GB2327739A

GB2327739A - Torsional vibration damper

Info

Publication number: GB2327739A
Application number: GB9813916A
Authority: GB
Inventors: Andreas Orlamunder; Martin Gerber; Eberhard Knaupp
Original assignee: Mannesmann Sachs AG
Current assignee: ZF Friedrichshafen AG
Priority date: 1997-07-02
Filing date: 1998-06-26
Publication date: 1999-02-03
Anticipated expiration: 2018-06-26
Also published as: ES2182601A1; FR2765648B1; DE19728241C2; GB9813916D0; ES2182601B1; DE19728241A1; GB2327739B; FR2765648A1

Abstract

A torsional vibration damper for the power train of a motor vehicle comprises mutually relatively rotatable input and output parts 5, 13, and a spring arrangement 17 torsionally elastically coupling the input part to the output part. The spring arrangement comprises at least one coiled spring element 61 with at least one turn 63 extending around the axis of rotation. The ends of the element make a torque transmitting connection to the input part and the output part. At least a proportion of the positions where the turn(s) intersect an axial longitudinal plane containing the axis of rotation 7 are mutually axially offset and may also be radially offset with uniformly decreasing turn diameters 63a, or stepped decreases (63b, figure 4).

Description

1 1 TORSIONAL VIBRATION DAMPER 2327739 The invention relates to a dual

mass flywheel and a torsional vibration damper for use in the power train of a motor vehicle.

A torsional vibration damper of known construction is described in DE 40 06 121 AI.

In this construction, the input and output components or parts are coupled together with spring arrangement composed of two entwined spiral springs. Each spiral spring has several spiral turns surrounding one another without a mutual axial offset. In the axial direction, the spiral springs enclose only a space corresponding to the axial thickness of their spring material. The spiral turns are responsible for the radial size of the spiral springs. They do not affect its axial size.

During a rotation of the input part and the output part in opposite directions, the turns of the spiral springs experience a radial constriction or expansion, that is a change in their mean diameter, corresponding to the relative direction of rotation. From the softest spiral turn, i.e. the spiral turn having the greatest mean diameter, to the hardest spiral turn, i.e. the spiral turn having the smallest mean diameter, each individual spiral turn experiences an incremental change in its mean radius. In the radial direction, the incremental changes of radius of the individual spiral turns accumulate beyond the natural radius of the spiral spring. The overall radial deformation of the spiral spring which is observed during a relative rotation of the input and output part must be allowed for when dimensioning the space allocated to the spiral spring.

1 1 2 The spiral springs known from DE 40 06 121 AI have a radial size large in comparison with their axial size, The deformation of the spiral turns occurring during a relative rotation of the input and the output part takes place in the same direction in which the spiral turns already require a large space, namely in the radial direction, and has to be additionally allowed for during fitting of the spiral springs. The space required for the spiral springs is therefore relatively great in the radial direction. In the case of motor vehicles, however, there is frequently only a very restricted radial space for components fitted into the power train. In particular in the case of cars with relatively small ground clearance, close limits are imposed on the radial size of these components.

An object of the invention is to provide a torsional vibration damper in which the radial spatial requirements of the springs is minimized and made more suitable for fitting into the power train of a motor vehicle.

According to the invention there is provided a torsional vibration damper usable in the power train of a motor vehicle; said damper comprising: an input component arranged rotatably round an axis of rotation, an output component arranged centrally relative to the axis of rotation and rotatable relative to the input part and a spring arrangement torsionally elastically coupling the input component to the output component, said spring arrangement comprising at least one spring element which is formed from coiled spring material, with at least one or a plurality of turns extending round the axis of 1 1 3 rotation and ends which establish a torque transmitting connection to the input component and the output component, wherein at least a proportion of the positions where the spring turn or turns intersect an axial longitudinal plane containing the axis of rotation, succeed one another in the circumferential direction of the coiled spring material and are axially mutually offset.

