Continuously Variable Transmission Device

This invention relates to an improved continuously variable transmission device.

As Global Warming gathers pace and world oil demand threatens to overtake a production capacity that appears to be reaching its peak, there is growing pressure to increase the energy efficiency of all types of vehicles. One approach is to replace reciprocating engines running on oil-based fuels with fuel cell 'engines' running on hydrogen. It has also been realised that significant energy is wasted by road vehicles during braking because their kinetic energy is thrown away as heat rather than being recovered and temporarily stored so that most of it can be used to provide subsequent acceleration. The ability to provide so-called regenerative braking is the major energy-saving feature of most of the new so-called 'hybrid' cars and trucks, which combine electrical and conventional power sources in the same vehicle. This invention addresses the requirement to provide regenerative braking and 'surge power' for acceleration in all forms of electric and hybrid vehicles, whether powered by fuel cells, reciprocating engines, gas turbines or other forms of prime mover.

Currently, the usually means of recovering a hybrid vehicle's kinetic energy is to run the vehicle's electric motor temporarily as a generator and store the electricity generated in a battery. During subsequent acceleration, energy to drive the electric motor is drawn from the battery. However, this process involvers four energy transformations (kinetic to electric, electric to chemical, chemical to electrical and finally electrical to kinetic), each of which wastes energy. For example, if each of these transformations has an individual efficiency of around 80%, a good figure at peak power, the overall efficiency, 'wheel-to-wheel', would only be around 40%, and this figure that has yet to be achieved in practice.

In principle, it is possible to achieve much higher levels of overall efficiency by avoiding any inherently wasteful energy transformations and keeping the energy in kinetic form throughout. This typically involves the use of one or more high-speed flywheels. A major technical challenge is then to provide a transmission which can

effectively and efficiently bridge the gulf between flywheel speeds measured in many ten of thousands of revolutions per minute and road wheel rotational speeds ranging from minus a few hundred rpm to plus two thousand rpm or more.

There are important differences between the requirements for a CVT (continuously variable transmission) to efficiently support direct connection of a conventional reciprocating engine to the driving wheels and one which supports a high-speed flywheel system. The former is almost continuously transmitting energy, in cruise as well as during acceleration, while only needing to accommodate mild engine braking when slowing down. For maximum efficiency, some form of lock-up drive is desirable with an engine when the vehicle has reached its cruising speed, but this may be irrelevant with a flywheel system, where it may be helpful if the CVT can disconnect the flywheel completely when the vehicle is cruising at speed. A high- capacity flywheel is an almost constant speed device over a short time interval; one result is that there is little or no requirement for fast ratio changes in the flywheel application, unlike the situation with an engine when the transmission may need to change from a low ratio to high in a fraction of a second, to accelerate at full power from low speeds. However, it will be shown below that, although this invention was stimulated by the specific requirements of kinetic energy surge power units and is optimized for this in its basic form, it has more general application, including conventional road vehicles and those powered by gas turbines, etc.

DE-A-560276 (Jacobsen) discloses several of the key concepts needed for an effective CVT, building on James Watt's original invention of the planetary gearbox for transmission purposes. In essence, Jacobsen' s solution combined most of the elements of a large rolling-contact bearing with those of a single-stage planetary gearbox. While this design may have worked well at low speeds, its torque capacity would have been relatively low, limited by the tribology and metallurgy of the time, and limited also in maximum rotational speed by several design details. Jacobsen also discloses a single stage reduction gearbox on the high-speed side. Jacobsen has only a single control, a hand wheel for changing the ratio by adjusting the axial separation of the two rims of the outer race, with the two rims of the inner race being brought into contact with the planetary rollers by axial coil springs. Consequently the forces

exerted by the races on the planetary rollers are limited and determined primarily by these springs.

US-A-3299743 (Stockton) discloses an infinitely variable transmission similar to Jacobsen. Note that, where Jacobsen used cone-shaped rollers and convex races,

Stockton introduced 'ovoid' rollers and concave races. Stockton also introduced the concept of using oil pressure to adjust the force exerted by the races on the planetary rollers, and hence the torque applied before slippage occurs. An initial spring preloading kept the rollers in constant contact with the races.

US-A-3516305 (Chery) discloses the concept of taking torque from and delivering it to the planetary rollers through the axial centre of the rollers, rather than taking it from one or both ends as proposed in earlier patents.

