DE102022110201A1 - Vibration damper - Google Patents

Abstract

Vibration damper with frequency-dependent control on a mechanical system (1, 2), comprising damping means (5) between the mechanical system (1, 2) and a damper mass (3), the damping means being a friction clutch mounted on the mechanical system with at least one on the Damper mass acting friction surface (12), wherein the contact pressure between the at least one friction surface and the damper mass can be adjusted and / or regulated in its strength.

Description

The invention relates to a vibration damper with frequency-dependent control on a mechanical system according to the first claim.

Vibration dampers of the type mentioned at the beginning are vibration absorbers. They usually consist of a damper mass that is connected to the mechanical system via a spring element and a damper element and is subjected to a vibration amplitude by this system. They serve to reduce vibration amplitudes, especially in the case of resonances in mechanical systems. In mechanical vibrating systems with vibration dampers, the design effort for resonance vibrations is reduced. This ultimately leads to light and efficient machines. These machines, in turn, lead to better product quality and financial savings. Furthermore, effective dampers ensure a longer service life and less downtime in machines, which ultimately also results in financial savings.

(c = Federsteifigkeit, m = Dämpfermasse). Eine wirksame Tilgung von Schwingungsamplituden in einem mechanischen System wird dann erzielt, wenn die Resonanzfrequenz des Schwingungstilgers gleich der Erregerfrequenz des mechanischen Systems gewählt und die Dämpfung d so klein wie möglich gehalten wird. Dies führt zu den geringsten Amplituden in einem Bereich, in dem die Resonanzfrequenz des Tilgers angepasst werden kann. Außerhalb dieses Bereiches ist die Wirksamkeit des Schwingungstilgers jedoch begrenzt.The resonance frequency Q of a vibration absorber is calculated for low damping $$\mathrm{\Omega }=\sqrt{\text{cm}}$$

(c = spring stiffness, m = damper mass). Effective cancellation of vibration amplitudes in a mechanical system is achieved when the resonance frequency of the vibration absorber is chosen to be equal to the excitation frequency of the mechanical system and the damping d is kept as small as possible. This leads to the lowest amplitudes in a range in which the resonance frequency of the absorber can be adjusted. However, outside this range, the effectiveness of the vibration absorber is limited.

[10] discloses an example of such a basic structure for a vibration absorber concept with a spring element, damper mass and friction clutch. The settings of the absorber components are not changed during operation.

A vibration absorber is adapted to a mechanical system by detecting the operating frequencies (frequency identification) of the system and then setting parameters, in particular the spring stiffness of the vibration absorber.

• - mittels einer Fourier Transformationen (vgl. [1]):
  • Hierbei wird die Erregerfrequenz durch die Messung von Systemzuständen und die Fast Fourier Transformation (FFT) identifiziert. Die Frequenzen der FFT-Peaks werden erkannt und aus diesen Frequenzen wird die entsprechende Resonanzfrequenz durch eine hinterlegte Logik ermittelt.
• - über eine Erfassung eines Abstands zwischen den Nulldurchgängen einer Schwingung (vgl. [2]):
  • Die Erregerfrequenz wird durch eine Messung der Zeit zwischen den Nulldurchgängen eines Systemzustandes erfasst. Die ermittelte Resonanzfrequenz ist durch den Zusammenhang Ω_id = 1/(2Δt) gegeben, worin Δt der zeitliche Abstand zweier aufeinander folgenden Nulldurchgängen darstellt.

Frequency identification can be done in several ways:

• - using a Fourier transformation (see [1]):
  • The excitation frequency is identified by measuring system states and the Fast Fourier Transformation (FFT). The frequencies of the FFT peaks are recognized and the corresponding resonance frequency is determined from these frequencies using stored logic.
• - via recording a distance between the zero crossings of an oscillation (see [2]):
  • The excitation frequency is recorded by measuring the time between zero crossings of a system state. The determined resonance frequency is given by the relationship Ω _id = 1/(2Δt), where Δt represents the time interval between two successive zero crossings.

Both approaches are adaptive control methods in which information about the state of the system is obtained via an additional measured variable and the parameters of the controller are changed with this, in the present approaches the stiffness of the absorber. The approaches differ in the way the information from the measurement is processed.

• - Arbeitspunktabhängige Verstärkungsmethode (gain scheduling): Diese Methode stellt eine Erweiterung einer klassischen Regelung mit zusätzlich einfließenden Information dar. [3] und [4] beschreiben hierzu Konzepte, die sich auf eine Messung gleich mehrerer Signale (Fußpunkt und den jeweiligen Schwingkörper m1 oder m2 in [3], Phasendifferenz aus der Geschwindigkeit eines Servomotors und einem Lenkmoment in [4]) in und die Bestimmung eines entsprechenden Phasenversatzes (zwischen Erreger- und Antwortsignal) stützen. Aus dem entsprechenden Phasenversatz werden die einzustellenden Parameter (Resonanzeigenschaften eines Feder-Masse-Systems bis die Tilgung maximal ist, [3] sowie Regelparameter einer Regeleinheit, [4]) ermittelt. In [3] werden dabei die Resonanzeigenschaften eines Feder-Masse-Systems so eingestellt, dass die Wirkung der Tilgung maximal wird. Abweichend hierzu werden in [4] die gewonnenen Informationen aus dem Phasenunterschied herangezogen, um Schwingungen in einer Servolenkung zu reduzieren.
• - Störgrößenaufschaltungsmethode:
  • Bei einer Störgrößenaufschaltung wird zunächst die Störung am System erfasst und daraus die entsprechenden Stellgrößen berechnet. Störgrößenaufschaltungsmethoden ist die Realisierung einer Steuerung mit einem invertierenden Vorfilter. Ein bekanntes Beispiel hierzu stellen bekannte Techniken zur Rauschunterdrückung in Kopfhörern dar (active noise cancelling). Beispielsweise wird in [5] diese Technik zur Schwingungsminderung in Papiermaschinen vorgeschlagen. [6] offenbart beispielhaft eine Störgrößenaufschaltung, um die Körperschallschwingungen eines Körpers zu absorbieren. Dies wird durch ein Koppelelement innerhalb einer elektrischen Spule umgesetzt. Das Koppelelement in der Spule dient als Sensor und Aktor in einem und führt zu geringem Bauraumbedarf.