With the solution according to the invention, a division is made between the direction of the succession of turns of the spring material coil and the deformation direction thereof during a relative rotation of the input and the output component. Whereas a possible constriction or expansion of the turns still takes place in the radial direction, the direction of the succession of turns has an axial component. Owing to this axial component, the overall space required for the spring element is not concentrated almost exclusively in one direction as in DE 40 06 121 AI (in the radial direction there), but is distributed in two directions, namely the axial and the radial directions. In this way, the spring element and therefore the torsional vibration damper can be better adapted to predetermined spatial conditions in the power train of the motor vehicle and, in particular, the damper can be fitted in positions where the radial spatial conditions are relatively restricted.

The axial offset of individual mutually succeeding spring turn intersection positions affords a further advantage. As already mentioned, the incremental changes in the mean radii of the individual spiral turns during a relative rotation of the input and the output part accumulate in the construction known from DE 40 06 121 AI, In contrast, the construction in accordance with the invention ensures that the space which must be made available at an axial 1 1 4 location along the axis of rotation for any deformations of the turns of the spring material coil in the radial direction is defined essentially only by the extent of the radial deformation of a single turn of the spring material coil. Accumulation in the radial direction of the incremental changes of the mean radius of individual mutually succeeding turns does not take place as such. Therefore, the radial space to be taken into consideration merely for possible deformations of the turns of the spring material coil is also considerably smaller than with the prior art construction known from DE 40 06 121 AI. Therefore, not only the radial size of the spring element but also the overall space to be allowed for is reduced, the axial component of the direction of the succession of the spring turns allowing a certain degree of flexibility for adaptation of the torsional vibration damper to the prevailing spatial conditions. These advantages are afforded, in particular, if all mutually succeeding spring turns of the spring element intersect the axial longitudinal plane at axially mutually offset positions.

At its ends, the spring element can be coupled axially non-movably to the input and the output component, for uniform loading of the spring element. For example, the ends of the spring element can be welded to the input and the output component. An interlocking connection established when the spring element is radially bent at its ends and is inserted into appropriate holding slots in the input and the output components is also conceivable. In another embodiment, at least one of the ends of the spring element is coupled axially movably to the associated input or output components. The two ends of the spring element can both be coupled axially movably to the input or the output component for uniform loading of the spring element. In such a case, a radial deformation of the turns of the spring element during a 4 1 relative rotation of the input and the output component in opposite directions could be reduced and absorbed at least in part by a change in the axial length of the spring element.

The spring element can be designed as a helical spring with a constant mean diameter of its turns. With a spring element of this type, all turns have essentially the same spring stiffness insofar as the cross section of the spring material is identical throughout. However, the cross section of the spring material must be able to change in the circumferential direction of the spring material coil in order to achieve different spring behaviour in individual turns or turn portions of the spring element.

In order to obtain a desired spring characteristic, for example a progressive spring characteristic, at least a proportion of the succeeding turns of the spring element can have different mean diameters in the relaxed state of the spring element. As the mean diameter of a turn affects its stiffhess, it is thus conceivable by this measure to achieve the desired spring behaviour of the torsional vibration damper. However, the foregoing measure also permits the torsional vibration damper to be adapted to constructional parameters by appropriate choice of the mean diameter of the turns of the spring element.

The mean diameter of the turns of the spring element can decrease from one end of the spring element to the other. The mean diameter of the turns can decrease from the input component end of the spring element to the output component end. However the mean diameter of the turns can also decrease from the output component end of the spring element to the input component end. In an expedient embodiment, the mean diameter of the turns 1 9 6 changes continuously from one end of the spring element to the other. In this case, the spring element can be designed as a conical spring with a uniformly decreasing mean diameter of its turns. However, it is also conceivable for the mean diameter of the turns of the spring element to change stepwise from one end of the spring element to the other.

If individual mutually succeeding turns of the spring element have different mean diameters, it is possible for these turns to have axially partially overlapping intersection positions through the axial longitudinal plane. A spring element which is very compact in the axial direction can be obtained in this way.

