阅读说明：本技术 扭振缓冲器 (Torsional vibration damper ) 是由 S·马伊恩沙因 于 2018-05-24 设计创作，主要内容包括：扭振缓冲器,其包括能围绕转动轴线转动的驱动件(4)并通过至少一个弹簧元件与该驱动件耦合的能抵抗所述弹簧元件围绕所述转动轴线相对于所述驱动件(4)扭转的缓冲器质量件(8),其中,实施为支腿弹簧(7)的弹簧元件以一端部与所述驱动件(4)抗扭转地耦合并且以另外的端部通过在至少一个导向轨(10、16)上滚动的耦合滚子(11)与所述缓冲器质量件(8)耦合,其中,所述导向轨(10、16)是不对称的,使得所述支腿弹簧(7)在从零位置(N)运动出来时根据所述耦合滚子(11)的运动方向而不同强度地变形。(Torsional vibration damper, comprising a driver (4) which can be rotated about a rotational axis and a damper mass (8) which is coupled thereto by means of at least one spring element and can be rotated relative to the driver (4) against the spring element about the rotational axis, wherein the spring element, which is embodied as a leg spring (7), is coupled at one end to the driver (4) in a rotationally fixed manner and at the other end to the damper mass (8) by means of a coupling roller (11) which rolls on at least one guide rail (10, 16), wherein the guide rails (10, 16) are asymmetrical such that the leg spring (7) deforms differently in dependence on the direction of movement of the coupling roller (11) when moving out of a zero position (N).)

1. A torsional vibration damper, comprising: a drive element (4) which can be rotated about a rotational axis, and a damper mass (8) which is coupled to the drive element by means of at least one spring element and can be rotated relative to the drive element (4) against the spring element about the rotational axis, characterized in that the spring element is designed as a leg spring (7) which is coupled with one end to the drive element (4) in a rotationally fixed manner and with the other end to the damper mass (8) by means of a coupling roller (11) which rolls on at least one guide rail (10, 16), wherein the guide rails (10, 16) are asymmetrical in such a way that the leg spring (7) deforms differently strongly depending on the direction of movement of the coupling roller (11) when moving out of a zero position (N).

2. Torsional damper according to claim 1, characterized in that the guide rail (10) is provided on the leg spring (7) and the coupling roller (11) is arranged on the damper mass (8) in a rotatable manner about a fixed axis of rotation (12).

3. Torsional damper as in claim 1, characterized in that the guide rail (16) is provided on the damper mass (8) and the coupling roller (11) is arranged on the leg spring (7) in a rotatable manner about a fixed axis of rotation.

4. Torsional damper as in claim 1, characterized in that guide rails (10, 16) are provided both on the leg spring (7) and on the damper mass (8), and in that the coupling roller (11) can roll freely rotatably on both guide rails (10, 16), wherein at least one guide rail (10, 16) is asymmetrical.

5. Torsional damper according to claim 4, characterized in that the two guide rails (10, 16) are asymmetrical.

6. The torsional damper of claim 5, wherein the two asymmetries are the same.

7. Torsional damper according to any of the preceding claims, characterized in that the deformation is greater in the case of a coupling roller (11) departing from the free end of the leg spring from the zero position (N) than in the case of a movement of the coupling roller (11) towards the free end of the leg spring.

Technical Field

The invention relates to a torsional vibration damper, comprising a drive part which can rotate about a rotational axis, and a damper mass part which is coupled to the drive part by means of at least one spring element and can be rotated relative to the drive part about the rotational axis against the spring element.

Background

Such torsional vibration dampers are used in torque transmission arrangements for damping or dampening possible torsional vibrations, which are referred to as so-called spring-mass dampers. For example, such torsional vibration dampers are arranged in the drive train of a motor vehicle for damping possible torsional vibrations which are in most cases inherently introduced into the drive train by the internal combustion engine.

