Vibration damper

Abstract

The invention relates to a vibration damper (16) with a fixing means (18) which fixes a damper mass (10) if a first limit frequency is reached or undershot and if a second limit frequency is reached or overshot. Here, the damper mass (10) is preferably fixed solely by the rotational-frequency-dependent centrifugal force which displaces, for example, one control mass (20) per damper mass (10) in the radial direction R counter to the force of a control spring (22). Resonances which lie outside the limit frequencies can be avoided by the vibration damper (16) being switched on and off. When used in the drive train of a motor vehicle, with the vibration damper (16) preferably being connected in the drive train here to a crankshaft of an internal combustion engine, the vibration behaviour of the crankshaft can be influenced in such a way that the amplitude of the vibrations of the crankshaft lies below a predefinable limit amplitude.

Description

[0001] Vibration damper

[0002] The present invention relates to a vibration damper, preferably for use in the drive train of a motor vehicle, in particular on a crankshaft of an internal combustion engine.

[0003] The crankshaft vibrations generated by combustion engines must be dampened. Various vibration dampers, or vibration absorbers, are available for this purpose. Vibration dampers, such as those described in DE 102019 128 149 A1, can reliably dampen first-order crankshaft vibrations, but they lead to the formation of additional resonances in the vibration spectrum, which can be damaging to the crankshaft.

[0004] Proceeding from this, the object of the present invention is to provide a vibration damper which reliably dampens the crankshaft vibration without causing further resonances in the vibrating system. This object is achieved with the features of the present independent claim. Dependent claims are directed to advantageous developments. The vibration damper according to the present invention comprises at least one damper mass carrier and at least one damper mass which is movably mounted on the damper mass carrier, wherein the vibration damper is rotatable about a shaft with a rotation axis, wherein the vibration damper comprises a locking device which fixes the damper mass relative to the shaft when a rotational frequency of the vibration damper is less than or equal to a first cut-off frequency or greater than or equal to a second cut-off frequency. The second cut-off frequency is greater than the first cut-off frequency.The cutoff frequencies enclose a frequency band within which the damper mass can oscillate freely. At rotation frequencies lower than the first cutoff frequency or higher than the second cutoff frequency, the damper mass is fixed and cannot oscillate. The free oscillation of the damper mass dampens vibrations in the system through destructive interference.

[0005] The rotation axis defines an axial direction. A radial direction is perpendicular to the rotation axis. The terms "radially inward" and "radially outward" used in this document refer to the radial direction and the rotation axis, respectively. A radially inward displacement of an element therefore means that the element moves radially in the direction of the rotation axis. Similarly, a radially outward displacement of an element means that the element moves radially away from the rotation axis.

[0006] Damping a vibration is understood in particular to mean the cancellation of a vibration, in particular through destructive interference and / or the dissipation of energy. In this respect, a vibration damper can also be understood as a vibration absorber. The vibration damper is preferably designed as a centrifugal pendulum, preferably comprising a centrifugal pendulum input for introducing a torque into the centrifugal pendulum and a centrifugal pendulum output for outputting a torque from the centrifugal pendulum, as well as at least one damper mass (pendulum mass) arranged between the centrifugal pendulum input and the centrifugal pendulum output, which is arranged at least on a first damper mass carrier (pendulum flange), and wherein each pendulum mass is movably guided in a guide assigned to it along a pendulum track relative to the pendulum flange.

[0007] Alternatively, it is also preferred that the vibration damper is designed as a spoke damper with a spoke spring having an outer ring for forming at least part of the damper mass. The spoke spring has an outer ring for forming at least part of the damper mass and a fastening disc for fastening to a shaft subject to torsional vibrations, the outer ring being connected to the fastening disc via essentially radially extending resilient spokes. The spoke damper preferably also has a rotationally rigid flange plate for forming a friction surface and a spring element that is motion-coupled to the outer ring of the spoke spring, the spring element acting with a spring force on the friction surface of the flange plate.

