An adjustable nonlinear dynamic vibration absorber and system

Abstract

The application discloses an adjustable nonlinear dynamic vibration absorber and system, and has the technical scheme that a tuning mass block is slidably connected to the surface of a support plate, and first scissor structures are rotatably connected between the tuning mass block and the two ends of the support plate; the first scissor structures are provided with second scissor structures and dampers, the second scissor structures and the dampers can drive the first scissor structures to stretch and contract through elastic belts, the second scissor structures are provided with elastic elements for driving the second scissor structures to stretch and contract; the stretching direction of the elastic elements is parallel to the sliding direction of the tuning mass block, and is perpendicular to the damping direction of the dampers, so that the first scissor structures and the second scissor structures can realize vibration absorption through geometric nonlinearity when moving. The stiffness and damping characteristics of the application can be adjusted, and the linear elements can be used to realize nonlinear stiffness, damping and super-low natural frequency, and the problem of Duffing system instability that cannot be eliminated by traditional nonlinear vibration absorbers is solved.

Description

TECHNICAL FIELD

[0001] The present application relates to the technical field of vibration suppression / absorption, and in particular to an adjustable nonlinear dynamic absorber and system. BACKGROUND

[0002] In many engineering practices, vibration control is a key task. Passive vibration control provides a more favorable solution for structural vibration control than active control methods, which is energy-saving, reliable, stable and economical. In some engineering practices, a dynamic vibration absorber (DVA) or a tuned mass damper (TMD) is an effective passive device that can suppress the vibration amplitude of a main structure / system (protected object) by transferring and dissipating the energy of the main structure. The TMD is an additional mass block connected to the main vibration system through a spring and a damper. The mathematical optimization method of the traditional linear tuned mass damper (LTMD) is the fixed point theory of the linear TMD system, and a closed-form solution is obtained. This fixed point theory lays the foundation for theoretical analysis of the TMD system and is widely used later.

[0003] Although the LTMD system is widely used for vibration suppression due to its simplicity, some shortcomings cannot be ignored. One of them is the excessively narrow suppression bandwidth in the resonance frequency range; the vibration suppression bandwidth can be doubled by using a softened nonlinear belleville spring. The optimization objectives can be divided into minimizing the maximum response amplitude of the main system resonance (i.e., optimization) or minimizing the total vibration energy in the full frequency range (i.e., optimization). For the LTMD system with a non-damped main structure, an exact analytical solution has been obtained. However, when the main system has damping, it is difficult to obtain an exact solution theoretically. When the main system contains damping elements, only numerical optimization solutions are obtained. For the nonlinear tuned mass damper (NTMD) system, it is more difficult to perform theoretical optimization analysis due to the nonlinear and coupled dynamics equations of the NTMD. The objective function of the optimization cannot be expressed in an explicit function before any simplification or linearization process is performed.

[0004] Despite the difficulties, the optimization analysis of NTMD has attracted much attention in recent years. The reason is that in engineering practice, most TMDs or main systems have inherent nonlinear elements (nonlinear stiffness / nonlinear damping). NTMD systems show obvious advantages in many cases. For example, as mentioned earlier, the softening nonlinear stiffness can double the vibration suppression bandwidth of the main structure in the frequency response curve. The jump and hysteresis phenomenon caused by the bifurcation of the saddle node may occur in the steady-state response of the nonlinear main vibration structure, and the LTMD cannot suppress these unstable bifurcations of the main structure. Compared with the LTMD, the optimal design of the NTMD can help suppress the large amplitude response and instability problems of the nonlinear main structure. For example, the vibration reduction performance, robustness and sensitivity of the NTMD coupled with the Duffing system are superior to the linear system. For example, when the excitation amplitude is large, strong nonlinear problems of the main system may be excited, such as bifurcation, instability and separated resonance curves. However, the current nonlinear absorber generally cannot eliminate the instability problem of the Duffing system. SUMMARY

[0005] In view of the deficiencies of the prior art, the purpose of the present application is to provide an adjustable nonlinear dynamic absorber and system, which can adjust the stiffness and damping characteristics, and can realize nonlinear stiffness, damping and ultra-low natural frequency using linear elements, thereby solving the instability problem of the Duffing system that cannot be eliminated by the traditional nonlinear absorber.