To avoid the risk of damaging the torsional vibration damper when strong torques are introduced, for example during abrupt surges of torque, mutually allocated stop means which limit the maximum rotational angle between the input and output components can be provided on these parts. The rotational angle can also be limited, at least in one of the two relative directions of rotation, in that the turns of the spring element are blocked at a specific rotational angle. i.e. the turns of the spring element come into contact with one another and block further rotation. Furthermore, the turns of the spring element must not be prevented from being blocked in the relaxed state so cushioned torque transmission takes place only in one direction of relative rotation between the input and output parts whereas the torsional vibration damper produces a nonrotatable coupling between the input and output component in the other direction of relative rotation. A further method of limiting the rotational angle involves providing, radially inside or/and radially outside the turns of the spring material coil, limiting faces on the input orland the output component against which the turns abut at a specific 7 amount of radial constriction or expansion owing to a rotation of the input and the output component relative to one another. Thus further rotation of the input and the output component relative to one another can be limited when the turns of the spring element are blocked by the limiting faces.

It has been found that high torques can be transmitted with large rotational angles with the design of the torsional vibration damper constructed in accordance with the invention. At the same time, the spring element can have comparatively low spring stiffhess so the output component can be readily decoupled from the input component in terms of vibration. With a preferred design, the spring element has a few turns, the number of turns being between one and a half and five, preferably between two and four. A wire material of circular cross section has proven suitable as spring material for the spring element, but materials with other cross-sectional shapes can obviously also be selected.

The spring element can be arranged in the region of the outer periphery of the input orland the output component. Particularly great diameters of the turns with correspondingly low spring stiffness can be achieved in this case, and this is advantageous for a spring element which is compliant throughout and acts over a great rotational angle However, the spring element can also be arranged in a radially central region of the input or/and the output component. The spring element can then be arranged between the input and the output component with protection from the exterior.

1 8 It has been found that a torsional vibration damper constructed in accordance with the invention can produce excellent spring characteristics with only a single coiled spring element, and this reduces the number of parts and the production and fitting costs. The spring characteristics of the torsional vibration damper can however be optimized by providing at least two coiled spring elements. These spring elements can have identical or different spring characteristics. In particular, it is conceivable to provide spring elements of which the spring effects lag and which are therefore deformed and develop their spring action only at a predetern-iined rotational angle of the input part relative to the output part. Such lagging spring elements can be combined with spring elements which act directly during rotation of the input and the output component from the basic rotational position. Therefore, any desired spring characteristic can be embodied, possibly even with different spring behaviour in compression and tension. Space can be saved, in particular, if two spring elements are coiled at least partially axially inside one another. However, it is also conceivable for two spring elements to be at least partially radially coiled inside one another.

A torsional vibration damper constructed in accordance with the invention can basically be fitted at any position in the power train of a motor vehicle. For example, it can be fitted directly in a clutch, in particular a friction clutch of a motor vehicle. In a preferred development of the invention, however, the input component is formed by a primary mass of a dual mass flywheel which can be fastened on the crankshaft of an internal combustion engine and the output part is formed by a secondary mass of the dual mass flywheel on which a pressure plate unit of a motor vehicle friction clutch can be fastened. A pressure plate unit of this type usually comprises a clutch housing, a clutch main spring and a pressure plate which 1 % 9 are combined to form a unit fastened on the secondary mass of the dual mass flywheel. When the torsional vibration damper is used in this way in a dual mass flywheel, the mean radius of the turns of the spring element is preferably greater in the relaxed state of the spring element than the external radius of a friction clutch side pressing face of the secondary mass. However, 0 the turns of the spring element can overlap the pressing face of the secondary mass in the radial direction or its mean radius can be smaller than the internal radius of the pressing face.

The invention may be understood more readily, and various other aspects and features of the invention may become apparent, from consideration of the following description.

Embodiments of the invention will now be described by way of example only with reference to the accompanying drawings, wherein:

Figure 1 is an axial longitudinal section through half of a dual mass flywheel with a torsional vibration damper constructed in accordance with the invention; Figure 2 is a sectional view of the dual mass flywheel in Figure 1, the section being taken along line II-II of Figure 1; Figure 3 is a schematic representation of a second embodiment of a torsional vibration damper constructed in accordance with the invention.

Figure 4 is a schematic representation of a third embodiment of a torsional vibration damper constructed in accordance with the invention; Figure 5 is a schematic representation of a fourth embodiment of a torsional vibration damper constructed in accordance with the invention; Figure 6 is a representation of a fifth embodiment of a torsional vibration constructed in accordance with the invention with two spring elements arranged radially inside one another; Figure 7 is a representation of a sixth embodiment of a torsional vibration damper constructed in accordance with the invention with two spring elements arranged axially inside one another., Figure 8 is a view corresponding to Figure 1 and representing a seventh embodiment of a torsional vibration damper constructed in accordance with the invention, and Figure 9 is a view corresponding to Figure 1 and representing an eighth embodiment of a torsional vibration damper constructed in accordance with the invention.

