Magnetoelastic coupling quasi-zero stiffness vibration reduction sliding shoe assembly

Abstract

A magnetoelastic coupling quasi-zero rigidity vibration reduction sliding shoe assembly comprises a sliding shoe body and a ball socket, and the ball socket is provided with a mounting part used for containing and fixing; a ball head of the plunger is detachably mounted in the ball socket of the sliding shoe body; the rigidity compensation mechanism is arranged on the mounting part of the sliding shoe body and comprises an elastic element for providing positive rigidity and a magnetic element for providing negative rigidity, and the positive rigidity and the negative rigidity can be mutually counteracted under a preset working load by adjusting the pre-compression amount of the elastic element and the initial acting distance of the magnetic element; therefore, the sliding shoe assembly presents the quasi-zero stiffness characteristic near the working point.

Description

Technical Field

[0001] This invention relates to the field of hydraulic pump vibration reduction technology, and in particular to a magnetoelastic coupling quasi-zero stiffness vibration reduction slipper assembly. Background Technology

[0002] In the process of increasing the pressure of axial piston pumps, vibration suppression of the slipper-swashplate pair has become a key bottleneck restricting technological breakthroughs. Periodic pressure pulsations in the piston chamber are transmitted to the slipper through the rigidly connected piston ball head, triggering high-frequency axial vibrations. This not only accelerates fatigue spalling on the swashplate surface but also induces resonance in the pump body structure, leading to a sharp drop in system reliability. Currently, traditional vibration reduction schemes have certain limitations. While the scheme using rubber gaskets and disc spring preload structures can attenuate vibration amplitude to some extent, its positive stiffness characteristics make it difficult to overcome the limiting contradiction between "high load-bearing stiffness and low vibration isolation flexibility."

[0003] While hydraulic damping designs can improve the dynamic response of the system to some extent, they are prone to instability under ultra-low speed conditions, thus affecting the vibration reduction effect. Although existing quasi-zero stiffness technology has applications in vibration isolation, it is difficult to apply to slip shoe pairs due to excessive space requirements, poor load adaptability, and insufficient pressure resistance. Summary of the Invention

[0004] To address the shortcomings of the prior art, the present invention aims to provide a magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly, which aims to offset the dynamic stiffness of the system during operation in real time through a positive and negative stiffness coordination mechanism, optimize the vibration energy transmission path of the slipper pair, suppress the axial vibration of the high-pressure plunger pump, and reduce the risk of swashplate wear and plunger fatigue fracture.

[0005] To achieve the above objectives, the present invention provides the following technical solution:

[0006] A magnetoelastically coupled quasi-zero stiffness vibration damping slipper assembly, comprising:

[0007] The skate body includes a ball socket, and a ball socket is formed inside it;

[0008] A plunger having a ball head rotatably received within the ball socket; and

[0009] A stiffness compensation mechanism, located within the ball socket, includes an elastic element and a magnetic element arranged in parallel. The elastic element provides positive stiffness, and the magnetic element provides negative stiffness.

[0010] By adjusting the preload of the elastic element and the initial spacing of the magnetic element, the positive stiffness and negative stiffness cancel each other out within the specified working load area, thereby causing the slipper assembly to exhibit quasi-zero stiffness characteristics in that area.

[0011] Furthermore, the ball socket is provided with a mounting part for installing the stiffness compensation mechanism, the mounting part including at least two mounting grooves that are radially distributed and concentric with each other.

[0012] Furthermore, the two mounting slots are respectively a second mounting slot located on the outer side and a first mounting slot located on the inner side. The second mounting slot is located at the center of the ball socket, the magnetic element is located in the first mounting slot, and the elastic element is located in the second mounting slot.

[0013] Furthermore, the second mounting slot and the first mounting slot are connected by a stepped structure, and the stepped structure is provided with a centering and non-magnetic limiting member.

[0014] Furthermore, the limiting member is a ring-shaped structure, and its inner diameter is smaller than the outer diameter of the magnetic element.

[0015] Furthermore, the inner peripheral wall of the second mounting groove is provided with an inwardly protruding limiting protrusion for axially fixing the elastic element.

[0016] Furthermore, the magnetic element includes at least two permanent magnets spaced apart from each other and arranged with their same poles facing each other.

[0017] Furthermore, the outer diameter of the permanent magnet is smaller than the inner diameter of the first mounting groove, and larger than the inner diameter of the limiting member.