[0006] In order to achieve the above-mentioned purpose, the present application is realized by the following technical solutions:

[0007] In a first aspect, the embodiments of the present application provide an adjustable nonlinear dynamic absorber, comprising a tuning mass block slidingly connected to the surface of a support plate, a first scissor structure rotatably connected between the tuning mass block and the two ends of the support plate; a second scissor structure and a damper are installed on the first scissor structure, the second scissor structure and the damper can drive the first scissor structure to extend and retract through the extension and retraction of the extension belt, and the second scissor structure is matched with an elastic element for driving the extension and retraction of the second scissor structure.

[0008] The extension and retraction direction of the elastic element is parallel to the sliding direction of the tuning mass block, and is perpendicular to the damping direction of the damper, so that the first scissor structure and the second scissor structure move through geometric nonlinear action to realize vibration absorption.

[0009] As a further implementation manner, the first scissor structure comprises at least one first X-shaped unit, the first X-shaped unit is installed with the second scissor structure on one side along the sliding direction of the tuning mass block, and the damper is installed on the other side.

[0010] As a further implementation manner, the second scissor structure comprises a plurality of second X-shaped units connected in rotation, and the elastic unit is symmetrically installed along two sides perpendicular to the sliding direction of the tuned mass damper.

[0011] As a further implementation manner, the second scissor structure is connected to one end of the first scissor structure close to the end of the support plate, and the damper is connected to one end of the first scissor structure close to the tuned mass damper.

[0012] As a further implementation manner, the first scissor structure is connected to the tuned mass damper and / or the support plate through a rotating shaft and a bearing installed outside the rotating shaft.

[0013] Alternatively, the first scissor structure is connected to the tuned mass damper and / or the support plate through a hinge.

[0014] As a further implementation manner, the second scissor structure is connected to the first scissor structure through a rotating shaft and a bearing installed outside the rotating shaft.

[0015] Alternatively, the second scissor structure is connected to the first scissor structure through a hinge.

[0016] As a further implementation manner, the support plate is provided with a guide rail, and the tuned mass damper is slidably connected to the guide rail.

[0017] As a further implementation manner, the support plate is in a U-shaped structure.

[0018] In a second aspect, the embodiments of the present application also provide an adjustable nonlinear dynamic vibration absorber system, comprising the vibration absorber.

[0019] As a further implementation manner, the vibration absorber is detachably connected to a main structure.

[0020] The beneficial effects of the present application are as follows:

[0021] (1) One or more embodiments of the present application can obtain an advantageous nonlinear damping and adjustable quasi-zero stiffness by providing a first scissor structure and a second scissor structure with X-shaped units, which can significantly improve the robustness of system parameters; the interaction between the first scissor structure and the second scissor structure not only realizes the nonlinear stiffness and damping effect, but also further amplifies the nonlinear stiffness and damping effect through mutual coupling, thereby realizing a stronger vibration absorption effect.

[0022] (2) The X-tuned mass damper of one or more embodiments of the present application can significantly widen the vibration suppression bandwidth and have lower resonance peaks by increasing the nonlinear LLS stiffness; this feature is beneficial for vibration suppression in a specific frequency range, such as at resonance; the X-tuned mass damper can successfully eliminate the instability problem of the Duffing primary structure, such as bifurcation and split resonance curves, which cannot be eliminated by traditional nonlinear tuned mass dampers.

