Shock absorber and rotor system

Abstract

The invention provides a shock absorber and a rotor system. The shock absorber comprises an annular inertia mass ring and a center assembly ring. The inertial mass ring and the central assembly ring are concentrically arranged; the radial size of the inner side of the inertia mass ring is larger than the radial size of the outer side of the center assembly ring. A plurality of springs are radially arranged between the inertial mass ring and the central assembly ring; one end of the spring is arranged on the inertia mass ring, and the other end of the spring is arranged on the center assembly ring. The technical scheme can be widely applied to the technical field of vibration reduction of rotor systems.

Description

Technical Field

[0001] The invention relates to the technical field of vibration and noise control, in particular to the technical field of vibration reduction of a rotor system. Background Art

[0002] Rotors or shafts are widely used in many engineering fields such as aerospace, ships, and vehicles. As an important part of mechanical equipment, the dynamic characteristics of rotating parts are closely related to the operating life, mechanical efficiency, vibration and noise of mechanical equipment. Ensuring the stable operation of the shaft is a prerequisite for the normal operation of engineering machinery.

[0003] Usually, the rotating shaft is affected by environmental factors during operation. For example, it is affected by flow field excitation load, axial load, collision and friction with the machine body, fatigue cracks caused by long-term high-intensity work, worn bearings and other factors. These external factors acting on the rotating shaft are collectively referred to as external excitation. External excitation will cause lateral vibration, longitudinal vibration or torsional vibration of the rotating shaft. These vibrations are very unfavorable to the stable operation of the rotating shaft under rated conditions, and also pose potential hazards to the safe operation of the overall equipment.

[0004] In order to reduce the abnormal vibration of the rotating shaft under external excitation, people have developed a large number of active or passive vibration absorbers and vibration isolators. However, most of the existing vibration absorbers and vibration isolators can only suppress a single vibration of the rotor, and cannot suppress the lateral vibration, longitudinal vibration and torsional vibration of the rotor machinery at the same time. At the same time, the existing vibration absorbers have complex structures, large volumes and masses, occupy a lot of mechanical equipment space, and have poor robustness. Summary of the invention

[0005] In order to solve the problems that the existing vibration absorber can only suppress a single vibration and has a complex structure, the present invention provides a vibration absorber and a rotor system.

[0006] The technical solution of the present invention is as follows:

[0007] A shock absorber comprises an annular inertial mass ring and a central assembly ring; the inertial mass ring and the central assembly ring are arranged concentrically; the radial dimension of the inner side of the inertial mass ring is greater than the radial dimension of the outer side of the central assembly ring; a plurality of springs are radially arranged between the inertial mass ring and the central assembly ring; one end of the spring is arranged on the inertial mass ring, and the other end of the spring is arranged on the central assembly ring.

[0008] Optionally, the number of the springs is an even number.

[0009] Optionally, the number of the springs is 8; the bending stiffness K of the springs is:

[0010]

[0011] Among them, the tension and compression stiffness k of the spring is:

[0012]

[0013] Among them, E0 is the Young's modulus of the spring, G0 is the shear modulus of the spring, v0 is the Poisson's ratio of the spring, θ0 is the helix angle of the spring, n0 is the number of turns of the helix of the spring, d0 is the diameter of the spring wire of the spring, r0 is half of d0, D0 is the spring median diameter of the spring, and R0 is half of D0.

[0014] Optionally, the potential energy V of the lateral vibration of the shock absorber α1 for:

[0015]

[0016] Among them, x a is the displacement of the inertial mass ring in the X direction in the XYZ coordinate system, y a is the displacement of the inertial mass ring in the Y direction in the XYZ coordinate system; the XYZ coordinate system is the coordinate system when the shock absorber is in a static state.

[0017] Optionally, the stiffness K of the shock absorber in the X direction during lateral vibration is x and the stiffness K in the Y direction Y for:

[0018]

[0019] Optionally, the potential energy V of the torsional vibration of the shock absorber α2 for:

[0020]

[0021] Wherein, d3 is the inner diameter of the inertial mass ring, d2 is the outer diameter of the central assembly ring, θ z is the rotation angle of the inertial mass around the positive direction of the Z axis.

