fixed to or part of 0 sleeve shaft 317 and is engaged with compound planet gears 329, 330.

Sleeve shaft 317 connects gear 331 to sun gear 332 which has 36 teeth and is engaged with planet gears 333, 334 carried by layshafts 312, 313 respectively and having 24 teeth. Layshafts 312, 313 connect gears 333, 334 to respective compound planet gears 335, 336 which have 30 teeth and are engaged with output gear 337 which is fixed to or part of the output shaft

304. The axes Z-Z are spaced at 75.00mm from axis X-X and the axes Y-Y at 25.00mm.

If a torque of IT is introduced via shaft 301 , then gears 320, 321 , 323 will load gear 325, in the same rotational direction as the input shaft with

I .5T, and at the same time, rearwardly load the main carrier 309 with 0.5T.

The I .5T is transmitted along sleeve shaft 318 to gear 326, this being engaged at a ratio of I : I with gear 327, so that the I .5T is then transmitted along the layshoft 314 directly to gear 329. Gear 329 is engaged at a ratio of 3 : I with gear 331. Therefore, gear 331 is forward loaded in the same direction as the input shaft with 4.5T. This 4.5T is now transmitted along sleeve shaft 1 17 to gear 332, thereby loading this gear the forward direction with the 4.5T.

Gear 332 is engaged with the output gear 337 by way of the compound planet gears 334, 336, and their ratios cause the 4.5T to be reduced (24 - 36 x I : I = 0.6666666 : I ) to 3T which is applied to the output gear 337 in the same rotational direction as input gear 320. This gives a bottom gear ratio of 3 : I .

Meanwhile, in order that 3T be applied to gear 337, the 4.5T has been divided into two parts: 3T applied directly to the output and I .5T to the main carrier 309 in the same direction as input shaft 301.

This means that layshaft 310 is applying 0.5T rearwards torque drive to carrier 309 and, at the same time, layshaft 313 is applying 1.5T in a forward direction to main carrier 309.

This imbalance is compensated by the off-set ratio of 3 : I in favour of the axis Z-Z operating radius (the 75.00mm off-sert of axis Z-Z from axis X-X as compared with the 25.00mm off-set of axis Y-Y).

Therefore, the 0.5T multiplied by the 3 : I off-set advantage, gives I .5T effective loading. This is a perfect balance with the I .5T (forward) as applied to the main carrier 309 by layshaft 313. If an overlap torque in a rearward direction is required, as in previous examples, then for example by extending the 75.00mm to 80.00mm, there would be on overlap in favour of rearward main carrier 309 rotation, of 0.IT.

This would, of course, require the diameters of the various gears to be adjusted, but not their relative ratios.

Figure 4 is, therefore, an embodiment of the invention which uses a single carrier device, shared by the two contradictory torques.

Figure 5 illustrates a fifth embodiment using the same off-set techniques, but in conjunction with fixed (case-held) annular gears. The overlap in this instance is in favour of forward rotation of the main carrier, rather than rearward rotation as previously described.

Referring to Figure 5, input shaft 401 extends into casing 402 through input journal 403 and output shaft 404 extends from the casing through output journal 405. First and second intermediate shafts 406 and 407 are independently rotatably located between input and output shafts 401 and 404 by means of locating studs 408, 409 and 410 received in corresponding locating bores in the input shaft, second intermediate shaft and output shaft.

Two first layshafts 41 I and 41 2 are rotatably carried by casing 402 for rotation about respective subsidiary axes Z-Z equally spaced from the main axis X-X of rotation of the input and output shafts. A pair of second layshafts 413 and 414 are carried by a carrier gear 415 rotatable on second intermediate shaft 407, the layshafts 413, 414 defining respective second subsidiary axes Y-Y. A main carrier 416 rotatable on shaft 407 presents two stub axles 41 7, 418 which are fixed to or part of the carrier 416. An annular carrier 41 is rotatable with the first intermediate shaft 406.