Figure 1 shows a dual mass flywheel 1 having, on the side of a crankshaft 3 indicated in broken lines of an intemal-combustion engine (not shown), a primary flywheel mass 5 which serves to introduce a starting torque. The primary mass 5 is fastened centrally relative to an axis of rotation 7 of the crankshaft 3 by fastening bolts 9 on the crankshaft 3 which are inserted in several holes 11 distributed peripherally round the axis of rotation 7 in the radially internal region of the primary mass 5.

On the side of the primary mass 5 remote from the crankshaft 3, the dual mass flywheel 1 has a secondary flywheel mass 13 which is rotatable round the axis of rotation 7 and is used for mounting a motor vehicle friction clutch 15. The secondary mass 13 is coupled torsionally elastically to the primary mass 5 via a spring arrangement 17. By means of a bearing arrangement 19, the secondary mass 13 is mounted rotatably relative to the primary mass 5 on the primary mass 5. The bearing arrangement 19 comprises a ball bearing 21 which supports the secondary mass 13 radially and axially on the primary mass 5 and is fixed by locking washers 23 and 25 axially on the secondary mass 13 or on the primary mass 5. The bearing arrangement 19 can alternatively comprise at least one sliding bearing.

A friction arrangement 27 with a fliction plate 29 and a spring washer 31 axially biasing the friction plate 29, acts between the primary mass 5 and the secondary mass 13, and serves to absorb torsional vibrations between the primary mass 5 and the secondary mass 13. The friction arrangement 27 together with the spring arrangement 17 forms a torsional vibration damper which cushions and absorbs torsional vibrations between the primary mass 5 and the secondary mass 13. The primary mass 5 forms an input component of the torsional vibration damper and the secondary mass 13 forms an output component of the torsional vibration damper.

12 The ffiction clutch 15 has a disc or plate 33, arranged centrally relative to the axis of rotation 7. The clutch disc 33 has a flange 47 and a hub part 35. A friction lining carrier 39 is fastened on the flange 47 by means of rivets 37. The hub part 35 has a hub 41 with a hub aperture 43 provided with internal teeth 45 for rotational engagement with a gearing input shaft (not shown). The fliction clutch 15 also comprises a pressure plate unit 49. This pressure plate unit 49 comprises a clutch housing 51 which is connected non-rotatably and axially rigidly to the secondary mass 13 and on which a pressure plate 53 is held non- rotatably but axially movably in a manner not shown in detail, for example by tangential springs. The pressure plate 53 is biased toward the secondary mass 13 by a clutch main spring 55, which takes the form of a diaphragm spring, held on the clutch housing 51. On its side facing the friction clutch 15 the secondary mass 13 has a face 57 against which ffiction linings 59 arranged on the ffiction lining carrier 39 are ffictionally pressed by the pressure plate unit 49 in the engaged state of the ffiction clutch 15.

The spring arrangement 17 comprises a spring element 61 composed of coiled wire material of circular cross section. Only one single spring element 61 of this type is shown in the illustrated embodiment but two or more spring elements 61 of this type can be adopted. The wire material of the spring element 61 is coiled round the axis of rotation 7 and forms at least one turn 63. In the example illustrated, the spring element 61 has between two and three turns 63, though the number of turns obviously does not have to be a whole number and can also be, in particular, 2.5. The turns 63 of the spring element 61 coaxially surrounding the axis of rotation 7 form a helical spring of which the mean turn diameter is substantially identical for all turns 63. The spring turns 63 intersect an axial longitudinal plane containing the axis of 13 rotation 7 i.e., the plane of the drawing, at axially offset positions in the circumferential direction of the coil.