[0023] (3) The scissors structure of one or more embodiments of the present application, in combination with dampers and elastic elements, can achieve advantageous nonlinear stiffness and damping, adjustable and optimally designed stiffness ratio and damping ratio to achieve optimal vibration absorption effect. BRIEF DESCRIPTION OF DRAWINGS

[0024] The accompanying drawings, which form a part of the present application, are used to provide further understanding of the present application, and the illustrative embodiments of the present application and their description serve the purpose of explaining the present application. They are not intended to limit the present application unduly.

[0025] Figure 1 is a perspective view of the present application according to one or more embodiments;

[0026] Figure 2 is a top view of the present application according to one or more embodiments;

[0027] Figure 3 is a front view of the present application according to one or more embodiments;

[0028] Figure 4 is a first scissors structure schematic view of the present application according to one or more embodiments;

[0029] Figure 5 is a second scissors structure schematic view of the present application according to one or more embodiments;

[0030] Figure 6(a) is a -50% damping coefficient sensitivity analysis diagram of the present application according to one or more embodiments;

[0031] Figure 6(b) is a -87.5% damping coefficient sensitivity analysis diagram of the present application according to one or more embodiments;

[0032] Figure 6(c) is a +25% damping coefficient sensitivity analysis diagram of the present application according to one or more embodiments;

[0033] Figure 7 is a vibration suppression bandwidth comparison diagram of the X-tuned mass damper (NTMD) and the linear tuned mass damper (LTMD);

[0034] Figure 8 is a vibration suppression bandwidth comparison diagram of the traditional nonlinear tuned mass damper and the X-tuned mass damper;

[0035] Figure 9 This is a comparison diagram of two nonlinear vibration absorbers acting on the Dufen main system;

[0036] Figure 10(a) is a time history curve of the X-type vibration absorber under random excitation on the main system;

[0037] Figure 10(b) shows the amplitude-frequency response of the X-type vibration absorber under random excitation;

[0038] Figure 11 This is a comparison chart of the amplitude-frequency curves with and without the X-type vibration absorber;

[0039] Figures 12(a)-12(d) These are time history response curves at different frequencies;

[0040] Figure 13(a) shows the impact excitation diagram of the X-type vibration absorber under optimized parameters;

[0041] Figure 13(b) shows the amplitude-frequency response of the X-type vibration absorber under optimized parameters;

[0042] Figure 14(a) is the impact excitation diagram of the spring-mass vibration absorber under optimized parameters;

[0043] Figure 14(b) shows the amplitude-frequency response of the spring-mass vibration absorber under optimized parameters;

[0044] Figure 15(a) is the impact excitation diagram of the main system without a vibration damper;

[0045] Figure 15(b) is the amplitude-frequency response diagram of the main system without a vibration damper;

[0046] Among them, 1. First scissor-type structure, 11. First X-shaped unit, 12. First connecting rod, 13. First rotating shaft, 2. Tuning mass block, 3. Elastic element, 31. First elastic element, 32. Second elastic element, 33. Third elastic element, 4. Damper, 41. First damper, 42. Second damper, 5. Second scissor-type structure, 51. Second connecting rod, 52. Second X-shaped unit, 53. Second rotating shaft, 6. Support plate, 7. Guide rail. Detailed Implementation

[0047] Example 1:

[0048] This embodiment provides an adjustable nonlinear dynamic vibration absorber, such as... Figures 1-3 As shown, it includes a first scissor structure 1, a tuning mass block 2, and a second scissor structure 5. One side of the tuning mass block 2 is slidably connected to the support plate 6, and the other side surface is symmetrically connected to two first scissor structures 1 along the sliding direction perpendicular to the tuning mass block 2. One end of the first scissor structure 1 is rotatably connected to the tuning mass block 2, and the other end is rotatably connected to the support plate 6.

[0049] Wherein, the first scissor structure 1 installs the second scissor structure 5 and the damper 4, the second scissor structure 5 is installed with the elastic element 3, through the coupling of the damper 4, the elastic element 3 and the first scissor structure 1, the second scissor structure 5, using the geometric nonlinearity between different directions of the scissor structure when moving, the advantageous nonlinear stiffness and damping, adjustable optimal stiffness ratio and damping ratio can be realized.

