WO2008145959A1

WO2008145959A1 - Clutch driven plates

Info

Publication number: WO2008145959A1
Application number: PCT/GB2008/001687
Authority: WO
Inventors: Paul Andrew Gallagher; Danilo Galeotti; Nicola Mazzanini
Original assignee: Automotive Products S.P.A.
Priority date: 2007-05-31
Filing date: 2008-05-16
Publication date: 2008-12-04

Abstract

A clutch driven plate (10) has a central hub (11) and an outer assembly (13) carrying at least one friction lining (19), the hub being rotationally displaceable relative to the outer assembly against the action of a plurality of circumferentially spaced spring means (36) carried by the outer assembly via a linkage (24) in order to damp rotational vibrations between the hub and outer assembly. Each spring means (36) is compressed between a fixed end cap (41) and a moveable end cap (38) which is pivotally connected (37) with an end of a link (25) of the linkage (24) so that the moveable end cap can pivot relative to the link to compress the spring means in as linear a manner as possible.

Description

CLUTCH DRIVEN PLATES

The present invention relates to clutch driven plates.

Known driven plates comprise a central hub for mounting on a drive shaft, such as an input shaft of an associated vehicle transmission, and an outer assembly arranged concentrically about the axis of rotation of the hub. Typically, the outer assembly carries at least one friction lining with the outer assembly being rotationally displaceable relative to the hub against the action of spring means in order absorb shocks which may occur during clutch take-up and to dampen torsional vibrations which arise in the associated vehicle drive line

Typically the spring means comprises a number of coil springs or sets of springs arranged circumferentially about the driven plate. The springs will normally be located in aligned windows in a flange which rotates with the hub and a pair of side plates arranged either side of the flange, which rotate with the outer assembly. The maximum range of relative rotational movement between the hub and the outer assembly is limited by the length of the springs which in turn is limited by the length of the windows. For a driven plate of any given diameter, the maximum length of window which can be used is determined by the dimensions of the driven plate and the need to maintain structural integrity. Thus the amount of relative rotational movement between the hub and outer assembly is limited leading to the provision of less progressive spring damping.

It is an objective of the present invention to provide an improved clutch driven plate.

Thus, in accordance with the present invention, there is provided a clutch driven plate comprising a central hub and an outer assembly carrying at least one friction lining the hub being rotationally displaceable relative to the outer assembly against the action of a plurality of circumferentially spaced spring means carried by the outer assembly via a linkage in order to damp rotational vibrations between the hub and outer assembly each the spring means being each compressed between a fixed end cap and a moveable end cap which is pivotally connected with an end of a link of the linkage so that the moveable end cap compresses the spring in as linear a manner as possible.

The moveable end cap and the end of the link preferably have cooperating abutments which limit the relative pivoting of the moveable end cap and link.

The outer assembly preferably includes a pair of spaced outer plates with the linkage mounted between the outer plates.

In a preferred construction each spring means is mounted in a pair of corresponding windows in the outer plates and a plurality of linkages connect the hub and the outer assembly, each linkage including a first link which moves a respective moveable end cap.

Conveniently each first link is pivotally mounted adjacent one end on the outer plates and acts at its other end on the associated moveable spring end cap, the first link being connected intermediate its ends with the hub by a second link pivoted adjacent one end on the first link and adjacent the other end on the hub so that as the hub rotates relative to the outer assembly the first link pushes on the moveable end cap to compress the spring means.

Abutments on the first link and hub come into contact to limit the relative rotation of the hub and outer assembly under high torque loads with the driven plate operating in a drive or overrun condition.

Preferably the spring means are coil springs and the spring end caps have peripheral flanges which surround the ends of the coils to retain the spring means on the end caps.

The end caps may also include central projections which extend within the spring coil means to further help retain the spring means on the end caps. The linkage is arranged so that the spring means is compressed in the same direction by the link no matter whether the driven plate is operating in a drive or overrun condition.