The two ends 65 of the helical spring 61, only one of which is shown in Figure 1, are rigidly coupled peripherally to the primary mass 5 and the secondary mass 13. As shown in particular in Figure 2 with reference to the primary mass end 65 of the helical spring 61, the spring end 65 can be bent radially inwards and engage radially in a holding slot 67 of the associated flywheel mass 5 or 13. In this way an interlocking engagement is guaranteed at least in the circumferential direction. The spring ends 65 of the helical spring 61 are in torque transmitting contact with the primary mass S and the secondary mass 13. In the axial direction, the ends 65 of the helical spring 61 can be fixed axially unmovably in their associated holding slot 67. It is conceivable to solder or to weld the spring ends 65 to the primary mass 5 and the secondary mass 13 for this purpose. However, the holding slots 67 can also be wider in the axial direction than the diameter of the wire material forming the helical spring 6 1, as indicated in broken lines at 69 in Figure 1. A degree of axial play of the ends 65 of the helical spring 61 in its associated holding slot 67 is produced in this way. Furthermore, a torquetransmitting coupling, affected by play, can be created between the ends 65 of the spring element 61 and the primary mass 5 or the secondary mass 13 in the circumferential direction, so the spring action of individual spring elements 61 can be delayed, in particular if several spring elements 61 are provided.

In the embodiment shown in Figure 1, the diameter of the spring wire of the spring element 61 is about 10% of the mean radius of the turns 63 of the spring element 61. With 1 14 respect to the size of the primary mass 5 and the secondary mass 13, the diameter of the spring wire of the spring element 61 is also about 10% of the radius of the primary mass 5 and the secondary mass 13. However, it is obvious that the spring material of the spring element 61 can be greater or smaller than this in particular the spring wire can have a substantially smaller radial and axial thickness according to the desired spring temper.

The axially succeeding turns 63 of the helical spring 61 have relatively slight spacing from one another which is considerably smaller than the diameter of the wire material of the helical spring 61. The mutual contact between the turns 63 of the helical spring 61 can also be weaker under certain circumstances. During a relative rotation of the primary mass 5 and the secondary mass 13 in opposite directions, the spring energy absorbed by the helical spring 61 leads to radial expansion or constriction of at least a proportion of the turns 63 of the helical spring 61, depending on the direction of relative rotation, as indicated in broken lines at 71 for the central turn 63 in Figure 1. This elastic deformation of the turns 63 accompanying an enlargement or reduction of the mean turn diameter causes a restoring torque which increases with the rotational angle and attempts to rotate the two flywheel masses 5 and 13 back into their starting position. The maximum constriction of the turns 63 can be limited radially inwardly by an external peripheral wall 73, forming an internal boundary, of at least one of the two flywheel masses 5 and 13 (shown on the primary mass 5 in Figure 1). The radial expansion of the turns 63 can be limited radially outwards, for example, by a peripheral housing wall surrounding the dual mass flywheel 1. However, such limits are not required in every case and can be dispensed with, in particular, if rotation is limited by co-operating stop means on the primary mass 5 and the secondary mass 13 The mean diameter of the turns 63 of the helical spring 61 is greater in the embodiment shown in Figure 1 than the external radius of the pressing face 57 of the secondary mass 13, the helical spring 61 being arranged in the radially externally located edge regions of the primary mass 5 and the secondary mass 13. However, it is also conceivable to select the mean diameter of the helical spring 61 such that the turns 63 of the helical spring 61 lie in the radial region of the pressing face 57, i.e. overlap it radially or even extend round the axis of rotation 7 radially inside the pressing face 57. As the torsional vibration damper according to the invention does not have any components which are rubbed or whetted against one another during a relative rotation of the primary mass 5 and the secondary mass 13, the frictional losses and the wear are correspondingly slight. In particular, a so-called dry-running dual mass flywheel can be formed which requires no lubricant in the region of the spring arrangement 17. The constructionally simple spring arrangement 17 allows, for example, rotational angles of up to about 65% in both directions of relative rotation. At the same time, it allows the transmission of great torques.

Figures 3 to 9 show further embodiments of torsional vibration dampers according to the invention, identical reference numerals to those in Figures 1 and 2, but supplemented by a small letter as a suffix, being used for identical or equivalent components. Unless otherwise stated hereinafter, reference is made to the foregoing description of Figures 1 and 2 to explain these components.

