This invention relates to friction clutches for motor vehicles, hi particular, but not exclusively, this invention relates to friction clutches for use in cars, more preferably high performance and/or racing cars. The invention also relates to a control mechanism for a friction clutch.

In a typical motor vehicle friction clutch, one or more friction plates or discs are positioned between a pressure plate and a flywheel or reaction plate. The flywheel and pressure plate are usually in driving connection with an output shaft of the engine and rotate about a common axis. One or more of the friction discs will be in driving connection with an output shaft of the clutch, which will often comprise an input shaft of an associated gearbox of the vehicle. Such clutches are typically engaged by means of a diaphragm spring that acts between a clutch cover and the pressure plate. The diaphragm spring biases the pressure plate towards a flywheel in order to clamp the friction discs between pressure plate and the flywheel. The amount of torque that can be transmitted by a clutch of any given diameter is determined in part by the clamp load that the diaphragm spring is able to exert. In order to increase the clamp load, and hence the torque that can be transmitted through a clutch of any given diameter, it is known to stack two or more diaphragm springs one on top of another. The diaphragm springs are all engaged and disengaged simultaneously and operate in effect as a single diaphragm spring means.

The types of friction clutches used in racing cars are such that pedal travel between a clutch release position and a clutch engaged position is very small. This in effect means that the clutch engagement is very sudden and sharp.

Effective control of clutch engagement is particularly important for a racing driver to ensure that power is transmitted to the driving wheels in a manner that enables them to pull away from the start line as quickly as possible. Sudden engagement of the clutch may result in excessive wheel spin or even stalling of the engine. Typically, a driver will wish to hold the clutch in a partially engaged condition during which time only part of the available engine torque is transmitted through the clutch, until the vehicle has gathered sufficient speed and the clutch can be fully

engaged. The position at which the clutch is held is sometimes referred to as the "bite point".

Holding a racing clutch on the bite point is difficult as only a relatively small amount of pedal travel can lead to a significant increase in clamp load and hence the torque transmitted through the clutch. This problem is exacerbated were the clutch incorporates so called carbon/carbon composite clutch discs, i.e. clutch discs formed from a carbon based matrix filled with a carbon based filer.

It is a property of carbon/carbon composite clutch discs that their coefficient of friction increases significantly as the temperature of the discs increases, until the discs reach a saturation temperature beyond which the coefficient of friction remains largely static or may even decrease slightly. It is a further property of carbon/carbon clutch discs that they expand as they heat up.

When a clutch is held in a partially engaged condition, the friction discs are not fully clamped and so slip relative to one another. As the discs rub against one another their temperature will increase. Where some or all of the friction discs are made of a carbon/carbon composite material, this increase in temperature will lead to an increase in the coefficient of friction of those discs and hence an increase in the amount of torque transmitted through the clutch, even if the clutch pedal is held stationary. At the same time, the carbon/carbon composite friction discs will expand as their temperature rises. As the discs expand they become more firmly clamped between the pressure plate and the flywheel and cause the clamp load exerted by the diaphragm spring to increase without any change in the clutch pedal position. The increase in coefficient of friction and expansion of the carbon/carbon friction discs happens very quickly and a driver may have insufficient time to react by releasing the clutch pedal in order to compensate for the increase in torque being transmitted through the clutch. This may lead to excessive wheel spin at the start of the race, to the engine being stalled if the engine speed is too low or even to a false start

Previous attempts to reduce the sudden engagement of the clutch have been disclosed in, for example, FR-A-2546591 in which three equi-spaced flexible spring plates bias a reaction plate away from a flywheel member to provide axial cushioning on clutch engagement. A further attempt is disclosed in the applicant's International

patent application WO 93/07400, in which a number of spring means are mounted in recesses in the flywheel to bias a friction disc away from the flywheel and cushion engagement the clutch.

Whilst cushioning of the clutch has been found to be effective, there is an ongoing need to improve the take-up of friction clutches.

