DESCRIPTION CONTINUOUSLY VARIABLE RATIO

TRANSMISSION DRIVE

The present invention relates to continuously variable ratio transmission devices

("variators") of toroidal-race, rolling-traction type, and in particular to means for applying the "end load" needed to provide traction between rollers and races in such variators.

Variators are used in transmissions - particularly but not exclusively in

motor vehicle transmissions - to provide continuous variation of transmission drive ratio. In a toroidal-race, rolling-traction type variator a pair of co-axially mounted circular races have respective semi-toroidally recessed surfaces which together define a generally toroidal cavity, within which is mounted a plurality of rollers. The rollers are sandwiched between the races, running upon their recessed faces, and serve to

transmit drive from one to the other. By changing the rollers' inclination to the races'

common axis (which will be referred to as the "variator axis") the relative speeds of the two races - and hence the variator ¹ s drive ratio - can be altered in a stepless manner.

One might envisage controlling roller inclination directly by application of a

suitable turning moment to the roller bearings, but this turns out to be impractical.

Instead, the rollers are caused to steer themselves to suitably vary their inclinations.

One way to achieve this (although not the only way, it should be noted) is to mount the rollers such that they each have some limited freedom to move back and forth

along a circumferential path about the variator axis. Displacement along this path

results in a steering moment upon the roller, changing its inclination, and in this way a relationship is created between (i) roller displacement and (ii) roller inclination.

Hence by moving the roller back and forth along its circumferential path, variator

drive ratio is adjusted.

Transfer of drive between rollers and races relies upon traction between them, and to provide this a force must be applied to bias these components toward each

other. This force will be referred to herein as the "end load". The rollers and races do not actually contact each other, being separated by a thin film of traction fluid which is maintained by continuously jetting the fluid onto these parts. Traction results from shear within the fluid layer, when subject to a suitably high pressure due to the end load. End load is typically provided by applying an axial force to one of

the races, urging it toward the other.

In what follows, reference will be made to torques applied to the variator races, which requires some clarification. External torques are of course applied to the variator input, through its coupling to the vehicle engine, and to the variator output,

through its coupling to the driven vehicle wheels. Each variator race is also subject

to an internal torque by the rollers. The net torque on both races due to the external and internal torques is small (or zero, at constant speed). However references herein to the torque applied to the race are to the externally applied torque.

The size of the end load has an important effect on variator performance and

longevity. Excessive end load reduces efficiency and causes premature wear, leading

eventually to variator failure. Inadequate end load can lead to excessive - and even catastrophic - slip between the rollers and the races. The end load required to prevent

excessive slip varies with the torques exerted upon the races. In principle a constant end load can be used, but this must be large enough to sustain traction when the variator is subject to the maximum expected torques, so that under all other

conditions the end load is larger than is necessary. Improved efficiency and variator

lifetime can be achieved by adjusting the end load in sympathy with the torques handled by the variator.

A known way to achieve this is described in some detail in Torotrak

(Development) Ltd's published International Patent Application WO 02/079675 (Application No. PCT/GB02/01551). This describes a relatively sophisticated variator in which the rollers are hydraulically controlled. Each roller is coupled to

a respective hydraulic actuator. A common pressure (the "reaction pressure") is applied to each actuator, and determines a force (the "reaction force") applied to each

roller, urging it along the aforementioned circumferential path. At equilibrium, this force is balanced by an equal and opposite force applied to each roller by the action of the races. Put simply, the effect of the reaction pressure is to urge the rollers one way (clockwise or anti-clockwise) while the races urge the roller the ^" other way.

Consideration of the variator's geometry shows that the force exerted by the races on each roller is proportional to the sum of the torques acting on the two races (the

"reaction torque"). Hence, at equilibrium, reaction torque is proportional to reaction pressure. The end load is provided by means of a hydraulic actuator exerting an axial force on one of the races, and the fluid pressure supplied to this end load actuator is equal to, or at least controlled by, the reaction pressure, so that the end load is

proportional to the reaction torque. Ih fact WO 02/079675 also explains how this

relationship could be adjusted to further refine end load control. However, proportionality of end load to reaction torque is, with some provisos, a highly

efficient mode of end load control since, for any given variator ratio, it provides a constant traction coefficient at the roller/race interface. The traction coefficient is defined as the ratio of (i) the force parallel to the interface to (ii) the force normal to

the interface. Note that (i) is proportional to the reaction force. The normal force,

item Qi), is proportional to end load although, since the direction of the normal to the interface varies with the roller inclination, it is also proportional to the cosine of the roller's angle to the variator axis.