[0022] Optionally, the torsional stiffness of the torsional vibration of the shock absorber is:

[0023]

[0024] Optionally, the potential energy V of the axial vibration of the shock absorber α3 for:

[0025]

[0026] Among them, z a is the displacement of the inertial mass ring in the Z direction in the XYZ coordinate system.

[0027] Optionally, the axial stiffness K of the shock absorber for axial vibration is Z for:

[0028]

[0029] A rotor system comprises a rotating shaft, on which the shock absorber as claimed in claim 1 is arranged, and the number of the springs is 8; two mass disks, a first mass disk and a second mass disk, are arranged on the rotating shaft, and the first mass disk and the second mass disk are respectively arranged on both sides of the shock absorber, and the dynamic equation of the rotor system is:

[0030]

[0031] in,

[0032]

[0033]

[0034]

[0035]

[0036] The length of the shaft is L, the radius of the shaft is R, the cross-sectional area of ​​the shaft perpendicular to the length direction is A, the density of the shaft is ρ; E is the elastic modulus of the shaft, G is the shear modulus of the shaft, I p is the polar moment of inertia of the shaft, I d is the equatorial moment of inertia of the rotating shaft, and c is the damping of the rotating shaft;

[0037] The xyz coordinate system is the three-dimensional coordinate system where the shock absorber is located after rotating around the rotation axis;

[0038] In the xyz coordinate system, the translational stiffness of the boundary constraints of the axis of rotation are K x0 , K xL , K y0 , K yL , K z0 , K zL The rotational stiffness of the shaft is k x0 , k xL , k y0 , k yL , k z0 , k zL The damping of the boundary constraints of the rotating axis are c x0 、c xL 、c y0 、c yL 、c z0 、czL , Ω is the angular velocity of the shaft rotating around the positive direction of the z-axis; m0 is the unbalanced concentrated mass of the shaft, r is the eccentricity of the concentrated mass; m d1 is the mass of the first mass disk, m d2 is the mass of the second mass disk, m d3 is the mass of the central assembly ring, m a is the mass of the inertial mass ring; J d1 is the equatorial moment of inertia of the first mass disk, J d2 is the equatorial moment of inertia of the second mass disk, J d3 J is the equatorial moment of inertia of the central assembly ring; p1 is the polar moment of inertia of the first mass disk, J p2 is the polar moment of inertia of the second mass disk, J p3 is the polar moment of inertia of the center assembly ring, J pa is the polar moment of inertia of the inertial mass ring; c a is the damping of the shock absorber; q x is the generalized coordinate of the rotation axis in the x-axis direction, q y is the generalized coordinate of the rotation axis in the y-axis direction, q z is the generalized coordinate of the rotation axis in the z-axis direction, q zθ w is the generalized coordinate of the rotation axis twisted along the positive direction of the z-axis; xa is the coordinate of the shock absorber on the x-axis, w ya is the coordinate of the shock absorber on the y-axis, w za is the coordinate of the shock absorber on the z-axis, θ za is the rotation angle of the shock absorber along the positive direction of the z-axis; δ is the Dirac function; z is the integrated variable, z1 is the coordinate of the first mass disk on the z-axis, z2 is the coordinate of the second mass disk on the z-axis, and z3 is the coordinate of the shock absorber on the z-axis; Ф x , Ф y , Ф z are the displacement functions of the rotor system in the x-axis, y-axis and z-axis directions respectively; s is the Timoshenko variable, μ is the shear coefficient; M0(t) is the torsional excitation, and F0(t) is the axial excitation.

[0039] The technical effects of the present invention are as follows:

[0040] The vibration damper of the present invention only includes an inertial mass ring, a center assembly ring and a spring, and has a simple structure. After being assembled to the rotating shaft, it can utilize the expansion and contraction and bending of the spring to absorb the vibration energy in three directions of the rotating shaft, thereby achieving the purpose of suppressing the lateral vibration, longitudinal vibration and torsional vibration of the rotating shaft. The rotor system of the present invention is equipped with the vibration damper of the present invention, and the vibration damper can be set according to the operating conditions of the rotor system, so as to more accurately suppress the lateral vibration, longitudinal vibration and torsional vibration of the rotating shaft. In summary, the technical solution of the present invention achieves the purpose of the present invention.