Summary of the Invention

This invention relates to a continuously variable transmission device of the type having planetary rollers in rolling contact with radially inner and outer races each comprising two axially spaced rims, with control means for varying the axial separation of the two rims of the outer race and thus the radial position of the planetary rollers in rolling contact with them. Such a transmission device also has control means to vary those forces exchanged between the planetary rollers and the races which act orthogonally to the contact surfaces between the rims and the planetary rollers. Energy is transmitted through the transmission by means of rolling traction.

According to the present invention there is provided a continuously variable transmission comprising a casing defining an axis of rotation, input and output shafts on said axis, respective inner race rims on said shafts, respective outer race rims about said inner race rims, a planet carrier rotatable on said axis, and planets carried by said carrier in rolling contact with said inner and outer race rims and adapted to transmit torque by rolling traction, said inner race rims and outer race rims being relatively movable along said axis, and said planets being relatively movable orthogonal to said axis to vary the speed ratio between said input and output shafts, and a controller

being provided to vary the axial separation of said race rims, wherein said carrier is substantially planar and orthogonal to said axis, and wherein said planets extend on either side of said carrier.

Such a transmission is compact and has a low number of moving parts.

In the preferred embodiment one outer race rim is fixed axially with respect to the housing, the other outer race rim being slidable axially of the housing and guided by an annular spigot thereof.

The planets are preferably ovoid, and define a circular channel in the periphery thereof to receive the planet carrier in torque transmitting relationship. In the preferred embodiment the planets comprise half-ovoids adapted for mounting about said carrier on an axle. The axle may be formed with one of said half-ovoids and be adapted to receive the mating half-ovoid by press-fitting. Other connection means are possible to the intent that each pair of half-ovoids are fixed to rotate as one in use, and define a channel for the planet carrier.

The planet carrier comprises a central aperture on said axis of rotation, and preferably comprising a rolling element bearing adapted to receive one of said input and output shafts. The carrier is adapted to carry a plurality of planets, typically in the range two to six, and is provided with a corresponding aperture or recess for each planet axle.

Each of said apertures is somewhat oval, with the long axis thereof radial to said axis of rotation. In the preferred embodiment, each aperture comprises straight substantially radial side portions linked by semi-circular end portions.

The ovalised apertures permit radial movement of the planets with respect to the planet carrier, so as to allow changes of speed ratio.

Each planet is preferably mounted for rotation in the carrier by a rolling element bearing, the inner race of which supports an axle of a respective planet. The outer race of said rolling element bearing is arranged in the plane of said carrier in a

respective aperture so as to be movable radially from the inner semi-circular end to the outer semi-circular end.

Each outer race is mounted in the carrier by two opposed bearing guides, each comprising a substantially η' section shoe having one pair of arms located over the sides of said outer race, and the other pair of arms located closely over the side of a respective carrier aperture adjacent said radial side portions. In this way each planet bearing is permitted to slide radially of the carrier whilst being restrained against axial movement.

Preferably the shoes are dimensioned for non-sliding engagement with said radial side portions, for example by abutment with the semi-circular end portions. In the preferred embodiment the arms associated with the carrier are closer together than the arms associated with the bearing outer race.

Preferably each planet aperture has semi-circular end portions of a radius substantially matching the outer radius of the corresponding rolling element bearing. Said end portions define radial stops for the planets, which are useful in defining a non-driving configuration as will be further explained.

The carrier preferably includes axially directed struts for transmission of torque from the plane of the centre of the planets to an end plane whereby torque can be transferred to/from an input/output shaft. In the preferred embodiment, the number of such struts is equal to the number of planets, and the circumferential edges of the struts may closely follow the peripheral contour of the planets.

Axial movement of the carrier must be accommodated during ratio changes, and for this purpose a plunging drive coupling is required for the corresponding input/output shaft. Preferably this coupling is incorporated within the transmission device, and may comprise a spline, preferably a ball spline, coupling.

End loading of the transmission is preferably provided by a diaphragm spring acting operatively between the transmission casing and the movable outer race rim. In the

preferred embodiment said spring comprises back to back dual spring, the outer diameters of the respective springs constituting the input/output. The dual spring preferably comprises a plurality of radially inwardly extending fingers, aligned in pairs and substantially co-extensive. The inner ends of said fingers may be connected.

The diaphragm spring may be adapted to exert a minimum pre-load on the movable outer race rim.