An input gear 420 having 25 teeth is fixed to or part of input shaft

401 and engages planet gears 42 1 and 422 having 40 teeth each and carrying stub axles 423, 424 which are fixed to or part of the gears 42 1 , 422. Carrier gear 425 receives the stub axles 423, 424 and engages planet gear 426 rotatable with layshaft 41 I and having 40 teeth. Compound planet gear 427 -(6-

is also rotatable with layshaft 41 1 and has 40 teeth, gear 427 engaging sun gear 428 having 100 teeth and fixed to second intermediate shaft 407. Sun gear 429 has 30 teeth and is also fixed to or part of intermediate shaft 407. Sun gear 429 is engaged with planet gears 430, 431 having 26 teeth and rotatable with layshafts 413, 414. Compound planet gears 432, 433 are rotatable with layshafts 13, 14, have 30 teeth and engage output gear 434 having 26 teeth and fixed to or part of output shaft 404.

Planet gears 421 , 422 engage annular gear 435 having 105 teeth which >s fixed to or part of carrier 419 which is itself fixed to the first intermediate shaft 406.

Sun gear 436 has 24 teeth and is fixed to or part of intermediate shaft 406 and engages planet gears 437, 438 having 20 teeth and engaged with a fixed annular gear 439 having 64 teeth. Planet gears 437, 438 are carried by stub axles 440, 441 which are fixed to or part of a sun gear 442 having 100 teeth. Gear 442 is rotatable on the first intermediate shaft 406 and engages planet gear 443 having 40 teeth and rotatable with layshaft 412. Compound planet gear 444 also has 40 teeth and is rotatable with layshaft 412. Gear 444 is engaged with sun gear 445 which has 100 teeth and is rotatable with main carrier 416.

Planet gears 446, 447 rotatable about stub axles 417, 418 have 70 teeth and engage a second fixed annular gear 448 having 224 teeth. Gears 446, 447 engage carrier gear 415.

The axes Z-Z are spaced at 77.00mm from axis X-X and the axes Y-Y at 22.00mm.

Assuming that the output sahft is loaded by introducing a torque of IT in a forward direction to input shaft 401 , input gear 420 will cause carrier gear 425 to be driven forward with 5.2T and annular gear 435 to be driven rearwards with 4.2T.

The 5.2T will be passed on to the layshaft 41 I via the engagement between gears 425 and 426, this being at a ratio of 0.4: 1 , and subsequently from layshaft 41 1 via gear 427 to gear 428 at a ratio of 0.25 : I . Therefore, geor 428 will be driven forward σt 5.2T.

Meanwhile, annular gear 435 will cause 4.2T to be applied to sun gear

436 in a rearward direction opposite to that of the input shaft, so that gear

5 436 will cause the fixed annular gear 439 to be loaded with 64 - 24 x 4.2T x

2.66C66C7 : I = I I .2T, and carrier gear 442 with I I .2T plus 4.2T = I 5.4T in the rearward direction.

The I 5.4T will be passed via a 0.4 : 1 /0.25 : I series of engagements ' ^υ to eventually drive gear 445 rearwards with 15.4T.

The 5.2T as now applied to gear 428 is transmitted along shaft 407 to sun gear 429, causing the engagement between gears 429 and 430 to diminish the torque upon layshaft 41 3 to 26 - 30 x 5.2T = 4.5066667T. This sum is ' -> then subjected to another loss via the engagement between gear 432 and the output gear 434 (26 - 30) finally arriving at shaft 404 as 3.9057765T. This is the output drive capabi lity of bottom gear.

The difference between 3.9057775T and the initial 5.2T represents 0 the forward drive torque (by-pass torque) applied to sun carrier gear 415 which is driven forward at the same time as gear 434 with 1.2942225T.