[0008] The present invention enables the vibration damper to be switched on and off, depending on the rotational frequency. When connected to a crankshaft of an internal combustion engine, the rotational frequency is determined by the rotational speed of the internal combustion engine. By adding a vibration damper to a shaft, for example and particularly preferably a crankshaft of an internal combustion engine, resonance of the crankshaft at an original resonant frequency is suppressed. However, the vibration damper typically causes the creation of two new resonances, one of which occurs at a first resonant frequency that is lower than the original resonant frequency and the other at a second, higher resonant frequency.By defining the first cutoff frequency between the first resonant frequency and the original resonant frequency, and the second cutoff frequency between the original resonant frequency and the second resonant frequency, the resonance at the original resonant frequency can be damped, while the switching on and off prevents the damper mass from oscillating at the first and second resonant frequencies, thus preventing this resonance from occurring. This ensures that none of the aforementioned resonant frequencies is passed through.

[0009] Preferably, an internal combustion engine with a crankshaft as the shaft is proposed.

[0010] Preferably, the damper mass is determined by centrifugal force. Centrifugal force is directly proportional to the rotational frequency. This proportionality can be easily used to define the cutoff frequencies. Furthermore, this is possible with a simple design, for example, by displacing a control mass or deforming a braking element through centrifugal force, which are then applied to the damper mass to determine its position. In particular, this eliminates the need for electronic deactivation of the vibration damper.

[0011] Preferably, the locking device provides a force-locking, form-locking, or force-and-form-locking fixation of the damper mass. This can be achieved, for example, by a frequency-dependent displacement or deformation of braking elements or similar components, which are displaced or deformed outward under the influence of centrifugal force as the rotation frequency increases.

[0012] The locking device preferably comprises two stops against which the damper mass can be applied when the frequency falls below the first cutoff frequency and when the second cutoff frequency is exceeded. In this way, a locking device can be designed in a simple manner. Examples of such stops are control masses or parts of control masses or brake plates. The locking device preferably comprises at least one control mass that is displaced by a centrifugal force against a spring force. Preferably, one or two control masses are formed per damper mass. The frequency-dependent displacement of the at least one control mass applies or releases it from the damper mass. The spring force, which is preferably applied by one or two control springs, ensures a restoring moment. In this way, the locking device can be implemented in a simple manner.The first and second cutoff frequencies can be determined via the respective spring constant, the mass of the control mass, and the geometric configuration. Preferably, the at least one control mass causes a force and / or form fit between the control mass and the damper mass when the first cutoff frequency is reached or exceeded, and when the second cutoff frequency is reached or exceeded.

[0013] In this context, the control mass particularly preferably has a recess in which a damper mass is at least partially formed. Thus, the geometry, i.e., the shape and extent, of the recess can be used to create a force and / or form fit between the damper mass and the control mass. Furthermore, the extent of the recess in the radial direction relative to the extent of the damper mass in the radial direction forms a further degree of freedom in determining the cutoff frequencies. Preferably, an inner edge of the recess in the radial direction forms a first stop, and an outer edge in the radial direction forms a second stop for the damper mass.In particular, the damper mass is fixed to the second stop when the rotational frequency of the vibration damper about the rotational axis is less than or equal to the first cutoff frequency, while the damper mass is fixed to the first stop when the rotational frequency is greater than or equal to the second cutoff frequency. In this context, it is particularly preferred that the recess has a radial extension that is greater than a radial extension of the damper mass. This allows the control mass to move in a radial direction and, with the aid of the aforementioned stops, to fix the damper mass. At the same time, the damper mass can oscillate freely when the control mass is not in contact with the damper mass.

