WO2008038043A2

WO2008038043A2 - Continuously variable transmission

Info

Publication number: WO2008038043A2
Application number: PCT/GB2007/050578
Authority: WO
Inventors: Christopher John Greenwood
Original assignee: Torotrak (Development) Limited
Priority date: 2006-09-26
Filing date: 2007-09-24
Publication date: 2008-04-03
Also published as: JP6926379B2; KR20090064590A; KR101440848B1; CN101535110A; WO2008038043A3; KR20140063777A; DE112007002280T5; JP2010505074A; JP2015007481A; GB2455030A; JP2017062039A; GB0904809D0; GB2455030B; JP2018200115A; CN101535110B; DE112007002280B4; GB0618929D0

Abstract

The invention is concerned with control of a continuously variable vehicle transmission. The transmission in question is to have a variator (10) which is coupled between a rotary transmission input (32) and a rotary transmission output (46), to provide continuous variation in a transmission drive ratio, via a launch device such as a clutch (62; 68). The variator and the launch device are constructed and arranged to exert a required torque, defined by an input signal. In accordance with the invention, the same control signal is used to control both the variator torque and the torque capacity of the launch device. In such an arrangement, vehicle launch (i.e. moving away from rest) can be controlled in a straightforward manner by progressively raising the relevant control signal, preferably a hydraulic pressure signal. Further independent claims are including for applying traction load to races (20, 22) of the variator and for a method setting torque of the launch device (62; 68) higher than torque of the variator (10).

Description

DESCRIPTION CONTINUOUSLY VARIABLE TRANSMISSION

The present invention is concerned with continuously variable transmissions. One aspect of the invention concerns control of a launch device and of a variator in such a transmission.

In any continuously variable transmission there is a device which provides for the stepless variation of transmission ratio. Such a device will be referred to herein as a "variator".

In a vehicle transmission, provision must be made for "launch" - that is, for accelerating the vehicle away from a standing start. In this connection, some transmissions rely upon use of a "launch device" such as a clutch. This serves to de-couple the engine from the driven vehicle wheels while the vehicle is stationary. To cause the vehicle to move away from a standing start, the transmission is placed in a low gear, the engine is set to produce a suitable torque, and the launch device is progressively engaged, raising the speed of the driven vehicle wheels. Management of this process is however potentially complex.

An alternative approach, well known in the continuously variable transmission art, is to apply the output of the variator to an epicyclic mixing gear, which makes it possible to achieve a condition referred to as "geared neutral" in which the transmission effectively provides an infinite speed reduction, without physically d e-coupling t he t ransmission o utput from t he t ransmission i nput. In this type of transmission, no launch device as such is required. Launch is achieved simply by moving the variator ratio away from the "geared neutral" value. However, of necessity such transmissions involve some constructional complexity in terms of gearing, and certain challenges relating to control.

It is useful to draw a distinction between variators which are "ratio controlled" and those which are "torque c ontrolled". Ratio controlled variators have some physical mechanism for adjusting their own ratio to achieve a set value. For example, known variators of the "half-toroidal" rolling-traction type typically utilise a valve having one part (e.g. the valve spool) which is operatively coupled to variator rollers whose position corresponds to variator ratio and another part (e.g. a movable sleeve forming the valve ports) which is moved to set variator ratio. The valve state depends on the relative positions of these two parts, and the valve controls a pressure applied to piston/cylinder arrangements acting on the variator rollers. The result is a hydro-mechanical feedback loop in which the valve constantly compares variator ratio to a desired value and adjusts it to achieve that value. Associated electronics select the desired variator ratio and send a signal representing this to the transmission.