[0041] The further effects of the above optional manner will be described below in conjunction with the specific implementation manner. BRIEF DESCRIPTION OF THE DRAWINGS

[0042] Figure 1 It is a front view of an embodiment of the shock absorber of the present invention.

[0043] Figure 2 for Figure 1 A side cross-sectional view of the illustrated embodiment.

[0044] Figure 3 Spring bending stiffness analysis Figure 1 .

[0045] Figure 4 Spring bending stiffness analysis Figure 2 .

[0046] Figure 5 Spring bending stiffness analysis Figure 3 .

[0047] Figure 6 Spring bending stiffness analysis Figure 4 .

[0048] Figure 7 This is a comparison chart between the calculation results of the spring bending stiffness formula and the finite element simulation results.

[0049] Figure 8 for Figure 1 Lateral vibration analysis diagram of the shock absorber shown.

[0050] Fig. 9 for Figure 1 Torsional vibration analysis diagram of the shock absorber shown.

[0051] Fig.10 for Figure 1 Axial vibration analysis diagram of the shock absorber shown.

[0052] Fig.11 It is a vibration analysis diagram of the rotor system of the present invention.

[0053] Fig.12 for Fig.11The vibration amplitude-frequency characteristic curve of the disk 2 in the rotor system is shown.

[0054] Fig.13 for Fig.11 The time domain response curve of the disk 2 in the rotor system shown in FIG. 1 during lateral vibration.

[0055] Fig.14 for Fig.11 The time domain response curve of the disk 2 in the rotor system during torsional vibration is shown.

[0056] Fig.15 for Fig.11 The time domain response curve of the disk 2 in the rotor system shown in FIG. 1 when it vibrates axially.

[0057] The symbols in the figure are explained as follows:

[0058] 1. shock absorber; 101. inertial mass ring; 102. spring; 103. center assembly ring. DETAILED DESCRIPTION

[0059] The technical solution of the present invention is described in detail below in conjunction with the accompanying drawings.

[0060] like Figure 1 and Figure 2 As shown, the shock absorber 1 includes an annular inertial mass ring 101 and a central assembly ring 103. The two annular inertial mass rings 101 and the central assembly ring 103 are arranged concentrically, that is, the geometric center of the inertial mass ring 101 coincides with the geometric center of the central assembly ring 103. The radial dimension of the inner side of the inertial mass ring 101 is greater than the radial dimension of the outer side of the central assembly ring 103, that is, the inertial mass ring 101 is located in the outer ring position relative to the central assembly ring 103. Eight telescopic coil springs 102 are radially arranged between the inertial mass ring 101 and the central assembly ring 103. The radial arrangement refers to an arrangement mode in which the telescopic direction of the spring 103 coincides with the radial direction of the inertial mass ring 101 and the central assembly ring 103. One end of the spring 102 along its telescopic direction is fixedly connected to the inertial mass ring 101, and the other end of the spring 102 along its telescopic direction is fixedly connected to the central assembly ring 103.

[0061] When the vibration absorber 1 is assembled to the rotor system (ref. Fig.11 ), the central assembly ring 103 is fixed on the rotating shaft of the rotor system, and the lateral vibration, axial vibration and torsional vibration generated by the rotor system generate corresponding vibrations in the lateral, axial and rotational directions through the rotating shaft excitation damper 1. Figures 8 to 9, the spring 102 on the shock absorber 1 absorbs the lateral vibration excitation from the cross-sectional direction of the rotating shaft through its own tensile stiffness and compressive stiffness; the spring 102 on the shock absorber 1 absorbs the bending vibration excitation from the cross-sectional direction of the rotating shaft through its own bending stiffness (bending stiffness); the spring 102 on the shock absorber 1 absorbs the torsional vibration excitation from the rotating shaft through its own bending stiffness. The shock absorber 1 absorbs the vibration excitation of the six degrees of freedom of the rotor system through the tensile stiffness, compressive stiffness and bending stiffness of the spring 102. The potential energy formula and stiffness formula of the shock absorber 1 in the corresponding direction can be further obtained through the tensile stiffness formula, compressive stiffness formula and bending stiffness formula of the spring 102, and then the vibration reduction effect of the shock absorber 1 in the rotor system can be accurately obtained.