1 16 Figure 3 shows a spring arrangement 17a torsionally elastically coupling an input part 5a to an output part 13a (for example in the form of the primary mass and the secondary mass of a dual mass flywheel), with a spring element 61a designed as a conical spring. The turns 63a of this conical spring 61 a have a mean diameter which decreases uniformly from the input part 5a to the output part 13a. The cone angle of the conical spring 61 a designated by cc is a small acute angle which is preferably not more than 45'. With this embodiment also, the positions where the turns 63a intersect the plane of the drawing in Figure 3 are mutually axially offset and can optionally partially overlap one another axially with appropriate mutual spacing. An even more compact construction of the conical spring 61 a could be achieved in this way.

Figure 4 shows a spring arrangement 17b with a spring element 6 1 b of which the turns 63b also have a mean diameter which decreases from the input part 5b to the output part 13b. However, the mean diameter of the turns 63b does not change uniforrnly as with the conical spring 6 1 a in Figure 3 but stepwise, the two central turns 63 b of this spring element 6 1 b having a substantially identical mean diameter. However, the positions where all turns 63b intersect the plane of the drawing in Figure 4, as with the helical spring 61 in Figures 1 and 2 and the conical spring 61a in Figure 3, are mutually axially offset. This embodiment, but also the embodiment in Figure 3, are merely to demonstrate how the mean turn diameter of the turns 63b of the spring element 61b can be varied to allow adaptation of the spring element 61b to predetermined spatial conditions, which are indicated schematically in Figures 3 and 4 by a corresponding contour of the input part 5a or 5b and the output part 13a or 13b.

17 Figure 5 shows an embodiment in a schematic view similar to Figures 3 and 4. A conical spring 61c which couples the input part 5c torsionally elastically to the output part 13c and of which the turn diameter increases from the input part Sc to the output part 13c is provided in this embodiment. The positions where the turns 63c of the conical spring 61c perforate the plane of the drawing in Figure 5 are in turn axially offset and can optionally partially overlap axially with an appropriately small pitch of the conical spring 61 c.

Figure 6 shows an embodiment of a spring arrangement 17d with two spring elements 61'd and 6Pd. The two spring elements 61'd, 61"d are designed as helical springs with a constant turn diameter. They are arranged radiallyinside one another, i.e. the turn diameter of the turns 63 "d of the helical spring 61 "d is smaller than the turn diameter of the turns 63'd of the helical spring 61'd. The positions where the turns 6Yd of the helical spring 61"d and the positions where the turns 63'd of the helical spring 61M intersect the plane of the drawing in Figure 6 are not axially offset or are only insignificantly mutually offset in the axial direction if the two helical springs 61'd, 61"d are compared. However, it is also conceivable that the positions where the turns 63M of the helical spring 61M intersect the plane of the drawing in Figure 6 are axially offset from the positions where the turns 63'd of the helical spring 61'd intersect the drawing plane in particular are located axially between two intersection positions of the turn of the helical spring 6 I'd A spring arrangement 17d with two spring elements 6 Vd, 6 1 " arranged radially inside one another can also be embodied by conical springs or step springs, as shown in the embodiments in Figures 3 to 5.

1 18 Figure 7 shows a further embodiment of a dual mass flywheel I e. With this dual mass flywheel le, a primary mass Se and a secondary mass 13e are torsionally elastically coupled to one another via two spring elements 6 Ve and 6 1 "e designed as respective helical springs. The two helical springs 6 Ve, 61 " e have the same turn diameter. They are arranged axially inside one another. When the positions where the turns 6Ye, 6Ye of the helical springs 6Ve, 6Ve intersect the plane of the drawing in Figure 7 are observed, an intersection position of the turns of the helical spring 6Ve and a perforation position of the turn of the helical spring 61e alternate. The turns 6Ye, 63"e of the helical springs 6Ve, 61"e have an approximately identical pitch It is also conceivable to use spring elements 6Ve, 6Ve with different pitches so two or more intersection positions of the turns of one respective spring element 6Fe, 61"e are located, for example, between two axially succeeding intersection positions of the turns of the other spring element 61"e, 6Pe.