[0050] The mass of the tuned mass damper 2 can be adjusted, and the mass ratio is designed according to the main structure and the vibration absorption demand, so as to realize the adjustable and optimized vibration absorption performance.

[0051] In the embodiment, the support plate 6 is provided in a U-shaped structure; specifically, the support plate 6 includes a support part and a connecting part perpendicular to both ends of the support part; the connecting part is used for connecting the first scissor structure 1. The surface of the support part is provided with a guide rail 7 in a direction perpendicular to the connecting part, and the tuned mass damper 2 is slidably connected with the guide rail 7.

[0052] Further, the first scissor structure 1 is arranged on the side of the tuned mass damper 2 away from the guide rail 7 and is installed in a direction parallel to the guide rail 7; one end of the first scissor structure 1 is rotatably connected to the central position of the surface of the tuned mass damper 2, and the other end is rotatably connected to the middle position of the connecting part. When the tuned mass damper 2 moves along the guide rail 7, the first scissor structure 1 on one side of the tuned mass damper 2 is stretched, and the first scissor structure 1 on the other side is compressed.

[0053] Further, the first scissor structure 1 includes at least one first X-shaped unit 11, and the embodiment takes the first scissor structure 1 including one first X-shaped unit 11 as an example. As shown in FIGS. 1 and 2, the first X-shaped unit 11 is connected with the first connecting rod 12 on both sides in a direction parallel to the guide rail 7, wherein one side of the first X-shaped unit 11 is connected to the middle position of the connecting part of the support plate 6 through two first connecting rods 12, and the other side is connected to the central position of the tuned mass damper 2 through two first connecting rods 12, forming a scissor structure. Figure 2 Figure 4 Further, the first scissor structure 1 includes at least one first X-shaped unit 11, and the embodiment takes the first scissor structure 1 including one first X-shaped unit 11 as an example. As shown in FIGS. 1 and 2, the first X-shaped unit 11 is connected with the first connecting rod 12 on both sides in a direction parallel to the guide rail 7, wherein one side of the first X-shaped unit 11 is connected to the middle position of the connecting part of the support plate 6 through two first connecting rods 12, and the other side is connected to the central position of the tuned mass damper 2 through two first connecting rods 12, forming a scissor structure.

[0054] In the embodiment, one end of the first connecting rod 12 is rotatably connected with the first X-shaped unit 11, and the other end is rotatably connected with another first connecting rod 12 and the support plate 6 or the tuned mass damper 2. The rotatable connection can be achieved by a hinge or a first rotating shaft 13 provided with a bearing.

[0055] Further, one end of the first X-shaped unit 11 close to the connecting part in the direction of the guide rail 7 is connected with the second scissor structure 5, and the other end close to the tuned mass damper 2 is connected with the damper 4; and the extension direction of the damper 4 is consistent with the extension direction of the second scissor structure 5. Through the cooperation of the damper 4 and the first scissor structure 1, nonlinear damping can be realized, and the robustness of the system is improved.​

[0056] In the embodiment, two dampers 4 are provided, i.e., a first damper 41 and a second damper 42, and the first damper 41 and the second damper 42 are symmetrically arranged with respect to the center of the tuned mass 2.

[0057] It can be understood that, in other embodiments, the first scissors structure 1 includes a plurality of first X-shaped units 11 connected in sequence, the first X-shaped unit 11 at one end is connected with the support plate 6 through the first connecting rod 12, and the first X-shaped unit 11 at the other end is connected with the tuned mass 2.

[0058] In the embodiment, the second scissors structure 5 is installed on the inner side (the side close to the surface of the tuned mass 2) of the first scissors structure 1, and the damper 4 is installed on the outer side (the side away from the surface of the tuned mass 2) of the first scissors structure 1.