[0062] The formulas for the tensile stiffness and compressive stiffness of the spring 102 belong to the prior art, and the bending stiffness formula is the key to obtaining the vibration reduction effect of the shock absorber of the present invention. Figures 3 to 6 This is the process of analyzing and deducing the bending stiffness of the spring 102 . Figure 4 for Figure 3 Left view, Figure 5 for Figure 3 A partial enlarged view from a top view.

[0063] like Figure 3 As shown, one end of the coil spring (hereinafter referred to as the spring) is fixed, and the other end of the spring is a free end. When a torque M0 is applied to the free end of the spring, the rotation angle of the center line of the spring (the axis of the spring's expansion and contraction direction) after the corresponding spring is bent is α. The total length of the spring is L0, the pitch of the spring is l0, and n0 is the effective number of turns of the spring (the number of turns that the spring can maintain the same pitch), then:

[0064] L0=n0l0 (1)

[0065] The helix angle of the spring is θ0, and the mean diameter of the spring (i.e. the diameter of the helix formed by the center line of the spring) is D0, then:

[0066]

[0067] The length of the spring wire (the length of the spring wire corresponding to the effective number of coils of the spring) is s0, then:

[0068]

[0069] like Figure 4 As shown in the figure, the black dashed arrow indicates the direction of the torque M0. For the convenience of calculation, the torque M0 is decomposed into M1 and M2. The decomposition process is as follows: Figure 6 As shown in the figure on the left. Decompose the torque M2 into M 21 and M 22 , the decomposition process is as follows Figure 6 As shown in the figure on the right. Figure 6The viewing angle of the left image is Figure 4 A view showing the spring coil. Figure 6 The middle right picture is Figure 4 The radial directions of the spring coils presented in are:

[0070]

[0071] M1 and M 22 The bending strain per unit length of spring wire, M 21 The torsional strain per unit length of the spring wire is generated. I in formula (5) da is the equatorial moment of inertia per unit length of spring wire, I pa is the polar moment of inertia of the spring wire per unit length, and the total strain energy V0 of the spring wire is:

[0072]

[0073] The total work W0 done by torque M0 is:

[0074]

[0075] Substituting formula (4) into formula (5) yields formula (7):

[0076]

[0077] The work done by the torque is equal to the strain energy of the spring, and the final bending stiffness of the spring obtained by the deformation formula (7) is shown in formula (8). In formula (8), E0 is the Young's modulus of the spring, G0 is the shear modulus of the spring, v0 is the Poisson's ratio of the spring, θ0 is the helical angle of the spring, n0 is the number of turns of the helical rotation of the spring, d0 is the diameter of the spring wire, r0 is the radius of the spring wire, D0 is the spring median diameter of the spring, and R0 is half of D0.

[0078]

[0079] The accuracy of formula (8) is verified by comparative experiments. The specific experimental method is to use formula (8) to calculate the bending stiffness of springs with different numbers of spring coils under the premise of determining the physical parameters such as the spring helix angle and Young's modulus. The calculated spring bending stiffness is compared with the spring bending stiffness obtained by simulation using the finite element software Ansys (the simulation platform is Ansys Workbench 2022R1). Figure 7 The following is a comparison chart of the calculation results of multiple springs with the same parameters using formula (8) and the simulation results using Ansys finite element software. Figure 7It can be seen that under different numbers of spring coils, the fitting degree of the spring bending stiffness curve is high, which proves the accuracy of formula (8).

[0080] In the above comparative experiment, the physical properties of the spring used are shown in Table 1.

[0081] Table 1

[0082]

[0083] After obtaining the bending stiffness formula of the spring, the potential energy formula and stiffness formula in the corresponding direction of the shock absorber 1 can be further obtained, thereby accurately predicting the vibration reduction effect of the shock absorber 1 in the rotor system. The specific derivation process of the potential energy formula and stiffness formula is described below.

[0084] Lateral vibration

[0085] Figure 8 (a) is a schematic diagram of the lateral vibration of the shock absorber. Figure 8 (b) is a schematic diagram of the coordinate transformation and force decomposition of the shock absorber. Figure 8 (c) is the lateral vibration analysis model diagram.