End load adjustment must in some situations be carried out with little time lag. Externally created events, such as rapid vehicle braking, can abruptly create a requirement for an increase of end load, and failure to respond with sufficient speed could result in catastrophic slip within the variator. The known hydraulic system provides the required speed of response.

It is conventional to provide a "pre-load", which is a minimum end load

provided ^" even when the ^" tbfqϋe-dependerit component of the ^" end load is zero, either

because the torque on which it is based is very low, or because (particularly at startup) the device providing it is inactive. This is achieved using a spring 19 acting upon

one of the races to provide a pre-load force P (see Figure Ia). Note that the end load force F and the pre-load force P are applied in opposite directions and to different

races. Consequently the variation of the total axial load has the form represented in

Figure 2, in which the total axial load is on the vertical axis and the reaction torque is on the horizontal axis. The total load is proportional to torque, except that because

of the pre-load spring the load does not fall to zero at very low torque, as indicated by the dotted line representing the end load contribution, but is instead sustained at

a minimum level PL by the pre-load spring. Outside this low torque area, the pre¬

load spring is overcome by the end load, and makes no contribution to the total axial load.

Some variators are not hydraulically controlled. The present invention has been devised in connection with a relatively simple variator in which roller displacement is instead controlled through a direct mechanical coupling. Hence

hydraulic end load control is not appropriate. In principle it would be possible to

control end load through some form of mechanical coupling which translates the reaction force applied to the roller's mountings into end load, but in reality this is not considered to be practical. Therefore some other way of suitably adjusting end load is required.

In the hydraulic system described above, end load is adjusted in dependence upon reaction torque. In principle, end load could instead be adjusted in dependence upon either input torque alone or output torque alone, and in fact it is known to vary end load in proportion to input torque. This is done in variators of the so-called "half toroidal" type but proves less appropriate in full toroidal variators. The difference

between half and full toroidal variators is illustrated in Figure 1. In the full toroidal variator 10 of Figure Ia, the races 12, 14 define an almost complete torus about the

variator axis 16 and the rollers lie along a diameter of the circle generating the torus.

The centres 17 of the rollers 18 lie upon, or close to, the centre of this circle, and the

effect of the end load force F is to keep the rollers in position, even while their

inclinations change as indicated by curved arrows in the drawing, hi the half-toroidal

variator 20 of Figure Ib, the centres 27 of the rollers 28 are radially inwardly offset from the center 29 of the circle generating the torus. The effect of end load force F

is to urge the rollers radially outwardly, and to resist this each roller requires a thrust bearing 30. The races 22, 24 only define the radially inner part of the torus.

Consideration of the geometry and performance of half-toroidal variators

shows that an end load which is proportional to torque T at the variator input shaft 32 provides efficient performance. However the same manner of end load control proves less efficient when applied to full-toroidal type variators.

It is an object of the present invention to provide a simple but efficient means and method of controlling the end load.

In accordance with a first aspect of the present invention, there is a continuously variable ratio transmission device ("variator") comprising an input race operatively coupled to a variator input shaft and an output race operatively coupled to a variator output shaft, the input and output races being mounted for rotation about

a common axis, the ^" races having respective part-toroidally shaped surfaces which together define a substantially toroidal cavity containing at least two rollers, each roller running upon the shaped surfaces of both races to transfer drive from one to the other and being provided with mountings which permit the roller's inclination to the common axis to vary, in order to vary the ratio of input shaft speed to output shaft

speed, and an end load arrangement for biasing the races into engagement with the

rollers to provide traction between them, the variator being characterised in that the end load arrangement comprises (a) a first end load device for applying a first end

load component which varies in accordance with torque applied to the input race and (b) a second end load device for applying a second end load component which is

independent of input and output torques, the first and second end load devices acting upon the same variator race so that the total end load is the sum of the first and second end load components.

In accordance with a second aspect of the present invention there is a continuously variable ratio transmission device ("variator") comprising an input race

operatively coupled to a variator input shaft and an output race operatively coupled

to a variator output shaft, the input and output races being mounted for rotation about a common axis, the races having respective part-toroidally shaped surfaces which together define a substantially toroidal cavity containing at least two rollers, each ^■ roller running upon the shaped surfaces of both races to transfer drive from one to the

other and being provided with mountings which permit the roller's inclination to the common axis to vary, in order to vary the ratio of input shaft speed to output shaft speed, and an end load arrangement for biasing the races into engagement with the rollers to provide traction between them, the variator being characterised in that the

end load arrangement comprises (a) a mechanical actuator which applies to one of the

variator races an axial force which varies according to the torque applied to the input

race and (b) a pre-stressed spring acting axially upon the same variator race.