Certain types of lubricant show a significant increase in their coefficient of friction when subjected to high pressure, such as occurs in the contact patches between the rims and rollers of the present invention. A lubricant with a pressure-sensitive coefficient of friction radically increases the torque capacity of a rolling traction transmission relative to a similar device using conventional lubricants, or none, and is preferred for use in this invention to lubricate the contact surfaces of the races and planetary rollers.

Brief Description of Drawings

Further features of the invention will appear from the following description of embodiments of the invention and will now be more particularly described by way of example with reference to the accompanying drawings, in which:-

Figure 1 is a side view of one embodiment of the invention with the highspeed shaft at one end and the low-speed shaft at the other;

Figure 2 is a sectioned perspective view of the embodiment of Fig. 1, showing the general arrangement of many of the major components;

Figure 3 is a sectioned view of the fixed outer rim, the axially slidable outer rim and the fixed outer rim piston and its associated diaphragm spring;

Figure 4 is a sectioned view of the high-speed shaft assembly, comprising the fixed inner rim, the axially slidable inner rim, and the associated hydraulic cylinder fixed to the high speed shaft;

Figure 5 is a sectioned view of one of the planetary rollers;

Figure 6 shows the planet carrier with two of four planetary rollers installed, with their associated bearings and bearing guides.

Figure 7 shows one of the planetary roller bearing guides; and

Figure 8 contrasts the relative positions of the axially slidable rims, the planetary rollers and the planet carrier, in low and high ratio transmission states.

Description of Preferred Embodiment

In this description the spatial relationship of the major components, and their orientation with respect to one another is important. The skilled person will understand that dimensions are adaptable to suit the intended application.

In a continuously variable rolling traction transmission device according to the invention, during acceleration of the vehicle in which a transmission such as the embodiment of the invention shown in Figure 1 is mounted, the energy input to the device and thence to the vehicle may be applied via a high-speed input shaft 1 at one end of the device, and the energy output from the device may be taken to the driving wheels (not shown) from a low-speed output shaft 2 at the other end of the device. Conversely, during regenerative braking energy flows in the opposite direction.

The high-speed shaft 1 and the low-speed shaft 2 share the same axis of rotation. The casing 3 of the transmission device can be considered fixed relative to the main structure of the host vehicle, although flexible mountings may be used to connect the casing 3 of the transmission to the host vehicle in order to minimize vibration and noise. The device shown in Figure 2 has a typical diameter of 300 mm, a length of 700 mm and is capable of transmitting over 200 kW, with a relative ratio range in excess of 5:1.

Figure 2 shows major components of the device. An outer race comprises outer rims 4, 5; an inner race comprises inner rims 7, 8. The outer rim 4 is fully fixed to the casing 3. A fixed outer piston 6 is fully fixed to the casing 3. The lowest transmission ratio is achieved with the axially slidable outer rim 5 of the radially outer race located at its position of maximum axial spacing from the fixed outer rim 4, and with the rims of the inner race moved as close to each other as practicable so that the planetary rollers 10 are forced to move to their greatest radial (orbital) position. The highest torque-multiplying transmission ratio is achieved with the axially slidable rim 5 being moved to its position of minimum axial distance from the fixed outer rim 4. The inner rim 7 is fully fixed to the high-speed shaft 1. The other inner rim 8 is fixed rotationally relative to the high-speed shat 1 but is able to slide axially relative thereto. The inner rim 7 and the inner rim 8 form the two halves of the inner race, and their contact surfaces which contact the planetary rollers 10 are mirror images of one another. There is similar symmetry between the outer rims 4, 5.

When the slidable outer rim 5 is moved closer to the fixed outer rim 4 in the direction of the arrow 16 (Fig. 3) the planetary rollers 10 are forced to move to a closer radial position, or lower orbit, which forces the inner rim 7 and the axially slidable inner rim 8 to move further apart. Intermediate ratios are achieved by adjusting the axial separation of the two rims of the outer race, which is rotationally fixed relative to the casing 3. Note that the principal axis shared by the high-speed shaft 1 and low-speed shaft 2 is also shared by the two outer rims 4 and 5, the two inner rims 7 and 8, the planet carrier and the outer rim piston 6 and its associated dual-plate diaphragm spring.

The planetary rollers 10 have a surface that conforms generally to the curvature of an ovoid, which allows the contact patch to maintain a more favourable contact angle than would be obtained with spherical planetary rollers. The contact surfaces of the planetary rollers 10 and the inner and outer races must be kept covered with a thin film of lubricant. The lubricant is typically of a type which responds to high pressure in a contact patch by momentarily taking on a semi-solid form, resulting in the ability to transmit greater torque through the contact patch than a conventional lubricant. The

details of the lubrication system have been omitted from the drawings, for clarity, but are conventional.