[0059] Further, the second scissors structure 5 includes a plurality of second X-shaped units 52 connected in sequence, and the second X-shaped units 52 are symmetrically installed with the elastic unit 3 on both sides perpendicular to the direction of the guide rail 7; wherein the extension direction of the elastic unit 3 is consistent with the direction of the guide rail 7. In the embodiment, the elastic unit 3 can be a spring. The elastic unit 3 is used in cooperation with the second scissors structure 5 to achieve nonlinear stiffness; further adjustment of the structural parameters can widen the vibration absorption frequency band and achieve better vibration absorption effect.

[0060] The number of the second X-shaped units 52 is set according to the actual vibration reduction requirement. In the embodiment, two second X-shaped units 52 are taken as an example, as shown in FIG. 5, three elastic units 3 are provided, i.e., a first elastic element 31, a second elastic element 32, and a third elastic element 33. Figure 5

[0061] The second X-shaped unit 52 is rotationally connected with the first scissors structure 1 through the second connecting rod 51, and the specific connection mode can be hinge connection or connection through the second rotating shaft 53 provided with a bearing. The interaction between the second scissors structure 5 and the first scissors structure 1 can not only achieve nonlinear stiffness and damping effect, but also further amplify the nonlinear stiffness and damping effect by using the mutual coupling effect, thereby achieving stronger vibration absorption effect.

[0062] The first scissors structure 1 and the second scissors structure 5 both include X-shaped units, which can provide favorable geometric nonlinearity. Compared with the traditional vibration absorber, the embodiment can significantly improve the robustness of the system parameters, expand the vibration suppression bandwidth, achieve ultra-low frequency vibration absorption / vibration suppression, and at the same time can eliminate the potential instability of the main system.

[0063] The assembly method of the embodiment is as follows:

[0064] ​S1: First, design the size of the X vibration absorber according to actual needs, select the number n of first X units in the first scissor structure 1 and the length L of the first connecting rod 12 according to the required size.

[0065] S2: Design the approximate size of the second scissor structure 5 according to the size of the first scissor structure 1.

[0066] S3: Optimize the size of the two scissor structures according to the vibration absorption requirements, mass ratio and other conditions.

[0067] In this embodiment, the optimization is completed by directly optimizing the structure parameters using the response function solved by the harmonic balance method (HBM).

[0068] S4: According to the mass ratio requirement, select the size and mass of the tuning mass block 2, and install the tuning mass block 2 on the guide rail 7 of the support plate 6.

[0069] S5: Install the second scissor structure 5 on both sides of the tuning mass block 2 using bearings or hinges, and install the elastic element 3 with selected stiffness in the second scissor structure 5.

[0070] S6: Install the first scissor structure 1 outside the support plate 6, and install the damper 4 selected in the connecting rod in the first scissor structure 1.

[0071] S7: Install the assembled X vibration absorber on the main structure (target object requiring vibration absorption / damping).

[0072] This embodiment is inspired by the geometry and motion of the leg bones of birds, and a bionic bird leg bone structure (LLS) or X-shaped structure is widely used for vibration protection; the X-shaped structure can provide flexible adjustable nonlinear stiffness and damping, and this adjustability can be used to design the optimal parameters of the NTMD.

[0073] This embodiment introduces a vibration absorber (X vibration absorber) with a bionic X-shaped structure, which has adjustable stiffness and damping characteristics, and can achieve nonlinear stiffness, damping and ultra-low natural frequency using linear elements. Compared with traditional linear vibration absorbers, it has good robustness. In addition, compared with traditional nonlinear stiffness vibration absorbers, the X vibration absorber has more comprehensive damping performance and can solve the problem of Duffing system instability that cannot be eliminated by traditional nonlinear vibration absorbers.

[0074] Embodiment two:

[0075] This embodiment provides an adjustable nonlinear dynamic vibration absorption system, which comprises the vibration absorber of embodiment one, and further comprises a main structure, wherein the vibration absorber and the main structure are detachably connected.