For mounting the secondary mass 13e on the primary mass 5e there is provided a split bearing arrangement 19e which comprises a radial bearing 77e radially supporting the secondary mass 13e on the primary mass 5e and an axial bearing 79e axially supporting the secondary mass 13e on the primary mass 5e. The radial bearing 77e and the axial bearing 79e are constructionally separated. The radial bearing 77e may be an antiftiction bearing. However, it can also be, for example, a sliding bearing formed by a plastic ring which is favourable to sliding. In the embodiment in Figure 7, the axial bearing 79e is designed as a sliding bearing and is formed by a disc 8 1 e of plastics material. This disc 81e has several peripherally distributed through-orifices 83e through which crankshaft fastening bolts 9e are inserted.

9 4.

19 Figure 8 shows a further embodiment of a dual mass flywheel If This dual mass flywheel If differs from the dual mass flywheel shown in Figure 1 essentially by the diffierent radial position of the turns 63 f of the spring element 6 1 f relative to the pressing face 5 7f of the secondary mass 13f. Whereas the turns of the spring element 61 are located radially outside the external radius of the pressing face in the embodiment shown in Figure 1, they are arranged so as to overlap the pressing face 57f radially in the embodiment shown in Figure 8.

The primary mass 5f is shortened radially relative to the secondary mass 13f which, in turn, projects radially externally beyond the clutch disc 33f and extends axially beyond the friction linings 59f. It is also conceivable for the primary mass 5f to project radially beyond the spring element 61f at the side and optionally to have a projection which extends axially from the crankshaft 3f, axially overlaps at least a proportion of the turns 63f of the spring element 61f and forms a radially external limit for the radial expansion of the turns 63f of the spring element 61 f Figure 9 shows a dual mass flywheel lg similar to the dual mass flywheels in Figures 1 and 8. In the dual mass flywheel lg, the radius of the turns 63g of the spring element 61g designed as a helical spring substantially corresponds to the internal radius of the ftiction clutch side pressing face 57g of the secondary mass 131g. The primary mass 5g extends radially past the spring element 61g at the side and has a flywheel ring gear 85g meshing with a starter pinion on its outer periphery. The primary mass 5g and the secondary mass 13g define a receiving chamber 87g in which the spring element 61g is received with protection. With regard to the radius of the turns 63 g of the spring element 6 1 g, it will be appreciated that it can also be smaller than the internal radius of the pressing face 57g so the spring turns 63g no longer overlap the pressing face 57g in the radial direction.

1 % 21

Claims

Claims

1. A torsional vibration damper fro the power train of a motor vehicle said damper comprising: an input component (5) arranged rotatably round an axis of rotation (7), an output component (13) arranged centrally relative to the axis of rotation (7) and rotatable relative to the input part (5) and a spring arrangement (17) torsionally elastically coupling the input component (5) to the output component (13), said spring arrangement comprising at least one spring element (61) which is formed from coiled spring material, with at least one turn (63) extending round the axis of rotation (7) and ends (65) which establish a torque transmitting connection to the input component (5) and the output component (13), wherein at least a proportion of the positions where the spring turn or turns (63) intersect an axial longitudinal plane containing the axis of rotation (7), succeed one another in the circumferential direction of the coiled spring material and are axially mutually offset.

2. A torsional vibration damper according to claim 1, wherein the spring element (61) has several turns (63) and all succeeding turns (63) of the spring element (61) have axially mutually offset positions where they intersect the axial longitudinal plane.

A torsional vibration damper according to claim 1 or 2, wherein at least one of the ends (65) of the spring element (61) is coupled axially nonmovably to the associated input or output component (5,13).

22 4. A torsional vibration damper according to claim 3, wherein the two ends (65) of the spring element (61) are both coupled axially non-movably to the input and output components (5,13).

5. A torsional vibration damper according to claim 1 or 2, wherein at least one of the ends (65) of the spring element (61) is coupled axially movably to the associated input or output component (65).

6- A torsional vibration damper according to claim 5, wherein the two ends (65) of the spring element (61) are both coupled axially movably to the input and output components (5,13).

7. A torsional vibration damper according to one of claims 1 to 6, wherein the spring element (6 1) is a helical spring with a constant mean diameter of its turns (63).