In accordance with a third aspect of the present invention there is a method of controlling end load in a continuously variable ratio transmission device ("variator") comprising an input race operatively coupled to a variator input shaft and

an output race operatively coupled to a variator output shaft, the input and output

races being mounted for rotation about a common axis, the races having respective

part-toroidally shaped surfaces which together define a substantially toroidal cavity

containing at least two rollers, each roller running upon the shaped surfaces of both races to transfer drive from one to the other and being provided with mountings

which permit the roller's inclination to the common axis to vary, in order to vary the ratio of input shaft speed to output shaft speed, and an end load arrangement for

biasing the races into engagement with the rollers to provide traction between them,

the method comprising applying to the same variator race (a) a substantially constant end load component and (b) an end load component which varies with variator input torque.

Preferably the first end load component is substantially proportional to the absolute value of the torque applied to the input race. "Absolute value" is used here to refer to the magnitude of the torque, regardless of its sign (direction).

The second end load component is preferably substantially constant. It may

in particular be provided by means of a pre-stressed spring, and so may be subject to some degree of variation as the relevant parts move somewhat.

It is especially preferred that the first end load device forms a coupling which

serves both to transmit torque from the variator input shaft to the input race and to apply the first end load component, the coupling comprising first and second parts,

which are capable of relative rotational movement, at least one ramp surface which

extends along the circumferential direction and is inclined to a plane perpendicular to the common axis, and a force-transmitting part which contacts the ramp surface and is caused to move along it by said relative rotational movement, and so to apply

between the first and second coupling parts an axial force which varies according to the torque applied to the input race.

The ramp surface could in principle be part of a screw thread. Preferably, however, the ramp acts upon a force-transmitting part in the form of a ball which

serves to urge the first and second coupling parts apart.

It is further preferred that the coupling comprises at least two ramp surfaces which are oppositely inclined to said plane perpendicular to the common axis, such that the axial force is applied in the same direction regardless of the direction of action of the torque applied to the input race.

Preferably the first and second coupling parts have respective ramp surfaces, the force-transmitting part being formed separately from both coupling parts and being retained between their respective ramp surfaces.

One of the coupling parts may be formed by the input race itself.

It is particularly preferred that the first and second end load devices are both anchored to the variator input shaft ,and the variator output race is mounted around and axially restrained relative to the input shaft, so that the end load force is borne in

tension by the variator input shaft. By arranging for the end load to be borne by the shaft, one dispenses with the need for thrust bearings to transmit the load to the variator casing.

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

Figures Ia and Ib are highly simplified sections, in an axial plane, through a full toroidal and a half toroidal variator respectively;

Figure 2 is a graph of the variation of axial load in a known variator;

Figure 3 is a graph of traction coefficient against ratio for variators using different end load control strategies;

Figure 4a is a section in an axial plane through a variator embodying the present invention;

Figure 4b is a section in a radial plane through the same variator;

Figure 5 is a view along a radial direction of a ball ramp device used in the same variator;

Figure 6 is a perspective illustration of a thrust disc of the ball ramp device;

Figure 7 is a perspective view of the rear of a race used in the same variator;

Figure 8 is a perspective view of the front of the same race;

Figure 9 is a graph of end load against input torque for a variator embodying the present invention; and

Figure 10 is a perspective illustration of a registration disc used in the ball ramp device of Figures 4 and 5.

The present invention has been devised following an analysis of full-toroidal variator performance, and in this connection reference is directed to Figure 3, which represents variation of variator traction coefficient, on the vertical axis, with ratio, on the horizontal axis, for a full-toroidal type variator. The calculation has been based on the assumption that variator output torque is constant. This assumption is

particularly appropriate for low speed vehicles, where output torque need not be limited by the power available from the vehicle engine, and the maximum wheel

torque (which is the torque that would cause the vehicle wheels to spin) is available

throughout the transmission's ratio range. Based upon this assumption, variator input torque can be calculated from the chosen constant output torque and the variator

ratio. In the resulting graph, line L ₀ represents the case of a constant end load, which

could for example be provided by a simple spring, and exhibits large (and undesirable) variation of traction coefficient with ratio. Recall that an inadequate end load (or equivalently an excessively high traction coefficient) can lead to excessive

slip between the variators and races. The constant end load would have to be chosen to avoid this even at the high traction coefficient end of the curve (at high negative ratios) and as a result the variator would be inefficient when working elsewhere on the curve, at lower ratios.