Based on the aforementioned prior art, the following approaches are known for adjusting the component properties of an absorber during operation:

• - Operating point-dependent amplification method (gain scheduling): This method represents an extension of a classic control with additional information. [3] and [4] describe concepts that relate to the measurement of several signals (base point and the respective oscillating body m1 or m2 in [3], phase difference between the speed of a servo motor and a steering torque in [4]) in and the determination of a corresponding phase offset (between the excitation and response signals). The parameters to be set (resonance properties of a spring-mass system until the cancellation is maximum, [3] and control parameters of a control unit, [4]) are determined from the corresponding phase offset. In [3] the resonance properties of a spring-mass system are adjusted so that the effect of the cancellation is maximized. In contrast to this, in [4] the information obtained from the phase difference is used to reduce vibrations in a power steering system.
• - Disturbance input method:
  • When a disturbance variable is applied, the disturbance in the system is first detected and then the corresponding manipulated variables are calculated. Disturbance input methods are the implementation of a control system with an inverting pre-filter. A well-known example of this is known techniques for noise suppression in headphones (active noise canceling). For example, this technique for reducing vibration in paper machines is proposed in [5]. [6] discloses an example of a disturbance variable feed-in to absorb the structure-borne sound vibrations of a body. This is implemented by a coupling element within an electrical coil. The coupling element in the coil serves as a sensor and actuator in one and requires little installation space.

Furthermore, [7] discloses investigations into the dynamics of a damper model with dry friction damping on a damper mass.

[8] also describes a dry lock-up damper and preloaded dry friction absorber masses. These systems are based on dry friction and exploit the stick-slip properties to reduce vibrations in a system. In a model view, the mechanical system to be damped is represented by a mass suspended relative to the environment. The absorber is attached to this system. A lock-up damper includes a dry friction point, a spring and a damper mass. The equation of motion of these systems, as well as numerical parameter studies and analytical solutions via averaging methods are presented. The damping mechanisms are then compared in transient simulations. The friction absorber consists of a damper mass pre-stressed between two wedges at the wedge angle. The preload force acting on the wedges is caused by a spring.

Proceeding from this, one object of the invention is to design a vibration damper of the type mentioned in such a way that it is suitable for reducing vibration amplitudes that occur with high energy efficiency by means of minimal dissipation.

Based on this, efforts are being made to propose a vibration damper that is characterized by frequency and amplitude-dependent control.

The tasks are solved with a vibration damper with the features of the first patent claim. Subclaims that relate to these represent advantageous embodiments.

To solve the problem, a vibration damper with frequency-dependent control on a mechanical system is proposed. The vibration damper includes a damper mass that is connected to the mechanical system via damping means. Preferably, a primary spring element is also inserted between the mechanical system and the damper mass, preferably connected in parallel to the damping means. The mechanical system is preferably and in the context of a model consideration formed by a mass to be calmed, which is suspended from the environment via a mechanical spring in an elastic manner and is therefore capable of oscillation.

However, an essential feature of the vibration damper is that the damping means include a friction clutch attached to the mechanical system with at least one friction surface acting on the damper mass, the force of the contact pressure between the at least one friction surface and a counter surface to this friction surface preferably being adjustable by the damper mass and/or can be regulated, preferably by means of adjusting means. Preferably, the contact force of the friction surface and a counter surface to this friction surface is generated by means of one (or more) secondary spring elements and can be adjusted and/or regulated by means of the adjusting means. The friction surface also preferably serves as a guide for the damper mass.

A preferred embodiment provides for the damper mass to be guided by guide elements for only one degree of freedom of movement. The guide element for the damper mass preferably comprises at least one of the friction surfaces, more preferably only the friction surfaces, more preferably at least three friction surfaces or at least two friction surfaces with guide grooves.

A particularly preferred embodiment provides a friction clutch with two mutually aligned friction surfaces that are rigidly connected to the mechanical system and are movable relative to one another, between which a damper mass is arranged. The damper mass can therefore only be moved parallel to the friction surfaces, whereby friction must be overcome, which then causes the damping. The secondary spring element elastically prestresses the friction surfaces against each other, with the damper mass being clamped between them with the aforementioned contact pressure. The adjusting means serve either to adjust the contact pressure, for example by varying the spring length and/or the spring travel.

In particular, embodiments are also proposed in which the topography and/or the angle of attack of the friction surface can be adjusted entirely or only in part, preferably in the edge regions, preferably in its angle of attack by the adjusting means. The counter surface to a friction surface, preferably on the side of the damper masses, can thus be faced with a local gradient or gradient on the friction surface, which has to be overcome during the damping of a lateral movement such as an oscillatory movement. An increase in particular in the angle of attack of the edge regions advantageously enables a local increase in the friction to be overcome due to an increase in gradient to be overcome in addition to the friction, on the one hand, and due to the increased spring force of the second spring element due to friction surfaces bent up from one another, on the other hand. When damping a vibration, an increase in the friction to be overcome with increasing vibration amplitude, which can be adjusted via the topography and the angle of attack of friction surface areas, takes place in an advantageous manner due to an incline in addition to the friction.