It is an objective of the present invention to provide an improved friction clutch in which engagement can be more easily controlled than in prior art clutches.

It is a further objective of the present invention to provide an improved actuation mechanism for a friction clutch having two or more diaphragm spring means, which enables engagement of the clutch to be more easily controlled than with known clutch actuation mechanisms.

In accordance with a first aspect of the invention, there is provided a friction clutch for a motor vehicle as defined in claim 1.

Further particulars of the first aspect of the invention can be found in the claims dependent on claim 1.

In accordance with a second aspect of the invention, there is provided a clutch for a motor vehicle as claimed in claim 19.

Further particulars of the second aspect of the invention can be found in the claims dependent on claim 19. hi accordance with a third aspect of the invention, there is provided a clutch control mechanism as defined in claim 22.

Further particulars of the third aspect of the invention can be found in the claims dependent on claim 22.

Several embodiments of the invention will now be described, by way of example only, with reference to the following drawings in which:

Figure 1 is an axial cross-sectional view through a first embodiment of a clutch in accordance with the invention;

Figures 2A to 2C are cross-sectional views of part of the clutch of Figure 1 in which the clutch is shown in a released, partly engaged, and fully engaged condition respectively;

Figure 3 is a graph showing clamp load against clutch actuator travel for the clutch of Figure 1;

Figure 4 is a graph similar to that of Figure 3 but showing clamp load against clutch actuator travel for a known friction clutch in which multiple diaphragm springs are engaged and released simultaneously;

Figures 5 A to 5 C show schematically an arrangement of diaphragm spring means and clutch release fulcrums in accordance with a further embodiment of the invention in released, partly engaged and fully engaged positions respectively;

Figures 6A to 6C are views similar to Figures 5A to 5C but showing a yet further embodiment of the invention;

Figure 7 is a view similar to that of Figure 6C but showing a still further embodiment of the invention;

Figure 8 is a view similar to that of Figure 6C but showing a further embodiment of the invention in which an outer diaphragm spring means includes two diaphragm spring members; and

Figure 9 is a view similar to that of Figure 6C but showing a further embodiment of the invention in which a pivot ring is provided between two diaphragm spring means.

With reference to initially to Figure 1 , there is illustrated a multi-plate clutch assembly 10 which comprises a flywheel 11, housing 12 and cover 13. The housing is in the form of a series of circumferentially spaced lugs formed integrally with and projecting axially from the flywheel, in a manner well known in the art. The cover 13 is mounted to the free end of the lugs by means of bolts 14. The flywheel 11 is adapted to be fixed to a crankshaft of an associated engine (not shown). Mounted axially between the cover 13 and flywheel 11, are four drive discs of friction material 15a, 15b, 15c, 15d and a pressure plate 16. The drive discs and pressure plate are

mounted so as to be substantially rotationally fast with the housing 12 but slidable axially relative thereto. Three driven friction discs 17a, 17b, 17c are interleaved between the drive discs 15a, 15b and 15b, 15c and 15c, 15d respectfully. The driven discs 17a, 17b, 17c are rotationally fast with a central hub 5 which has a splined bore 6 by means of which drive can be transmitted between the hub and an output shaft (not shown) which may be an input shaft of an associated gear box (also not shown). Two diaphragm or belleville spring means 18, 19 bias the pressure plate towards the flywheel, so as to clamp the drive and driven discs together between the pressure plate and the flywheel. In the present embodiment, each of the diaphragm spring means 18, 19 comprises a single diaphragm spring member. However, in certain embodiments, any one or both of the diaphragm spring means 18, 19 may include two or more diaphragm spring members.