[0014] Preferably, the spring force acts in a radial direction, and the control mass is connected to the damper mass in a force-locking, form-locking, or both force-locking, manner at a rotational frequency less than or equal to the first cutoff frequency or greater than or equal to the second cutoff frequency. In this case, the spring force acts parallel but opposite to the centrifugal force, i.e., radially inward. This can be achieved by using tension springs or compression springs extending in a radial direction. Alternatively, tension springs or compression springs can be used that have at least a portion of the spring force acting in a radial direction.

[0015] The locking device preferably comprises at least one brake plate, which is resilient and deformable depending on the frequency under the influence of a centrifugal force. The resilient brake plate builds up a spring force that serves as a restoring force and ensures that the brake plate returns when the centrifugal force decreases. The brake plate is preferably made of spring steel. The brake plate is preferably formed as a tongue made of a carrier that is non-rotatably connected to the shaft. The respective cutoff frequency is defined by the length and thickness of the brake plate, as well as the spring constant of the brake plate.A configuration is also preferred in which a brake plate is combined with a control mass, so for example the brake plate causes the damper mass to be fixed when the second limit frequency is reached or exceeded, while the control mass causes the damper mass to be fixed when the second limit frequency is reached or undershot.

[0016] The locking device preferably has a first brake plate which bears against the damper mass at frequencies greater than or equal to the second cut-off frequency, and a second brake plate which bears against the damper mass at frequencies less than or equal to the first cut-off frequency. This can be achieved in particular by the first brake plate extending radially outside the damper mass and the second brake plate extending radially inside the damper mass. As the rotational frequency increases, the centrifugal force increases and the brake plates each bend radially outwards. When the first cut-off frequency is exceeded, the second brake plate detaches from the damper mass. The damper mass can oscillate freely until the second cut-off frequency is reached. Once the second cut-off frequency is reached, the first brake plate is bent so far that it bears against the damper mass and fixes it preferably in a force-fitting or form-fitting manner.In this context, it is preferred that the first brake plate and the second brake plate are secured to a carrier, with the first brake plate being secured further inside the carrier in a radial direction relative to the rotation axis than the second brake plate. The carrier, in turn, is preferably connected to the shaft in a rotationally fixed manner.

[0017] The invention and the technical environment are explained in more detail below with reference to the accompanying figures. It should be noted that the invention is not intended to be limited by the exemplary embodiments shown.

[0018] In particular, unless explicitly stated otherwise, it is also possible to extract partial aspects of the facts explained in the figures and combine them with other components and findings from the present description and / or figures. It should be noted in particular that the figures, and in particular the proportions shown, are only schematic. The same reference numerals denote the same objects, so that explanations for other figures can be used as a supplement if necessary. The following schematically show:

[0019] Fig. 1 shows a centrifugal pendulum known from DE 10 2019 128 149 A1 as an example of a vibration damper;

[0020] Fig. 2 to 4 show a first embodiment of the present invention at three different resonance frequencies in section;

[0021] Figs. 5 to 7 show a second embodiment of the present invention at three different resonance frequencies in section;

[0022] Fig. 8 to 10 show a third embodiment of the present invention at three different resonance frequencies in section; and

[0023] Fig. 11 shows an example of the vibration behavior of the vibration damper according to the present invention mounted on a crankshaft.

[0024] Fig. 1 shows a vibration damper 1 known from DE 10 2019 128 149 A1 and designed as a centrifugal pendulum 2 for speed-dependent

[0025] Torsional vibration damping within a drive train 4 of a motor vehicle (not shown here), comprising a centrifugal pendulum input 6 for introducing a torque into the centrifugal pendulum 2 and a centrifugal pendulum output 8 for outputting a torque from the centrifugal pendulum 2, as well as a plurality of damper masses 10 (pendulum masses) arranged between the centrifugal pendulum input 6 and the centrifugal pendulum output 8, which are arranged at least on a first damper mass holder 12 designed as a pendulum flange. In the illustration shown, the centrifugal pendulum 2 is arranged externally on a dual-mass flywheel (not designated in more detail). Of course, other arrangements of the centrifugal pendulum 1 or other configurations of the vibration damper 1 are also possible. Fig. 1 is intended to represent merely one exemplary possibility for a better understanding of the invention.The centrifugal pendulum 2 rotates around a rotational axis 14 on a shaft 15, which is only schematically indicated and represents, for example, the crankshaft of an internal combustion engine. A radial direction R is defined perpendicular to the rotational axis 14. An axial direction A is defined in the direction of the rotational axis 14.