[0086] like Figure 8 As shown in (a), when the shaft (i.e., the shaft connected to the center assembly ring) produces lateral vibration, the vibration absorber is excited to produce lateral vibration, i.e. Figure 8 (a) Vibration generated by the shock absorber along its radial direction. Figure 8 The phantom in (a) shows the original position of the shock absorber when no vibration occurs. The extension and compression of the spring in the shock absorber absorbs the energy of the lateral vibration.

[0087] like Figure 8 As shown in (b), the coordinates of the rotation center of the shock absorber in the fixed coordinate system (the two-dimensional coordinate system in the initial state when the shock absorber spring does not undergo elastic deformation) XOY are converted to the coordinates in the rotated two-dimensional coordinate system xoy:

[0088]

[0089] Among them, Ω is the rotation speed of the shaft, and Ωt is the rotation angle of the coordinate system xoy.

[0090] Decompose the force exerted by the shock absorber on the rotating shaft into the force F along the oy direction in the xoy coordinate system y and the force F along the ox direction x :

[0091]

[0092] The force F exerted by the shock absorber on the rotating shaft in the XOY coordinate systemX and F Y From the force F in formula (2), x and F y Express:

[0093]

[0094] Substituting formula (1) and formula (2) into formula (3), we can get F X and F Y The expression in XOY coordinate system:

[0095]

[0096] Then, the potential energy V of the shock absorber in lateral vibration α1 for:

[0097]

[0098] The stiffness K of the shock absorber in the OX direction during lateral vibration x and the stiffness K in the OY direction Y for:

[0099]

[0100] Torsional vibration

[0101] Fig. 9 (a) is a schematic diagram of torsional vibration of the shock absorber. Fig. 9 (b) is the geometric model diagram of the shock absorber during torsional vibration. Fig. 9 (c) is the stiffness analysis diagram of the shock absorber during torsional vibration.

[0102] like Fig. 9 As shown in (a), when the rotating shaft (i.e., the rotating shaft connected to the center assembly ring) generates torsional vibration, the vibration absorber is excited to generate torsional vibration, i.e., the vibration absorber in 9 (a) rotates around the rotation center of the center assembly ring. Fig. 9 The phantom in (a) shows the original position of the shock absorber when no torsional vibration occurs. The bending of the spring in the shock absorber absorbs the energy of the torsional vibration. Fig. 9 The red dotted line is an auxiliary line, d3 is the inner diameter of the inertial mass ring, d2 is the outer diameter of the center assembly ring, θ z is the torsion angle of the shock absorber, and β is the torsion angle of a single spring.

[0103] from Fig. 9 From the analysis of (c), we can see that:

[0104]

[0105] Then the torque M of the shock absorber during torsional vibration is:

[0106]

[0107] The potential energy V of the shock absorber during torsional vibration α2 for:

[0108]

[0109] The stiffness of the shock absorber during torsional vibration G Z for:

[0110]

[0111] Axial vibration

[0112] Fig.10 (a) is a schematic diagram of the axial vibration of the shock absorber. Axial vibration refers to the vibration of the shock absorber along the axis of the shaft on which it is located. Fig.10 The virtual shadow in (a) shows the original position of the shock absorber when no axial vibration occurs. Fig.10 (b) is the geometric model diagram of the shock absorber during axial vibration. Fig.10 (c) is the stiffness analysis diagram of the shock absorber during axial vibration.

[0113] like Fig.10 As shown in (a), when the rotating shaft (i.e. the rotating shaft connected to the center assembly ring) generates axial vibration, the shock absorber is excited to generate axial vibration. When the axial vibration is generated, the bending of the spring in the shock absorber absorbs the energy of the axial vibration. Fig.10 The red dotted line in the middle is the auxiliary line, and γ is the torsion angle of a single spring in the shock absorber.

[0114] from Fig.10 From the analysis of (c), we can see that:

[0115]

[0116] When the shock absorber is in axial vibration, the torsion angle of the spring in the shock absorber is:

[0117]

[0118] Potential energy V of the shock absorber in axial vibration α3 for:

[0119]

[0120] Among them, z a is the displacement of the inertial mass ring in the Z direction in the XYZ coordinate system. The XYZ coordinate system is a three-dimensional coordinate system in the initial state when the shock absorber spring does not undergo elastic deformation.