In torque- controlled variators, there is no such physical arrangement for adjusting variator ratio to a desired value. Instead, the variator receives a control signal which represents a torque to be created, hi the case of a known full-toroidal type variator such as the one described in international patent application PCT/GB2005/03098, publication number WO 2006/027540 (Torotrak Development Limited), this signal takes the form of a hydraulic pressure. In response to it, the variator creates the required torque at its input/output. The actual drive ratio of the variator is permitted to change automatically, to accommodate speed changes resulting from the application of this torque to the relevant inertias. Thus, at the engine/input side of the transmission, the torque created by the variator and the engine output torque sum to determine a net torque acting upon the rotary inertia of the engine and related parts, and so determine the engine acceleration. At the wheel/output side of the transmission, torque created by the variator sums with externally applied torques, due to braking, road gradient etc, to determine the net torque available to accelerate the vehicle itself. Consequent speed changes at both input and output involve changes in variator ratio, and the variator automatically accommodates these.

In known full-toroidal rolling-traction variators of torque controlled type, the variator serves to create a "reaction torque" which corresponds to the control signal. Reaction torque is the sum of the torques at the input and output of the variator. Equivalently, it can be defined to be the torque which must be reacted to the variator' s mountings, in order to prevent it from spinning.

Variators typically rely upon traction between rotating parts for transfer of drive. For example, in the case of toroidal-race rolling-traction variators, rollers engage fractionally w ith t oroidally-recessed v ariator r aces a nd i t i s t hrough t his frictional engagement that drive is transferred from the variator input to its output at variable ratio. To provide traction between the rollers and races, they must be biased toward each other. The biasing force used to create traction in a variator will be referred to herein as the "traction load". In principle, a fixed traction load could be used. However, this would need to be set at a level high enough to avoid excessive slip between the rollers and races under all conditions. The chosen traction load value would consequently be excessive for most conditions, resulting in poor energy efficiency and premature wear to the rolling parts. Hence it is conventional to vary the traction load in sympathy with applied torque. More specifically in a torque controlled variator, traction load is typically varied in proportion to reaction torque. This has the advantage that it provides a constant traction coefficient. Adjustments to traction load must sometimes be made very quickly, to prevent slip in the event of sudden "transient" events such as emergency braking. This is done in some existing systems by using hydraulics to apply the traction load. Specifically, a hydraulic pressure supplied to control pistons coupled to the variator rollers is also led to a hydraulic actuator used to create the traction load, so that the force applied to the variator rollers and the traction load vary in sympathy.

In hydraulic systems of this type, it is usual to provide hydraulic "end stops" to limit the motion of the variator rollers, and so prevent them being driven off the races. This can be done for example by arranging that a fluid outlet from a cylinder containing one of the aforementioned pistons is closed by the piston itself when it reaches the end of its intended travel, the resultant pressure increase in the cylinder serving to arrest the motion of the piston. The increased pressure is also applied to the traction load actuator, as it must be if the change in reaction torque due to the action of the end stops is to be matched by a corresponding change in traction load, which is necessary if slip is not to be created when the end stops act.

The present invention is intended to provide an improved CVT. More specifically (but not necessarily exclusively) it is intended to provide a CVT which is simple in its construction and in its manner of control.

According to a first aspect of the present invention, there is a continuously variable vehicle transmission comprising a rotary input connectable to a rotary driver, a rotary output connectable to vehicle wheels, a variator coupled between the rotary input and the rotary output to provide for stepless variation in drive ratio, and a launch device arranged to selectively couple/de-couple the rotary input and the rotary output, the variator being constructed and arranged to exert a required torque, and the launch device being constructed and arranged to provide a required torque capacity, the transmission being characterised in that it comprises a control arrangement which applies the same control signal to the variator, to set the required torque, and to the launch device, to set its torque capacity.

Torque capacity of the launch device - which device may take the form of a clutch - is the maximum torque it is able to transfer to the variator and is determined by its degree of engagement - e.g. by the applied fluid pressure, in a hydraulically driven clutch. By controlling both clutch and variator torque using a single signal, not only is the equipment required for transmission control greatly simplified, but control of the launch process is also facilitated.