[0026] Figures 2 to 10 show exemplary embodiments of vibration dampers 16 according to the present invention. Only the position of the damper masses 10 is shown. The remainder of the vibration damper 16 can be designed as shown in Fig. 1, i.e., as a centrifugal pendulum with two damper masses 10 arranged in the axial direction A on opposite sides of the pendulum flange 12 and optionally connected to one another, or as a centrifugal pendulum with damper masses 10 arranged only on one side of the pendulum flange 12. In both described combinations, several damper masses 10 can be distributed over the circumference of the centrifugal pendulum. Other vibration dampers 16 are also possible, in which a damper mass can execute vibrations against a damper mass holder.

[0027] Figures 2 to 4 show a first embodiment of the present invention at three different rotational frequencies of the vibration damper 16 about the rotational axis 14 shown in Figure 1, each as sectional views. The rotational frequency increases continuously from Figure 2 to Figure 4. All three figures are described together below unless explicitly referenced to a figure.

[0028] In addition to the damper mass 10, which is mounted so as to be displaceable against a damper mass holder (not shown here), the vibration damper 16 also has a locking device 18, which in the present example comprises a control mass 20, which can be displaced in a frequency-dependent manner by a centrifugal force in the radial direction R against a spring force applied via a control spring 22. The control spring 22 in this exemplary embodiment is a tension spring and is fixed at its end opposite the control mass 20 in a rotationally fixed manner to the shaft 15 (cf. Fig. 1 ), about which the vibration damper 16 is rotatable. In this example, the control mass 20 has a recess 24 in which the damper mass 10 is partially received. A radial extension 26 of the recess 24 is greater than a radial extension 28 of the damper mass 1 (cf. Fig. 2).The extension of the control mass 20 in the circumferential direction relative to the axis of rotation is preferably smaller than the extension of the damper mass 20. The control mass 20 is displaceable in the radial direction R. A centrifugal force 30 acts on the control mass 20 in an outward radial direction R (see Fig. 3) depending on the frequency and, when connected to a drive train of a motor vehicle, also depending on the speed of an internal combustion engine configured there. In the inward radial direction R, a spring force 32 acts through the control spring 22 (see Fig. 3), which causes a restoring moment. Centrifugal force 30 and spring force 32 are in equilibrium at a constant rotational frequency. If the rotational frequency increases, the centrifugal force 30 increases and the control mass 20 is displaced radially outward. At the same time, the spring force 32 increases until equilibrium is again reached between the centrifugal force 30 and the spring force 32.If the rotation frequency is reduced, the centrifugal force 30 decreases and the control mass 20 shifts inward in the radial direction R under the action of the spring force 32. At the same time, the spring force 32 decreases until a new equilibrium is reached between the centrifugal force 30 and the spring force 32.

[0029] The movement of the control mass 20 is limited by stops, as described below. Fig. 2 shows a situation in which the rotational frequency of the centrifugal pendulum 16 is lower than a first limit frequency. In this case, the spring force 32 of the control spring 22 acting inward in the radial direction R causes a radial outer side 34 of the recess 24 to bear against a radial outer contour 36 of the damper mass 10. The radial outer side 34 of the recess 24 and the radial outer contour 36 of the damper mass 10 correspond at least in some areas in shape and dimension, so that a positive and non-positive connection occurs between the damper mass 10 and the control mass 20. In this case, a longitudinal side 40 of the damper mass 10 directed in the axial direction A towards an inner wall 38 of the recess 24 does not bear against the inner wall 38. A gap remains between the inner wall 38 and the long side 40, which is always present regardless of the rotation frequency (see Fig.3 and 4). The radial outer side 34 acts as a first stop 35, which limits the movement of the control mass 20 relative to the damper mass 10.