[0121] Axial stiffness K of the shock absorber in axial vibration Z for:

[0122]

[0123] After obtaining the potential energy formula and stiffness formula of the shock absorber 1 in the corresponding direction, the dynamic equation of the rotor system after the shock absorber 1 is assembled to the rotor system can be further obtained. Fig.11 An example of such a rotor system is shown in Figure 1. Fig.11 As shown, the rotor system includes two rigid rotating disks arranged on the rotating shaft, and a vibration absorber 1 arranged on the same rotating shaft as the two rigid rotating disks and arranged between the two rigid rotating disks. The general boundary conditions of the rotor system include translational and rotational support for lateral vibration, torsional support for torsional vibration, and axial support for axial vibration. The vibration absorber 1 will absorb the energy of the lateral vibration, torsional vibration and axial vibration of the rotor system. The dynamic equations of the rotor system include the lateral vibration equation, the torsional vibration equation and the axial vibration equation. The following reference Fig.11 The dynamic equations of the rotor system are deduced in detail.

[0124] The parameters involved in the following derivation process are as follows: the length of the shaft is L, the radius of the shaft is R, the cross-sectional area of ​​the shaft perpendicular to the length direction is A, and the density of the shaft is ρ; E is the elastic modulus of the shaft, G is the shear modulus of the shaft, I p is the polar moment of inertia of the shaft, I d is the equatorial moment of inertia of the rotating shaft, and c is the damping of the rotating shaft;

[0125] The xyz coordinate system is a three-dimensional coordinate system after the shock absorber spring undergoes elastic deformation;

[0126] In the xyz coordinate system, the translational stiffness of the boundary constraints of the axis of rotation are K x0 , K xL , K y0 , K yL , K z0 , K zL The rotational stiffness of the shaft is k x0 , k xL , k y0 , k yL , k z0 , k zL The damping of the boundary constraints of the rotating axis are c x0 、c xL 、c y0 、c yL 、c z0 、c zL, Ω is the angular velocity of the shaft rotating around the positive direction of the z-axis; m0 is the unbalanced concentrated mass of the shaft, r is the eccentricity of the concentrated mass; m d1 is the mass of the first mass disk, m d2 is the mass of the second mass disk, m d3 is the mass of the central assembly ring, m a is the mass of the inertial mass ring; J d1 is the equatorial moment of inertia of the first mass disk, J d2 is the equatorial moment of inertia of the second mass disk, J d3 J is the equatorial moment of inertia of the central assembly ring; p1 is the polar moment of inertia of the first mass disk, J p2 is the polar moment of inertia of the second mass disk, J p3 is the polar moment of inertia of the center assembly ring, J pa is the polar moment of inertia of the inertial mass ring; c a is the damping of the shock absorber; q x is the generalized coordinate of the rotation axis in the x-axis direction, q y is the generalized coordinate of the rotation axis in the y-axis direction, q z is the generalized coordinate of the rotation axis in the z-axis direction, q zθ w is the generalized coordinate of the rotation axis twisted along the positive direction of the z-axis; xa is the coordinate of the shock absorber on the x-axis, w ya is the coordinate of the shock absorber on the y-axis, w za is the coordinate of the shock absorber on the z-axis, θ za is the rotation angle of the shock absorber along the positive direction of the z-axis; δ is the Dirac function; z is the integrated variable, z1 is the coordinate of the first mass disk on the z-axis, z2 is the coordinate of the second mass disk on the z-axis, and z3 is the coordinate of the shock absorber on the z-axis; Ф x , Ф y , Ф z are the displacement functions of the rotor system in the x-axis, y-axis and z-axis directions respectively; s is the Timoshenko variable, μ is the shear coefficient; M0(t) is the torsional excitation, and F0(t) is the axial excitation.

[0127] Substitute the kinetic energy T and potential energy V of the rotor system and the external force work W into the Lagrangian operator L a .

[0128] L a =T-V+W (13)

[0129] Then, the second kind of Lagrange equation is used to solve the total vibration matrix equation of the rotor system (14):

[0130]

[0131] The matrices M, C, J, and K in formula (15) are s , F and the generalized coordinate q are shown in formula (16) respectively.