In accordance with a second aspect of the present invention, there is a variator comprising at least one pair of part-toroidally recessed races which together define a generally toroidal variator cavity and which are mounted for rotation on a common variator axis, and at least two rollers disposed between the races to run upon their part-toroidally recessed faces and so transfer drive between them at a variator drive ratio, the rollers being mounted in a manner which permits them to tilt to change the inclination of the roller axes to the variator axis and so permit stepless changes in variator drive ratio, the variator being characterised in that one of its races is coupled to a connection shaft through a mechanical traction loading arrangement which serves both to transmit torque between the connection shaft and the variator race, and to exert on the race a traction load force which is a function of the torque transmitted, the traction loading force urging the variator races into engagement with the rollers to provide the traction needed for transfer of drive, and in that it comprises mechanical abutments which limit roller inclination.

The combination of a mechanical traction load device (in place of a hydraulic device), and mechanical end stops (in place of hydraulic end stops) is highly advantageous. The mechanically generated traction load is able to vary with the necessary speed. Because it is generated in response to the torque acting on the relevant variator race, and not in response to the force applied to the rollers, the changes in variator torque resulting from the action of the end stops automatically result in suitable changes in traction load, without the need for the end stops themselves to be operatively coupled to the traction loading device.

According to a third aspect of the present invention, there is a method of controlling a continuously variable vehicle transmission comprising a rotary input connectable to a rotary driver, a rotary output c onnectable to vehicle wheels, a variator coupled between the rotary input and the rotary output to provide for stepless variation in drive ratio, and a launch device arranged to selectively couple/de-couple the rotary input and the rotary output, comprising controlling the variator to provide a desired reaction torque and controlling torque capacity of the launch device in sympathy with the variator reaction torque, such that torque applied to the launch device by the variator is always smaller than the torque capacity of the launch device. The coordinated control of the reaction torque and the torque capacity of the launch device is thus simplified. Furthermore, this method facilitates a highly advantageous method of managing launch, by progressively increasing variator reaction torque and torque capacity of the launch device, the torque capacity of the launch device always exceeding the torque applied to it by the variator at least until slipping of the launch device ceases, so that until that point the transmission is maintained at its minimum ratio by the torque referred to it through the launch device.

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

Figure 1 is a schematic representation of a continuously variable transmission ("CVT") constructed in accordance with the present invention;

Figure 2a is a more detailed representation of a traction load device used in the CVT, viewed along a radial direction;

Figure 2b is a perspective illustration of a variator race, showing its rear face;

Figure 3 is a schematic representation of a hydraulic control arrangement ofthe CVT; and

Figure 4 illustrates certain components of the variator used in the CVT, viewed along an axial direction.

Figure 1 illustrates a CVT utilising a variator 10 of toroidal-race rolling- traction type. More specifically, this is a twin cavity, full-toroidal variator. It has first and second input races 12, 14 having respective faces 16, 18 which are semi- toroidally recessed. Between the input races are first and second output races 20, 22, and these too have respective semi-toroidally recessed faces 24, 26, so that between the first input and output races 12, 20 is formed a first toroidal cavity 28, and between the second input and output races 22, 14 is formed a second toroidal cavity 30. The races have a common rotational axis defined by a main shaft schematically indicated at 32 about which they rotate.

Each cavity 28, 30 contains a respective set of rollers 34, 36. Each roller is mounted for rotation about a roller axis such as 38 and runs upon the toroidal faces of its associated input and output races to transfer drive from one to the other. The rollers' mountings (not seen in figure 1, but to be described shortly) also permit them to change their inclination — i.e. to change the angle between the roller axis 38 and the main shaft 32 - in accordance with changes in variator drive ratio. The main shaft 32 serves as the rotary input to the variator, and is coupled (either directly or through intermediate gearing, not shown) to a rotary driver such as an engine, which in this particular embodiment takes the form of an internal combustion engine schematically represented at 40. The invention could equally well b e i mplemented u sing a different type of rotary driver such as an electric motor, an external combustion engine etc. The input races 12, 14 of the variator are secured to the main shaft 32 so that they rotate along with it, and so are driven by the engine 40. The output races 20, 22 are able to rotate relative to the main shaft 40. In the illustrated embodiment this is provided for by means of roller bearings 42, 44 through which the output races are respectively mounted upon the main shaft 32. Drive is transmitted from the input races 12, 14 to the output races 20, 22 ( or v ice v ersa, i n an " over-run" condition) through the rollers 34, 36 at variable drive ratio. The output races 20, 22 are able to be operatively coupled to a final drive 46 leading to the vehicle wheels, and this will be described shortly.