[0030] If the rotational frequency is lower than the first cutoff frequency, the spring force 32 is greater than the centrifugal force 30, since the spring force 32 cannot be further reduced by the contact of the control mass 20 with the damper mass 10, but the centrifugal force 30 is further reduced. At the first cutoff frequency, the spring force 32 corresponds to the centrifugal force 30. In these cases, the positive and frictional engagement of the radial outer contour 36 of the damper mass 10 with the radial outer side 34 of the recess 24 results in the damper mass 10 being fixed to the control mass 20. This case is shown in Fig. 2.

[0031] If the rotation frequency exceeds the first limit frequency, the control mass 20 detaches from the damper mass 10 and is displaced outwards in the radial direction R. As long as the rotation frequency does not exceed the second limit frequency, the damper mass 10 can move freely relative to the

[0032] Damper mass holders 12 (see Fig. 1) oscillate and are used to dampen vibration irregularities. The situation shown in Fig. 3 applies here.

[0033] If the rotational frequency also reaches or exceeds a second limit frequency, a force and form fit occurs between a radial inner side 42 of the recess 24 and a radial inner contour 44 of the damper mass 10.

[0034] The inner side 42 and the inner contour 44 are configured to correspond to one another. The damper mass 10 is thus fixed to the control mass 20 in a force-fitting and form-fitting manner. This situation is shown in Fig. 4. The inner side 42 of the recess 24 acts as a second stop 43, which limits the movement of the control mass 20 relative to the damper mass 10.

[0035] Due to the design of the centrifugal pendulum 16 shown in Figures 2 to 4, the control mass 20 is connected to the damper mass 10 in a force-locking and form-locking manner when the rotational frequency of the centrifugal pendulum 16 is less than or equal to the first cutoff frequency or greater than or equal to the second cutoff frequency. The first cutoff frequency and the second cutoff frequency are determined by the selection of the control mass 20, the control spring 22, the radial extension 26 of the recess 24, and the radial extension 28 of the damper mass 10.

[0036] Control mass 20 is generally preferably made of a metallic material and has an additional friction lining, particularly on the surfaces to which the damper mass 10 can be applied, in order to increase the friction force between damper mass 10 and control mass 20.

[0037] Figures 5 to 7 show a second embodiment of the present invention at three different rotational frequencies, each in cross-section. The rotational frequency increases continuously from Figure 5 to Figure 7. Figures 5 to 7 are described jointly below unless explicitly referenced to a figure. To avoid repetition, only the differences from the first embodiment are described; otherwise, reference is made to the above description of the first embodiment.

[0038] In contrast to the first embodiment, the locking device 18 of the centrifugal pendulum 16 in the second embodiment does not have one control mass, but two control masses. A first control mass 46 is connected in a rotationally fixed manner directly or indirectly to the shaft 15 (see Fig. 1) about which the vibration damper 16 is rotatable via a first control spring 48, which is designed as a tension spring. While a second control mass 50 is also connected in a rotationally fixed manner directly or indirectly to the shaft 15 via a second control spring 52, which is designed as a compression spring.

[0039] Both the first control mass 46 and the second control mass 50 are fixed with respect to the shaft 15 and thus rotate with it about the rotation axis 14 (see Fig. 1). In the radial direction R, the second control mass 50 is located further out than the first control mass 46, so that the influence of the radius on the centrifugal force is smaller for the first control mass 46 than for the second control mass. If the mass of the first control mass 46 is identical to the mass of the second control mass 50, a greater centrifugal force is exerted on the second control mass 50.

