WO2003031840A1

WO2003031840A1 - Torsional vibration dampers

Info

Publication number: WO2003031840A1
Application number: PCT/GB2002/004476
Authority: WO
Inventors: Richard David Maitland Cooke
Original assignee: Autmotive Products Group Limited
Priority date: 2001-10-10
Filing date: 2002-10-03
Publication date: 2003-04-17
Also published as: EP1434954A1; GB0124246D0

Abstract

A torsional vibration damper (10) for a vehicle driveline which includes an input member (11) for connection with a vehicle engine and an output member (12) for connection with a vehicle drive line. The input and output members are relatively rotatable against a damping means to absorb/alternate torsional vibrations emanating from the engine. The damping means includes a friction damping device (60) comprising two or more friction members (61, 62) axially biased into frictional contact by a belleville spring (64) having a generally flat load/deflection characteristic over its operational range, the belleville spring being capable of applying axial loading to the friction members whilst positioned on both sides of its flat condition.

Description

TORSIONAL VIBRATION DAMPERS

This invention relates to vehicle driveline torsional vibration dampers and in particular to a torsional vibration damper (of the type specified) which includes an input member for connection with a vehicle engine and an output member for connection with a vehicle drive line, the input and output members being relatively rotatable against damping means to absorb/attenuate torsional vibrations emanating from the engine. The damping means may take numerous forms and may include, for example, one or more of the following:- bob- weight damping linkages which connect the masses for the transmission of drive torque and which provide a speed depending torsional damping effect, springs loaded in compression and fluid damping devices. Normally the damping means also includes some form of friction damping device.

One form of torsional vibration damper of the type specified is a twin mass flywheel in which the input and output members each comprise significant flywheel masses. In other forms the input and output member are of the relatively light construction being simply strong enough to transmit the necessary loads.

One of the problems associated with torsional vibration dampers of the type specified which employ a friction damping device as part of the damping means is that such friction damping devices normally comprise two or more friction members axially biased into friction contact by a belleville spring means and the frictional damping effect of such a damper is very sensitive to the level of spring bias. This causes problems by requiring accurate manufacture and assembly of the parts. Also the spring bias may vary significantly as the friction members wear during use of the device.

It is an object of the present invention to provide a torsional vibration damper of the type specified which includes a friction damping device which at least mitigates the above problems. Thus according to a first aspect of the present invention there is provided a torsional vibration damper of the type specified which includes a friction damping device comprising two or more friction members axially biased into frictional contact by a belleville spring having a generally flat load/deflection characteristic over its operational range, the belleville spring being capable of applying axial loading to the friction members whilst positioned on both sides of its flat condition.

With such an arrangement by using a belleville spring with a generally flat load/deflection characteristic not only can the spring be arranged to apply a substantially constant axial force to the friction members over the entire operating life of the damper but also it is less important to control the operating angle of the belleville accurately during assembly of the damper (since it applies a substantial constant force over a range of angles) hence it is not necessary to accurately shim the belleville. This greatly simplifies manufacture and assembly of the damper.

The friction damping device may comprise a plurality of friction members assembled in an annular stack with alternate members of the stack respectively operatively connected for co- rotation with the input member or the output member.

The operative connection between at least one friction member and its associated input or output member may allow a predetermined amount of relative rotation between the friction member and input or output member(so-called lost motion) in order that the friction damping effect does not become effective until said relative rotation has taken place. This effect can be used to provide so-called "phased damping" when the level of friction damping varies at different angles of relative rotation between the input and output members.

The belleville spring preferably applies the bias force to the friction members via a pressure plate which has a first axially extending abutment on a side of the pressure plate facing the belleville against which a first peripheral portion of the belleville applies its bias force thus allowing the belleville to lie on both sides of the flat condition relative to the first abutment, a second peripheral portion of the belleville reacting against a second axially extending abutment which again allows the belleville to lie on both sides of its flat condition relative to the second abutment.