The planetary rollers 10 are carried in the planet carrier shown in Figure 6 by rolling element bearings 22 mounted in bearing guides 21 fixed in apertures 20 cut into the carrier flange 11. The planet carrier comprises the carrier torque disc 13, multiple carrier struts 12 and the carrier flange 11. The carrier flange 11 is optionally stabilized relative to the high-speed shaft via a rolling element bearing 23 with its outer race connected fixedly to the centre of the carrier flange 11 and with its inner race able to slide axially on the high-speed shaft 1. The torque exerted by or on each planetary roller 10 is transferred to or from the carrier flange 11 via the bearings 22. Figure 7 shows a planetary bearing guide 21 in detail. Planetary bearing guides 21 are mounted in pairs, one on each side of each bearing aperture 20. The bearing guides 21 are shaped to keep the principal plane of the associated bearing 22 in the principle plane of the carrier flange 11 while allowing the bearing 22 and the associated planetary roller 10 to change orbit, the distance from the principal axis of the planetary roller 10 to the principal axis of both the high- and low-speed shafts. The slots and bearing guides 21 permit each planetary roller 10 and its bearing 22 to slide radially relative to the carrier flange 11 as the transmission ratio changes while ensuring that the principal plane of the carrier flange 11 continues to run through the centres of the planetary rollers as the orbits of the planetary rollers change with ratio changes.

Torque is transferred from and to the carrier flange 11 to and from the earner disc 13 via carrier struts 12 fixed to the carrier flange 11 at one end and fixed at the other end to the carrier disc 13 located towards the low-speed shaft 2 end of the transmission. The diameter of the torque disc 13 is typically less than the maximum diagonal dimension of the torque flange 11. The torque disc 13 and torque flange 11 are typically of similar thickness if of the same material, as they carry similar loads. Each carrier strut 12 is located between two planetary rollers 10, with its long axis set just outside the orbital range of the rolling axes of the planetary rollers 10, and approximately parallel to the principal axis of the high-speed and low-speed shafts. Consequently, the number of such struts will equal the number of rollers. Each carrier

strut 12 may be additionally reinforced, particularly at the ends (by e.g. splaying), to take the high maximum torque the device is capable of generating.

The low-speed shaft 2 is on the same rotational axis as the high-speed shaft 1 and is supported in the transmission casing by two or more rolling element bearings which locate the low-speed shaft 2 axially and radially. At the end of the low-speed shaft 2 which is closer to the torque disc 13 a ball spline joint 24 is fitted to connect the torque disc 13 and hence the rest of the planet carrier axially slidably to the low-speed shaft 2. The core of the ball spline joint 24 is attached to the torque disc 13 at the centre of the disc, with the outer casing of the ball spline joint being coupled to a mating feature of the low-speed shaft 2, as shown in Figure 2. Note that Figure 2 shows the slow-speed shaft 2 detached from the rest of the transmission device in order to expose the core of the ball spline 24.

The high-speed shaft 1 can be externally attached at the same end of the transmission as the low-speed shaft 2 or at the opposite end. In the first case, the high-speed shaft 1 runs through the hollow low-speed shaft 2 (as in a quill shaft) to connect the rims of the inner race to the flywheel(s) or engine, supported by rolling element bearings at each end, with optional additional support possible from plain or needle roller bearings acting between the low-speed and high speed shafts. In the second case, with access to the low-speed shaft 2 at the opposite end of the transmission to access to the high-speed shaft 1 explained in detail here, the inner end of the high-speed shaft 1 is supported by a rolling element bearing mounted within the inner end of the low-speed shaft 2. This technique is commonly found in manual gearboxes fitted to rear-wheel- drive motorcars. The outer end of the high-speed shaft 1, on the other side of the inner race, is supported by one or more rolling element bearings mounted in the transmission casing 3 and which locate the high-speed shaft 1 radially.