As gear 41 5 is engaged with planet gears 446, 447 and via these gears with fixed annular gear 448, the I .2942225T in a forward direction loads the ⁵ annular gear 448 with 3.45 1 2600T (224 - 84 = 2.6666667 : 1 ) and the main carrier 416, in a forward direction with 4.7458425T (3.4512600T + I .2942225T). It wi ll be remembered that the main carrier 4 16 is already subjected to a rearward loading of I 5.4T and, therefore, the 4.7458425T forward drive must be now adjusted by way of multiplication of the off-set 0 ratio 4.7458425T x 3.5 : I . Therefore, the avai lable 4.7458425T becomes an effective 16.6 10449T as far as the I 5.4T of rearward loading is concerned, creating an overlap torque of 1.2 I 04488T in favour of forward rotation of the main carrier 41 .

5 it will be understood, therefore, that with gears 435 and 425 able to rotate forward in unison with the carrier gear 442 and main carrier 41 6 rotating forward in unison but at a slower rate (as a result of the reduction between sun and annulus), there is a I : I ratio between input shaft 401 and output shaft 404. This is indicative of the top-gear situation.

If braking were applied to layshaft 41 1 , a reverse gear drive could be provided without recourse to any other alteration.

This transmission can, therefore, provide fully variable, constant- mesh (load sensitive) forward drive with a range of I : I to 3.9057775 : I and reverse gear. Furthermore, by allowing one or both annular gears to be braked to case in a gradual fashion, a clutching action can be included.

Therefore, it is seen that this transmission offers exceptional versatility.

Figure 6 represents a departure from the basic principles, in that, the off-set loading of a carrier component is replaced by an off-set resulting from a disparity in related compound diameters; i.e. two gears of differing size and tooth-count.

This, in itself, may not seem very different from usual gear ratio characteristics, but the differences become apparent when full considerations is given to tooth loadings at specific pitch-circle diameters.

Referring to Figure 6, input shaft 501 extends into casing 502 through input journal 503 and output shaft 504 extends from casing 502 through output journal 505. An intermediate shaft 506 is located between the input and output shafts by locating shaft portions 507, 508 borne in corresponding locating bores of the input and output shafts.

An annular gear carrier 509 is fixed to or part of a first sleeve shaft 510 which is rotatably received on intermediate shaft 506 and lies adjacent a second sleeve shaft 507 also rotatable on shaft 506.

Input gear 520 has 26 teeth and is rotatable with input shaft 501 and engages idler gears 521 , 522 having 39 teeth and mounted on stub axles 523, 524 which are fixed to or part of a planet gear carrier 525. Gears 521 , 522 are engaged with annular gear 526 having 104 teeth and fixed to or part of carrier 509. Sleeve shaft 510 carries a sun gear 527 having 32 teeth and engaged with planet gear 528 having 20 teeth. Gear 528 is rotatable with lαyshαf t 529 which is borne by the casing 502 and carries a compound planet gear 530 having 16 teeth. Sleeve shaft 507 carries a sun gear 531 having 36 teeth and engaged with planet gear 530.

Sun gear 532 has 144 teeth, is rotatable with sleeve shaft 507 and engages planet gear 533 carried by layshaft 534 borne by the casing 502. Compound planet gear 535 is also rotatable with the layshaft 534 and has 80 teeth. Gear 535 is engaged with carrier gear 536 having 40 teeth and carrying layshafts 537 and 538. Gear 536 is rotatable on the intermediate shaft 506.

Sun gear 539 is fixed to or part of intermediate shaft 506 and has 36 teeth. Gear 539 is engaged with planet gears 540, 541 which are rotatable with layshafts 537, 538 respectively and have 36 teeth. Compound planet gears 542, 543 are rotatable with layshafts 537, 538 have 40 teeth and engage output gear 544 which is fixed to or part of output shaft 504.

If a torque of IT is applied to input shaft 501 in a forward direction and a load is applied to output shaft 504, then input gear 520 will cause, via its engagement with gears 521 and 522 torque of 4T to be applied to the annular gear 526, thereby also loading gear 527 in a rearward direction.