In the illustrated variator 10 traction load is provided by means of a mechanical (non-hydraulic) traction loading device 48 which serves to b ias the variator races 12, 14, 16, 18 into engagement with the variator rollers 34, 36 with a force (the "traction load") which is proportional to the output torque of the variator. It does so by urging the two innermost races (which in the illustrated embodiment are the output races 20, 22) away from each other. The traction load is transmitted through the rollers 34, 36 to the outermost races, which in the present embodiment are the input races 12, 14, and these in turn refer the force to the main shaft 32, which is thus placed in tension. By referring the traction load to the main shaft 32 in t his way, any n eed for thrust bearings to withstand the traction load is avoided.

The traction loading device 48 uses a simple ramp arrangement to transmit the output torque, and this ramp arrangement creates a traction load along the axial direction which is a function of (and more specifically, in the present embodiment, proportional to) the torque transmitted.

Figures 2a and 2b make the construction of the traction loading device 48 clear. An output drive gear 50 is fixedly secured to the rear face of output race 22. On its face remote from the output race 22, the output drive gear 50 has a set of ramp-like recesses, seen in phantom at 52 in figure 2a. On its own rear face, the output race 20 has a corresponding set of ramp-like recesses 54, best seen in figure 2b. The recesses 52, 54 have a part-circular section, when viewed along a circumferential direction as in figure 1, to receive rollers 56, formed in this embodiment as spherical balls. When viewed along a radial direction, the recesses 52, 54 are seen to have a shallow "V" shape. Separation of the output drive gear 50 and the output race 20 is minimised when, as in figure 2a, the deepest regions of the recesses 52, 54 are aligned, so that the balls 56 position themselves in these regions. However, consider what happens when the variator output is required to sustain a torque. Note that the output race 20 is able to rotate relative to the output drive gear 50 by virtue of the bearings 42, 44. As the output torque causes relative rotation of these parts, the deepest regions of the recesses 52, 54 become misaligned and the balls 56 thus ride up the "V" shaped recesses, forcing the output race 20 away from the output drive gear 50 and so creating the required traction load. T his relative rotation c eases when the resulting traction load is balanced by the torque being transmitted. Hence traction load is a function of output torque, as previously stated. The precise nature of this function depends on the formation of the recesses 52, 54, but in the illustrated embodiment one is proportional to the other.

The illustrated transmission is capable of providing both forward and reverse gears - that is, it is able to reverse the direction of rotation of the final drive 46. This is achieved by providing two routes for power take off from the output races 20, 22. The first of these routes is via a first set of teeth 58 formed upon the output drive gear 50, which drive a chain gear 60 through a drive chain, which is omitted from figure 1 for the sake of clarity but which runs upon teeth 58 and chain gear 60. The chain gear 60 in its turn is operatively coupled to one side of a forward clutch 62 whose other side is operatively coupled to the final drive 46. The second route for power take off is through a second set of teeth 64 formed upon the output drive gear 50. These teeth 64 mesh with a gear wheel 66 which is operatively coupled to one side of a reverse clutch 68, whose other side is in turn operatively coupled to the final drive 46. Note that the first route 58, 60, 62 for power take off provides no reversal of the direction of rotation, due to its use of a chain drive. The second route 64, 66, 68 for power take off does create a reversal of direction, due to its use of a pair of gears. Hence engagement of the forward clutch 62 produces rotation of the final drive in one direction, and engagement of the clutch 68 produces rotation of the final drive in the opposite direction. Note that in this particular embodiment the forward and reverse clutches 62, 68 are constructed in a manner which prevents both from being simultaneously engaged.

The final drive 46 comprises gearing 70 leading ultimately to the vehicle wheels, which are not shown.