With the aforementioned friction clutch with adjustable contact pressure, not only a parameter adjustment of the friction damping that is dependent on the amplitude but also on the frequency of the vibration to be dampened can be realized, namely by means of an adjustability between two system dynamics, a static friction and a sliding friction, in the friction clutch as a function from the contact force. The contact force changes the frictional force while the mass remains unchanged, and with it the inertial forces that occur in the vibration damper and thus also the breakaway torque when there is a transition from static and sliding friction. The friction clutch can also be adjusted and regulated to the vibration process to be dampened. In an advantageous manner, not only can the oscillation amplitudes that occur in the system be reduced to the maximum, but at the same time they can also only be dissipated with the energy necessary for this. The combination of these two features is the basis for an effective, energy-efficient damper. A basic idea of the invention therefore also consists in a frequency-dependent control or regulation of the damping, in particular via an adjustable contact pressure on the friction surface, which shifts the resonances in the system through the stick-slide change.

What is therefore required is a device and a method with which damping acting on mechanical friction between two elements can be semi-actively regulated, the preload on the friction surfaces being able to be changed overall and also locally depending on an applied frequency or amplitude. This is preferably done via a closed control loop, comprising a sensor for detecting the frequency, a control or regulating unit in which the detected frequency is processed into at least one control signal, and an adjusting means, preferably an actuator or other means, for regulating the contact pressure and or other parameters for setting the friction state (stick-sliding change) of the friction clutch as a receiver of the at least one control signal.

In order to energetically optimize vibration damping or cancellation of a mechanical system, a vibration damper is proposed which, on the one hand, reduces larger vibration amplitude heights, but does not permanently dissipate energy from the system. Damping of vibrations outside the resonance frequencies and the associated significantly reduced vibration amplitudes is often not absolutely necessary. An energy-efficient damper therefore only dissipates energy when this is necessary to fulfill its function or protect the system.

In the context of a preferred embodiment, a secondary spring element is specifically proposed for generating the contact force on the at least one friction surface, which is further preferably adjusted in its preload and/or in its spring travel length by an adjusting means of the aforementioned type, more preferably with magnetic, piezoelectric or motor actuators implemented, adjustable. It is preferably proposed that the spring element and the adjusting means each act separately on the friction surfaces or the counter surfaces rubbing on them, i.e. the adjusting means do not directly adjust the spring element, for example in its initial position or in the spring travel.

One embodiment provides for the adjusting means to be arranged in the damper mass and designed in such a way that the surfaces of the damper mass facing the friction surfaces, i.e. the friction partners to the friction surfaces, are movable and can thus be positioned towards and/or away from each other on the friction surfaces .

The friction clutch preferably has at least two, preferably two, three or four arm elements arranged on the mechanical system on the mass to be calmed, which extend around the damper mass and are pressed elastically against one another by the secondary spring element while clamping the damper mass. Further preferably, adjusting means are connected to the arms and are suitable for moving the friction surfaces towards and/or apart and/or changing their topography in the manner mentioned above.

The friction surfaces on the arms also serve to guide the damper mass in the friction clutch and essentially determine its degrees of freedom of movement. For the damper mass and the mechanical system, these are preferably arranged aligned with one another or differ from one another in their orientation by a maximum of 30°. What is essential here are the vectorial components of the two degrees of freedom with the same orientation, which enable mutual influence and thus damping in this orientation.

In the context of a particularly preferred embodiment, the friction surfaces each have slope areas around a central area on one or both sides of the edge and preferably continuously increasing towards the edge. The middle area preferably extends over the potential rest positions of the damper mass without attacking forces, caused by the movement of the technical system to be damped, on the friction surface. The middle area is preferably flat, so that the damper mass does not change its position in this area without acting forces. The breakaway force that has to be overcome for a frictional movement between the friction and counter surface from static to sliding friction depends on the contact pressure between the friction and counter surface. The two slope areas are preferably arranged symmetrically to the central area and are adjustable and have the same structure, design and friction surface properties.

In a preferred embodiment, the edge-side slope areas are elastically flexible in their slope, which means that, in particular, shock-like relative movements between friction and counter surfaces can be absorbed more smoothly during the transition into the slope areas.

The slope regions are preferably connected to the central region via elastic bending elements and, together with this, form a continuous friction and thus surface, the transitions to the slope regions more preferably not being sharp-edged, but in a likewise elastically bendable region. The risk of tilting of the counter surfaces on the friction surfaces, especially when transitioning into the slope areas, is thus significantly reduced.

The slope areas must be overcome by the counter surface against the contact pressure within the friction clutch. The greater the gradient, the greater the counterforce on the damper element when climbing these gradient areas; the rising movement is oriented against the contact pressure. Consequently, the damping effect of the vibration damper is greater in the slope areas than in the middle area. The larger the vibration amplitude impressed by the mechanical system and to be dampened by the vibration damper, the more the movement of the counter surface extends into the edge slope areas. As the amplitude increases, the rising movement to be overcome and thus the damping also increase disproportionately.

The slope ranges are adjustable in their slope and thus the damping of the vibration damper from the moment the vibration amplitude is reached in the slope ranges by the counter surface, more preferably by means of motor or piezoceramic driven actuators, preferably self-locking actuators (worm drive, piezoceramic actuators).

Within the scope of a preferred embodiment, the two gradient areas can be adjusted synchronously in their angle and/or topography, with the gradient areas further preferably having a uniform adjustability of the gradient over their entire extent.