A low friction buffer is preferably provided radially outward of the spring means to limit radially outward deformation of the spring means under centrifugal forces when the driven plate is rotating.

The present invention also provides a driven plate for a friction clutch comprising a central hub and an outer assembly carrying at least one friction lining the hub being rotationally displaceable relative to the outer assembly against the action of a spring means carried by the outer assembly via a linkage in order to damp rotational vibrations between the hub and outer assembly, the spring means comprising coil springs whose axial spacing between adjacent coils varies along the length of each spring to provide a non-linear load/deflection characteristic.

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

Figure 1 is an axial view of a driven plate in accordance with the invention taken in the direction of arrow A on figure 2;

Figure 2 is a cross-sectional view of the driven plate of figure 1 taken on the line X- X;

Figure 3 is an enlarged view of the circled portion of figure 2;

Figure 4 shows part of figure 1 with part of one of the side plates cut away to show details of the internal linkage used;

Figures 5, 6 and 7 show the linkage of figure 4 in its zero torque, drive and overrun conditions respectively; Figures 8 and 9 show abutments which limit the relative rotation between the hub and outer assembly of the driven plate;

Figures 10 and 11 show abutments which control the pivoting of spring and caps and the associated links;

Figures 12 and 13 show details of a low friction buffer used to control radially outward deflection of the damping springs under centrifugal forces.

Figure 14 shows details of part of a linkage provided with self-lubricating bushes on its pins, and

Figure 15 shows a section on the line Y-Y of figure 14.

With reference to the drawings a driven plate for a friction clutch of a motor vehicle is indicated generally at 10. The driven plate comprises a hub 11 and an outer assembly 13 arranged concentrically about the axis of rotation 14 of the hub.

The hub 11 has a central bore with internal splines 15 for mounting to a shaft (not shown) of an associated transmission of the vehicle. The outer diameter of the hub 11 has a central radial flange 17 which has three radially extending, equi-spaced, lugs 18.

The outer assembly 13 comprises a pair of friction linings 19 riveted at 20. to an annular spring segment 21 which is itself secured to a pair of side plates 22a, 22b by rivets 23 which pass through holes 22c and 22d in side plates 22a and 22b respectively. Side plate 22a is supported on hub 11 via a plastics bush 27 whilst side plate 22b is directly supported on hub 11.

Each of the side plates 22a, 22b is generally annular in shape and has three circumferentially spaced windows 35 which align with corresponding windows in the other of the side plate for receiving spring means 36. The outer assembly 13 is connected to the hub 11 by means of three linkages 24 arranged between the two side plates 22a, 22b, one of which is visible in Figure 4 to 6. All of the linkages are the same and so only one linkage will be described in detail.

Each linkage 24 comprises a first link member 25 that is pivotally mounted between the side plates 22a, 22b by means of a rivet 26 (see Figures 2 and 3). The first link member 25 is connected to the hub 11 by means of a pair parallel links 28 which are arranged one on either side of the first link member 25. A first end of the links 28 is pivotally attached to first link member 25 by means of a pin 29, whilst the other end of the links 28 is similarly attached, by means of a pin 30, to one of the radially extending lugs 18 on the hub. As can be seen most clearly from Figure 3, spacer washers 31 , and 32 located between links 28 and the side plates 22a and 22b, and shims 31a and 32a between links 28 and link 25 hold pin 29 in position. Pin 30 is similarly located by the bush 27, the spacer washer 33 and the shims 33a and 27a.

Each first link member 25 is also connected via a pin 37, with a moveable spring end cap 38 which slides in windows 35 and which engages a first end of an adjacent spring means 36. The pin 37 engages slots 39 in ears 40 on end cap 38. The other end of the spring means 36 is in engagement with a fixed end cap 41 received in the end portions of aligned windows 35 in the side plates 22a, 22b. Each of the spring means 40 comprising a pair of concentric spring coils 42, 43.