Such torsional vibration dampers usually consist of a drive part which can be rotated about an axis of rotation and is loaded with a torque, and a damper mass part which can be rotated relative to the drive part about the axis of rotation, wherein the drive part and the damper mass part can be rotated relative to one another, against the action of a spring, by means of one or more helical compression springs which are distributed around the circumference. Due to the moment of inertia, the damper mass part is displaced against the action of the helical compression spring when torsional vibrations are introduced by the drive part, so that energy is drawn off by the spring deformation and the relative movement of the vibrations and is again fed to the system with a delay in time. This damping makes it possible to stabilize the torque profile, so that the torque generated, for example, to be transmitted to the downstream transmission is more uniform. An exemplary embodiment for torsional vibration dampers of this type is described in DE 19840664 a1 or DE 102014223308 a 1.

Internal combustion engines (i.e., internal combustion engines) generally have a periodic excitation in which the torque magnitude in the direction of rotation is greater than the opposite torque magnitude. However, the known torsional vibration dampers do not allow direction-specific damping, so that vibrations or the amplitude of the moments cannot be optimally damped.

Disclosure of Invention

The invention is therefore based on the task of: an improved torsional damper is proposed with respect to this.

In order to solve this problem, according to the invention, in a torsional vibration damper of the type mentioned at the outset, provision is made for: the spring element is embodied as a leg spring, which is coupled at one end to the drive part in a rotationally fixed manner and at the other end to the damper mass via a coupling roller which rolls on at least one guide rail, wherein the guide rail is asymmetrical such that the leg spring deforms with different strength depending on the direction of movement when moving out of the zero position.

The invention provides an asymmetrically designed torsional vibration damper in which the damping action depends on the direction in which the vibration is situated, i.e. whether the vibration is generated in the direction of rotation or vice versa. Thus, a direction-sensitive damping of the torque amplitude can be achieved for this. For this purpose, according to the invention: the spring element is embodied as a leg spring which is coupled to a drive by means of which vibrations are introduced into the damper. The leg spring is coupled to the damper mass via a coupling roller, so that a relative movement of the leg spring with respect to the damper mass (i.e. thus a vibration-sensitive movement in and against the direction of rotation) is possible. The coupling rollers roll on the at least one guide rail, so that the coupling position of the leg spring to the damper mass in the circumferential direction changes accordingly as a function of the degree of torsion of the leg spring relative to the damper mass in relation to the zero position (i.e. unloaded state) and the leg spring is deformed accordingly as a function of the degree of torsion, as a result of which the actual damping occurs.

Now, according to the invention: the guide rail is designed asymmetrically, so that the leg spring, starting from the zero position, deforms to a different extent in a movement in the direction of rotation than in a movement opposite to the direction of rotation. As a result, an asymmetrical characteristic curve of the damper is achieved with respect to the zero position (ultimately with respect to the spring action in the circumferential direction) by means of the guide rail (which is a curved rail), so that different rigidity or rigidity changes occur in the leg spring depending on the direction of movement and the degree of deformation. Preferably, a greater rigidity or rigidity change is active when the damper reacts to the greatest moment amplitude (i.e. when the leg spring is twisted in the direction of rotation).

The torsional vibration damper is therefore characterized by an asymmetrical spring characteristic which is realized by an asymmetrical guide rail or running rail on which the coupling rollers roll. Since the roller contact point moves in the direction of the leg or the attachment region of the leg spring to the drive element when the leg spring is twisted relative to one another in the direction of rotation, the lever arm of the leg spring is shorter and the stiffness is increased, i.e. the damper acts here against a greater or maximum moment amplitude. When moving counter to the direction of rotation, the bearing point or the roller contact point moves in the other direction to the leg end, the lever becomes longer and the rigidity decreases, i.e. the damper opposes a lower moment amplitude here.

Different inventive variants arise with regard to the coupling of the leg spring and the damper mass and therefore with regard to the arrangement of the guide rail. According to a first variant, the guide rail can be arranged on the leg spring, and the coupling roller is arranged on the damper mass in a rotatable manner about a fixed axis of rotation. That is, the coupling roller is positionally fixed relative to the leg spring and its guide rail, which move past it. Embodiments of the coupling roller as a step roller (which is optionally supported on the damper mass in one piece or, however, rotatably about a bearing journal) can be envisaged. A corresponding slotted receptacle is provided on the damper mass part, in which the coupling roller is received, the leg spring being positioned with its guide rail, viewed radially, below the receptacle.