Gear 527, being engaged with gear 528 at a ratio of 20 - ' 32 = 625 : I . causes gear 528 and layshaft 529 to rotate in a similar direction to the input shaft 501 with 2.5T.

Gear 530 being fixed to laysahft 529 is also provided with this 2.5T, and also being engaged with compound sun gear 531 at a ratio of 36 16 =

2.25 : I , provides gear 531 with 2.5T x (36 - 16) = 5.625T in the rearward direction. Therefore, assembly 529,507 is loaded with 5.625T in a reverse direction.

Meanwhile, the 5T output from the input epicyclic unit delivers its torque (via the intermediate shaft 506) to sun gear 539, causing this gear to be forward loaded with 5T.

Gear 539 being engaged at I : I with planet gears 540, 541 causes 2 0

layshafts 537, 538 to be provided with 5T in reverse direction, thereby causing planet gears 542, 543 to be driven with 5T.

Planets 542, 543 engage output gear 544 at a ratio of 32 - 40 = 0.8 : I . Therefore, only 4T is delivered to the output shaft 504. This represents the bottom gear capability of the transmission.

The remaining, forward by-pass IT capability is, of course, driving the arner gear 536 in the same direction as the input, and as it is engaged with planet gear 536 in the same direction as the input, and as it is engaged with planet gear 535 at a ratio of 80 ^- 128 = 0.625 : I , the IT is diminished to 0.625 : I x IT = 0.625T at shaft 534, thereby causing rotation of compound planet 533 in a direction opposite to input shaft 501 with a capability of 0.625T.

Gear 533 is engaged, as shown, iwth compound sun gear 532 at a ratio of 144 - 64 = 2.25 : I . Therefore, the 0.625T must now be multiplied by this figure in order to establish the forward loading placed upon gear 532, i.e. I.40625T.

^■ As the point of engagement Z between gears 533 and 532 is 72.00mm from the main axis X-X of rotation of assembly 532, 507, 531 and the point of engagement Y between gears 530 and 531 is only 18.00mm radially out from the main axis then regardless of the ratio, which has already been taken into account, an off-set situation exists, i.e. I .40625T must be multiplied by the leverage advantage offered to it by the considerable disparity in radial engagement. Thus, the I .40625T multiplied by the off -set ratio of 4 : I becomes 5.625T effective at the point of engagement between gears 530 and 531 , and, it will be remembered, there is only 5.625T opposing this capability at the same precise point of engagement. Therefore the assembly 531 , 507, 532 must be held in a stationary position, providing output drive without slipping.

This perfectly balanced situation can, of course be altered by adjusting the indicated radial contact points. For example, by extending the

72.00mm engagement radially outwards by say, 6.00mm and retaining the same rotational ratios between all gears concerned, the off -set ratio will be increαsed to α new figure of 4.3333333 : I , thereby increasing the capability of the I .40625T to 6.09375T at the point of engagement between gears 530 and 531. The result is a rearward overlap torque in favour of a forward rotation of the compound assembly 531 , 507, 532. The overlap in this 5 example is (6.09375T - 5.625T)=0.46875T.

It will be seen from the indicated included ratios that an en masse forward rotation of the input epicyclic unit 520, 521 , 522, 526 will result in a I : I input to output ratio. 10

Throughout this specification, the overlap torque has been explained in terms related to the direct multiplication of one torque loading by the ratio of the off-set operating datums, i..e datum "Z" being the most mechanically advantageous application factor. It is realised that the various

' 5 loadings can be assimilated in various ways, and the figures quoted can be subjected to further modification. However, the basic contentions hold good, in that differing lever lengths supplied with unequal loadings can be used to compensate for imbalance and, if required, provide a small torque capability with more purchase than a much larger torque loading. 0

In the description of Figure 3 it is suggested that 2.444444T of forward drive is applied to the main carrier 209 at the point represented by datum "Y" by way of the bearings responsible for the isolation of sleeve shaft 21 5 from axle 213. It is also suggested that I .0694444T is applied to 5 the same main carrier in a rearward direction at the point represented by datum "Z", the bearings responsible for the isolation of layshaft 216 from axle 214 applying this rearward load.