The surface of a planetary roller 10 in contact with the rims of the inner and outer races approximates closely to much of a spheroid lengthened in the direction of a polar diameter, where the axis about which the spheroid spins is parallel to the principal axis of the high- and low-speed shafts. While the curvature of the convex contact surfaces of the planetary rollers 10 in the planes running through the principal

axis of each planetary roller 10 is likely to be more complex than an arc of a circle and to vary from application to application, it is also likely to approximate to an arc of a circle. The corresponding contact surfaces of both pairs of rims will have complementary concave surfaces with axial curvatures approximating to an arc of a circle but with a radius somewhat larger than that of an associated planetary roller 10. The contact surfaces of the planetary rollers 10 and of the rims of the races are always circular in cross-section in the plane orthogonal to the principle axis of the shafts. An equatorial channel 29 is provided in each planetary roller 10 to allow each pair of bearing guides 21 to engage the rolling element bearing 22 mounted at the centre of each planetary roller, with the principal axis of the bearing 22 coincident with the principal axis of the planetary roller 10.

As shown in Figure 5, a typical planetary roller 10 consists of three main features, namely two identical roller ends 18 and a roller shaft 19. The axis of revolution of the contact surface of each roller end 18 is the axis of rotation 30 of the planetary roller 10. The principal axis of the roller shaft 19 is coincident with the axis of rotation of the planetary roller 10. The associated bearing 22 fits between the two roller ends 18 on the roller shaft 19. The inner race of the planetary roller bearing 22 is typically an interference fit on the roller shaft 19. The rotor ends 18 may be an interference fit on the roller shaft 19. One method of assembly will now be described to illustrate construction. One of the two bearing guides 21 associated with the specific planetary roller 10 is mounted in the appropriate slot 20 in the carrier flange 11. Each slot 20 is shaped to allow the second bearing guide 21 to be inserted and then moved into its operating position once the bearing 22 has been fitted. The planetary roller bearing 22 is engaged in the fixed bearing guide 21 and moved to its outer orbital position. The second bearing guide 21 can then be moved into position and fixed to the carrier flange 11. The roller shaft 19 is forced into the shaft hole in one of the rotor ends 18. The roller shaft 19 is then forced into its operating position with its mid-point inside the inner race of the planetary roller bearing 22. The second roller end 18 is then forced into position. Note that one option is to adopt a roller shaft length which results in the two roller ends 18 clamping the inner race of the planetary roller bearing 22 between them.

The degree of conformity between the curvature of the contact surfaces of the planetary rollers and the curvature of the rims reflects a trade-off durability and torque capacity versus efficiency. Low conformity, with rim surfaces significantly larger in radius than the radii of the planetary rollers, is likely to result in excellent efficiency in energy transmission, at the expense of transmission life and peak torque capacity. Conversely, where the radii of rims and rollers are close, efficiency will be lower. When the present transmission is used as part of a kinetic energy system in which the drive may be transmitting significant energy for less than 20% of running time in surges of high power, the specific implementation is likely to be biased towards high conformity. Conversely, if the present invention is being used in association with a Brayton cycle engine, for example, to supply continuous power, the specific implementation is likely to be biased towards low conformity.

An advantage of the present invention in controlling the flow of energy through the transmission is that only one component and one associated variable needs to be controlled by an electrical, mechanical, hydraulic or other form of actuator to control the instantaneous ratio of the transmission, and only one other component and associated variable requires similar actuation to control the instantaneous torque transmitted by the transmission. The first variable is the axial distance between the two rims of the outer race, which determines the instantaneous ratio through the core of the transmission. To avoid the expense of supporting and controlling the axial movement of both rims, one of the rims can be completely fixed relative to the transmission casing, and only the other outer rim is required to move, and then only axially. Practical considerations favour making the outer rim closer to the planet carrier the adjustable outer rim, and this rim will be referred to as the ratio control rim or the slidable outer rim 5 in what follows.

In an implementation controlling a flywheel, setting a precise ratio in the transmission is relative unimportant. The important parameters are the rate of change of ratio and its direction. For example, if the rims of the outer race are currently half the maximum distance apart, then moving the ratio control outer rim 5 towards the high ratio position, i.e. closer to the fixed outer rim 4, will force the planetary rollers 10 radially inwards, forcing one or both rims of the inner race on the high-speed shaft 1

to move outwards. It will be seen that, if the outer rim further from the planet carrier is fixed, the inner rim nearer the planet carrier needs only a small amount of axial movement to accommodate the full range of positions of the planetary rollers 10, in contrast to the other inner rim which will need to move through a considerable axial range. If the radii of the planetary rollers 10 and the rims are well-chosen, the inner rim nearer to the planet carrier can be connected fixedly to the high-speed shaft, with the bearings of the high-speed shaft 1 accommodating the slight axial movement occasioned by the axial movement of the planetary rollers 10. Note that significant axial movement is normally constrained by the planetary rollers 10 and races themselves, acting as one large, self-locating, rolling element bearing. In an implementation with the high-speed and low-speed shafts emerging at opposite ends of the transmission, the inner end of the high-speed shaft 1 can have a rolling element bearing fitted in the axially inner end of the low-speed shaft 2 to support the near end of the high-speed shaft 1 when the forces between the planetary rollers 10 and rims are low.