Alternatively, the second variant provides for: the guide rail is formed on the damper mass part, and the coupling roller is arranged on the leg spring in a rotatable manner about a fixed axis of rotation. In this case, i.e., in the reverse configuration, the coupling rollers are arranged in a stationary manner on the leg springs, and the guide rail is located on the damper mass, wherein the coupling rollers can also be embodied, for example, as step rollers.

Finally, the third inventive variant provides: the guide rails are formed both on the leg spring and on the mass of the damper, and the coupling roller is guided in a freely rotatable manner on the two guide rails, wherein at least one guide rail is asymmetrical, but preferably the two guide rails are asymmetrical. The coupling roller is therefore freely movably received between the leg spring and the damper mass, which involves a simple roller which is received between the two mass sections and is axially supported there, if appropriate by lateral support projections or cone noses. In order to achieve an asymmetrical damper characteristic or spring characteristic, it is sufficient if only one guide rail is asymmetrical. It is practical, however, that the two guide rails are asymmetrical, wherein the asymmetries correspond to one another. This also makes it possible to increase the degree of rigidity change or deformation of the leg spring, in particular in the case of a corresponding deflection in the direction of rotation, more significantly, since the two guide rail geometries or guide rail asymmetries "add up" to some extent.

The two asymmetries can be identical here, i.e. the guide rails have the same geometry. However, it is also conceivable for reasons of space to have different configurations of the asymmetry. In this case, the spring stiffness and thus the corresponding adjustment or variation of the spring characteristic curve can be achieved over a relatively wide range.

As already explained, it is practical for the deformation to be greater when the coupling roller is moved away from the free end of the leg spring starting from the zero position than when the coupling roller is moved toward the free end of the leg spring, since the moment amplitude is greater than this opposite moment amplitude, as viewed in the direction of rotation, and a correspondingly greater damping is required, resulting in a correspondingly greater increase in rigidity in the deformed leg spring.

Drawings

The invention is explained below with reference to the figures according to embodiments. The figures are schematic representations and show:

fig. 1 is a schematic representation of a clutch device with an integrated torsional vibration damper according to the invention, fig. 2 is a schematic representation of a torque curve of the vibrations generated in the system to be damped by the torsional vibration damper,

figure 3 is a schematic view of a torsional damper according to a first inventive configuration,

FIG. 4 is a schematic illustration of the kinematic coupling of the leg spring and the damper mass of the torsional damper of the second embodiment, an

Fig. 5 is a comparative illustration of the coupling without the asymmetry according to the invention inside the guide rail.

Detailed Description

Fig. 1 shows an example of an application for a torsional vibration damper 1 according to the invention in a schematic representation. The torsional vibration damper is arranged in the clutch housing 2 and is connected downstream of the clutch device 3 for closing the force flow. The clutch device 3 is coupled in a known manner to a driven shaft of an internal combustion engine, not shown, the torque of which is transmitted in a known manner to the torsional vibration damper 1 via the clutch device 3. For this purpose, the torsional vibration damper has a drive element 4 which can be rotated about a common axis of rotation and which is placed on a corresponding shaft, which is guided to a transmission, not shown, by means of an internal toothing. Downstream of the torsional vibration damper 1, a torque converter device 5 is connected, which is also only shown here in a basic manner. The basic construction of such an arrangement and its function are sufficiently known.

Inherently, the system introduces torque vibrations via the internal combustion engine, not shown, into the arrangement shown in fig. 1, which should be damped by the torsional vibration damper 1. A schematic representation of such a profile of the torque oscillation or torque amplitude is shown in fig. 2. Along the abscissa, the time t is taken, along the ordinate, the differential torque Δ M, which varies as a function of the oscillation in the vicinity of the zero position, wherein the amplitude (+) in the direction of rotation of the drive shaft is shown in the upward direction along the ordinate and the opposite amplitude (-) is shown in the downward direction along the ordinate. As fig. 2 shows, the amplitude in the direction of rotation is significantly greater than the opposite amplitude, i.e., an asymmetrical torque profile results.