The diaphragm spring means 18, 19 each have a solid annular outer region 20 with a plurality of spaced fingers 18a, 19a projecting radially inwardly from the inner diameter of the outer annular region. An outer surface of the outer diaphragm spring means 19 contacts a fulcrum ring 22 on the cover 13 at position close to the outermost diameter of the springs 18, 19. An inner surface of the inner diaphragm spring means 18 contacts a fulcrum ring 23 on the pressure plate 16 at a position radially inwards from the fulcrum ring 22 on the clutch cover 13. The arrangement is such that when the clutch is fully engaged, the diaphragm spring means 18, 19 act between the clutch cover 13 and the pressure plate 16 to bias the pressure plate towards the flywheel 11 to clamp the drive discs 15a, 15b, 15c, 15d and the driven discs 17a, 17b, 17c together between the pressure plate 16 and flywheel 11 in a manner well known in the art. A control mechanism for the clutch includes a clutch control sleeve 24 mounted concentrically about a shaft (not shown) that is in driving connection with the driven discs 17a, 17b, 17c via the hub 5. The control sleeve has two clutch release fulcrums 25, 26. A first release fulcrum 25 engages on an inner surface of the fingers 18a of the inner diaphragm spring means 18 whilst a second release fulcrum 26 engages on an inner surface of the fingers 19a of the outer diaphragm spring means 19.

The clutch control sleeve 24 is operatively connected with a clutch control actuator (not shown) such that movement of the actuator causes the sleeve 24 to move axially relative to the flywheel 11 to move the diaphragm spring means between an engaged position, in which a full spring or clamp load is applied to the pressure plate 16 by the diaphragm spring means 18, 19, and a released position, in which the spring load is removed from the pressure plate, as will be described in more detail later. In an alternative embodiment, the control sleeve 24 may be an integral part of the actuator.

The clutch control actuator may be a hydraulic slave cylinder operated by a driver of a vehicle to which the clutch assembly is mounted by means of a clutch pedal and a hydraulic master cylinder, in a manner well known in the art. However, the actuator may be any suitable type of actuator and may be operated by mechanical means such as a cable or may be an electronic actuator. The actuator may also be controlled by or through an electronic/programmable control system rather than directly by the driver. As shown in Figure 1, the radial fingers 18a of the inner diaphragm spring means 18 are shorter than the fingers 19a of the outer diaphragm spring means 19. Accordingly, the inner or leading release fulcrum 25 is spaced radially outwardly from the outer or trailing release fulcrum 26, so as to contact the fingers 18a on the inner diaphragm spring means 18. The two release fulcrums 25, 26 on the control sleeve 24 are spaced axially by a distance X which is greater than the axial thickness of the inner diaphragm spring means 18, for reasons that will be explained below.

Operation of the clutch assembly 10 will now be described with reference to Figures 2A to 2C.

Figure 2 A shows the clutch 10 with the diaphragm spring means 18, 19 in a fully released position. The control sleeve 24 has been withdrawn by the clutch actuator in a clutch disengagement direction as indicated by arrow A, so that the fulcrums 25, 26 engage with the fingers 18a, 19a of their respective diaphragm spring means 18, 19 to lift the springs away from the pressure plate 16. In this condition, no clamp load is applied to the pressure plate by the diaphragm spring means 18, 19. It will be noted that due to the axial spacing of the control sleeve release fulcrums 25,

26, there is a gap between the two diaphragm spring means except at their outermost

diameter region where the inner diaphragm spring means 18 contacts the outer diaphragm spring means 19, which in turn is in contact with the clutch cover fulcrum ring 22.