In order to establish the balance, or indeed overlap between these 0 two torque components, the I .0694444T has simply been multiplied by the off -set ratio acting in favour of datum "Z", and the resultant I .0694444T x 2.326241 I = component of 2.444444T and, in the event, provide an overlap torque of 0.043341 I T capable of providing rearward drive emphasis to the main carrier. 5

This method has been used throughout the specification as means of demonstrating the basic principles of operation, and explanations of being σble to vary the extent of the lever disparity without altering the gear ratios, has been clearly indicated, however, an alternative method of indicating the overlap or balance capabilities of the off-set ratio would have been to include some indication of the direct effect upon both datums "Z" and "Y" by each in respect of the other. Thus, if, as already demonstrated with reference to Figure 3, the multiplication of I .0694444T by the off-set ratio of 2.326241 1 : I produces 2.4877855T at datum "Y", then it might be reasonable to suggest that the 2.4444444T, already opposing this loading should be divided by the same off-set ratio, i.e. 2.4444444T - 2.326241 1 providing 1.050813T at datum "Z", 1.050813T being 0.0186314T less than the

I .0694444T that can be demonstrated as being present at datum "Z".

This indicates that the 2.4444444T forward loading of datum "Y" produces, via the off -set ratio of I : 2.326241 1 a forward torque component of 1.050813T at datum "Z" and, at the same time, the rearward torque loading of datum "Z" of I .0694444T produces, via a reversal of this same of off-set ratio a rearward torque loading of datum "Y" of 2.4877855T. Therefore, if the two possibilities are assimilated into one overall condition, we find that 2.4444444T divided bythe off-set ratio is 1.050813T, i.e. 0.0186314T less than would be required to balance the I .0694444T present

(rearward) at datum "Z" and I .0694444T multiplied by the off-set ratio of 2.326241 I : I is 0.04334 1 I T more than the 2.4444444T required to balance the forwared torque component present at datum "Y".

These two resultant figures can be played off, one against the other.

If all other factors are negated, it is clear that the output from datum "Y" is incapable of balancing the output of datum "Z" by a shortfall of 0.0186314T, and the output of datum "Z" is capable of balancing the output of datum "Y" with an excess of 0.043341 I T. This would seem to indicate that the total of the shortfall and the excess, added together, might offer a total rearward drive emphasis to the main carrier of 0.06 1 725T. However, in any event, the indications are irrefutable, - in that, there will be a rearward drive capability provided for the main carrier, thereby providing a means of changing overall ratio from input to output from bottom to top gear.

In answer therefore, to the question "how does the transmission know when to change gear", the answer is simple; i.e. it does not know, it simply tries to change the whole time and is only prevented from doing so by the rotational status of the output gear/shaft combination; i.e. as this accelerates the overlap torque can begin to cause rearward rotation of the main carrier and thereby alter the whole input to output ratio envelope.

Gears are, in all respects, simply repetitive levers, and as such they must conform to the characteristics of levers. Therefore the various items shown in this specification are viewed in light of two basic factors, i.e. the diametric pitch of the point of contact (the establishment of the leverage one upon another) and the number of teeth contained peripherally, in relation to another so engaged (the establishment of the interrelated speeds). These two, quite separate factors are quite often, and quite wrongly, interpreted as being each the direct result of the other. However, as rotational speeds are often directly quantifiable in relation with resultant torques, the confusion is understandable. It is more clearly demonstrated however by mechanisms utilising more than one gear component, i.e. a standard differential, than by two engaged gear items. However, in the case of a corrected gear, i.e. one including a spurious number of teeth for a given diametric pitch, the fact that the speed relationship can, and probably has been, altered does not affect the direct leverage, one upon the other, if the diametric pitch, i.e. the point of immediate engagement, has not been altered in any way from that of an uncorrected gear of similar pitch-circle radius.