[0018] Furthermore, the elastic element is a butterfly spring, which is engaged with the limiting boss of the second mounting groove.

[0019] Furthermore, the friction surfaces of the sliding shoe and the plunger that rotate relative to each other are provided with a solid lubricating coating.

[0020] The beneficial effects of this invention are:

[0021] 1. The present invention proposes a magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly, which achieves quasi-zero stiffness characteristics near the working load point by parallel setting of elastic elements and magnetic elements. This makes the slipper assembly exhibit extremely high static stiffness to support the load when bearing the main working load, while having extremely low dynamic stiffness, which can efficiently isolate and attenuate high-frequency vibrations and pressure pulsations transmitted by the plunger.

[0022] 2. The present invention proposes a magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly, which provides a precise and reliable mounting foundation for the elastic element and magnetic element of the stiffness compensation mechanism by setting a coaxial mounting groove. This ensures that the elastic force and magnetic force act symmetrically on the plunger ball head along the same axis, avoiding additional torque and wear caused by eccentric force, and ensuring the accuracy and stability of the stiffness cancellation effect.

[0023] 3. The present invention proposes a magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly, which uses a pair of permanent magnets with the same poles facing each other as magnetic force elements to generate a nonlinear repulsive force that increases sharply with the decrease of distance. Combined with the linear positive stiffness of the elastic element, it is easy to achieve force balance and stiffness cancellation at a specific point. Attached Figure Description

[0024] To more clearly illustrate the technical solutions in the embodiments of the invention or the prior art, the drawings used in the description of the embodiments or the prior art will be briefly introduced below. Obviously, the drawings described below are only some embodiments of the invention. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0025] Figure 1 This is a structural cross-sectional view of a magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly according to the present invention;

[0026] Figure 2 This is a cross-sectional view of the slide body of a magnetoelastic coupling quasi-zero stiffness vibration damping slide assembly according to the present invention.

[0027] Figure 3 This is a cross-sectional view of the plunger of a magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly according to the present invention.

[0028] Figure 4 This is a schematic diagram of the plunger of a magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly according to the present invention;

[0029] Figure 5 This is a schematic diagram illustrating the working principle of a magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly according to the present invention.

[0030] In the figure, 10 is a sliding shoe; 20 is a plunger; 30 is an elastic element; 40 is a magnetic element; 50 is a limiting element; 60 is a limiting protrusion; 701 is a first mounting groove; and 702 is a second mounting groove. Detailed Implementation

[0031] The following is combined Figures 1-5 The present invention will be described in detail below.

[0032] A magnetoelastically coupled quasi-zero stiffness vibration damping slipper 10 component, such as Figure 1-2 As shown, including

[0033] The skate 10 body includes a ball socket, and a ball socket is formed inside it;

[0034] Piston 20, having a ball head rotatably received within a ball socket; and

[0035] The stiffness compensation mechanism, located inside the ball socket, includes an elastic element 30 and a magnetic element 40 arranged in parallel. The elastic element 30 is used to provide positive stiffness, and the magnetic element 40 is used to provide negative stiffness.

[0036] By adjusting the preload of the elastic element 30 and the initial spacing of the magnetic element 40, the positive stiffness and negative stiffness cancel each other out in the specified working load area, so that the slipper 10 assembly exhibits quasi-zero stiffness characteristics in this area.

[0037] The slide shoe 10 has a ball socket machined inside, and the plunger 20 has a ball head at one end. The ball head is accommodated in the ball socket in a swingable manner, forming a ball joint connection. The stiffness compensation mechanism is integrated inside the ball socket and consists of an elastic element 30 and a magnetic element 40 arranged in parallel. The elastic element 30 provides positive restoring force and exhibits positive stiffness characteristics when compressed, while the magnetic element 40 provides negative stiffness characteristics through the repulsive electromagnetic force it generates. By pre-adjusting the pre-compression amount of the elastic element 30 and the initial air gap between the two poles of the magnetic element 40, the positive and negative stiffness of the parallel mechanism can be almost completely canceled out when the slide shoe 10 is subjected to a specific working load. This achieves extremely low stiffness, i.e., quasi-zero stiffness characteristics, in a load range near the working point, effectively isolating the vibration transmission from the plunger 20.