[0040] The applied forces are shown as an example in Fig. 7. A first centrifugal force 54 and a first spring force 56 act on the first control mass 46. These are opposite to one another because the first control spring 48 is designed as a tension spring. The first centrifugal force 54 points outwards in the radial direction R, while the first spring force 56 points inwards in the radial direction R. A second centrifugal force 58 acts on the second control mass 50 and points outwards in the radial direction R. In addition, a second spring force 60 acts and points inwards in the radial direction R. At the second cut-off frequency, the magnitude of the first centrifugal force 54 is equal to the magnitude of the first spring force 56, as shown in Figure 7. If the rotational frequency exceeds the second cut-off frequency, the first centrifugal force 54 increases further, while the first spring force 56 remains the same.However, due to the frictional connection between damper mass 10 and first control mass 46, this does not lead to a further displacement of the first control mass 46.

[0041] The first control mass 46 and the second control mass 50 have inclined surfaces 62 facing the damper mass 10, which correspond in terms of inclination to two inclined surfaces 64 of the damper mass 10. If one of the inclined surfaces 64 of the damper mass 10 rests against one of the inclined surfaces 62 of a control mass 48, 50, the damper mass 10 is non-positively secured to the respective control mass 48, 50, and vibration of the damper mass 10 is prevented.

[0042] Figures 5 to 7 show the situation of damper mass 10 and control masses 46, 50 at different rotational frequencies. Fig. 5 shows the situation at a rotational frequency less than or equal to the first cut-off frequency. Here, the damper mass 10 is non-positively attached to the second control mass 50, and the corresponding inclined surface 64 of the damper mass 10 is non-positively connected to the inclined surface 62 of the second control mass 50. Fig. 6 shows the situation at a rotational frequency between the first cut-off frequency and the second cut-off frequency. Here, the damper mass 10 can oscillate freely; it is not in frictional connection with one of the control masses 46, 50. Fig. 7 shows the situation at a rotational frequency greater than or equal to the second cut-off frequency. Here, the damper mass 10 is in frictional connection with the first control mass 46 and therefore rotates together with the shaft 15 (cf. Fig. 1). The damper mass 10 cannot therefore oscillate.The inclined surfaces 62 of the control masses 46, 50 and / or the inclined surfaces 64 of the damper mass 10 are preferably provided with a friction lining.

[0043] The first cutoff frequency and the second cutoff frequency are determined by the selection of the control masses 46, 50, the control springs 48, 52, the positioning of the control masses 46, 50 in the radial direction R and the radial extension 28 of the damper mass 10.

[0044] Figures 8 to 10 show a third embodiment of the present invention at three different rotational frequencies, each very schematically in section. The rotational frequency increases from Figure 8 to Figure 10. Figures 8 to 10 are described jointly below unless explicitly referenced to a figure. To avoid repetition, only the differences from the first and second embodiments are described; otherwise, reference is made to the above description of the first and second embodiments.

[0045] Figures 8 to 10 show a third embodiment of a vibration damper 16 at three rotational frequencies. In contrast to the first and second embodiments, the locking device 18 in the third embodiment does not have a control mass actuated by a control spring. Rather, the pendulum mass 10 is fixed at a rotational frequency less than or equal to the first cutoff frequency (Fig. 8) and at a rotational frequency greater than or equal to the second cutoff frequency (Fig. 10) via a first brake plate 66 and a second brake plate 68. It should be noted that, in addition to the three embodiments shown here, embodiments are also possible in which a brake plate 66, 68 is combined with a control mass 46, 50 with a corresponding control spring 48, 52.