As noted above, the mountings for the rollers 34, 36 are omitted from Figure 1. One suitable form of mounting is illustrated in Figure 4. In this drawing can be seen one of the variator races 12, 14, 20 or 22 and also two rollers 34 or 36. Their position is influenced through a control lever 72 which is pivotally mounted upon a fulcrum 74 received in a slot 76 of the control lever. The control lever has a generally radially projecting lever arm 78 integrally formed with a cross-piece 80 to form an inverted "T" shape. Ball couplings 82, 84 at opposite extremities of the cross-piece 80 couple it to respective roller-bearers 86, 88 which carry and rotatably mount respective rollers. Note also (although it cannot be seen in Figure 4) that the two ball couplings 82, 84 do not lie in a common radial plane. In Figure 1, the radial plane at the centre of toroidal cavity 30 is indicated by a dotted line 90. The two ball couplings 86, 88 are each displaced from this centre plane 90 and lie on opposite sides of it, so that a line from the centre of each ball coupling to the centre of each roller 34, 36 is inclined to the radial plane. This inclination is referred to as the "castor angle". When the control lever 72 is moved, it will be apparent from the drawing that both rollers correspondingly move either clockwise or anti-clockwise about the axis of the main shaft 32. When they do so, they are subject (in a manner well known to those skilled in this art) to a steering e ffect by the variator races. B oth rollers consequently tilt about the aforementioned line/axis passing through the centres of the ball couplings and rollers. The steering effect upon the rollers always tends to bring them to a tilt angle in which the axes of the rollers intersect the axis of the main shaft 32. By virtue of the castor angle, they are always able to find a tilt angle which provides this intersection. The result is that the tilt of the rollers - and hence the variator ratio - is a function of the position of the control lever 72.

It is important that the loads borne by individual rollers are equal and in the Figure 4 arrangement, movement of the control lever 72 along the generally radial direction defined by the slot 76 permits equalisation of roller load.

An actuator 92 is used to apply a controllable biasing force to the lever arm 78. hi the present embodiment, the actuator 92 is a hydraulic device which is double-acting. That is, it receives two opposed hydraulic pressures, the force it exerts being determined by the difference in these two pressures so that it can either be to the left or to the right in Figure 4. Note also that a single actuator in this embodiment controls the respective levers 72 of both variator cavities 28, 30. Although the second control lever cannot be seen in Figure 4, it should be understood that a bar 94 leads from one control lever 72 to the other, and the piston 96 of the actuator 92 is pivotally coupled to the mid-point of this bar. Hence the position of the piston 96 corresponds to the position of the mid-point of the bar, but the relative positions of the two control levers can change slightly, as needed to equalise roller loading between the two cavities.

The hydraulics used to control the CVT will now be described with reference to Figure 3. In this drawing, box 98 schematically represents an arrangement for providing hydraulic fluid at adjustable pressure. Suitable means for achieving this will be known to those skilled in the art. This pressure is led to a variator crossover valve 100 through which it can be applied to either side of the piston 96, to urge the control levers 72 in one direction or the other. In Figure 3, an exhaust from the low pressure side of the piston is shown leading to a sump 102, although in practice, to avoid the relevant chamber being emptied altogether, it may instead be led to a low pressure source. The actuator 92 applies a force to both control levers 72 whose magnitude is determined by the pressure supply 98 and whose direction is controlled by the variator crossover valve 100. It is by adjustment of this force that control is exercised over variator reaction torque.

The pressure from source 98 is also led to a clutch selection valve 104. This valve serves to apply the aforementioned hydraulic pressure selectively either to the forward clutch 62 or to the reverse clutch 68. The inactive clutch is exhausted to the sump 102 through the same valve. Hence the clutch selection valve 104 determines whether the transmission operates in forward or reverse, and the pressure supply 98 determines the force with which the active clutch is engaged, and hence its torque capacity. An isolation valve 105 between the pressure supply 98 and the clutch selection valve 104 serves to selectively disconnect these parts when the vehicle is in neutral.