A preferred embodiment has measuring sensors for detecting a frequency (via an electrical measurement signal) on the mechanical system, preferably on or above the mass to be calmed, as well as a control or regulating unit for converting the detected frequency into at least one control signal for the adjustable friction clutch is. The damping means between the mechanical system and a damper mass with the sensor, the control and regulation unit and the friction clutch are part of a closed control circuit.

The measuring sensors preferably include piezoelectric or resistive vibration measuring sensors placed directly on the mass to be calmed or inductive, capacitive or optical measuring sensors arranged stationary and spaced from the calming mass, preferably position measuring systems, each for the quantitative detection of frequencies and amplitudes of a vibration of the mechanical system. Alternatively, measuring sensors are available on the elastic suspension of the mass to be calmed, the spring of the mechanical system to record the vibration state, e.g. using strain gauges applied to the spring.

The damping mass is preferably spherical or cylindrical or at least has counter surfaces oriented towards the friction surfaces with rounded edge areas. The risk of tilting of the mating surfaces on the friction surfaces, particularly in the aforementioned slope ranges, is thus significantly reduced.

The aforementioned vibration damper is characterized by the fact that variable absorption is not implemented in a wide frequency range, but only at a definable operating frequency. Frequencies that deviate from this, as well as smaller vibration amplitudes or components, are only dampened to a significantly smaller extent, which in turn leads to energy savings. The aim is to dampen certain frequencies and larger vibration amplitudes as selectively as possible. The interaction of static and sliding friction (deletion) between the friction surface and the counter surface ensures a relative movement in only one selectable frequency range. This means that the necessary energy is only dissipated in this area. In contrast, conventional vibration dampers with variable absorber frequencies permanently have a damped relative movement and thus also a permanent energy dissipation. Since dissipation cannot be completely suppressed, this approach has a negative impact on efficiency.

The basic principle of the invention is therefore based on a fundamental change in the dynamics of the vibration damper. By changing between static and sliding friction, the degrees of freedom in the system change. This change brings with it a change in the structural resonances in the system. As the contact force between the sliding surfaces and the countersurfaces varies, the friction properties and thus also the aforementioned structural resonance frequencies change. They can thus be circumvented in an advantageous manner.

The proposed disturbance variable application methods also differ in that they represent active methods for selective damping of predeterminable frequencies and amplitudes due to the changeability of contact force and pitch angle. They generate the exact counterforce required to compensate for vibrations.

• 1 ein mechanisches System mit einem konventionellen ungeregelten Schwingungsdämpfer (Vibrationstilger, Stand der Technik),
• 2 ein mechanisches System mit einem Schwingungsdämpfer gemäß 1, jedoch mit einem zusätzlichen trockenen Reibungselement,
• 3 beispielhaft einen Schwingungsdämpfer mit angewinkelten Steigungsbereichen als Reibfläche zur Dämpfungsmasse hin,
• 4a und b zwei weitere Ausführungsbeispiele des vorgeschlagenen Schwingungsdämpfers,
• 5 eine graphische Darstellung der Amplitudengänge über die Frequenz bei Haftreibung und Gleitreibung des in 4a dargestellten Schwingungsdämpfers, sowie deren ideale Systemantwort,
• 6 einen beispielhaften grundlegenden Aufbau und Verschalten einer Steuer- und Regeleinheit als Teil,
• 7 die Struktur des Frequenzidentifikationsmoduls aus 6 im Detail,
• 8 eine graphische Darstellung der Amplitudengänge über die Frequenz bei Haftreibung und Gleitreibung nach einer Sweep-Simulation durch ein frequenzbasiertes Regelungsverfahren,
• 9a bis d beispielhafte Ausgestaltungen der Belastungsvorrichtungen für eine einstellbare Anpresskraft (über Wegverschiebung Δl) der Reibflächen gegen die Dämpfermasse im Schwingungsdämpfers: Piezoaktorvarianten (a und b) Magnetvariante (c) Seilvariante (d),
• 10a und b prinzipielle Darstellungen von Ausgestaltungsbeispielen von Stellgliedern zur Verstellung des Anstellwinkels der randseitige Steigungsbereiche der Reibfläche - Schraubenvariante zur Winkeländerung. (a) Seil-Drehfeder-Variante zur Winkeländerung (b) - sowie
• 11a und b prinzipielle Darstellungen von Ausgestaltungsbeispielen für eine Reibfläche, umfassend mehrere Reibflächensegmente mit ebenen und/oder gewölbten randseitigen Steigungsbereiche über, auf oder in einer separaten tragenden Struktur.

The invention is explained in more detail using exemplary embodiments of the vibration damper, the following figures and descriptions. All features shown and their combinations are not only limited to these exemplary embodiments and their configurations. Rather, these should be viewed as combinable as representatives of other possible embodiments that are not explicitly shown as exemplary embodiments. Show it

• 1 a mechanical system with a conventional unregulated vibration damper (vibration absorber, state of the art),
• 2 a mechanical system with a vibration damper 1 , but with an additional dry friction element,
• 3 For example, a vibration damper with angled pitch areas as a friction surface towards the damping mass,
• 4a and b two further embodiments of the proposed vibration damper,
• 5 a graphical representation of the amplitude responses over the frequency for static friction and sliding friction of the in 4a vibration damper shown, as well as their ideal system response,
• 6 an exemplary basic structure and interconnection of a control and regulation unit as part,
• 7 the structure of the frequency identification module 6 in detail,
• 8th a graphical representation of the amplitude responses over frequency for static friction and sliding friction after a sweep simulation using a frequency-based control method,
• 9a until d exemplary configurations of the loading devices for an adjustable contact force (via displacement Δl) of the friction surfaces against the damper mass in the vibration damper: piezo actuator variants (a and b) magnet variant (c) rope variant (d),
• 10a and b basic representations of design examples of actuators for adjusting the angle of attack of the edge slope areas of the friction surface - screw variant for changing the angle. (a) Rope torsion spring variant for changing the angle (b) - as well
• 11a and b basic representations of design examples for a friction surface, comprising several friction surface segments with flat and/or curved edge slope areas above, on or in a separate supporting structure.