[0076] This example utilizes the tunable parameter nonlinear TMD (X absorber) designed by bionic X structure, and further carries out the multivariable parameter optimization analysis for the X absorber, the analysis of tunable stiffness and damping characteristics, the research on nonlinear influence and vibration reduction performance.

[0077] (1) Optimization objective H ∞ optimization and sensitivity analysis:

[0078] For the vibration absorption system of this example, the optimization objective can be the minimization of the motion of the main structure mass, for example, the minimization of the maximum displacement, velocity and acceleration in the full frequency range, and the optimization objective can also be the maximization of the system energy consumption. In addition to energy consumption, the optimization objective can also be to expand the vibration suppression bandwidth in the resonance frequency range.

[0079] In this example, the optimization objective of the X absorber is to minimize the maximum amplitude / displacement of the main structure in the most dangerous resonance frequency range. This is a typical H ∞ optimization. The challenge of H ∞ optimization of the X absorber system is how to optimize the structural parameters and ultimately affect the system response. Through Taylor expansion, the relationship between the structural parameters and the damping coefficients is an explicit polynomial function. Therefore, the optimization can be completed by directly optimizing the structural parameters using the response function solved by the harmonic balance method (HBM). The logical program of the optimization analysis in this example is:

[0080] 1) Description of the optimization objective.

[0081] 2) Multi-parameter optimization analysis using HBM and multi-layer nested loop programming in MAPLE.

[0082] 3) Optimal result analysis and comparison under different parameter conditions.

[0083] 4) Comparison with linear damping absorber under system misadjustment.

[0084] Firstly, the optimization objective function is defined as minimizing the maximum amplitude of the main structure mass. This is a minimization of maximum problem, which can be expressed as:

[0085]

[0086] where G = |x1 / A0| is the dimensionless amplitude ratio, which can also be called the amplitude amplification factor, A0 = f0 / K1 is the input amplitude, α i is each damping, and Ω is the dimensionless external excitation frequency, which ranges from [Ω1, Ω2].

[0087] The optimization objective including the structural parameters can be expressed as:

[0088]

[0089] The linear and nonlinear damping tuned mass damper system mass. The results show that when the damping coefficient of the TMD changes, the performance of the X absorber in the case of misadjustment is better than that of the LTMD. From Figures 6(a)-6(b) The results show that when the damping coefficient decreases, the X absorber with nonlinear damping maintains a relatively low amplitude in the resonance region; in Fig. 6(c), when the damping coefficient increases slightly, both systems exhibit over-damping, and both peaks disappear. In summary, when the system damping coefficient decreases, the X absorber with LLS damping can maintain a relatively low amplitude for both the main structure and the tuned mass damper compared to the LTMD. When the system damping coefficient increases slightly, both the X absorber and the LTMD show over-damping.

[0090] (2) Absorber band analysis:

[0091] As shown in Figure 7 , when the damping is small (linear damping ratio ξ L = 0.005, the nonlinear damping ratio is reduced to the same proportion), the X absorber with nonlinear stiffness can at least triple the vibration suppression bandwidth. Compared with the linear system, softening the stiffness can double the suppression bandwidth. This phenomenon shows that the X absorber has the potential to expand the vibration suppression bandwidth. In addition, under the same damping coefficient, nonlinear damping can help suppress amplitude peaks. Therefore, the X absorber with both nonlinear stiffness and damping has better performance in expanding the vibration suppression bandwidth.

[0092] As shown in Figure 8 , increasing the nonlinear stiffness (cubic stiffness) ratio of the traditional nonlinear absorber α2 to 100, it is obvious that the vibration suppression bandwidth of the system can be increased by increasing α2. However, the disadvantage is obvious: high cubic stiffness (α2 = 100) can cause bifurcation and instability problems. In contrast, by adjusting the structural parameters of X (increasing nonlinear stiffness), the X absorber can effectively increase the vibration suppression bandwidth. More importantly, unlike the traditional cubic stiffness absorber, the vibration system with the X absorber is more stable, and with the increase of nonlinear stiffness, it can suppress bifurcation and other instability phenomena, which is desirable in practice.