The arrangement is such that when the hub 11 rotates relative to the outer assembly 13 from a neutral position (as shown in Figures 4 and 5), the link member 25 is pivoted about the rivet 26 in a clockwise direction (see arrow C) so as to compress the spring means 36 (see figure 6). The geometry of the linkages is arranged so that initially, a large relative movement between the outer assembly and the hub away from the neutral position results in a comparatively small compression of the spring means 36. Hence the spring will be loaded only lightly. However, the amount by which the spring is compressed for a given amount of movement between the outer assembly and the hub increases as the outer assembly and hub move further relative to one another from the neutral position. Consequently the loading on the spring increases non-linearly as the outer assembly moves relative to the hub further from the neutral position.

As will be appreciated the hub 11 can rotate relative to the outer assembly 13 in either direction from the neutral position depending on whether the driven plate is being operated under drive or overrun conditions. Thus, if the normal direction of rotation of the driven plate is in the direction of arrow D as shown in Figure 1 then under drive conditions, that is when torque is being transmitted from the outer assembly to the hub, the hub will tend to rotate in a clockwise direction (as shown in Figure 6) relative to the outer assembly. However, when the driven plate is being operated in overrun conditions, that is when torque is being transmitted from the hub to the outer assembly, the hub will tend to rotate anti-clockwise (as shown in Figure 7) relative to the outer assembly.

As can be seen from Figure 6 and 7 the spring means 36 is compressed in the same sense (see arrow E) no matter whether the driven plate is operating under drive or overrun conditions.

It is an important feature of the driven plate that under drive conditions relative rotation between the outer assembly 13 and the hub 11 is stopped by contact between the sides 18a of projections 18 and the edges 25a of link members 25 (see Figure 8) and in the overrun condition the sides 18b of the projection 18 contact the edges 25b of the link members 25 (see Figure 9). This arresting of the relative rotation of the outer assembly 13 and the hub 11 is arranged to occur before the spring means 36 are completely compressed so that the characteristics of the spring means remain predictable and the durability of the spring means is increased. Thus any very high torque loads are transmitted directly between the outer assembly to the hub by contact between sides 18a and edges 25a or sides 18b and edges 25b.

In accordance with the present invention the moveable spring end cap 38 is free to pivot to a limited extent (see angle X of Figure 10) relative to link member 25 about means 36 is compressed by link member 25 in as linear a manner as possible, [link 25 applies the force to spring means 36 as parallel to the axis of spring means 36 as possible]. This pivoting of end cap 38 relative to link 25 also avoids excessive friction between end cap 38 and the windows 35 in side plates 22a and 22b.

The end 50 of link member 25 has an abutment surface 50a (see Figure 10) which contacts the surface 38a on end cap 38 to limit the pivoting of the end cap 38 relative to link 25 about pin 37 during use of the driven plate (see Figure 11). This is particularly useful under low torque and high rotational speeds to avoid excessive deformation of the springs and unwanted contact with the side plates. Spring end caps 38 and 41 both have flanges 38b and 41b which surround the outer spring coil

42 and projections 38c and 41c which project inside inner coil 43. This again helps with the linear compression of the spring means 36 and helps to reduce unwanted contact between to springs and the side plates.

Additionally, as shown in Figures 12 and 13, each spring means 36 is provided with a low friction spring buffer 45 in the form of a block of plastics material or a PTFE coated metal strip or similar which is secured between the side plates 22a and 22b and which has a curved inner surface 45a which contacts the outer surface of the coils 42 of the spring means if the spring means tend to bow outwardly under centrifugal forces during use of the driven plate. By preventing excessive bowing of the springs unwanted hysterisis and permanent setting of the springs is avoided. The buffer allows compression of the spring coils by link 25 without excessive rubbing or binding of the coils on the buffer.