1 represents the basic structure of a conventional vibration damper, placed on a mechanical system. In the simplified model example shown, the mechanical system is represented by a mass 1 (mass m ₁ ) to be calmed by the vibration damper, which is elastically suspended from the environment, in the example by a mechanical spring 2 (Spring stiffness c ₁ ). The vibration damper itself comprises a damper mass 3 (mass m ₂ ), which is placed on the mass to be calmed via a parallel connection consisting of a primary spring element 4 (spring stiffness c ₂ ) and damping means 5 (damping constant d ₂ ). Conventional systems like these often have a constant damping constant, which dampens all vibrations with all amplitudes regardless of frequency.

gebracht, die durch eine harmonische Kraft mit der Erregeramplitude F und der Erregerkreisfrequenz Q modelliert wird. Die Dämpfungsmittel mit der Dämpfungsmasse greifen an der zu beruhigenden Masse direkt an, womit die Systemdynamik verbessert und die Auswirkungen der Störung minimiert werden sollen.The mechanical system is caused to vibrate by a disturbance $$\text{F}=\text{sin}\mathrm{\Omega }\text{t}$$

brought, which is modeled by a harmonic force with the exciter amplitude F and the exciter angular frequency Q. The damping means with the damping mass directly attack the mass to be calmed, which is intended to improve the system dynamics and minimize the effects of the disturbance.

2 shows that in 1 system shown, supplemented by an additional dry friction element 6, which is inserted parallel to the aforementioned damping means 5 (damping constant d ₂ ) and primary spring element 4 (spring stiffness c ₂ ) between the mass 1 to be calmed and the damper mass 3. The friction element supports the damping effect of the damping means and basically enables static and sliding friction.

3 shows the arrangement of a vibration damper on an aforementioned mechanical system (1, 2). Like conventional systems, the vibration damper here also has a damper mass 3. The damping means comprise a friction clutch 7 fastened on the mass to be calmed, with two friction surface areas acting on both sides of the damper mass 3 at an angle of attack α as edge slope areas 8 without a central area arranged in between. The damper mass has contact with all four gradient areas shown and is also fixed in its rest position by them. A primary spring element between the mass 1 to be calmed and the damper mass 3 is not provided and is not required. The friction surfaces are each arranged on an arm element 9 and clamp the damper mass, with the contact force being introduced via a secondary spring element 10 via the ends of the arm elements.

4a and b, on the other hand, show two exemplary embodiments in which the damper mass 3 is connected to the mass 1 to be calmed by means of a primary spring element 4 and is separated from it between two friction surfaces ( 4a) or on a friction surface ( 4b) is pressed with a contact force via a secondary spring element 10. When the primary spring element is relaxed, the damper mass is in a central position, preferably in a central region 11 of one of the friction surfaces 12 between the two edge slope regions 8 angled at an angle of attack α.

The damping element therefore comprises a damper mass 3, a primary spring element 4 and one or two friction surfaces 12 ( 4b or. 4a) , which are pressed onto counter surfaces of the damper mass by a secondary spring element. The friction contacts preferably consist of two segments angled at the angle of attack α (pitch areas 8) and a flat central area 11 of length 2Δ _l between them. The individual areas of the friction surface are preferably connected to one another by means of elastic joints, with a preferably kink-free and continuous transition being ensured.

Depending on the preload of the secondary spring element and thus the amount of contact force as well as the current state of motion, there is static or sliding friction between the friction surface and the counter surface on the damper mass. 5 shows a graphical representation of the amplitude characteristics for static friction and sliding friction of the in 4a Vibration damper shown, as well as their ideal system response. The oscillation amplitude of the main mass _A For very large preloads (contact forces, adjustable spring travel length Δl = F _Δl /c ₃ towards infinity) there is a sticking system dynamic 13 with a first resonance peak 15. Whereas a reduction in the preload (contact forces, Δl tends towards 0) results in a sliding system dynamic 14 with two (second) resonance peaks 16 sets. If Δl = 0 is chosen, the system also has an amortization frequency at Q = (c2/m2) ^1/2 . At the extinction frequency, the oscillation amplitudes of the mass 1 of the mechanical system to be calmed disappear almost completely. An ideal control of the contact force and thus the state of friction occurs when a control and regulation unit is used to switch back and forth between the sliding or sticking system dynamics in such a way that the system dynamics with the curve with the lowest vibration amplitudes are applied (ideal curve 17 with switching at excitation frequency Ω ₁ and Ω ₂ ). The frequency-based control method aims to reproduce this ideal curve.

The entire system structure of the control and regulation unit is in 6 presented and exists from three assemblies, the mechanical system 18, a frequency identification module 19 and detection of a frequency on the mechanical system as well as a control or regulating unit 20 for converting the detected frequency as measurement signals to at least one control signal 21 for actuators 22 engaging in the mechanical system. For influencing In the example of the mechanical system, two parameters can be varied, namely the angle of attack α and the preload length Δl (cf. 4a and b). These preferably serve as control signals and thus as input variables for the mechanical system.