As noted above, some means are normally provided to limit the minimum and maximum variator ratio. In the absence of such means, there would be the danger of the rollers 34, 36 tilting so far as to leave the variator races 12, 14, 20, 22, with potentially catastrophic results. As noted above, such "end stops" are typically hydraulically implemented in the prior art. However, in the illustrated embodiment of the present invention, simple mechanical stops are placed upon the movement of the rollers. Specifically, these stops limit motion of the single actuator 92, 96 used to control all of the rollers. They could in principle take any number of different forms, but are shown in Figure 3 as buffers 106, 108 within the actuator 92 which simply abut the piston 96 when it reaches the end of its travel.

The area of piston 96, and the areas of pistons within the forward and reverse clutches 62, 68 (the latter pistons not being seen in the drawings, although the construction of suitable clutches will be well known to those skilled in the art) are chosen to ensure that the torque capacity of the active clutch 62, 68 exceeds the output torque of the variator, both of course receiving the same hydraulic pressure from source 98. Consider therefore what happens during a vehicle launch. Prior to launch, pressure to the clutch selection valve 104 is relieved by the isolation valve 105. Neither clutch is engaged, and the vehicle wheels are thus de-coupled from the variator, To initiate launch, the clutch selection valve 104 is set to provide either forward or reverse, the pressure supply 98 is set to a suitably low value, and the state of the isolation valve 105 is then changed to apply this pressure to the relevant clutch. Because the torque capacity of the clutch always exceeds the output torque of the variator, the variator will initially be forced to adopt its minimum ratio, as determined by the end stop buffer 106. This will happen whatever the state of the variator crossover valve 100, but in fact to avoid any "clunk" created by the clutch torque driving the variator to the end of its ratio range, the crossover valve 100 is set initially also to urge the variator to its minimum ratio.

Engagement of the active clutch applies torque to the driven wheels and the vehicle thus begins to accelerate. At some point during the launch the state of the variator crossover valve 100 is changed, so that the applied hydraulic pressure tends to urge the piston 96 away from its buffer 106, to increase variator ratio. The timing of this change is not critical, so long as it takes place while the active clutch is slipping, since during this time the variator will in any event be kept at its minimum ratio by the torque exerted by the active clutch 62 or 68. The pressure from source 98 is progressively increased as the vehicle accelerates, and at some point slipping of the active clutch ceases. Thereafter, continued increase of the hydraulic pressure is able to move the piston 96 off its end buffer 106, so that variator ratio is able to increase as the vehicle accelerates.

During subsequent acceleration and braking, slipping of the active clutch is not to be expected since the load applied to it by the variator is smaller than its torque capacity.

The effect is to provide a transmission in which management of launch can be c ontrolled in a p articular Iy s traight forward manner, and which represents a considerable simplification over known CVTs in terms of its hydraulics. Modern motor vehicles typically use electronics to implement a coordinated strategy for control of the transmission and the engine. The CVT under consideration here would be controlled in this way. The two most basic quantities to be controlled, in the present example, are variator reaction torque (set by means of the pressure supply 98) and engine output torque, set by means of a torque demand supplied to an engine controller.

The aforegoing embodiments are presented purely by way of example and it will be apparent to the skilled reader that the invention could in practice be implemented in many different ways. For example, the illustrated embodiments utilise mechanical abutments to limit the travel of the rollers and hence the ratio of the variator. However, it is known in the art to use instead a hydraulic arrangement in which outlet ports from the actuator 92 are formed in the sides of its cylinder, so that excessive travel of the piston in either direction closes the outlet ports and so provides an end-stop function. The same type of arrangement could be used in implementing the present invention. Also, whereas the illustrated embodiment uses a mechanical ball and ramp arrangement for providing traction load, this function could also be carried out in other embodiments by hydraulics. It is well known, for example, to supply the same pressure both to the actuators 92 and to a hydraulic piston/cylinder arrangement acting upon one of the variator races to provide the end load, and the same could be done in embodiments of the present invention.