[0132]

[0133] The specific composition of the mass matrix M of the rotor system is shown in formula (17).

[0134]

[0135] The specific composition of the damping matrix C of the rotor system is shown in formula (18).

[0136]

[0137] The specific composition of the gyro matrix J of the rotor system is shown in formula (19).

[0138]

[0139] The stiffness matrix K of the rotor system s The specific composition of is shown in formula (20).

[0140]

[0141] The specific composition of the external force matrix F of the rotor system is shown in formula (21).

[0142]

[0143] The following simulation experiment compares the rotor system without a vibration damper and the rotor system with a vibration damper (such as Fig.11 The dynamic response of the rotor system (shown) is measured to verify the vibration reduction effect of the shock absorber.

[0144] Fig.12 Given Fig.11 The amplitude-frequency characteristic curves of the rigid rotating disk in the rotor system shown in the figure for lateral vibration, torsional vibration and axial vibration are obtained with and without vibration absorbers. Fig.12 It can be seen that in lateral vibration and axial vibration, the vibration absorber can significantly suppress the vibration of the main resonance area. In torsional vibration, the rotor system with the vibration absorber moves the first-order resonance peak and provides good vibration suppression in a wider frequency band.

[0145] Fig.13 The time domain response curves of the rigid turntable in the rotor system during lateral vibration (with and without vibration absorbers) are shown. Fig.13 It can be seen that the damper reduces the steady-state vibration amplitude of the rotor system at 110 Hz and shortens the time required for the system to reach steady-state motion.

[0146] Fig.14 The time domain response curves of the rigid turntable in the rotor system during torsional vibration (with and without vibration absorbers) are shown. Fig.14 It can be seen that the damper reduces the steady-state vibration amplitude of the rotor system at 210 Hz.

[0147] Fig.15 The time domain response curves of the rigid turntable in the rotor system under axial vibration (with and without vibration absorbers) are shown. Fig.15 It can be seen that the damper reduces the steady-state vibration amplitude of the rotor system at 106 Hz and shortens the time required for the system to reach steady-state motion.

[0148] From the vibration time domain diagrams of the rotor system under the above three vibration modes, it can be verified that the vibration absorber allows the rotor system to recover to a low-amplitude motion state at a faster speed under the premise of displacement excitation, and in the stable vibration state, the amplitude of the rotor system containing the vibration absorber is low.

[0149] The vibration damper and the rotor system of the present invention can quantify the vibration damping effect, and the vibration damper can be set according to the operating conditions of the rotor system, so as to more accurately suppress the lateral vibration, longitudinal vibration and torsional vibration of the rotating shaft.

[0150] It is worth noting that the above description is only a preferred embodiment of the present invention, and does not limit the scope of patent protection of the present invention. The present invention can also be replaced by equivalent technologies. Therefore, any equivalent changes made by using the description and illustrations of the present invention, or directly or indirectly applied to other related technical fields, are included in the scope of the present invention.

Claims

1. A shock absorber, characterized in that: The invention comprises an annular inertial mass ring and a central assembly ring; the inertial mass ring and the central assembly ring are arranged concentrically; the radial dimension of the inner side of the inertial mass ring is greater than the radial dimension of the outer side of the central assembly ring; a plurality of springs are radially arranged between the inertial mass ring and the central assembly ring; one end of the spring is arranged on the inertial mass ring, and the other end of the spring is arranged on the central assembly ring.

2. A vibration absorber according to claim 1, characterized in that: The number of the springs is an even number.

3. A vibration absorber according to claim 2, characterized in that: The number of the springs is 8; the bending stiffness K of the spring is: Among them, the tension and compression stiffness k of the spring is: Among them, E0 is the Young's modulus of the spring, G0 is the shear modulus of the spring, v0 is the Poisson's ratio of the spring, θ0 is the helix angle of the spring, n0 is the number of turns of the helix of the spring, d0 is the diameter of the spring wire of the spring, r0 is half of d0, D0 is the spring median diameter of the spring, and R0 is half of D0.