Figure 2B shows the clutch in a partially engaged state in which the control sleeve 24 has been moved towards the flywheel 11 in a clutch engaging direction (as indicated by the arrow B) by the clutch actuator. The inner diaphragm spring means 18 has been brought into contact with the fulcrum ring 23 on the pressure plate to apply a spring or clamp load to the pressure plate. Figure 2B illustrates the clutch 10 at a point at which the full spring load of the inner diaphragm spring means 18 is being applied to the pressure plate 16. At this stage, the leading release fulcrum 25 associated with the inner diaphragm spring means 18 is no longer supporting the inner fingers 18a of the inner diaphragm spring means 18. However, because the trailing release fulcrum 26 associated with the outer diaphragm spring means 19 is spaced axially from the leading fulcrum 25 by a distance greater than the thickness of the inner diaphragm spring means 18, the fingers 19a of the outer diaphragm spring means 19 are still supported on the trailing fulcrum 26 and there is a gap between the two diaphragm spring means 18, 19 at the radius of the fulcrum ring 23 on the pressure plate. As a result, the outer diaphragm spring means 19 does not apply a significant spring load to the pressure plate 16. Figure 2C shows the clutch 10 fully engaged. The clutch control sleeve 24 has moved closer to the flywheel 11 in the clutch engaging direction B to bring the outer diaphragm spring means 19 into engagement with the inner diaphragm spring means

18 so that the outer diaphragm spring means 19 applies a spring load to the pressure plate. Figure 2C illustrates a point at which the outer diaphragm spring means 19 is applying its full spring load to the pressure plate 16 and the clutch is fully engaged.

To disengage the clutch, the above sequence is reversed.

It will be appreciated from the above that the two diaphragm spring means 18,

19 operate sequentially during engagement of the clutch, each applying its spring load o the pressure plate 16 over different phases of engagement of the clutch.

During an initial phase of engagement, as the control sleeve 24 moves from the position shown in Figure 2A to the position shown in Figure 2B, the inner diaphragm spring 18 means is brought into contact with the fulcrum ring 23 on the pressure plate and its spring load is gradually applied to the pressure plate. This continues until the position shown in Figure 2B is reached where the pressure plate 16 is subject to the full spring load of the inner diaphragm spring means 18. During this phase, the gap between the two diaphragm spring means 18, 19 ensures that no significant spring load is applied to the pressure plate, 16 by the outer diaphragm spring means 19. As the control sleeve 24 continues to move in the clutch engagement direction

B from the position shown in Figure 2B towards the position shown in Figure 2C, there is an intermediate phase during which the gap between the two diaphragm spring means 18, 19 is taken up. During this intermediate phase, the full spring load of the inner diaphragm spring means 18 continues to be applied to the pressure plate 16 but no significant spring load from the outer diaphragm spring means 19 is applied. Thus, there is a period of travel of the clutch control sleeve 24, and hence of travel of the clutch actuator, during which the pressure plate 16 is subjected to a relatively constant spring load from the diaphragm spring means.

Once the gap between the two diaphragm spring means 18, 19 has been taken up, the spring load of the outer diaphragm spring means 19 begins to be applied to the pressure plate 16. This represents the start of a further, and final in this case, phase of engagement as the spring load from the outer diaphragm spring means 19 is gradually applied to the pressure plate 16 until the clutch reaches the fully engaged position as shown in Figure 2C, in which the full spring load of both diaphragm spring means 18, 19 is applied to the pressure plate.

Figure 3 is a typical graph showing the clamp load exerted by the diaphragm spring means 18, 19 against the clutch actuator travel between a clutch released position at A to a fully engaged position at D. Line B indicates the position shown in Figure 2B at which the full clamp load of the inner spring means 18 is applied to the pressure plate 16 and line C indicates the position at which the gap between the two spring means has been closed and the spring load of the outer spring means 19 starts

to be applied to the pressure plate. The three phases of clutch engagement described above can be seen from the graph with the initial phase extending between points A and B, the intermediate phase between points B and C and the further or final phase between points C and D. It will be noted that the clamp load on the pressure plate 16 is not strictly constant between points B and C. This is believed to be due to the interaction of the two spring means 18, 19 where they contact one another at the outer diameter. Nevertheless, the rate of increase in spring load applied to the pressure plate by the diaphragm spring means over the intermediate phase is low compared with the initial and further phases. A more constant spring load over the intermediate phase may be achieved by separating the diaphragm spring means 18, 19 by means of a fulcrum ring.