[0038] In this embodiment, as Figure 3 As shown, the ball socket is provided with a mounting section for installing the stiffness compensation mechanism. The mounting section includes at least two radially distributed and concentric mounting grooves. Specifically, the mounting section includes at least two radially distributed and strictly concentric annular mounting grooves; these mounting grooves can be grooves, steps, or nested structures formed by machining, the purpose of which is to provide a precise positioning and mounting base for the elastic element 30 and the magnetic element 40, ensuring that the force axes of the two are collinear with the force transmission axis of the plunger 20, thereby achieving symmetrical force transmission and precise matching of stiffness characteristics.

[0039] In this embodiment, the two mounting slots are a second mounting slot 702 located on the outer side and a first mounting slot 701 located on the inner side. The second mounting slot 702 is located at the center of the ball socket, the elastic element 30 is located in the second mounting slot 702, and the magnetic element 40 is located in the first mounting slot 701. The mounting structure includes a second mounting slot 702 located on the outer side and a first mounting slot 701 located on the inner side and concentrically. The second mounting slot 702 has a larger radial dimension to accommodate and provide space for the elastic element 30 to undergo elastic deformation within the design range. The first mounting slot 701 is located at the center of the bottom of the ball socket, and its size matches that of the magnetic element 40 to precisely fix and position the magnetic element 40, ensuring that the magnetic field lines it generates act in a predetermined direction.

[0040] In this embodiment, the second mounting groove 702 and the first mounting groove 701 are connected by a stepped structure, and a centering and non-magnetic limiting member 50 is provided at the stepped structure. A limiting member is installed at the stepped transition; the limiting member is preferably made of a non-magnetic material, and its functions are: first, to serve as a machining reference to ensure the coaxiality of the inner and outer grooves; second, to serve as a mounting limit or stop for the magnetic element 40 to prevent the magnetic element 40 from deflecting during installation or operation; and third, to avoid the formation of additional magnetic paths that would interfere with the preset magnetic field distribution of the magnetic element 40.

[0041] In this embodiment, the limiting member 50 is an annular structure with an inner diameter smaller than the outer diameter of the magnetic element 40. The limiting member is typically annular, such as a washer or spacer; its inner diameter is smaller than the outer diameter of the magnetic element 40, thus limiting the magnetic element 40 radially inward and preventing it from shifting inward or detaching during operation due to vibration or magnetic force, ensuring that the magnetic element 40 always remains in the preset working position. An annular groove is machined at the junction of the second mounting groove 702 and the first mounting groove 701 for embedding the non-magnetic limiting member 50 made of polyetheretherketone (PEEK).

[0042] In this embodiment, the inner peripheral wall of the second mounting groove 702 is provided with an inwardly protruding limiting protrusion 60 for axially fixing the elastic element 30. The inwardly protruding limiting protrusion 60 is machined on the inner peripheral wall of the second mounting groove 702; this limiting protrusion 60 can be a continuous annular shoulder or multiple protrusions evenly distributed circumferentially; its function is to constrain and position the elastic element 30 placed thereon axially, preventing it from tilting or popping out during pre-compression or operation, and ensuring that the force center of the elastic element 30 is consistent with the load line of the plunger 20.

[0043] In this embodiment, the magnetic element 40 includes at least two permanent magnets spaced apart from each other and arranged with their same poles facing each other. The magnetic element 40 includes a pair of permanent magnets that are spaced apart from each other and arranged with their same poles facing each other, i.e., two N poles facing each other or two S poles facing each other, so that a repulsive magnetic force is generated between the two magnets. This repulsive force increases nonlinearly as the distance decreases, thereby providing the nonlinear negative restoring force required to achieve the negative stiffness characteristics.

[0044] In this embodiment, the outer diameter of the permanent magnet is smaller than the inner diameter of the first mounting groove 701, but larger than the inner diameter of the limiting member 50. The permanent magnet is made of high-performance neodymium iron boron material, machined into a thin ring shape, and its surface is nickel-plated to improve corrosion resistance. Its outer diameter is 2.9 mm, slightly smaller than the outer diameter of the first step inside the ball socket of the slipper 10.

[0045] In this embodiment, the elastic element 30 is a butterfly spring, which is snapped into the limiting boss of the second mounting groove 702. The elastic element 30 is not limited to a specific type and can be various elastic bodies that provide positive stiffness, such as Bass springs, wave springs, or helical springs. By selecting different elastic elements 30, the magnitude of positive stiffness, linearity, and load capacity can be adjusted, thereby providing flexibility for matching with magnetic elements 40 of different specifications to achieve quasi-zero stiffness characteristics.