The mechanical system also has measuring sensors for recording its vibration states and passes these on as raw data 23 to the frequency identification module. In this, the respective oscillation frequency Ω _id of the mechanical system is identified using a Fourier transformation. As a rule, this is equal to the angular frequency Ω, caused by an excitation with F sin (Ωt). This must be identified because in practice one does not know in advance which stimulus will affect the system. This identified frequency is then passed on to the control or regulating unit 20, which sets the aforementioned parameters α and Δl via the aforementioned actuators 22.

To determine the control signals for α and Δl in the control or regulating unit, the setpoints α _d and Δl _d must first be determined. These, in particular the upper interval limit for the setpoint α _d and the lower interval limit for the setpoint Δl _d , are selected depending on the identified frequency Ω _id of the mechanical system, where Ω _id is between Ω ₁ and Ω ₂ (cf. 5 ) lies. The controlled variables must then be set to these parameters.

• 1.Erfassung des Schwingungssignals X₂(t), 23 an der Dämpfungsmasse, wobei die letzten N Punkte 24 auf einer Zeitachse 25 zusammengefasst werden.
• 2. Fast Fourier Transformation (FFT) 26 von X₂(t).
• 3. Identifizierung 27 von möglichen identifizierten Frequenzen Ω_id, i aus den Frequenzen der FFT-Peaks.
• 4. Die identifizierten Frequenzen Ω_id, i mit dem geringsten Abstand zum gleitenden Mittelwert wird als die nächste identifizierte Frequenz Ω_id festgelegt.

7 shows an example of the structure of the frequency identification module 19 6 in detail as a block diagram. Based on the structure, the identification process can essentially be described in the following steps:

• 1.Detection of the vibration signal X ₂ (t), 23 at the damping mass, with the last N points 24 being summarized on a time axis 25.
• 2. Fast Fourier Transform (FFT) 26 of X ₂ (t).
• 3. Identification 27 of possible identified frequencies Ω _id , i from the frequencies of the FFT peaks.
• 4. The identified frequencies Ω _id , i with the closest distance to the moving average are set as the next identified frequency Ω _id .

A simulation of the described method with a so-called sweep excitation is presented in 8th shown. It is shown as in 4 a graphical representation of the amplitude responses over the frequency for static friction 28 and sliding friction 29 and an ideal curve 30 after a sweep simulation using the aforementioned frequency-based control method. By changing the sizes α and Δl, the sliding and static friction areas in the vibration damper can be adjusted within the ideal curve mentioned. α and Δl are changed when the frequencies Ω ₁ and Ω ₂ (constant parameters, switching points or switching frequencies between sliding and static friction areas) are exceeded. These are the switching points / switching frequencies between the frequencies Ω ₁ and Ω ₂ (cf. 4 ) there is a sliding friction, outside there is a static friction between the damper mass and the friction surfaces. In this way, even small oscillation amplitudes in the sliding friction area can be adjusted and at the same time large amplitudes of the adhesive system can be avoided.

The ideal curve calculated according to 4 shows a trouble-free transition in the transitions between the sliding and static friction areas, which in 8th has clear interference signals. Although the curve shown for the actively controlled system is not congruent with the ideal curve 30, particularly in the transitions between the sliding and static friction areas, the goals of the control are still achieved. Outside the interval range between Ω ₁ and Ω ₂ (cf. 5 ) there is static friction and therefore no sliding relative movement between the damper mass and the friction surface; Accordingly, no energy is dissipated there.

9a until d show exemplary embodiments of the contact pressure of the loading devices of the friction surfaces against the damper mass in the vibration damper, which can preferably be adjusted via the aforementioned parameters Δl.

9a shows a first embodiment with a piezo actuator 31, which is inserted in a two-part damper mass 3 and the distance between the two countersurfaces pressing on the central region 10 of the friction surface and thus the contact pressure is designed to be variable. The piezo actuator is the actuator for adjusting the contact force and thus Ω ₁ and Ω ₂ as well as the transitions between sliding and static friction. The design is characterized by a particularly high level of rigidity, which means that particularly high forces can be generated. The two-part one However, damper mass fundamentally complicates its vibration behavior.

9b shows a second embodiment with a pieoz actuator 31, which is arranged away from a preferably one-piece damper mass 3. The damper mass no longer needs to be divided. The piezo actuator is used to place additional load on the friction surfaces compared to the counter surfaces. It is proposed to design the piezo actuator as shown as a hollow cylinder with tie rods, so that when it expands, it causes the arm elements 9 to contract and thus increase the contact force on the damper mass 3.

Alternative configurations provide magnets for changing the preload. 9c represents such an alternative embodiment with an electromagnetic actuator 32 in series connection with a permanent magnet 33 instead of a piezo actuator according to 9b . To change the bias voltage, two (or more) magnets are preferably required to be used, with at least one being a controllable electromagnet. The contact force is modulated via the electrical voltage that can be applied to it. Further refinements provide for the position of the magnets to be swapped or the permanent magnet to be replaced by a second electromagnet.

Alternatively - as exemplified in 9d shown - regulate the contact pressure of the friction surfaces on the damper element 3 via the arm elements 9 with the help of cable elements 34 and several deflection rollers 35. The length and tension of the cable elements can be changed using a motorized cable winch 36 as an actuator. The use of several deflection rollers distributed over the arm elements ensures that force is applied evenly and counteracts tilting of the friction surfaces.

To adjust the angle of attack α of the slope areas of a friction surface, two in 10a and b suggested concepts. Shown are an arm element 9 with a friction surface 12 with a stationary inner region 11 and on both sides of it two edge-side slope regions 8, which are connected to the central region via elastic bending elements 37. Preferably, the surface areas of the elastic bending elements between the inner area and the edge-side slope areas are part of the friction surface, as shown.