8. A torsional vibration damper according to one of claims 1 to 6, wherein at least a proportion of the succeeding turns (63a, 63b. 63c) of the spring element (61a; 61b; 61c) have different mean diameters in the relaxed state of the spring element (61 a; 6 1 b; 6 1 c).

9. A torsional vibration damper according to claim 8, wherein the mean diameter of the turns (63a.. 63b; 63c) of the spring element decrease from one end of the spring element (61 a; 61b; 61 c) to the other.

1 % 23 10. A torsional vibration damper according to claim 9, wherein the mean diameter of the turns (63a. 63b) decreases from the input component end of the spring element (61a; 61b) to the output component end.

11. A torsional vibration damper according to claim 9, wherein the mean diameter of the turns (63c) decrease from the output component end of the spring element (61c) to the input component end.

12. A torsional vibration damper according to one of claims 8 to 11, wherein the mean diameter of the turns (63 a; 63 c) changes continuously from one end of the spring element (61 a., 6 1 c) to the other.

13. A torsional vibration damper according to claim 12, wherein the spring element (61a; 61 c) is as a conical spring with a uniformly decreasing mean diameter of its turns (63a; 63c).

14. A torsional vibration damper according to one of claims 8 to 11, wherein the mean diameter of the turns (63a) of the spring element changes stepwise from one end of the spring element (6 1 b) to the other.

15. A torsional vibration damper according to one of claims 8 to 14, wherein mutually succeeding turns (63b) of the spring element (61b) of the different mean diameter have positions where they intersect the axial longitudinal plane which partially overlap axially.

24 16, A torsional vibration damper according to one of claims 1 to 15, wherein the spring element (6 1) has only a few turns.

17. A torsional vibration damper according to claim 16, wherein the spring element (61) has between one and a half and five turns (63), preferably between two and four turns (63).

18. A torsional vibration damper according to one of claims 1 to 17, wherein the spring material of the spring element (61) is formed by a wire material of circular cross section.

19. A torsional vibration damper according to one of claims 1 to 18, wherein the spring element (61) is arranged in the region of the outer periphery of the input and/or the output component (5,13).

20. A torsional vibration damper according to one of claims 1 to 18 wherein the spring element (61 g) is arranged in a radially central region of the input or the output component (5g, 13g).

2 1. A torsional vibration damper according to one of claims 1 to 20, where only one single coiled spring element (61) is provided.

22. A torsional vibration damper according to one of claims 1 to 20, wherein at least two coiled spring elements (61'd, 6Pd, 6Ve, 61"e) are provided.

1 h, 23. A torsional vibration damper according to claim 22, wherein two spring elements (6Ve, 61 "e) are coiled axially inside one another at least with a proportion of their turns (6Ye, 63 "e).

24. A torsional vibration damper according to claim 22 or 23, wherein two spring elements (6 Vd, 61 "d) are coiled radially inside one another at least with a proportion of their turns (63'd, 63"d).

25. A torsional vibration damper according to one of claims 1 to 24, wherein the input component (5) is formed by a primary mass (5) of a dual mass flywheel (1) which can be fastened on the crankshaft (3) of an internal-combustion engine and the output component (13) is formed by a secondary mass (13) of the dual mass flywheel (1) which a pressure plate unit (49) of a motor vehicle friction clutch (15) can be fastened.

26. A torsional vibration damper according to claim 25, wherein mean radius of the turns (63) of the spring element (61) in the relaxed state of the spring element (61) is greater than the external radius of a ffiction clutch side pressing face (57) of the secondary mass (13).

27. A torsional vibration damper according to claim 25, wherein the turns (63f) of the spring element (61f) in the relaxed state of the spring element (61f) lie in the radial range of a friction clutch side pressing face (5 7f) of the secondary mass ( 13 f).

11 26 28. A torsional vibration damper according to claim 25, wherein the mean radius of the turns (63g) of the spring element (61g) in the relaxed state of the spring element (61g) corresponds roughly to the internal radius of a ffiction clutch side pressing face (57g) of the secondary mass ( 13 g) or is smaller than this internal radius.

29. A dual mass flywheel incorporating a torsional vibration damper or a torsional vibration damper or a dual mass flywheel with a torsional vibration damper incorporated with a friction clutch substantially as described with reference to and as illustrated in any one or more of the Figures of the accompanying drawings.