If, therefore, the exact contact radius remains constant, but the number of teeth is varied, then the related speed will change but not the leverage, i.e. torque transmitted.

If, even in a corrected-gear situation, the contact pitch-circle radius is altered in any way, then the leverage characteristics will also change, the fact that this change may be of a different order to that of an uncorrected gear of altered tooth count, helps to emphasis the basic contentions of this specification.

The fact that it is very difficult to run gears of related diameters with unrelated numbers of teeth, does give rise to hard and fast acceptance 2 4

of the belief that such matters are inviolate.

It is, however, only when the compounding of two gears is undertaken that such "laws" can be shown to be incomplete.

Figure 6 uses a compound gear section as a means of providing an off-set leverage situation without incurring contradictory tooth ratio problems. By giving the fixed layshaft datums widely differing centre distances from the main datum "X" (the rotational centre of layshaft 534 is 104.00mm radially out from datum "X" and the rotational centre of layshaft

529 is only 26.00mm radially out from datum "X") and then compensating for the difference by retaining similar rotational gear ratios regardless of huge differences in relative gear sizes, the feed to gear 531 from gear 530 is exactly the same as the feed to gear 532 from gear 533. Thus the exact point of engagement between 533 and 532 the applied torque is no different than that applied at the exact engagement point created between gear 530 and gear 531. If IT were applied at the engagement point between 533 and 532 at a ratio of 144 - 64 = 2.35 : I , the assembly 532, 507, 531 will receive 2.25T of rotational drive. Likewise, if IT were applied to gear 531 at the exact point of engagement between gear 530 and gear 531 at a ratio of 36 -

16 = 2.25 : I , again the assembly 532, 507, 531 will receive 2.25T. This would seem to indicate that the compound gear assembly is perfectly balanced. However, reference to Figure 7 would indicate otherwise.

The two lever arm lengths represented by X-532 and X-531 clearly show that, regardless of gear-tooth ratios, the radial arm length will have a very important effect upon the applied torque loadings. In this case of Figure 6, these were, of course, 5.625T at engagement 530/531 upon 531 and 1.40625T at engagement 533/532 upon 532.

The fact is that gear rotational speed ratios remain constant. (Thus, if assembly 531 , 507, 532 were itself rotated one revolution in either direction) then we would find that gear 532 would cause gear 533 to rotate 1 1 6 - 2.25 times, and this multiplied by the number of teeth present upon gear 535 (80) would be the number of teeth fed to gear 536, and gear

536 having 128 teeth would be caused to rotate 1.40625 times, in a direction similar to assembly 531 , 507, 532 (144 -^* 64 _ 2.25 revolutions x 80 teeth - 128 teeth = 1.40625 revolutions of gear 536).

One revolution of assembly 531 , 507, 532 would also cause gear 531 to rotate once, thereby causing gear 530 to rotate 36 j 16 = 2.25 times. This would cause gear 509 to rotate 2.25 times, i.e. providing 2.25 x 20 = 45 teeth to be applied to gear 527. Gear 527 having 32 teeth would cause it to rotate 1.40625 times (36 16 = 2.25 revolutions x 20 teeth - 32 = 1.40625 revolutions of gear 527. This is a I : I rotational situation, but in no way a I : I torque situation.

The points at which the contradictory torque loadings are applied, i.e. 72.00mm and 18.00mm radially opposed, causes the off-set ratio to be effective as a means of creating an imbalance so far as torque is concerned, but the gear tooth ratios retain their rotational speed ratio of I : I , thereby demonstrating that the two aspects of torque/speed transmission via normal spur gears in not necessarily a reciprocal function.