Claims

1. A continuously variable vehicle transmission comprising a rotary input connectable to a rotary d river, a rotary output c onnectable to vehicle wheels, a variator coupled between the rotary input and the rotary output to provide for stepless variation in drive ratio, and a launch device arranged to selectively couple/de-couple the rotary input and the rotary output, the variator being constructed and arranged to exert a required torque, and the launch device being constructed and arranged to provide a required torque capacity, the transmission being characterised in that it comprises a control arrangement which applies the same control signal to the variator, to set the required torque, and to the launch device, to set its torque capacity.

2. A continuously variable transmission as claimed in claim 1 in which the construction of the variator and launch device is such that the torque capacity of the clutch always exceeds output torque exerted by the variator.

3. A continuously variable vehicle transmission as claimed in claim 1 or claim 2 in which the control signal is a hydraulic pressure.

4. A continuously variable transmission as claimed in any preceding claim in which the launch device comprises a clutch or brake.

5. A continuously variable transmission as claimed in claim 3 in which the hydraulic pressure is applied to a hydraulic actuator of the variator and determines a force applied to a movable torque transfer part of the variator.

6. A continuously variable transmission as claimed in claim 3 in which the variator has movable rollers running upon a pair of races to transfer torque from one to the other, a hydraulic actuator being coupled to the rollers to exert a force upon them, and the said hydraulic pressure being applied to the hydraulic actuator to determine the said force.

7. A continuously variable transmission as claimed in claim 7 in which the variator further comprises a traction loading device coupled to one of the said races to transmit torque to/from it, the traction loading device being constructed and arranged to apply to the races a traction load which is determined by the torque it transmits.

8. A continuously variable transmission as claimed in claim 7 in which the traction loading device comprises first and second parts mounted for rotation about a common axis and shaped such that rotation of one relative to the other causes axial displacement of one relative to the other.

9. A variator comprising at least one pair of part-toroidally recessed races which together define a generally toroidal variator cavity and which are mounted for rotation on a common variator axis, and at least two rollers disposed between the races to run upon their part-toroidally recessed faces and so transfer drive between them at a variator drive ratio, the rollers being mounted in a manner which permits them to tilt to change the inclination of the roller axes to the variator axis and so permit stepless changes in variator drive ratio, the variator being characterised in that one of its races is coupled to a connection shaft through a mechanical traction loading arrangement which serves both to transmit torque between the connection shaft and the variator race, and to exert on the race a traction load force which is a function of the torque transmitted, the traction loading force urging the variator races into engagement with the rollers to provide the traction needed for transfer of drive, and in that it comprises mechanical abutments which limit roller inclination.

10. A variator as claimed in claim 9 in which the traction loading device comprises first and second parts mounted for rotation about a common axis and shaped such that rotation of one relative to the other causes axial displacement of one relative to the other.

11. A variator as claimed in claim 10 in which said first part of the traction loading device comprises at least one cam surface, which extends circumferentially and is inclined to the radial plane, and at least one follower which rides upon the said cam surface and bears upon the second part.

12. A variator as claimed in any of claims 9 to 11 in which the rollers are operatively coupled to a hydraulic piston/cylinder arrangement and the mechanical abutments are positioned to abut the piston to limit its travel and so limit roller inclination.

13. A method of controlling a continuously variable vehicle transmission comprising a rotary input cormectable to a rotary driver, a rotary output comiectable to vehicle wheels, a variator coupled between the rotary input and the rotary output to provide for stepless variation in drive ratio, and a launch device arranged to selectively couple/de-couple the rotary input and the rotary output, comprising controlling the variator to provide a desired reaction torque and controlling torque capacity of the launch device in sympathy with the variator reaction torque, such that torque applied to the launch device by the variator is always smaller than the torque capacity of the launch device.

14. A method as claimed in claim 13 which comprises applying the same control signal to the variator and to the launch device.

15. A method as claimed in claim 13 or claim 14 for a vehicle transmission, comprising managing vehicle launch by progressively increasing variator reaction torque and torque capacity of the launch device, the torque capacity of the launch device always exceeding the torque applied to it b y the variator at least until slipping of the launch device ceases, so that until that point the transmission is maintained at its minimum ratio by the torque referred to it through the launch device.