In a preferred implementation, the inner rim closer to the planet carrier is completely fixed to the high-speed shaft 1. The other inner rim is fixed rotationally to the highspeed shaft 1 but is free to slide axially. Because of the friction characteristics of the high-mu lubricant, a simple splined joint between the sliding inner rim and the highspeed shaft may not be optional, so a preferred solution is to use a ball spline, as shown in Figure 4. The precise axial position of the torque control inner rim 8 does not need to be controlled but the force applied by the torque control rim 8 on the planetary rollers 10 does need controlling. Ideally this force should be applied by action and reaction between the two inner rims, the force carried internally by the high-speed shaft 1. In a preferred implementation this is achieved by making part of the axially slidable torque control rim 8 a piston within a cylinder 9 mounted fixedly to the high-speed shaft 1. The cylinder 9 is fed with fluid under pressure via a passage bored through the centre of the high-speed shaft 1 from its outer end. The hydraulic circuit may use the high-mu lubricant or may be a completely separate system using a different fluid.

The rate at which energy is transferred to and from the flywheel system is a function of the rate of change of ratio within the transmission core. The direction of energy flow is a function of whether the distance between the two rims of the outer race is increasing or reducing. The maximum torque rating is a function of the strength of the mechanical components, the friction characteristics of the friction surfaces of the rims and the characteristics of the lubricant used in the core of the transmission, and the compressive forces applied to the contact patches between the planetary rollers 10 and the races by the controlling rims. At a basic level, what the transmission will do depends on the difference between the forces being applied to the planetary rollers by the outer rims and the forces being applied by the inner rims. If these forces are in balance, the ratio will remain fixed and little, if any, energy will flow unless the speed of either shaft is altered externally. However, if the forces acting on the planetary rollers from the outer rims are less than the forces from the inner rims, the planetary rollers will tend to move to a lower ratio position, as the inner rims move closer together and the outer rims are forced further apart. This will increase the speed of the low-speed shaft 2, accelerating the vehicle if the low-speed shaft 2 is connected to its drive wheels, and slowing down the high-speed shaft 1, drawing energy out of the flywheel system, if that is what is connected. Conversely, if the forces acting on the planetary rollers 10 from the outer rims are greater than the forces form the inner rims, the planetary rollers will tend to move to a higher ratio, as the inner rims are forced further apart and the outer rims move closer together. This is illustrated in Figure 4 by the movement, relative to the other components of the torque control rim 8 in the direction of the arrow 17. This will reduce the speed of the low-speed shaft 2, slowing the vehicle if connected, and accelerate the high-speed shaft 1, feeding energy into the flywheel system, if connected.

During normal operations, the control system for the transmission will receive from the vehicle control system a series of signals indicating the instantaneous level of acceleration or deceleration required, and other commands. If such signals cease, the transmission controller will normally be programmed to bring the vehicle gently to a halt. The transmission controller will act on the commands from the vehicle controller by adjusting the absolute and relative values of the forces acting on the planetary rollers from the rims of the inner and outer races. In a preferred implementation of the

present invention, the axial forces applied to the rims of the inner race are applied by a hydraulic system. The torque-controlling inner rim 8 is free to move axially relative to the high-speed shaft 1 on which it is mounted but it is relatively fixed rotationally. A cylindrical space is formed by a part of the torque control rim 8 acting as a piston and by a hollow cylinder 9 with its principal axis coincident with that of the inner race and the high-speed shaft 1. The hollow cylinder 9 is open at the end closer to the inner rims and is closed at the other end and is connected fully fixedly at its closed end to the high-speed shaft 1. A high-mu lubricant or some other liquid acts as the hydraulic fluid and is fed into the chamber formed by the piston element of the axially slidable inner rim 8 and the open cylinder 9 fixed to the high-speed shaft 1. The cylindrical wall of the piston component of the torque control rim 8 is a close fit with the inner cylindrical wall of the open cylinder 9. The forces applied to the axially slidable torque control rim 8 are balanced by the reacting forces from the other inner rim 7 fixed to the high-speed shaft when the planetary rollers 10 have settled in their new positions determined by the position of the ratio control outer rim 5. A slidable inner rim 8 can be achieved using a multiple spline joint with the splines parallel to the shaft or preferably a ball spline joint as shown, to avoid the possibility of the high- mu lubricant locking the joint under high torque. Light bias springs 28 are provided, as illustrated.