Figure 4 is a graph similar to that of Figure 3 but showing clamp load against clutch actuator travel for a conventional clutch with multiple diagram spring means that are engaged simultaneously. Position A indicates the clutch when fully released and D the clutch fully engaged. Line B indicates a typical bite point at which a driver will attempt to hold the clutch in order to feed the drive to wheels effectively during a racing start. As can be seen from the graph, the rate of change in clamp load for a given actuator travel is very high and hence even a slight movement in actuator travel from line B to line C results in a significant increase in clamp load.

In a clutch in accordance with the invention, the ability to apply the spring loads of the two spring means 18, 19 over different phases of the clutch engagement, and in particular to be able to provide an intermediate phase during which there is little or no increase in clamp load for a given period of clutch actuator travel, provides for greater control in clutch take up. By appropriate selection of the properties of the inner diaphragm spring means 18, it can be arranged that the clamp load applied over the intermediate phase is substantially equal to the bite position at which the driver wishes to hold the clutch in order to feed the engine power to the wheels in the most effective way. hi addition, by varying the axial spacing between the clutch release fulcrums 25, 26 and hence the gap between the diaphragm spring means 18, 19 when the clutch is released, the length of the intermediate phase can be varied to provide the

driver with a larger target area of pedal movement in which to hold the clutch on the bite point. This makes clutch control easier and more predictable.

In the clutch shown in Figures 1 to 3, the fulcrum ring 22 on the clutch cover is positioned close to the outside diameter of the first and second diaphragm spring means 18, 19 where the spring means contact one another. In a modified embodiment, the outer diameter of the first and second diaphragm means may lie further outside the fulcrum ring 22 such that when the clutch is disengaged, the point of contact between the first and second diaphragm spring means lies outside the radius of the clutch cover fulcrum ring. Figures 5 to 9 illustrate further embodiments of the invention. Only the diaphragm spring means and the release fulcrums are shown in these drawings for clarity.

Figures 5A to 5C illustrate an alternative release fulcrum arrangement, in which the fulcrums 25', 26' are located on a common radius. To accommodate this arrangement, the fingers 18a' of inner diaphragm spring means 18' are dished inwardly. Figures 5A, 5B, and 5C show the spring means and fulcrums in released, partly engaged and fully engaged positions respectively, corresponding to Figures 2A to 2C described above. Engagement and disengagement of the clutch is effected in a manner similar to that as described above in relation to the clutch 10 shown in Figures 1 to 3.

In the embodiments described so far, the release fulcrums 25, 26; 25', 26' are provided on a common clutch control sleeve 24, 24' for simultaneous movement.

Figures 6A to 6C illustrate an embodiment in which the fulcrums 25", 26" are provided on separate control sleeve members 27, 28 so as to be capable of independent movement. hi Figure 6A the clutch is in a released condition and both control sleeves 27, 28 have been withdrawn in a clutch disengaging direction so that the two diaphragm spring means 18", 19" apply no spring load to the pressure plate.

In Figure 6B, the control sleeve 27 for the inner diaphragm spring means 18" has been moved forward in a clutch engaging direction so that the full spring load of the inner spring 18" is applied to the pressure plate.

In Figure 6C, the control sleeve 28 for the outer diaphragm spring means 19" has been moved forward in a clutch engaging direction so that the full spring load of the outer spring means 19" is also applied to the pressure plate and the clutch is fully engaged.

By arranging for an appropriate delay between the inner spring means 18" being fully engaged with the pressure plate and the outer spring means 19" being brought into full engagement with inner spring 18", an intermediate, generally constant spring load phase of clutch engagement can be provided for.