[0046] The Bausch disc spring is a standard bowl-shaped, unsupported disc spring with a base diameter of 8mm, a top diameter of 4mm, a thickness of 0.5mm, and a standard stroke of 0.3mm. Two disc springs are used, placed back-to-back in the second mounting groove 702 of the ball socket in the slip shoe 10. The limiting boss in the second mounting groove 702 axially limits the Bausch disc spring and prevents radial movement.

[0047] In this embodiment, a solid lubricating coating is provided on the friction surfaces of the relative rotation between the slipper 10 and the plunger 20. This coating, which can be a graphite-based or polymer material coating, is applied to the friction pair formed between the inner surface of the ball joint of the slipper 10 body and the outer surface of the ball head of the plunger 20. Its function is to reduce the coefficient of friction between the ball head and the ball joint, reducing stick-slip and wear, thereby achieving vibration reduction while ensuring a long service life and stable working performance for the slipper 10 assembly.

[0048] In this embodiment, the contact surface between the slipper 10 and the ball head of the plunger 20, as well as the surface of the ball head of the plunger 20 itself, are treated with an ultra-hard, low-friction graphite coating.

[0049] The Bainckian disc spring provides positive stiffness in the compression phase, primarily supporting the foundation and bearing static or average loads to prevent system instability. The permanent magnet provides negative stiffness in the repulsive phase. Its function is twofold: when the external load increases rapidly, the repulsive force of the magnetic ring decreases, dynamically offsetting part of the load increase; when the external load decreases rapidly, the repulsive force of the magnetic ring increases, dynamically offsetting part of the load decrease. The Bainckian disc spring provides positive stiffness k1 under compression conditions, and the permanent magnet provides negative stiffness k2 under repulsive conditions. The positive and negative stiffnesses approximately cancel each other out under the working load, achieving quasi-zero stiffness characteristics during system operation, converting vibrational energy into the elastic potential energy of the Bainckian disc spring and the potential energy of the spacing change of the permanent magnet.

[0050] The present invention provides a magnetoelastic coupling quasi-zero stiffness vibration damping slipper 10, the vibration damping principle of which is as follows:

[0051] When the axial piston 20 pump is running, the mechanism connected in series between the ball socket of the slipper 10 and the ball head of the piston 20 enters a quasi-zero stiffness working state. At this time, the Bainstein disc spring is compressed under load and generates a positive support force, while the permanent magnet with the same pole opposite contributes a negative stiffness effect due to magnetic repulsion. The two are precisely matched to achieve force balance. The externally input vibration energy drives the slipper 10 to produce a significant micro-displacement. Part of the vibration energy is converted into the elastic potential energy of the Bainstein disc spring and the potential energy of the distance change of the permanent magnet. The remaining energy is efficiently dissipated into heat energy through the micro-motion friction of the graphite coating on the bottom surface of the slipper 10.

[0052] like Figure 4 The diagram shown is a force analysis diagram of the plunger 20 of the present invention. The force conditions of the plunger 20 during the oil suction stroke and the oil discharge stroke are different. Therefore, force analysis is performed separately for the two working strokes.

[0053] The suction stroke of the axial piston 20 pump is generally from the top dead center (where the axial displacement of the piston 20 is the maximum) to the bottom dead center (where the axial displacement of the piston 20 is the minimum), at which point the piston 20 completes the suction action. At this time, the piston 20 is dragged out of the cylinder by the central force spring through the pressure plate and the slipper 10, creating a negative pressure in its bottom cavity to draw in oil. The resultant inertial force of a single piston 20 and slipper 10... for ,in and The masses of plunger 20 and slipper 10 are respectively. Let be the relative acceleration of the i-th plunger 20. The maximum value of the total inertial force between the plunger 20 and the slipper 10. for:

[0054] in, The total mass of plunger 20 and slipper 10, The radius of the distribution circle for plunger 20 is given. Let ω be the angular velocity of the plunger 20 and the slipper. The swashplate inclination angle. This is the inertial force coefficient, whose value is related to the number z of the 20 plungers, as shown in the table below:

[0055] Table of plunger number coefficient

[0056] z 5 7 8 11 13 15 <![CDATA[C i ]]> 1.62 2.25 2.88 3.51 4.15 4.78

[0057] As can be seen from the above, the amplitude of the inertial force changes periodically with the crankshaft angle and is significantly affected by the number of plungers 20 z. When the phase superposition of the inertial forces of multiple plungers 20 in the pump is not coordinated, a low-frequency excitation force will be formed, which will induce the overall vibration of the cylinder block.