[0093] (3) Analysis of the vibration reduction effect of the absorber on the Duffing system:

[0094] When the main structure contains nonlinear stiffness, such as a typical Duffing system, the comparison results of the X absorber and the traditional nonlinear absorber are as follows Figure 9The X-vibration absorber and the conventional nonlinear vibration absorber are both effective when the input amplitude A0= 0.1. When the input amplitude increases to A0= 0.13, a separate resonance curve appears in the conventional nonlinear vibration absorber system, which means that the strong nonlinear phenomenon such as instability and bifurcation problem of the system is triggered when the amplitude is large. In contrast, the X-vibration absorber can effectively suppress these instability phenomena and maintain the effectiveness of vibration suppression in the case of such a large vibration amplitude (A0= 0.13).

[0095] The X-vibration absorber of the embodiment can provide advantageous nonlinear damping and adjustable quasi-zero stiffness, which can significantly improve the robustness of system parameters; as Figure 7 and Figure 8 As shown in FIG. 6, compared with the linear vibration absorber and the conventional nonlinear vibration absorber, the X-vibration absorber can expand the width of the vibration suppression frequency band by adjusting the structural parameters while maintaining the stability of the main structure; as Figure 9 As shown in FIG. 7, when coupled with a nonlinear structure (for example, a Duffing system), the X-vibration absorber can successfully suppress the instability situation (for example, bifurcation, resonance curve separation) due to the triggered strong nonlinearity, thereby showing stable and reliable vibration reduction performance, and therefore, the X-vibration absorber can more effectively improve the overall vibration suppression performance.

[0096] The experimental results of the embodiment are as follows:

[0097] Under random excitation, the vibration absorption effect of the X-vibration absorber tested in the experiment is shown in FIG. 10(a) and FIG. 10(b), wherein FIG. 10(a) is a time history signal of random excitation, and 10(b) is an amplitude-frequency curve of the main structure and the X-vibration absorber. Obviously, the vibration energy is transferred from the main system to the X-vibration absorber, so that the main system has a vibration suppression effect in the resonance frequency, and the X-vibration absorber has an obvious energy absorption peak.

[0098] The comparison of the vibration responses of the main system with and without the X-vibration absorber is shown in FIG. 11, and obviously, compared with the case without the X-vibration absorber, the X-vibration absorber can effectively suppress the resonance peak of the main system mass. The effectiveness of the X-vibration absorber is verified. Figure 11

[0099] The vibration absorption effect of the X-vibration absorber at different frequencies is shown in FIG. 12; at this time, the real test is a constant frequency and amplitude excitation. It can be seen that the X-vibration absorber can effectively absorb the vibration energy of the main system and has obvious vibration absorption effect in the resonance frequency (1.4 Hz-1.6 Hz). Figures 12(a)-12(d)

[0100] Figures 13(a)-13(b) ​​The vibration absorption effect of the X-vibration absorber on the main system under impact excitation is shown in the optimal state, wherein Fig. 13(a) shows the time history signal of the response of the X-vibration absorber and the response of the main system. Obviously, the vibration absorption effect of the X-vibration absorber is significant. For the main system, the vibration energy is absorbed by the X-vibration absorber, so that the vibration amplitude decays rapidly. After four periods, the main system almost stops vibrating. The amplitude-frequency response curves of the main system and the X-vibration absorber are shown in Fig. 13(b) respectively. It is worth noting that there is an equal peak in the amplitude-frequency curve of the main system mass at this time, which means the optimal vibration absorption effect.

[0101] Further comparative experimental results are shown in Fig. 14(a), Fig. 14(b) (traditional spring-mass vibration absorber), Fig. 15(a), Fig. 15(b) (without vibration absorber), Fig. 14(a), Fig. 14(b) are the system response and vibration suppression effect of the spring-mass vibration absorber system under the optimized parameters. Fig. 15(a), Fig. 15(b) is the response of the main system mass without vibration absorber under impact excitation.