According to the invention, such asymmetrical torque fluctuations can now be damped in a targeted manner by means of the torsional vibration damper 1.

Fig. 3 shows a schematic representation of a torsional vibration damper 1 according to the invention in a first embodiment. A drive 4 is shown, which is coupled to the shaft already described by means of a corresponding internal toothing 6. Attached to the drive element 4 is a leg spring 7 which projects radially from the latter, wherein the leg spring 7 can be fixedly and integrally fixed to the drive element 4, but can also be rotatably supported on the drive element about a hinge or the like to a small extent.

The leg spring 7 has a leg 9 which is directed toward the surrounding damper mass 8 and on which a guide rail 10 is formed, on which a coupling roller 11, which in the exemplary embodiment shown is mounted on the damper mass 8 so as to rotate about a fixed axis of rotation 12 (see also fig. 1 for this purpose), moves. The damper mass 8 (see fig. 1) has for this purpose two axial disks 13, between which a mass ring 14 is arranged. The coupling rollers 11 are received with respective bearing journals 15 in respective bearing recesses in the axial disks 13 and are mounted so as to rotate about a fixed axis of rotation. In fig. 3, the two axial discs are not shown for reasons of clarity.

The drive element 4 is therefore coupled via the leg spring 7 and the coupling roller 11 to the damper mass 8, which can be rotated relative thereto. In this case, the leg spring 7 or the driver 4 can be rotated relative to the damper mass 8, i.e., the relative position of the leg spring 7 relative to the damper mass 8 can be changed. This change results in a change in the contact point of the coupling roller 11 on the guide rail 10, which is decisive for the variation of the degree of rigidity of the leg spring 7 and thus for the degree of damping.

For this purpose, the guide rail 10 is embodied asymmetrically, as is clearly shown in fig. 3. It is assumed that fig. 3 shows a zero position, i.e. the torque is free of vibrations and acts constantly. The contact point of the coupling roller 11 to the guide rail 10 is located to a certain extent (viewed in the circumferential direction) in the region of the middle of the rail.

Now, if the driver 4 is twisted in the rotational direction (indicated by arrow P1) relative to the damper mass 8 as a result of the vibrations, a corresponding positive vibration amplitude of + Δ M occurs, as indicated by arrow P2 in fig. 3. I.e. on the positive amplitude branch according to fig. 2.

This relative twisting in the direction of rotation causes the coupling roller 11 to move to the right on the guide rail 10. In this section 10a of the guide rail 10, the rail geometry is visibly raised, i.e. the coupling rollers 11 move to some extent on the raised rail. Because of the presence of such a rail elevation adjacent or close to the attachment region of the leg spring 7 on the drive element 4, a shortening of the free lever of the leg 9 or of the leg spring 7 takes place, and a stronger deformation of the leg 9 or of the leg spring 7 and thus a significant increase in rigidity is caused by the rail elevation, which is currently required, since the torque amplitude in the case of a torsion in the direction of rotation is clearly increased more strongly than in the opposite case, see fig. 2.

In the case of a load reduction, i.e. when the relative rotation is reversed and the return occurs, the coupling roller 11 moves in the opposite direction from the guide section 10a onto the guide rail section 10b extending towards the free leg end, which guide rail section 10b is visibly flatter, i.e. is arched significantly less, than the guide rail section 10 a. I.e. the contact point or bearing point of the coupling roller 11 on the guide rail 10 is moved to the left, the free lever or lever length is increased significantly. The degree of deformation is significantly less in this lever section and therefore the stiffness change is also significantly less. This corresponds to a lower torque amplitude (corresponding to fig. 2) in the case of a rotation counter to the direction of rotation. That is, the degree of damping must be lower, which is achieved by different rail geometries in the guide rail section 10 b.