The first abutment is preferably formed as a circumferentially extending and axially projecting raised rib pressed into the pressure plate. The second abutment may conveniently comprise an axially extending ring one end of which reacts against one of the flywheel masses and the other end of which reacts against the second peripheral portion of the belleville. The second abutment may be integral with one of the flywheel masses.

The belleville arrangement of the present invention can be used with all types of additional damping means (e.g. bob weights, springs or fluid damping) and can be used when the torsional vibration damper is a twin-mass flywheel or when the input and output members do not have significant mass.

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 sectional view of a twin mass flywheel taken in the direction of arrow B of Figure 2;

Figure 2 is a sectional view taken along the line Z-Z of Figure 1;

Figure 3 shows a sectional view, on a larger scale, of the friction damping device shown in Figure 2.

Figure 4 shows diagrammatically part of the friction damping device of Figure3;

Figure 5 shows the generally flat load/deflection characteristic of the belleville spring used in the friction device;

Figure 6 shows diagrammatically part of a spurring arrangement used in the friction device, and Figure 7 shows an alternative belleville spring arrangement for loading the friction damping device.

With reference to figures 1 to 3 there is illustrated a twin mass flywheel 10 which is formed by two flywheel masses 11 and 12. One flywheel mass 11 (also known as the input flywheel mass) is fixed to a crankshaft 16 of an internal combustion engine byway of a central hub 20 using bolts (not shown) which extend through bores 15. A friction clutch (not shown) is secured to the second flywheel mass 12 (also known as the output flywheel mass) to connect the second flywheel mass with the input shaft (not shown) of an associated gearbox.

The flywheel mass 11 comprises central hub 20, an inner portion 21, an outer annular portion 22, a radially inwardly extending flange or cover plate 23, and a starter ring 21a. Cover plate 23 is spot welded at its outer periphery to outer portion 22.

The output flywheel mass 12 comprises an output plate 30, and a pivot plate 32 rotationally fast with each other. The output mass 12 is mounted for rotation relative to the input mass 11 by an L-shaped plain bearing member 50 supported for rotation with mass 12 within a bore 52 in plate 30.

The radially inner surface 53 of a plain bearing member 50 is coated with a plain bearing material such as a fluoro-polymer which rotates relative to the outer surface 54 of hub 20. Alternatively, bearing member 50 could be formed from PTFE impregnated with a sintered metal or could be a solid polymeric bearing member. The bearing 50 and plate 30 are held in position on hub 20 against axial displacement away from input mass 11 by a retaining ring 55 held in place by the bolts which pass through bore 15. Retaining ring 55 contacts the radially outwardly extending portion 50a of bearing 50.

Relative rotation between two flywheel masses 11 and 12 is controlled by a damping means which primarily comprises a plurality of pivotal linkages 40. The damping means also comprises a plurality of spring units 34D,34E carried on the input mass 11 via pins 34G which extend between and secure together disc 21 and cover plate 23, and also includes a friction damping device 60. The relative rotation of the flywheel masses is limited by contact between stop surfaces 32C on pivot plate 32 and stops 33 A mounted on the input mass 11 via pins 34G. All these components assist in controlling the relative rotation of the two flywheel masses 11 and 12 at specific relative angular positions or in specific angular ranges.

Each pivotal linkage 40 comprises a first link 41 (also known as a bobweight link) pivotally mounted between a centre hub portion 35 of the output plate 30 and pivot plate 32 by way of a first pivot 43, and a second flexible link 42 pivotally mounted on the output flywheel mass 12 byway of a second pivot 44 which extends between and secures together disc 21 and cover plate 23. The two links 41 and 42 pivotally connected to each other by means of a third pivot 45. It will be noted from Figure 1 that the first pivot 43 is positioned radially inwardly of the second and third pivots 44 and 45.