One in 10a The first embodiment shown provides two counter-rotating screw drives 38, preferably driven via a common motor drive 39, for each of the pitch ranges for their angle of attack adjustment. The two screws are inserted into internal threads in the arm elements 9 and each engage under the edge slope areas via a ball joint 40. By rotating the screw, it is axially displaced and the angle of attack is adjusted. In the embodiment shown, the screw rotation takes place by means of an axial displacement of a rack 41, which engages in the screw head designed as a gear 42. The motor drive drives a spindle drive 43 that is firmly connected to the rack. This embodiment represents a positioning drive with precise guidance of the pitch areas in both travel directions, ie additional means for generating a counterforce that counteracts the adjustment force are not required. A particular advantage lies in the high positional rigidity of the edge slope areas, even in the face of higher contact pressures

One in 10b The second embodiment shown provides for an adjustment of the angle of attack a cable winch 36 for two cable elements 34, which engage via deflection rollers 35 on two ball joints 40 under the edge slope areas 8 and attack downwards. Since the rope elements are naturally slack, they are only suitable for transmitting a tensile force. A counterforce required for this is generated by prestressed elastic bending elements 37 and/or by inserted return torsion springs 44. The counterforce must be via the contact force generated by the secondary spring element (10, cf. 3 , 4a , 4b as well as 9a to d), dominate, since otherwise the edge slope areas will be pushed down undesirably towards the arm elements 9.

Preferably, all friction surfaces, including any edge slope areas of a vibration damper as well as the countersurfaces on the damper mass, are identical to one another in their geometric dimensions and are arranged around the damper mass. This advantageously results in a preferably uniaxial guidance of the damper mass in the vibration damper in only one degree of freedom of movement. Friction surfaces and countersurfaces are preferably arranged symmetrically about this degree of freedom of movement.

In a preferred embodiment, the damper mass itself is spherical, at least partially, preferably at least in the area of the counter surfaces. The counter surfaces are further preferably formed by the surfaces of the damper masses in a rotationally symmetrical manner about the degree of freedom of movement.

Further preferably, the edge slope areas and any bending elements of each friction surface are designed symmetrically on both sides around the inner area.

Preferably, an adjustment or adjustment of the angle of attack α of all edge-side slope areas, i.e. also all friction surfaces of a vibration damper, is carried out synchronously with one another, preferably with a uniform adjustability of the angle of attack α.

However, the invention basically also includes other configurations, in particular designs with group or individual adjustment and adjustability of the angle of attack α of the edge regions of the friction surface. A preferred group-wise setting and adjustability provides for a synchronous quantitative change in the angle of attack of the edge areas, which are passed over at the same time when the damper mass is displaced over the friction surfaces.

11a and b show two further examples that represent embodiments in which the friction surfaces of the vibration damper are placed on a supporting structure 45, preferably glued flat onto or into it. The elastic bending elements 37 are part of these supporting structures and not the friction surface, with the setting and adjustment of the angle of attack of the edge-side slope regions 8 via the supporting structures in these examples preferably, but not necessarily, by means of counter-rotating screw drives, as previously shown using 10a are described in detail. The edge slope areas are rigid elements and can be tilted as a whole. 11a shows edge slope areas which, apart from a transition area to the central area 11, are flat and whose setting and adjustment of the angle of attack can be tilted as a whole. 11b shows the same arrangement, in which, however, the edge-side slope regions 8 are formed by concave edge-side slope regions 46 without a uniform pitch angle which preferably increases towards the edge of the friction surface, which are preferably designed to be rigid and can be tilted as a whole.

In the 11a and b further represent embodiments in which the friction surface is formed by different friction surface segments, preferably with different material surfaces. Preferably, the coefficient of friction of the friction pairing between the central area to the counter surface is set to be significantly lower than that of the friction pairing between the edge slope areas to the counter surface. This, together with the angle of attack, advantageously causes significantly increased friction and thus damping at larger vibration amplitudes when the counter surface swings into the edge slope areas. There is a pronounced energy dissipation selectively only in the oscillation maxima, ie the upper regions of the oscillation curve, while with a significantly lower friction when swinging through the middle region there is significantly less friction and thus damping. Even small vibration amplitudes, in which the counter surface does not even reach the edge slope areas due to the oscillation curve, only experience lower friction and damping and thus only low energy dissipation.

The friction surface or parts of it must be in 11a and b shown embodiment are not necessarily fixed flat to the supporting structure. Rather, an optionally provided relative movement between the friction surface and the supporting structure opens up additional friction and thus leads to a higher energy dissipation in the vibration damper.

Literature:

• [1] Lee, CY et al.: Structural Vibration Control using a Tunable Hybrid Shape Memory Material Vibration Absorber; J. Intelligent Mat. Sys. and Structures, 23(15), 1725-1734 (2012).
• [2] Weber, F., Maslanka, M.: Frequency and Damping Adaptation of a TMD with controlled MR Damper; Smart Mat. and Struct., 21(5), 055011 (2012).
• [3] DE 198 32 697 C2
• [4] DE 10 2010 021 634 A1
• [5] DE 199 14 627 B4
• [6] DE 10 2016 223 864 A1
• [7] Aramendiz, J., Fidlin, A., & Baranowski, E. On the Dynamics of a Prestressed Sliding Wedge Damper. PAMM, 19(1), e201900326 (2019).
• [8th] Tan, AS, Aramendiz, J., Ross, KH, Sattel, T., & Fidlin, A. Comparative study between dry friction and electrorheological fluid switches for tuned vibration absorbers. J. Sound and Vibration, 460, 114874 (2019).