A further important feature of the driven plate design is that the spring coils 42 and

43 are preferably arranged to have a non-linear compression characteristic or spring rate. This is achieved by arranging the spacing between the coils of the springs to be non-linear. This non-linearity is used to further reduce the initial spring rate whilst ensuring that at higher torque loads the necessary spring damping is still achieved. In a typical example of a driven plate with a total torque capacity of say 240 Nm (for use with an engine which will deliver 200 Nm maximum) the outer spring coil 42 of each spring would have:-

Free length 62mm

Outer Diameter 24 mm

Wire Diameter 4.8 mm

Stiffness for the first part of the deflection of the spring of 9.6 kg/mm up to a

Linear deflection of the spring of 10.8 mm, equivalent to 27° relative rotation when fitted in the driven plate of the present invention

Stiffness for the second part of the deflection of the spring of 2 kg/mm up to a

Linear deflection of the spring of 13.2 mm, equivalent to 33° relative rotation when fitted in the driven plate.

The inner spring coil 43 of each spring would have:- Free length 62 mm Outer Diameter 13.7 mm Wire Diameter 2.9 mm

Stiffness for the first part of the deflection of the spring of 4 kg/mm up to a Linear deflection of the spring of 10.8 mm, equivalent to 27° relative rotation when fitted in the driven plate

Stiffness for the second part of the deflection of the spring of 9 kg/mm up to a Linear deflection of the spring of 13.2 mm, equivalent to 33° relative rotation when fitted in the driven plate.

The two springs together have a total stiffness in the first part of the deflection of the spring of 13-6 kg/mm (up to 27°) and in the second part of the deflection of 31 kg/mm (from 27° to 33°).

Using springs with non-linear stiffness characteristics the ratio of the stiffness in the first part of the deflection to the stiffness in the second part of the deflection is typically 1 :2 or 1 :2-5. In an alternative arrangement instead of having springs whose stiffness varies non- linearly with deflection the inner spring can be shorter than the outer spring so that in the first part of the deflection only the outer spring is compressed and in the second part of the deflection both springs are compressed to provide a higher stiffness.

In a typical example of such an arrangement for a driven plate with a total torque capacity of say 240 Nm the outer spring would have:-

Free length 62 mm

Outer Diameter 24 mm

Wire Diameter 4.9 mm

Stiffness of the spring 10 kg/mm up to a Linear deflection of the spring of

14.3 mm, equivalent to 35° of relative rotation when fitted in the driven plate of the present invention.

The inner spring would have:- Free length 49 mm Outer Diameter 13.5 mm Wire Diameter 3.2 mm

Stiffness of the spring 30 kg/mm with a linear deflection of the spring of 1.3 mm, equivalent to the last 3° of relative rotation (from 32° to 35°) when fitted to the driven plate.

The two springs together therefore have a total stiffness of 10 kg/mm of relative rotation (up to 32°) and 40 kg/mm for the last 3° of relative rotation (32° to 35°).

Using a shorter inner spring the ratio of the stiffness in the first part of the deflection to the stiffness in the second part of the deflection can be larger up to say 1 :3-5.

By appropriate selection of the rate of the spring means 36 and the geometry of the linkages 24 it can be arranged that the damping torque is low over an initial range of deflection between the hub 11 and outer assembly 13, particularly at low speeds of rotation, to dampen out low level vibrations which occur at idle speeds. The damping torque can also be arranged to increase as the deflection increases to provide the higher levels of damping required to damp clutch take up and the higher shock loads which can occur at higher engine speeds.

Also the linkage 24 allow larger circumferential rotation between the hub 11 and outer assembly 13 than would occur if the spring means simply operated in windows associated with the hub and outer assembly respectively. This larger rotation between hub and outer assembly allows more progressive control of the variation of the damping torque provided by the spring means.