Under no-load conditions with the clutch 4 disengaged and the flywheel driven by the crankshaft in the direction of arrow E of Figure 1, centrifugal force acts on the pivotal linkages 40 and particularly on the first bobweight link 41 and urges the linkages in a radially outward direction with pivot 45 adopting a position radially outboard of pivot 43 as shown in fig 1 (this position is regarded as the neutral position between the drive and over-run directions of relative rotation of the flywheel masses). At higher rotational speeds the centrifugal force is greater and whilst this does not affect the configuration under no-load conditions it greatly affects the force required to move the flywheel mass 12 relative to the flywheel mass 11 i.e. the flywheel torsional stiffness.

If the clutch is engaged and power is transmitted in the drive direction from flywheel mass 11 to flywheel mass 12 there is a tendency for the two masses to rotate relative to each other (flywheel mass 11 rotates clockwise relative to flywheel mass 12 when viewing figure 1). At relatively low speeds when the influence of centrifugal force is smaller the flywheel masses move readily relative to each other i.e. the flywheel torsional stiffness is relatively low. However at relatively high speeds the influence of centrifugal force is much greater and relative rotation of the flywheel masses requires greater force i.e. the flywheel torsional stiffness is relatively high. Thus the flywheel torsional stiffness is speed sensitive.

If the clutch is engaged and power is transmitted in the over-run direction from flywheel mass 12 to flywheel mass 11 the effects are similar to the above except that the direction of relative rotation is reversed (flywheel mass 11 rotates anticlockwise relative to flywheel mass 12 when viewing figure 1) and in the embodiment shown in Figure 1 the first link 41 folds up relative to the second link 42.

Spring units 34D are only compressed when the flywheel masses rotate relative to each other in the drive direction. This compression occurs when arms 32E on the pivot plate contact the ends 34F of the springs after a given amount of relatively rotation determined by circumferential clearance P.

Similarly rubber block springs 34E come into operation towards the latter part of relative rotation between the masses in the drive direction when abutments 32D' on the pivot plate 32 contact the rubber springs and in the overrun direction when abutments 32D" on the pivot plate contact the rubber springs.

h accordance with the present invention the friction damping device 60 comprises a stack of interleaved annular friction members 61 and 62 (see Figure 3) with members 61 being splined to output plate 30 on splines 63 and members 62 being splined to hub 20 and hence input mass 11 via splines not visible in the plane of figures 2 and 3.

The friction members 61 and 62 are axially loaded against a shoulder 20a on hub 20 by a belleville spring 64 via an annular pressure plate 65. Belleville 64 also reacts against input flywheel 11 via an axially extending ring-like abutment 66. Pressure plate 65 is provided with an abutment in the form of a raised circumferentially extending rib 67 against which the belleville spring 64 acts. By arranging the belleville 64 to act on the pressure plate 65 and input mass 11 via raised rib 67 and axially extending ring 66 respectively, the belleville 64 can act on either side of its flat condition shown in full lines in figure 4 at 64a. Thus, for example, the belleville spring can assume the attitude shown in dotted lines 64b in figure 4 in which it is concave to the left as viewed in figure 4 or the attitude shown in dotted lines 64c in which it is concave to the right as viewed in figure 4.

By choosing the load deflection characteristic of belleville 64 to have a relatively flat form, as shown in figure 5, the load applied by the belleville over its entire operating range R (i.e. between say 0.5 mm and 1.5 mm deflection) remains substantially constant and it is therefore possible for the belleville to adopt a whole range of angles between positions 64a and 64c (the flat condition corresponding to a deflection of 1.0 mm in Figure 5) and apply a substantially constant axial load to pressure plate 65 and hence to the friction device 60. Not only does this ensure a substantially constant operation characteristic for the friction device 60 but it also obviates the need for high manufacturing accuracies in the components 66, 64, 65, 67 etc. since a substantially constant axial load is provided for a whole range of belleville cone angles.

hi the arrangement shown in figures 2 and 3, the friction members 61 and 62 are splined to their respective relatively rotatable components 30 and 20 without any circumferential play on their splinings other than the necessary clearance to ensure axial sliding. Thus the friction device 60 operates immediately there is any relative rotation between components 30 and 20.