4. The vibration absorber according to claim 3, characterized in that: The potential energy V of the lateral vibration of the shock absorber α1 for: Among them, x a is the displacement of the inertial mass ring in the X direction in the XYZ coordinate system, y a is the displacement of the inertial mass ring in the Y direction in the XYZ coordinate system; the XYZ coordinate system is the coordinate system when the shock absorber is in a static state.

5. The vibration absorber according to claim 4, characterized in that: The stiffness K of the shock absorber in the X direction during lateral vibration x and the stiffness K in the Y direction Y for:

6. The vibration absorber according to claim 5, characterized in that: The potential energy V of the torsional vibration of the shock absorber α2 for: Wherein, d3 is the inner diameter of the inertial mass ring, d2 is the outer diameter of the central assembly ring, θ z is the rotation angle of the inertial mass around the positive direction of the Z axis.

7. The vibration absorber according to claim 6, characterized in that: The torsional stiffness of the shock absorber for torsional vibration is:

8. The vibration absorber according to claim 7, characterized in that: The potential energy V of the axial vibration of the shock absorber α3 for: Among them, z a is the displacement of the inertial mass ring in the Z direction in the XYZ coordinate system.

9. The vibration absorber according to claim 3, characterized in that: The axial stiffness K of the shock absorber for axial vibration Z for:

10. A rotor system, including a rotating shaft, characterized in that: The vibration absorber as claimed in claim 1 is arranged on the rotating shaft, and the number of the springs is 8; two mass discs, a first mass disc and a second mass disc, are arranged on the rotating shaft, and the first mass disc and the second mass disc are arranged on both sides of the vibration absorber respectively, and the dynamic equation of the rotor system is: in, The length of the shaft is L, the radius of the shaft is R, the cross-sectional area of ​​the shaft perpendicular to the length direction is A, the density of the shaft is ρ; E is the elastic modulus of the shaft, G is the shear modulus of the shaft, I p is the polar moment of inertia of the shaft, I d is the equatorial moment of inertia of the rotating shaft, and c is the damping of the rotating shaft; The xyz coordinate system is the three-dimensional coordinate system where the shock absorber is located after rotating around the rotation axis; In the xyz coordinate system, the translational stiffness of the boundary constraints of the axis of rotation are K x0 , K xL , K y0 , K yL , K z0 , K zL The rotational stiffness of the shaft is k x0 , k xL , k y0 , k yL , k z0 , k zL The damping of the boundary constraints of the rotating axis are c x0 、c xL 、c y0 、c yL 、c z0 、c zL , Ω is the angular velocity of the shaft rotating around the positive direction of the z-axis; m0 is the unbalanced concentrated mass of the shaft, r is the eccentricity of the concentrated mass; m d1 is the mass of the first mass disk, m d2 is the mass of the second mass disk, m d3 is the mass of the central assembly ring, m a is the mass of the inertial mass ring; J d1 is the equatorial moment of inertia of the first mass disk, J d2 is the equatorial moment of inertia of the second mass disk, J d3 J is the equatorial moment of inertia of the central assembly ring; p1 is the polar moment of inertia of the first mass disk, J p2 is the polar moment of inertia of the second mass disk, J p3 is the polar moment of inertia of the center assembly ring, J pa is the polar moment of inertia of the inertial mass ring; c a is the damping of the shock absorber; q x is the generalized coordinate of the rotation axis in the x-axis direction, q y is the generalized coordinate of the rotation axis in the y-axis direction, q z is the generalized coordinate of the rotation axis in the z-axis direction, q zθ w is the generalized coordinate of the rotation axis twisted along the positive direction of the z-axis; xa is the coordinate of the shock absorber on the x-axis, w ya is the coordinate of the shock absorber on the y-axis, w za is the coordinate of the shock absorber on the z-axis, θ za is the rotation angle of the shock absorber along the positive direction of the z-axis; δ is the Dirac function; z is the integrated variable, z1 is the coordinate of the first mass disk on the z-axis, z2 is the coordinate of the second mass disk on the z-axis, and z3 is the coordinate of the shock absorber on the z-axis; Ф x , Ф y , Ф z are the displacement functions of the rotor system in the x-axis, y-axis and z-axis directions respectively; s is the Timoshenko variable, μ is the shear coefficient; M0(t) is the torsional excitation, and F0(t) is the axial excitation.

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