This principle can have important ramification, in that, the usual acceptance that a given torque into a shaft wil l, depending entirely upon tooth ratio couplings with said shaft, always provide the torque output as so dictated by this single consideration. However, this is not the case and attention can now be drawn to the device depicted in Figure 6 as illustrating another aspect. If there is 4T placed upon the annular gear 526 at the point of engagement between, for example, gear 52 1 and gear 526, at a radial point 52.00mm out from datum "X" i.e. (26 j 2 = 13 + 39), then 4T output from gear 527 could only be possible at exactly 52.00mm radially out from datum "X", i.e. gear 527 would require to have a diameter at the point of subsequent engagement with a further gear of 52 x 2 = 104.00mm.

If this subsequent engagement was at I : I , then in order that the subsequent drive shaft output from the subsequent engagement be driven at 4T, this subsequent I : I gear would also require being 104.00mm in diameter. However, the 4T output from gear 526 (as drawn) does not return to a similar radius as gear 526 and, therefore, the real torque output at engagement point 527, 528 must be higher than that indicated that is, as gear 527 is only 32.00mm in diameter, then the 4T applied over 104.00mm - 2 = 52.00mm, will have to be recalculated by an off-set ratio taking into considerαtion that the second lever arm is only 32 - 2 = 1 .00mm in length, the off-set ratio being 52 - 16 = 3.25 : I . Therefore the 4T will be multiplied by 3.25 : I, resulting in I 3T being present at the point of engagement between gear 527 and gear 528.

This fact also holds good for the output potential of the planet gear carrier 525 and its subsequent torque engagements. However, it is clear that certain diametric observations and considerations will have to be included in the designs in order that these off-set factors can be used without disadvantage.

These principles have not been icnluded in the descriptive text of the specification as a whole, as to minimise confusion, however, they are real factors which must be considered when finalising a set of parameters to be used as a basis of design.

This innovative approach may, in some ways, explain why, after all the many years of experiment and apparent understanding, many gearboxes suffer failure for inexplicable reasons.

Claims

- -CLAIMS:

1. A transmission device comprising an input shaft rotatable about a main axis of rotation, an output shaft, an input gear rotatable with the input shaft, an intermediate rotary assembly mounted for rotation about the main axis and engaged at respective first and second points of engagement at

-> different radial distances from the main axis by first and second gear assemblies driven by the input gear, the first gear assembly transmitting the input torque applied to the input gear to the intermediate assembly in one rotational direction and the second gear assembly transmitting the input torque to the rotatable assembly in the other rotational direction and being 0 coupled to transmit torque to the output shaft.

2. A transmission device according to claim I , wherein the intermediate rotary assembly comprises a carrier body rotatable about the main axis and provided with first and second axles defining respective first and second 5 subsidiary axes of rotation which are parallel to the main axis and are at different distances from the main axis, the first and second gear assemblies transmitting torque to the carrier body at said subsidiary axes.

3. A transmission device according to claim 2, wherein the first gear 0 assembly comprises an input differential gear assembly mounted on an intermediate shaft rotatable about the main axis and driven by the input gear, and a pair of reference planet gears mounted on the carrier body for rotation about the first subsidiary axis, the reference planet gear being driven by an output gear of the differential assembly and the other being 5 engaged with a stationary gear centred on the main axis.

4. A transmission device according to claim 3, wherein the input differentia! gear assembly comprises a diffeential carrier rotatable with the intermediate shaft, and a pair of idler bevel gears rotatable on respective 0 stub axles presented by the differential carrier and engaged by the input gear, the differential output gear being a bevel gear rotatably mounted on the intermediate shaft and engaging the idler bevel gears.

5 5. A transmission device according to claim 3, wherein the input differential gear assembly comprises an epicyclic annular gear carrier rotatably mounted on the intermediate shaft and provided with a sun gear driving the one reference planet gear, the annular gear carrier being driven be epicyclic planet gears engaged with the input gear and carried by an epicyclic planet carrier rotatable with the intermediate shaft.