Reference symbol list

1

to calm mass, m ₁

2

mechanical spring, c ₁

3

Damper mass, m ₂

4

primary spring element, c ₂

5

dampening agent

6

dry friction element

7

Friction clutch

8th

edge slope area

9

Arm element

10

secondary spring element, c ₃

11

middle area

12

friction surface

13

adherent system dynamics

14

sliding system dynamics

15

first resonance peak (with persistent system dynamics)

16

second resonance peaks (with sliding system dynamics)

17

Ideal curve

18

mechanical system

19

Frequency identification module

20

Control or regulating unit

21

Control signal

22

actuator

23

Vibration signal X ₂ (t)

24

N points

25

Timeline

26

Fast Fourier Transform (FFT)

27

identification

28

Amplitude response over frequency with static friction

29

Amplitude response over frequency with sliding friction

30

Ideal curve (with transitional disturbances between sliding and static friction)

31

Piezo actuator

32

electromagnetic actuator

33

Permanent magnet

34

Rope element

35

pulley

36

Winch

37

bending element

38

Screw drive

39

motor drive

40

Ball joint

41

Rack

42

gear

43

Spindle drive

44

Return torsion spring

45

supporting structure

46

concave edge slope area

QUOTES INCLUDED IN THE DESCRIPTION

This list of documents listed by the applicant was generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Cited patent literature

• DE 19832697 C2 [0067]
• DE 102010021634 A1 [0067]
• DE 19914627 B4 [0067]
• DE 102016223864 A1 [0067]

Non-patent literature cited

• Lee, C.Y. et al.: Structural Vibration Control using a Tunable Hybrid Shape Memory Material Vibration Absorber; J. Intelligent Mat. Sys. and Structures, 23(15), 1725-1734 [0067]
• Weber, F., Maslanka, M.: Frequency and Damping Adaptation of a TMD with controlled MR Damper; Smart Mat. and Struct., 21(5), 055011 [0067]
• Aramendiz, J., Fidlin, A., & Baranowski, E. On the Dynamics of a Prestressed Sliding Wedge Damper. PAMM, 19(1), e201900326 [0067]
• Tan, A. S., Aramendiz, J., Ross, K. H., Sattel, T., & Fidlin, A. Comparative study between dry friction and electrorheological fluid switches for tuned vibration absorbers. J. Sound and Vibration, 460, 114874 [0067]

Claims (19)

Vibration damper with frequency-dependent control on a mechanical system (1, 2), comprising a) Damping means (5) between the mechanical system (1, 2) and a damper mass (3), where b) the damping means comprise a friction clutch fastened on the mechanical system with at least one friction surface (12) acting on the damper mass, the force of the contact pressure between the at least one friction surface and the damper mass being adjustable and/or controllable. Vibration damper Claim 1 , characterized in that a primary spring element (4) is provided between the mechanical system (1, 2) and the damper mass (3). Vibration damper Claim 1 or 2 , characterized in that a secondary spring element (10) is provided to generate the contact pressure. Vibration damper Claim 3 , characterized in that the secondary spring element (10) is adjustable in its preload and/or in its spring travel length, the preload and/or the spring travel length of the secondary spring element being adjustable by adjusting means (31-36, 38-44). Vibration damper Claim 4 , characterized in that the adjusting means are arranged in the damper mass (3) and are suitable for moving the surfaces of the damper mass facing the friction surfaces (12) towards and/or from one another. Vibration damper according to one of the Claims 3 until 5 , characterized in that the friction clutch has at least two arm elements (9) arranged on the mechanical system (1, 2), which are arranged around the damper mass (3) and are pressed elastically against each other by the secondary spring element (10) while clamping the damper mass . Vibration damper Claim 6 , characterized in that the adjusting means (31-36, 38-44) are connected to the arm elements (9) and are suitable for moving the friction surfaces (12) towards and/or from one another. Vibration damper according to one of the Claims 4 until 7 , characterized in that the adjusting means (31-36, 38-44) comprise magnetic, piezoelectric or motor actuators. Vibration damper according to one of the preceding claims, characterized in that the damper mass (3) is guided by guide elements for only one degree of freedom of movement. Vibration damper Claim 9 , characterized in that the guide element for the damper mass (3) comprises at least one friction surface (12). Vibration damper according to one of the preceding claims, characterized in that the degrees of freedom of movement for the damper mass (3) and the mechanical system (1, 2) are arranged in alignment with one another or differ from one another in their orientation by a maximum of 30°. Vibration damper according to one of the preceding claims, characterized in that the at least one friction surface (12) has edge slope regions (8) on one or both sides around a central region (11) around a position of the damper mass (3) when the mechanical spring (2) is relaxed. Vibration damper Claim 12 , characterized in that the edge slope areas (8) are elastically flexible in their slope. Vibration damper Claim 12 or 13 , characterized in that the edge slope areas (8) are connected to the central area (11) via elastic bending elements (37) and together form a continuous surface. Vibration damper according to one of the Claims 12 until 14 , characterized in that the slope areas (8) on the edge are adjustable in their slope. Vibration damper Claim 15 , characterized in that the incline is motor-adjustable. Vibration damper according to one of the preceding claims, characterized in that a measuring sensor is provided for detecting a frequency on the mechanical system (1, 2) and a control or regulating unit for converting the detected frequency into at least one control signal for the adjustable friction clutch. Vibration damper Claim 17 , characterized in that the measuring sensor comprises piezoelectric, piezoresistive measuring sensors, placed on, on or above a mass of the mechanical system to be calmed, and / or inductive, capacitive or optical measuring sensors spaced from the mass to be calmed. Vibration damper Claim 17 or 18 , characterized in that the damping means (5) between the mechanical system (1, 2) and a damper mass (3) with the sensor (19), the control and regulation unit (20) and the friction clutch are part of a closed control circuit.

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