Line L ₁ represents the alternative strategy in which end load is proportional to variator ^" input torque. Ih this case traction coefficient is greatest at low negative variator ratios and the constant of proportionality of end load to input torque would

have to be chosen to avoid excessive slip at this end of the ratio range, but again the curve shows large traction coefficient variation with variator ratio.The result would necessarily be poor variator efficiency when running at higher ratios.

Line LR represents the strategy where end load is proportional to reaction

torque. The range of variation of traction coefficient with ratio is relatively small, and consequently the ratio of end load to reaction torque can be chosen to provide good variator efficiency throughout the ratio range. This is highly desirable and can be achieved, as already explained, in a hydraulically controlled variator.

Unfortunately, again as noted above, this strategy is difficult to achieve in less sophisticated variators using mechanical control.

However, refer now to line L _INV , which represents the variation of traction coefficient achieved using a strategy in accordance with the present invention. Note that it approximates to line L _R , and shows a similarly small range of traction

coefficient variation. This is achieved using an end load which is the sum of (a) a

first component proportional to the variator input torque and b) a second, constant component.

The physical construction of a variator which implements this end load control strategy will now be described with reference to Figures 4 to 8

This is a low cost variator intended to handle moderate torques. It has been developed for use in a ride-on lawnmower. However the invention could be applied

to transmissions for other types of vehicle and for handling higher power. The variator has an input race 100 with a part toroidal face 102 _f and an output race 104

with a part toroidal face 106. Together the races define a substantially toroidal cavity 108 about a common axis 110, defined by an input shaft 112. The input race 100,

best seen in Figures 5, 7and 8 has a through-going bore 114 through which it is

rotatably mounted upon a hollow spigot 116 of a thrust disc 118. The thrust disc 118

is splined to input shaft 112 and is prevented from moving axially by a flange 113 upon the shaft.

The variator' s output race 104 is mounted on a collar 111 which in turn is mounted upon the input shaft 112 though a bearing 120 . and rotates freely relative

to the shaft. Power take-off from the output disc is through a gear 113 formed as part

of the collar 111. In a complete transmission this gear typically couples to an epicyclic gear train through which drive is taken to motor vehicle wheels. However

such gear trains are well known in the art and will not be described herein.

This particular variator has two rollers controlled by a simple mechanical

lever mechanism, which will now be described for the sake of completeness.

However it is to be understood that other means of roller control can be adopted in

practice and the mechanism in question is presented by way of example and not limitation. The mechanism comprises a lever 130 whose outer end 132 extends outside variator casing 134 to couple to a control mechanism. Variator ratio is determined by the position of this lever. A flexible diaphragm 136 serves as a cover for an opening in the casing through which the lever emerges. The lever is pivoted around a pin 136 mounted in a web 138 formed as part of the casing. By virtue of a

slot 140 through which is engages with the pin, the lever is also able to move radially, for reasons which will shortly be explained. The lever has laterally extending arms 142, 144 on either side of the pin 136 which carry respective carriages 146, 148 in

which variator rollers 150, 152 are journalled through bearings 154, 156. The rollers

and their associated carriages are also able to turn as indicated by curved arrows in Figure 4b to change their inclinations to the variator axis, and thereby change variator drive ratio. As is well known,the rollers always seek an inclination at which then-

axes 158 intersect the variator axis 110. Moving the lever 132 moves the rollers circumferentially about the variator axis and transiently takes the roller axes away

from intersection with the variator axes. The rollers 150, 152 are consequently subject to a steering effect by the races 100, 104, changing their inchnations to restore

intersection of the aforementioned axes and producing a change in variator drive

ratio. As a result drive ratio is a function of lever position. A well known design issue

with toroidal variators, known as "equalisation" is addressed by virtue of the slot 140.

To explain, the rollers need some freedom to move relative to each other to equalise

the ratios at which the rollers are running (to allow e.g. for minor manufacturing tolerances in the mechanism) and so equalise the load the rollers carry. This freedom is provided in the illustrated mechanism by virtue of the freedom of the lever 130 to move radially through a small distance.

The illustrated variator has three different devices for biasing the races 100, 104 toward each other to provide roller/race traction:

(1) a ball ramp device 200;

(2) an end load spring 202 and

(3) a pre-load spring 204. These will now be considered in turn.