[0102] The experimental test results verify the effective vibration absorption / suppression effect of the X-vibration absorber under random, sinusoidal and impact excitation conditions. Especially in the case of optimal design parameters (Fig. 13(a), Fig. 13(b)), the X-vibration absorber can effectively absorb the resonance energy from the main structure, while the resonance peak amplitude of the main structure can be greatly reduced.

[0103] The embodiment optimizes the structural parameters (θ1, L1, β) of the X-vibration absorber based on the HBM and nested loop optimization program, and obtains the optimal stiffness and damping value to adapt to different vibration environments. ∞ The optimization target is used to optimize the structural parameters (θ1, L1, β) of the X-vibration absorber to obtain the optimal stiffness and damping value to adapt to different vibration environments.

[0104] Compared with the linear damping vibration absorber, the nonlinear damping of the X-vibration absorber can help improve the robustness of the system. Compared with the linear vibration absorber, by increasing the nonlinear LLS stiffness, the X-vibration absorber can significantly widen the vibration suppression bandwidth and has a lower resonance peak. This feature is beneficial to vibration suppression in a specific frequency range, for example, at resonance.

[0105] For strong nonlinear cases, the traditional nonlinear tuned vibration absorber with cubic stiffness cannot eliminate the instability problems of the Duffing main structure, such as bifurcation and separated resonance curves. On the contrary, the X-vibration absorber can successfully eliminate these potential instabilities, which is a very important safety factor in practice. Experimental tests verify the effectiveness of the X-vibration absorber under different excitations compared with the traditional spring-mass vibration absorber.

[0106] The above descriptions are only the preferred embodiments of the present application, and are not intended to limit the present application. The present application can have various modifications and changes for those skilled in the art. Any modifications, equivalent replacements, improvements, etc. made within the spirit and principle of the present application shall be included in the protection scope of the present application.

Claims

1. An adjustable nonlinear dynamic vibration absorber, characterized by, The tuning mass block is slidingly connected to the surface of the support plate, and the first scissor structure is rotatably connected between the tuning mass block and the two ends of the support plate. The second scissor structure and the damper can drive the first scissor structure to stretch and contract through the elastic element. The elastic element is parallel to the sliding direction of the tuning mass block and perpendicular to the damping direction of the damper, so that the first scissor structure and the second scissor structure realize vibration absorption through geometric nonlinearity when moving.

2. The adjustable nonlinear vibration absorber of claim 1, wherein, The first scissor structure includes at least one first X-shaped unit, and the first X-shaped unit is installed with the second scissor structure on one side along the sliding direction of the tuning mass block and with the damper on the other side.

3. The adjustable nonlinear vibration absorber of claim 1, wherein, The second scissor structure is connected to one end of the first scissor structure close to the end of the support plate, and the damper is connected to one end of the first scissor structure close to the tuning mass block. The first scissor structure is connected to the tuning mass block and / or the support plate through a rotating shaft and a bearing installed outside the rotating shaft.

4. The adjustable nonlinear vibration absorber of claim 1, wherein, Alternatively, the first scissor structure is connected to the tuning mass block and / or the support plate through a hinge. The second scissor structure is connected to the first scissor structure through a rotating shaft and a bearing installed outside the rotating shaft.

5. The adjustable nonlinear vibration absorber of claim 1, wherein, Alternatively, the second scissor structure is connected to the first scissor structure through a hinge.

6. The adjustable nonlinear vibration absorber of claim 1 or 5, wherein, The support plate is installed with a guide rail, and the tuning mass block is slidingly connected to the guide rail.

7. An adjustable nonlinear dynamic vibration absorber system, characterized by, The support plate is a U-shaped structure.

8. An adjustable nonlinear vibration absorber system according to claim 7, wherein The vibration absorber includes any one of claims 1-6. The vibration absorber is detachably connected to the main structure.

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