6. A transmission device according to any one of claims 3 to 5, wherein the secured gear assembly comprises a sun gear rotatable with the intermediate gear and engaged with a pair of planet gears one of which is engaged with the sun gear and the other of which is coupled to the output shaft, the planet gears being rotatably carried by a carrier gear rotatably mounted on the intermediate shaft and engaged with one of a pair of output planet gears mounted on the carrier body for rotation about the second subsidiary axis, the other output planet gear being engaged with a second stationary gear centred on the main axis.

7. A transmission device according to any one of claims 3 to 5, wherein the second gear assembly comprises an output differential gear assembly having a differential carrier rotatable with the intermediate shaft, a pair of idler bevel gears rotatable in stub axles presented by the differential carrier, a first output bevel gear engaged with the idler gears and rotatable on the intermediate shaft, and a second output bevel gear coupled to the output shaft, the first output bevel gear engaging one of a pair of output planet gears mounted on the carrier body for rotation about the second subsidiary axis, the other output planet gear engaging a second stationary gear centred on the main axis.

8. A transmission device according to claim 2, wherein the first and second gear assemblies comprises a pair of first planet gears mounted on the carrier body for rotation about the first subsidiary axis, one of the first planet gears being engaged with the input gear and the other being engaged with a first sun gear rotatable with an outer sleeve shaft rotatable on the output shaft and carrying a second sun gear, a pair of second planet gears mounted for rotation about a fixed axis defined by a stationary casing of the device and transmitting drive from the first sun gear to a second sun gear mounted on an inner sleeve shaft which is rotatable on the output shaft and cαrries α third sun gear, ond a pair of third planet gears mounted on the carrier body for rotation about the second subsidiary axis and transmitting drive from the third sun wheel to the output shaft.

9. A transmission device according to claim I , wherein the intermediate rotary assembly comprises a carrier body mounted for rotation about the main axis and a stationary casing of the device defines first and second subsidiary axes at different distances from the main axis.

10. A transmission device according to claim 9, wherein the first gear assembly comprises a differential gear assembly mounted on a first intermediate shaft rotatable about the main axis and comprising an annular gear carrier rotatable with the first intermediate shaft, a first sun gear rotatable with the first intermediate shaft and driving first planet gears engaged with a stationary annular gear and carried by a first planet carrier rotatable on the first intermediate shaft, a pair of second planet gears rotatable about the first subsidiary axis and transmitting drive from the first planet carrier to a sun gear rotatable with the carrier body about a second intermediate shaft, which carrier body carries a pair of third planet gears engaged with an output carrier gear rotatable on the second intermediate shaft.

I I . A transmission device according to claim 10, wherein the second gear assembly comprises a pair of fourth planet gears engaged with a planet gear rotatable on the input shaft and carrying planet gears of the differential gear assembly, a second sun gear rotatable with the second intermediate shaft and driven by the fourth planet gears, a third sun gear rotatable with the second intermediate shaft and driving fifth planet gears carried by the output carrier gear and coupled to drive the output shaft.

12. A transmission device according to calim I , wherein the intermediate rotary assembly comprises a first sleeve shaft which is rotatable about an intermediate shaft rotatable about the main axis and which carries first and second sun gears. -30-

13. A transmission device according to claim 12, wherein the first gear means comprises an epicyclic differential gear assembly including an annular gear carrier rotatable with a second sleeve shaft on the intermediate shaft and applying torque to the first sun gear via a third sun gear on the second sleeve shaft and first planet gears carried for rotation about a first fixed axis defined by a stationary casing of the device.

14. A transmission device according to claim 13, wherein the second gear means comprises a planet gear carrier of the epicyclic differential gear assembly which is rotatable with the intermediate shaft, a fourth sun gear rotatable with the intermediate gear and driving the second sun gear wheel via second planet gears carried for rotation about a second fixed axis defined by the stationary casing, the first and second fixed axes being at different distances from the main axis.

15. A transmission device substantially as hereinbefore described with reference to any one of Figures I to 6.

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