[0058] The suction force required for a single plunger 20 to draw in oil for:

[0059]

[0060] in, The plunger has an outer diameter of 20 mm. This refers to the vacuum level in the hydraulic pump's suction line.

[0061] If we assume that the number of plungers connected to the intake oil chamber is (z+1) / 2, the total intake force is:

[0062]

[0063] The oil discharge stroke is the process by which plunger 20 moves from the bottom dead center to the top dead center position, expelling the drawn-in oil through the oil discharge window. The working resistance of plunger 20 is then... for:

[0064]

[0065] in, The rated output pressure of the hydraulic pump. The return spring force of each plunger 20.

[0066] As can be seen from the above two equations, when the plunger 20 suddenly changes from the oil suction stroke to the oil discharge stroke, the pressure inside the cavity rises sharply, forming a hydraulic water hammer effect, which leads to increased vibration and wear of the plunger 20 ball head.

[0067] During the oil discharge stroke, from a purely mechanical perspective, the frictional force between the cylinder bore and the plunger 20 is related to its radial reaction force. The resultant forces N1 and N2 of the reaction force are located at positions L1 / 4 and L2 / 4 from the contact edge, respectively. and They are respectively:

[0068]

[0069]

[0070] in, is the coefficient of friction between the cylinder bore and the plunger 20.

[0071] As can be seen from the above, during the oil discharge stroke, the plunger 20 is subjected to asymmetrical support reaction force. When the pressure changes abruptly, the frictional resistance it experiences will change drastically, causing the plunger 20 to move radially.

[0072] In summary, the fretting and vibration caused by the force imbalance of the plunger 20 will be directly transmitted and amplified through the ball joint of the plunger 20 and the slide shoe 10. This not only increases the risk of fatigue fracture of the plunger 20 itself, but also disturbs the posture of the slide shoe 10, causing uneven wear on the bottom surface of the slide shoe 10, and ultimately inducing systemic vibration of the pump body.

[0073] like Figure 5 As shown, when the axial plunger 20 pump is running, the dynamic axial load caused by the pressure pulsation in the plunger 20 chamber acts on the plunger 20 pair of the slide shoe 10, triggering the coordinated response of the magnetoelastic coupling quasi-zero stiffness mechanism: during the instantaneous increase of load, the ball head of the plunger 20 pushes the slide shoe 10 to move axially, forcing the Bausch mandrel to compress and generate a linearly increasing positive stiffness support force. At the same time, the spacing between the same pole permanent magnets embedded in the ball socket of the slide shoe 10 decreases, stimulating a nonlinearly increasing magnetic repulsion negative stiffness effect. The dynamic coupling of the two forms a cancellation mechanism, making the increment of the disc spring support force and the increment of the magnetic ring repulsion force approximately zero after superposition, significantly suppressing the downward movement trend of the slide shoe 10. During the instantaneous decrease of load, the Bausch mandrel rebounds and releases the stored elastic potential energy, causing the support force to decrease linearly, while the increase in the spacing between the permanent magnets causes a nonlinear sharp decrease in the magnetic repulsion force. The negative stiffness characteristic makes the magnetic force decay rate lower than that of the mechanical system, thereby dynamically compensating for the load reduction and constraining the upward movement of the slide shoe 10. In this process, most of the vibration energy is converted into a reversible storage system of the elastic potential energy of the Bain disc spring and the magnetic field energy of the permanent magnet. The remaining small portion of energy is efficiently dissipated as heat energy through the micro-motion friction damping mediated by the ultra-hard graphite coating on the contact surface of the slipper 10 and the plunger 20.

[0074] Furthermore, the specific principle for calculating quasi-zero stiffness is as follows:

[0075] The stiffness of a Bainck disc spring under axial load is based on its geometry and material properties. According to Almen-Laszlo theory, the load-deformation relationship of a single disc spring is:

[0076]

[0077] in: It is the axial load of the disc spring, and the unit is Newton (N). It is the elastic modulus, and its unit is Pascal (Pa). Poisson's ratio; This is the thickness of the disc spring, in millimeters (mm). It is the outer diameter of the disc spring, in millimeters (mm). It is the height of the inner truncated cone, in millimeters (mm). It is the axial compression, and the unit is millimeters (mm). It is a geometric coefficient. ,in This refers to the inner diameter of the disc spring, in millimeters (mm).