The present invention offers two different means of adjusting the axial distance between the two rims of the outer race. When controlling a high-speed flywheel the preferred means to adjust the axial position of the ratio control rim 5 is likely to be similar to the means already disclosed to adjust the position of the inner rims. One or more hydraulic chambers are hollowed into the ratio control rim 5, the simplest form being a chamber bounded by the walls of the hollow chamber or chambers in the ratio control rim and a piston 6 or pistons fixed to the transmission casing 3, which is capable of carrying reactive axial forces to and from the other outer rim 4, also fixed to the transmission casing 3. If the transmission casing 3 and associated elements are circular in cross-section, then axial splines between the transmission casing 3 or outer piston 6 and the ratio control rim 5 can be used to constrain relative rotation but permit axial movement of the ratio control rim 5. In some applications, the contact forces in the splines will not be high enough to cause the splines to lock, but ball

splines may be used if they are. Another possibility is to adopt a cross-section for the control-rim/transmission-case interface which allows axial movement of the rim but prevents relative rotation. However, the risk of 'ratio lock' must be considered if a high level of torque is to be transmitted, keeping in mind the characteristics of the lubricant.

The device is adapted to transmit a high torque between the axially slidable outer rim 5 and the transmission casing 3 while permitting smooth axial movement of the slidable outer rim 5, with the minimum of frictional resistance to the axial movement of the slidable outer rim 5. As shown in Figure 3, this is achieved by using a dual- plate diaphragm spring connecting the slidable outer rim 5 to the fixed outer piston 6 connected fully fixedly to the transmission case 3. The diaphragm spring allows considerable axial movement of the slidable outer rim 5 while offering only low, consistent and smooth resistance to axial movement of the slidable outer rim 5. The larger plate 14 of the dual-plate diaphragm spring is fixed at its outer circumference to the fixed transmission casing 3 via the fixed annular piston 6. The outer circumference of the smaller plate 15 of the diaphragm spring is fixed to the slidable outer rim 5. The two plates are joined together around a common inner circumference large enough in diameter to clear the low-speed shaft 2. The common inner circumference is not contiguous and is provided with deep radial cut-outs to ensure that there is little resistance to axial movement of the outer circumferences of the spring plates. The spring plates may not necessarily be flat in the plane orthogonal to the principal axis but may be pre- formed to cause the slidable outer rim 5 to attempt to move to a particular rest position. For example, in the event of loss of hydraulic pressure in the chamber formed by the slidable outer rim 5 and the fixed annular piston 6, it may be desirable for the slidable outer rim 5 to move to a default position, which might be a high-ratio, low-output-speed position.

Another means (not shown) of adjusting the axial position of the ratio control rim is to use a mechanical actuator operated by an electric motor controlled by the transmission controller. The mechanical actuator could take the form of a screw jack arranged to exert axial force to drive the ratio control rim 5 towards the fixed outer rim 4. Forces from the inner rims through the planetary rollers 10 will drive the slidable ratio control

rim 5 back when the jack retracts. This form of ratio control may prove more appropriate where the high-speed shaft 1 is connected to a variable speed prime mover such as a reciprocating engine, when the ratio may need to remain fixed for a period of time.

The present invention provides an integrated clutch function within the core of the transmission, capable of operating at all speeds of both the high- and low-speed shafts. It will be observed that the invention as described already allows a de-clutching function at all medium and high speeds by allowing the pressure in the hydraulic system controlling the separation of the rims of the inner race to fall to zero, which then allows the friction forces between the planetary rollers 10 and the inner rims to fall to the point where slip occurs and little or no energy can pass from the inner rims on the high-speed shaft 1 to the planetary rollers 10 connected to the low-speed shaft 2 via the planet carrier, and vice versa. Such an arrangement allows a high speed flywheel to be disconnected on demand.