This means that depending on the direction of movement of the leg spring 7 or the leg 9 relative to the damper mass 8 and thus on the direction of movement of the coupling roller 11 on the guide rail 10 in the direction of rotation or counter to the direction of rotation, a changing stiffness is produced in the leg spring 7 as a result of a changing deformation of the leg spring 7, which deformation can be matched very well to the actual profile of the moment amplitude, which can be correspondingly effectively damped here.

In the configuration according to fig. 3, the coupling rollers 11 are mounted on the damper mass 8 so as to be rotatable about a stationary rotational axis 12, while fig. 4 shows a configuration in which the coupling rollers 11 are received freely movably between the axial disks 13, i.e. they are movable in the circumferential direction between said axial disks.

The leg spring 7 is shown in part with its guide rail 10, which is also here clearly strongly asymmetrical. The guide rail section 10a is again shown, which is thus strongly raised, and the guide rail section 10b is joined to the guide rail section 10a toward the free end. The radial variation of the contact point of the coupling roller 11 as a function of the deflection direction or the direction of movement is also illustrated by way of example in this figure. The zero position N is shown on the one hand. If the coupling roller 11 is moved to the right, i.e. is twisted in the direction of rotation, the contact point moves onto the strongly raised guide rail section 10a, the contact point K being shown by way of example ₁I.e. the coupling roller 11 moves relatively far to the right with respect to the stretch a.

Fig. 4 also shows the contact point K in the case of a leftward movement of the coupling roller 11 by the same distance a into the guide rail section 10b ₂That is, the following is also shown: twisting is performed in the opposite direction. It can be seen that the contact point K, as shown by Δ K ₁In radial direction with respect to the contact point K ₂Clearly further outside.

Fig. 4 also shows a damper mass 8 or mass element 14, on which a second guide rail 16 is formed. The second guide rail is arranged diametrically opposite the first guide rail 10, but it also has two different rail sections 16a and 16 b. The rail sections 16a and 16b preferably correspond in their geometry to the rail sections 10a and 10b, but need not be.

Since the coupling roller 11 is freely movably received between the guide rails 10 and 16a, it moves both on the rail section 10a and on the rail section 16a when the leg spring 7 is twisted in the direction of rotation. Due to the movement on the rail section 10a, the deformation or bending of the leg spring 7 takes place radially inwards. This deformation is supported by the movement on the inwardly directed rail section 16a, i.e. the two geometries of the rail sections 10a and 16a "add up" to some extent, which is significantly greater or doubled compared to the configuration according to fig. 3.

In the case of a rotation against the direction of rotation, the coupling rollers 11 move on the rail sections 10b and 16b to a significantly lower extent, and the stiffness change is correspondingly smaller.

Finally, fig. 5 shows the case in which the guide rails 10, 16 are designed symmetrically on the leg spring 7 or the damper mass 8, for comparison purposes only. It can be seen that the contact point K is respectively deflected for the distance a starting from the zero position N ₁And K ₂On the same level, that is to say the change in rigidity or deformation of the leg spring 7 in the case of movement in both directions away from the zero position is the same. Since the guide rail 16 is also embodied symmetrically, no changes in the deformation or rigidity behavior are produced by the movement of the coupling roller 11 on the guide rail 16.

Finally, it is conceivable that, instead of the asymmetrical embodiment of the guide rails 10 and 16 shown in fig. 4, only one of the two is embodied asymmetrically, but the coupling rollers 11 are still freely guided on them.

Alternatively, it is naturally also conceivable for fig. 3 for the guide rail 10 to be configured symmetrically and for the guide rail 16 to be configured asymmetrically.

List of reference numerals

1 torsional vibration damper

2 Clutch housing

3 Clutch device

4 driving piece

5 torque converter device

6 internal tooth part

7 leg spring

8 buffer mass

9 legs

10 guide rail

10a guide rail section

10b guide rail section

11 coupling roller

12 axis of rotation

13 axial disc

14 mass ring

15 bearing journal

16-rail section

16a rail section

16b rail section

P1 arrow

P2 arrow

N zero position

K ₁Contact point

K ₂A point of contact.