[0078] For a two-disc spring assembly placed back-to-back, neglecting contact surface friction, its total stiffness is approximately twice that of a single-disc spring, and its total axial load... It is also approximately twice that of a single disc spring.

[0079]

[0080] Stiffness is defined as the derivative of force with respect to displacement, then normal stiffness... The expression is:

[0081]

[0082] For two disc springs arranged back-to-back, the total stiffness for:

[0083]

[0084] Repulsive force between two axially magnetized permanent magnets with opposite poles Calculations were performed using an equivalent magnetic charge model:

[0085]

[0086] in, It is the vacuum permeability, approximately 4π × 10⁻⁶. -7 H / m; It is the magnetization intensity; the magnetization intensity of NdFeB is taken as... A / m; It is the magnetic ring spacing, in millimeters (mm). It is an axial unit vector.

[0087] This example will be simplified as follows:

[0088]

[0089] in, It is remanence; It is the area of ​​the magnetic ring, in square millimeters (mm). 2 ); It is the thickness of the permanent magnet.

[0090] negative stiffness The expression is:

[0091]

[0092] Overall Stiffness of the System for:

[0093]

[0094] To achieve quasi-zero stiffness, then we have The pre-compression of the Bausch disc spring and the initial spacing of the permanent magnet are set according to these conditions.

[0095] The permanent magnet used in this embodiment has an outer diameter of 2.9 mm, an inner diameter of 1 mm, and a thickness of 0.5 mm. Substituting these dimensions into the calculation yields the pre-compression amount of the Bausch mandrel disc spring. =0.106mm, initial spacing of permanent magnets =0.182mm.

[0096] The above embodiments are only for illustrating the technical concept and features of the present invention, and are intended to enable those skilled in the art to understand and implement the present invention. They should not be construed as limiting the scope of protection of the present invention. All equivalent changes or modifications made in accordance with the spirit and essence of the present invention should be covered within the scope of protection of the present invention.

Claims

1. A magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly, characterized in that, include The skate itself has a ball-shaped cavity inside; A plunger, including a ball head rotatably received within the ball socket; as well as A stiffness compensation mechanism, located within the ball socket, includes an elastic element and a magnetic element arranged in parallel. The elastic element provides positive stiffness, and the magnetic element provides negative stiffness. By adjusting the preload of the elastic element and the initial spacing of the magnetic element, the positive stiffness and negative stiffness cancel each other out within the specified working load area, thereby causing the slipper assembly to exhibit quasi-zero stiffness characteristics in that area.

2. The magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly as described in claim 1, characterized in that, The ball socket is provided with a mounting part for installing the stiffness compensation mechanism, and the mounting part includes at least two mounting slots that are radially distributed and concentric with each other.

3. The magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly as described in claim 2, characterized in that, The two mounting slots are a second mounting slot located on the outer side and a first mounting slot located on the inner side. The second mounting slot is located at the center of the ball socket. The magnetic element is located in the first mounting slot, and the elastic element is located in the second mounting slot.

4. The magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly as described in claim 3, characterized in that, The second mounting slot is connected to the first mounting slot by a stepped structure, and the stepped structure is provided with a non-magnetic limiting member.

5. The magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly as described in claim 4, characterized in that, The limiting member is a ring-shaped structure, and its inner diameter is smaller than the outer diameter of the magnetic element.

6. The magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly as described in claim 3, characterized in that, The inner peripheral wall of the second mounting groove is provided with an inwardly protruding limiting protrusion for axially fixing the elastic element.

7. The magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly as described in claim 5, characterized in that, The magnetic element includes at least two permanent magnets that are spaced apart from each other and arranged with the same poles facing each other.

8. The magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly as described in claim 7, characterized in that, The outer diameter of the permanent magnet is smaller than the inner diameter of the first mounting groove, but larger than the inner diameter of the limiting member.

9. A magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly as described in claim 6, characterized in that, The elastic element is a butterfly spring, which is engaged with the limiting protrusion of the second mounting groove.

10. A magnetoelastic coupling quasi-zero stiffness vibration damping slipper assembly as described in claim 1, characterized in that, The friction surfaces of the slipper and the plunger that rotate relative to each other are provided with a solid lubricating coating.

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