[0046] The brake plates 66, 68 are connected to a carrier 70, which in turn is rotationally fixedly connected to the shaft (not shown here) (see Fig. 1). The brake plates 66, 68 are preferably formed as tongues from the carrier 70 or are connected to it, for example, via rivets or welds (not shown). The brake plates 66, 68 are designed as spring plates that can be deformed against a restoring spring force. The deformation is caused by centrifugal force. This increases at each of the brake plates 66, 68 from Fig. 8 to Fig. 10 with increasing rotation frequency. Depending on the frequency, a bending of the spring plates 66, 68 occurs - only indicated in Figures 8 to 9 - so that the free ends 72 of the spring plates 66, 68 shift outwards (upwards in Figures 8 to 10) in the radial direction R (see Figure 8) with increasing resonance frequency. At a rotation frequency less than or equal to the first cutoff frequency (Figure 8).8), the second brake plate 68 rests against an inclined surface 64 of the damper mass 10, and the damper mass 10 is fixed to the second brake plate 68 in a force-locking and form-locking manner. If the resonance frequency rises above the first cutoff frequency, the second brake plate 68 detaches from the inclined surface 64 of the damper mass 10 and releases the damper mass 10. The damper mass 10 can oscillate freely (Fig. 9). If the rotational frequency of the vibration damper 16 reaches or exceeds the second cutoff frequency, the increasing deformation of the first brake plate 66 causes it to rest against the corresponding inclined surface 64 of the damper mass 10 and fix it in a force-locking manner (Fig. 10).

[0047] In the third embodiment, the first and second cutoff frequencies are determined by the length, thickness and spring constant of the brake plates 66, 68, as well as the positioning in the radial direction R relative to the damper mass 10.

[0048] Fig. 11 schematically shows various vibration curves. The rotation frequency f is plotted on the abscissa axis, while the vibration amplitude, i.e., the vibration angle of a crankshaft, is plotted on the ordinate axis. A limit amplitude 74 is shown, which can be specified and should not exceed the vibration amplitude. A first vibration curve 76 indicates the vibration behavior of a crankshaft without a vibration damper. This exhibits a single resonance with a resonance frequency, referred to below as the original resonance frequency 78, whose amplitude lies significantly above the limit amplitude 74. If a vibration damper, such as a centrifugal pendulum absorber known from the prior art, is connected to the crankshaft as shown in Fig. 1, the second vibration curve 80 results.This now has two resonances at a first resonance frequency 82 and a second resonance frequency 84, at which vibration amplitudes are present that lie above the limit amplitude 74. If, instead of the centrifugal pendulum shown in Fig. 1, a vibration damper 16 according to the present invention is connected to the crankshaft, the third vibration curve 86 results. The first limit frequency 88 is selected such that it lies between the first resonance frequency 82 and the original resonance frequency 78 of the first vibration curve 76. The second limit frequency 90 is selected such that it lies between the original resonance frequency 78 of the first vibration curve 76 and the second resonance frequency 84.As a result, the vibration damper 16 according to the present invention enables the elimination of the resonance of the crankshaft without the vibration damper 16 adding new resonances whose amplitude lies above the limit amplitude 74. The invention relates to a vibration damper 16 with a locking device 18 that fixes a damper mass 10 when a first limit frequency is reached or undershot and when a second limit frequency is reached or overshot. The fixing of the damper mass 10 is preferably achieved solely by the rotational frequency-dependent centrifugal force, which, for example, displaces a control mass 20 per damper mass 10 in the radial direction R against the force of a control spring 22. By switching the vibration damper 16 on and off, resonances outside the limit frequencies can be avoided.When used in the drive train of a motor vehicle, wherein the vibration damper 16 is preferably connected to a crankshaft of an internal combustion engine in the drive train, the vibration behavior of the crankshaft can be influenced such that the amplitude of the vibrations of the crankshaft is below a predeterminable limit amplitude.