The ball ramp device 200 is formed by the thrust disc 118 and the input race 100. As Figures 5 and 7 show, the rear face of the input race 100 has a set of

circumferentially extending recesses 206. The illustrated embodiment has four such

recesses. The front face of the thrust disc has a corresponding set of recesses 208, best seen in Figure 6. Viewed in a sectional plane containing the races' common axis (Figures 4a and 8) the recesses are part-circular, to receive respective balls 210. Viewed along the radial direction (Figure 5) they each have a shallow "V" form. Hence each recess can be thought of as being formed by two helical regions or ramps

212, 214 of opposite pitch.

The input shaft 112 is driven from some rotary power source, which is typically an internal combustion engine but could in principle be an electric motor,

external combustion engine etc. If zero torque is applied to the input shaft 112, then the rotational position of the input race 100 relative to the thrust disc 118 is such that

the deepest regions 216, 218 of the respective recesses are aligned. The balls 210 lie

in these deepest regions and the separation of the input race from the thrust disc is minimised.

If a torque is then applied to the input shaft 112, this is directly transmitted to the thrust disc 118, but it is only transmitted to the input race 100 through the balls 210. The race 100 is caused to rotate through a small angle relative to the thrust disc

118. Because the deepest regions of the recesses are now mis-aligned, as in Figure 5, the ball 210 is forced away from these regions and rides up the ramp-like surfaces of the recesses, forcing the input race 100 away from the thrust disc 118. The thrust disc 118 cannot move because it is fixed upon the shaft 112. Hence the input race 100

is urged toward the output race 104, contributing to the end load. It will be apparent that the force applied by the ball ramp device 200 is related to the input torque. The

precise relationship depends on the shapes chosen for the recesses, but in the present embodiment, the force is substantially proportional to the input torque. Note also that by virtue of the "V" section of the recesses, input torque in either rotational direction (clockwise or anti-clockwise) creates a force in the same axial direction (toward the output race 104).

To keep the balls 210 properly located relative to each other, a registration disc 211 is positioned between the thrust disc 118 and the input race 100, receiving the balls in respective through-going holes 213.

The end load spring 202 acts upon the input race 100 to urge it toward the

output race 104. Its contribution to the end load is thus added to that of the ball ramp

device 200. In the present embodiment the end load spring 202 is mounted between the thrust disc 118 and the input race 100, serving to urge these two parts away from

each other. To accommodate relative rotation of the thrust disc 118 and the input race 100, the end load spring 202 acts upon the thrust disc through a thrust bearing 203. To accommodate relative rotation of the thrust disc 118 and the race 100, the

end load spring acts against a bearing 203. The end load spring 202 is formed as a resilient metal disc, known to those skilled in the art as a Belleville washer and widely commercially available. The characteristic of force against axial length of

such a spring typically does not follow Hooke's law closely but instead has a region

where force is substantially constant despite changes of length, and the spring is arranged to operate in this region so that its force can be taken to be approximately constant.

Of course, the end load spring 202 tends to separate the parts making up the ball ramp device - the input disc 100 and thrust disc 118. In the absence of sufficient

input torque to cause the ball ramp device to overcome the end load spring, this could cause the balls 210 to be left loose, creating unwanted backlash in the device. The pre-load spring 204 prevents this. It acts upon the output race 104 to urge it toward the input disc 100. Hence the pre-load spring 204 acts in opposition to the ball ramp

device 200 and the end load spring 202. As Figure 4 shows, it is mounted upon the

collar 111, being retained between a shoulder 205 formed upon the sleeve and the rear face of the output race 104. Note that the forces from the ball ramp device 200,

the end load spring 202 and the pre-load spring 204 are referred to the shaft, which

is thus in tension.

The force of the pre-load spring 204 is sufficient to overcome that of the end

load spring 202 acting on its own. Hence when input torque is zero, or has a sufficiently low value, the pre-load spring dominates, keeping the parts of the ball

ramp device 200 in engagement with each other. When sufficient input torque is applied to the variator, the total force exerted together by the ball ramp device and the end load spring 202 exceeds the force from the pre-load spring, and consequently the pre-load spring then contributes nothing to the end load.

The resultant end load/input torque characteristic is represented in Figure 9, where input torque is on the graph's horizontal axis and force on the vertical axis.

Line L _INP represents the end load component contributed by the ball ramp device 200. Dotted line L _ELS represents the end load component contributed by the end load spring 202. The solid line L _EL represents the actual end load. At low input torque the minimum end load value is determined by the pre-load spring at a value PLS. Away

from this range, end load increases monotonically with end load. Note however that this characteristic is not equivalent to that shown in Figure 2, since the end load outside the low input torque region is vertically offset by the force ELS contributed

by the end load spring. It is by virtue of this offset that the desirable traction coefficient/variator ratio characteristic already described with reference to Figure 3 is achieved.