At medium and high planet carrier rotation speeds, centrifugal effects will keep the planetary rollers 10 in contact with the outer race and disengaged from the inner race. However, at low speeds of the low-speed shaft 2 and planet carrier, each planetary roller 10 will tend to fall towards the inner race in the upper half of its orbit, pick up kinetic energy from the inner race and then transmit some of it to the planet carrier and low-speed shaft 2 when the planetary roller 10 next makes contact with the outer race. To prevent this, the apertures 20 in the carrier flange 11 which constrain the bearings 22 supporting the planetary rollers 10 are limited in their innermost radial extent to ensure that when the inner rims are furthest apart the planetary rollers 10 are prevented from contacting the inner rims with more than minimal force. Similarly, the outermost radial extent of the slots 20 in the carrier flange 11 limit the maximum orbit of the planetary rollers 10 so that there may be little or no contact between the planetary rollers 10 and the rims of the outer race when the outer rims are at maximum axial separation. This may be useful when the vehicle is cruising and losses through the core of the transmission need to be minimised. If the host vehicle does not need to receive any energy from the kinetic energy storage system, the hydraulic pressure in both the ratio control system and the torque control system can be allowed to drop to

zero, allowing both the torque control rim 8 and the ratio control rim 5 to be pushed by the unconstrained movement of the planetary rollers 10 into positions where there is little or no pressure at the contact patches. In this condition, transmission losses are minimal.

While the difference between the forces applied to the inner and outer rims determines the rate of ratio change, it is the absolute value of the forces between the inner rims and the planetary rollers 10 and between the planetary rollers 10 and the outer rims which will determine whether slippage will occur between these components, for any given geometry and lubricant. With the proper level of computer co-ordination, this permits a sophisticated form of traction control capable of delicate application of cadenced torque in particularly slippery conditions such as snow, ice or sand, while also providing conventional traction control ¹ at all levels of power. In many circumstances, the engine will be able to remain at full power because the CVT and associated flywheel system will be capable of absorbing and temporarily storing the excess energy until traction is fully restored.

According to another aspect of the present invention, an electric motor (not shown) is mounted with its stator fixed to the transmission casing and its rotor connected fixedly to the low-speed shaft. The motor can serve a number of purposes. One is the provision of a reverse capability on the low-speed shaft. With the inner rims at maximum separation and the slots 20 in the carrier flange 11 holding the planetary rollers 10 away from the inner rims, the electric motor is free to turn the low-speed shaft 2 in the reverse direction. The electric motor can also provide fine control of the vehicle in both forward and reverse directions at low speeds, particularly below the forward speed above which the planetary rollers 10 engage with both races without slip when the ratio control rim 5 is set for the highest ratio.

There are typically a minimum of two and a maximum of six planetary rollers in the present invention. Generally, the greater the number of rolling planets, the higher the torque capacity of the transmission, given a fixed size of outer race. The number of planetary rollers dictates the number of carrier struts, carrier slots, being guides and bearings supporting the planetary rollers. For example, a four roller transmission (as

illustrated) will require four carrier struts, four carrier slots, eight bearing guides and four bearings for the planetary rollers.

Figure 8 contrasts the positions of most of the major components at the two extremes of ratio to illustrate the fundamental functions of the transmission. The top half of the figure shows the situation with the transmission set for its lowest ratio and the bottom half of the figure shows the situation with the transmission set for its highest ratio. Note that the planet carrier has deliberately been omitted, in the interest of visibility of the other components. The position of the plant carrier can be inferred from the position of the planetary rollers 10. The fixed outer rim 4 and the outer rim piston 6 remain fully fixed at all times relative to the transmission casing 3 (not shown in this figure). It would appear that most of the components of the high-speed shaft assembly have not moved axially either, but they will have moved fractionally, probably sequentially in both axial directions, during the transition from low to high ratio. The high-speed shaft 1 itself, the inner rim cylinder 9 and the inner rim 7 fully fixed to the high-speed shaft 1 always move as one. Clearly the major axial moves are made by the ratio control outer rim 5, the planetary rollers 10 and the torque control inner rim 8. It should also be clear why the axial position of the ratio control outer rim 5 determines the radial and axial positions of the planetary rollers 10 and why the inner rims slide to adopt the axial positions they do.

Figure 8 shows how the radii of the contact patches between the components vary as the axial position of the ratio control outer rim changes, resulting in changes of ratio. For example, in the upper half of the figure the effective radius inside the planetary rollers 10 of the contact patches with the inner rims 7 and 8 is relatively small compared to the effective radius of the contact patches between the same components in the lower half of the figure. Conversely, the effective radius within the inner rims 7 and 8 of the contact patches between the inner rims and the planetary rollers 10 is larger in the upper half of the figure than in the lower half. It is the combination of these and other changes in radii which delivers a wide and continuously variable range of transmission ratios.