[0049] Reference symbol

[0050] 1 vibration damper

[0051] 2 centrifugal pendulums

[0052] 4 Drivetrain

[0053] 6 Centrifugal pendulum input

[0054] 8 Centrifugal pendulum output

[0055] 10 Damper mass

[0056] 12 damper mass holders

[0057] 14 Rotation axis

[0058] 15 Wave

[0059] 16 vibration dampers

[0060] 18 Locking device

[0061] 20 tax mass

[0062] 22 Control spring

[0063] 24 recess

[0064] 26 radial extension of the recess

[0065] 28 radial extension of the damper mass

[0066] 30 Centrifugal force

[0067] 32 spring force

[0068] 34 radial outside of the recess

[0069] 35 first attack

[0070] 36 radial outer contour of the damper mass

[0071] 38 Inner wall of the recess

[0072] 40 Long side of the damper mass

[0073] 42 radial inside of the recess

[0074] 43 second attack

[0075] 44 radial inner contour of the damper mass

[0076] 46 first tax mass

[0077] 48 first control spring

[0078] 50 second tax mass

[0079] 52 second control spring

[0080] 54 first centrifugal force

[0081] 56 first spring force

[0082] 58 second centrifugal force

[0083] 60 second spring force

[0084] 62 Inclined surface of the control mass

[0085] 64 Inclined surface of the damper mass

[0086] 66 first brake plate

[0087] 68 second brake plate

[0088] 70 carriers

[0089] 72 free end of the brake plate

[0090] 74 Limit amplitude

[0091] 76 first oscillation curve

[0092] 78 original resonance frequency

[0093] 80 second oscillation curve

[0094] 82 first resonance frequency

[0095] 84 second resonance frequency 86 third oscillation curve

[0096] 88 first cutoff frequency

[0097] 90 second cutoff frequency A axial direction f rotation frequency

[0098] R radial direction

Claims

Claims 1 . Vibration damper (16), comprising at least one damper mass carrier (12) and at least one damper mass (10) which is movably mounted on the damper mass carrier (12), wherein the vibration damper (16) is rotatable about a shaft (15) with a rotation axis (14), characterized in that the vibration damper (16) comprises a locking device (18) which fixes the damper mass (10) relative to the shaft when a rotational frequency of the vibration damper (16) is less than or equal to a first cut-off frequency or greater than or equal to a second cut-off frequency.

2. Vibration damper (16) according to claim 1, wherein the fixing of the damper mass (10) is effected by the centrifugal force (30,54,58).

3. Vibration damper (16) according to claim 1 or 2, wherein the locking device (18) effects a non-positive, positive or non-positive and positive fixing of the damper mass (10).

4. Vibration damper (16) according to one of the preceding claims, in which the locking device (18) comprises two stops (35, 43) against which the damper mass (10) can be applied when the first limit frequency is undershot and when the second limit frequency is exceeded.

5. Vibration damper (16) according to one of the preceding claims, in which the locking device (18) comprises at least one control mass (20, 46, 50) which is displaced by a centrifugal force (30, 54, 58) against a spring force (22, 48, 52).

6. Vibration damper (16) according to claim 5, wherein the control mass (20) has a recess (24) in which a damper mass (10) is at least partially formed.

7. Vibration damper (16) according to claim 6, wherein the recess (24) has an extension (26) in the radial direction (R) which is greater than a radial extension (28) of the damper mass (10).

8. Vibration damper (16) according to one of the preceding claims, in which the locking device (18) comprises at least one brake plate (66, 68) which is resilient and is deformable in a frequency-dependent manner under the action of a centrifugal force (30).

9. Vibration damper (16) according to claim 8, in which the locking device (18) a first brake plate (66) which bears against the damper mass (10) at frequencies greater than or equal to the second cutoff frequency, and a second brake plate (68) which bears against the damper mass (10) at frequencies less than or equal to the first cutoff frequency.

10. Vibration damper (16) according to claim 9, wherein the first brake plate (66) and the second brake plate (66) are fixed to a carrier (70), wherein the first brake plate (66) is fixed further inside on the carrier (70) in a radial direction (R) relative to the axis of rotation (14) than the second brake plate (68).

Images