Figures 14 and 15 show views of parts of the linkage 24 in which the pins 29 and 30 are provided with self-lubricating bushes 29a,29b and 30a, 30b to reduce the friction between the pins and links 25 and 28 etc. to lower the hysteresis of the system. These bushes can be made, for example, from so-called DU or DP4 material supplied by Glacier Girlock Bearings and typically have a steel support layer of say 1mm thick and a porous bronze sintered layer which is overlaid and impregnated with a PTFE coating of several microns thickness which contains inorganic fillers and polymer fibres and which should give a coefficient of friction of between 0.1 and 0.2.

These bushes could be replaced by surface treatment of the pins using, for example, a carbon based coating such as that commercially known as Trio Bond 40 supplied by IonBond.

Claims

1. A clutch driven plate comprising a central hub and an outer assembly carrying at least one friction lining, the hub being rotationally displaceable relative to the outer assembly against the action of a plurality of circumferentially spaced spring means carried by the outer assembly via a linkage in order to damp rotational vibrations between the hub and outer assembly, each spring means being each compressed between a fixed end cap and a moveable end cap which is pivotally connected with an end of a link of the linkage so that the moveable end cap can pivot relative to the link to compress the spring means in as linear a manner as possible.

2. A driven plate according to claim 1 in which the moveable end cap and the end of the link have cooperating abutments which limit the relative pivoting of the moveable end cap and link.

3. A driven plate according to claim 1 or 2 in which the outer assembly includes a pair of spaced outer plates with the linkage mounted between the outer plates.

4. A driven plate according to claim 3 in which each of the spring means is mounted in a pair of corresponding windows in the outer plates and a plurality of linkages connect the hub and the outer assembly, each linkage including a first link which moves a respective moveable end cap.

5. A driven plate according to claim 4 in which each first link is pivotally mounted adjacent one end on the outer plates and acts at its other end on the associated moveable spring end cap, the first link being connected intermediate its ends with the hub by a second link pivoted adjacent one end on the first link and adjacent the other end on the hub so that as the hub rotates relative to the outer assembly the first link pushes on the moveable end cap to compress the spring means.

6. A driven plate according to claim 5 in which one or more pins which pivotally interconnect the linkage are provided with self lubricating bushes or coating to reduce the hysteresis of the linkage.

7. A driven plate according to claim 5 or 6 in which abutments on the first link and hub come into contact to limit the relative rotation of the hub and outer assembly under high torque loads with the driven plate operating in a drive or overrun condition.

8. A driven plate according to any one of claims 1 to 7 in which the spring means are coil springs and the spring end caps have peripheral flanges which surround the ends of the coils to retain the spring means on the end caps.

9. A driven plate according to claim 8 in which the end caps include central projections which extend within the spring coil means to further help retain the spring means on the end caps.

10. A driven plate according to any one of claims 1 to 9 in which spring means are coil springs and the axial spacing of adjacent spring coils varies along the length of each spring to provide a non-linear load/deflection characteristic.

11. A driven plate according to any one of claims 1 to 9 in which the spring means are coil springs arranged in pairs one inside the other with one spring or each pair being shorter than the other and only being compressed at higher angles of relative rotation between the hub and outer assembly to provide a higher spring stiffness.

12. A driven plate according to any one of claims 1 to 11 in which the spring means is compressed in the same direction by the link no matter whether the driven plate is operating in a drive or overrun condition.

13. A driven plate according to any one of claims 1 to 12 in which a low friction buffer is provided radially outward of the spring means to limit radially outward deformation of the spring means under centrifugal forces when the driven plate is rotating.

14. A driven plate for a friction clutch comprising a central hub and an outer assembly carrying at least one friction lining the hub being rotationally displaceable relative to the outer assembly against the action of a spring means carried by the outer assembly via a linkage in order to damp rotational vibrations between the hub and outer assembly, the spring means comprising coil springs whose axial spacing between adjacent coils varies along the length of each spring to provide a non-linear load/deflection characteristic.

15. A clutch driven plate constructed and arranged substantially as hereinbefore described with reference to and as shown in the accompanying drawings.