This embodiment provides for a greater degree of control over clutch engagement as it will be possible to adjust the length of the various phases of clutch engagement. However, the clutch actuation system will be more complex as two actuators will be required, one for each control sleeve 27, 28. Where the actuators are hydraulic actuators, movement of the two control sleeves may be controlled by means of a hydraulic control system. The actuators may be, for example, in the form of two concentric hydraulic actuators. However, any suitable form of actuator may be used. This arrangement may be particularly suited to applications in which engagement of the clutch is controlled by means of an electronic and/or programmable controller.

Figures 7 to 9 illustrate some further alternative embodiments, all are shown in the engaged position only. In all these embodiments, independent control sleeves are provided and it should be understood that these embodiments operate in a manner similar to that described in relation to Figures 6A to 6C above. It should also be understood that the arrangements shown in Figures 7 to 9 could be adapted for use with a single release sleeve if desired.

Figure 7 shows an embodiment similar to that shown in Figure 6C except the fingers 18a'" of the inner spring means 18'" are shorter than the fingers 19a'" of the outer spring means 19'" and are not dished. The release fulcrum 25'" for the inner

spring means operates at a larger radius than the release fulcrum 26'" for the outer spring means 19'", in a manner similar to that of the clutch 10 shown in Figure 1.

Figure 8 illustrates an embodiment in which the outer diaphragm spring means

19"" comprises two diaphragm spring members 29, 30 acting together as a single diaphragm spring means. It should be appreciated either of the inner or outer diaphragm spring means could comprise two or more diaphragm spring members operating in this way in any of the embodiments disclosed. Where a diaphragm spring means includes more than one diaphragm spring member, it may be the case that only one of the members has radially inwardly directed fingers for co-operation with the clutch release fulcrum.

Figure 9 shows an embodiment similar to that described above in relation to

Figures 6A to 6C, except that an additional fulcrum ring 31 is provided between the inner diaphragm spring means 18'"" and the outer diaphragm spring means 19'"".

Again, it should be understood that a fulcrum ring can be used between the diaphragm spring means in any of the embodiments described above.

Although the invention has been described in relation to multi-plate clutches for use in racing cars, the invention is not limited to this application but can be employed in any type of friction clutch in which the clamp load is provided by more than one diaphragm spring means. For example, the invention could be applied to clutches having only a single driven disc or plate or to clutches having a twin driven disc or plate arrangement. In this regard, it should be noted that the terms "flywheel" and "clutch cover" as used herein, including in the claims, are not intended to imply any particular constructional limitation. Accordingly, the term flywheel should be understood as encompassing any suitable reaction member operatively connectable with an output shaft of an engine or motor regardless of whether it is used to store energy from the engine or motor. Similarly, the term clutch cover should be understood as encompassing any suitable member fixed axially relative to the flywheel and rotationally fast therewith and against which the diaphragm spring means can react. Whilst the embodiments described herein have all been pull type clutches, by suitable adaptation, the invention may also be applied to push type clutches.

The invention provides a clutch arrangement in which spring load to actuation travel characteristics can be varied in a number of different ways to suit a variety of applications. Whilst the provision of an intermediate phase of generally constant spring load (or at least a having a low rate of change of spring load) is considered desirable for many applications, there may other applications where no intermediate phase is required. In this case, the clutch may be arranged so that the spring load of the outer diaphragm spring means is applied to the pressure plate as soon as the full spring load of the inner diaphragm spring means has been applied or even before this.

Whereas the invention has been described in relation to what are currently considered to be the most practicable and preferred embodiments, it should be understood that the invention is not limited to the disclosed arrangements but rather is intended to cover various modifications and equivalent constructions included within the scope of the invention as defined in the claims. For example, the invention is not limited to application with clutches for racing cars but can be used in any vehicle where it is desirable to control clutch take up in the manner provided.

Where the terms "comprise", "comprises", "comprised" or "comprising" are used in this specification, they are to be interpreted as specifying the presence of the stated features, integers, steps or components referred to, but not to preclude the presence or addition of one or more other feature, integer, step, component or group thereof.