It will be understood that the splining of components 61 and 62 could be arranged, as shown diagrammatically in figure 6, so that there is circumferential play 'q' in the spline connections to provide differing amounts of delay in the operation of the friction device dependant on the size of the play 'q'. It is possible for all the friction members to have the same amount of circumferential play or for different friction members to have different amounts of circumferential play so that the friction device can be phased into operation over a wide range of angles of relative rotation.

Again the number of friction members provided in the friction device 60 can be varied from two (i.e. one member 61 and one member 62) up to as many as are necessary to provide the total required friction damping effect. Also the ring-like abutment 66 could be formed integrally with, for example, component 21. Figure 7 shows an alternative arrangement for loading the belleville spring 64 onto the stack of friction members 61, 62 via an axially bent rim portion 68 of pressure plate 65 and a circlip 69 which engages in a groove 70 formed in hub 20.

The belleville springs 64 described above may be of a thicker form and may have their inner or outer peripheries provided with a series of generally radially extending slots to provide a series of circumferentially spaced spring fingers to provide the necessary spring characteristic in a more predictable and therefore more easily manufactured form.

The present invention thus provides a friction damping device which is less sensitive to manufacturing inaccuracies and inaccuracies in assembly and is able to provide a substantially constant damping effect over the entire life of the torsional vibration damper in which the friction device is fitted.

Although described above in relation to a twin mass flywheel which uses bob weights 41 to provide a speed dependent torsional stiffness, the present invention can be used with any form of torsional vibration damper of the type specified in which, for example, the damping means comprises only compression springs and a friction damping device.

Claims

1) A torsional vibration damper of the type specified which includes a friction damping device comprising two or more friction members axially biased into frictional contact by a belleville spring having a generally flat load/deflection characteristic over its operational range, the belleville spring being capable of applying axial loading to the friction members whilst positioned on both sides of its flat condition.

2) A torsional vibration damper according to claim 1 in which the friction damping device comprises a plurality of friction members assembled in an annular stack with alternate members of the stack respectively operatively connected for co-rotation with the input member or the output member.

3) A torsional vibration damper according to claim 2 in which the operative connection between at least one friction member and its associated input or output member allows a predetermined amount of relative rotation between the friction member and input or output member in order that the friction damping effect does not become effective until said relative rotation has taken place.

4) A torsional vibration damper according to any one of claims 1 to 3 in which the belleville spring applies the bias force to the friction members via a pressure plate which has a first axially extending abutment on a side of the pressure plate facing the belleville against which a first peripheral portion of the belleville applies its bias force thus allowing the belleville to lie on both sides of the flat condition relative to the first abutment, a second peripheral portion of the belleville reacting against a second axially extending abutment which again allows the belleville to lie on both sides of its flat condition relative to the second abutment. 5) A torsional vibration damper according to claim 4 in which the first abutment is formed as a circumferentially extending and axially projecting raised rib pressed into the pressure plate.

6) A torsional vibration damper according to claim 4 in which the first abutment is formed by forming an axially bent rim on the outer periphery of the pressure plate.

7) A torsional vibration damper according to claim 4, 5 or 6 in which the second abutment comprises an axially extending ring one end of which reacts against one of the flywheel masses and the other end of which reacts against the second peripheral portion of the belleville.

8) A torsional vibration damper according to any one of claims 1 to 7 in which additional damping means act between the input and output members.

9) A torsional vibration damper according to any one of claims 1 to 8 in the form of a twin mass flywheel.

10) A torsional vibration damper constructed and arranged substantially as hereinbefore described with reference to and as shown in figures 1 to 6 or 7 of the accompanying drawings.