Variable load quasi-zero stiffness vibration isolator based on electromagnetic actuator

Abstract

The invention provides a variable load quasi-zero stiffness vibration isolator based on an electromagnetic actuator, and belongs to the field of low-frequency vibration isolation. Comprising a supporting structure composed of a shell and a bearing platform, a positive stiffness element, a negative stiffness element and an electromagnetic actuator, the positive stiffness element, the negative stiffness element and the electromagnetic actuator are coaxially installed in the supporting structure, load gravity on the bearing platform is detected in real time through a force sensor, and output force of the electromagnetic actuator is dynamically adjusted through a control circuit according to load changes. And the rigidity of the system at the balance position is zero so as to complete self-adaptive variable load compensation. According to the parallel structure, the dynamic stiffness is remarkably reduced while high static stiffness is guaranteed, and the problems that a traditional quasi-zero stiffness vibration isolator is non-adjustable in load and narrow in zero stiffness interval are solved.

Description

Technical Field

[0001] The present invention belongs to the field of low-frequency vibration isolation, and particularly relates to a variable-load quasi-zero stiffness vibration isolator based on an electromagnetic actuator. Background Technique

[0002] Vibration isolation design is an important guarantee for the stable operation of on-vehicle and on-airborne precision instruments. Traditional linear vibration isolation systems can effectively isolate medium and high-frequency vibrations, while low-frequency vibration isolation is a difficult problem in the field of vibration control. According to classical vibration isolation theory, the vibration isolation performance of a vibration isolation system can only be exerted when the excitation frequency is greater than times the natural frequency of the system. For a linear system, under the condition of determining the vibration isolation mass, in order to reduce the natural frequency of the system, the system stiffness must be reduced. However, the reduction of stiffness leads to an increase in the static deformation of the system, which is an inherent contradiction of the linear vibration isolation system. The development of nonlinear vibration isolation theory provides an effective way to solve this contradiction. In recent years, some nonlinear vibration dampers have emerged. Among them, the quasi-zero stiffness vibration isolation system can not only meet the requirements of low-frequency vibration isolation, but also does not require input energy, has a strong load-bearing capacity, and has excellent development prospects.

[0003] At present, the zero stiffness interval of the existing quasi-zero stiffness vibration isolator is small, and it only has zero stiffness near the equilibrium position. When the amplitude is large, the stiffness of the vibration isolator will increase sharply, which seriously weakens the vibration isolation performance of the vibration isolator. And traditional quasi-zero stiffness vibration isolators are all designed according to specific load parameters. Due to the characteristics of high static and low dynamic, they are very sensitive to changes in load, and once they deviate from the equilibrium position, they no longer have the zero stiffness characteristic, and the vibration isolation effect will be significantly reduced. This greatly limits the engineering application of quasi-zero stiffness vibration isolators.

[0004] The zero stiffness range of traditional quasi-zero stiffness vibration isolators is small. When the amplitude is large, the stiffness of traditional vibration isolators will increase sharply, resulting in a decrease in the vibration damping effect of the vibration isolator. Moreover, traditional quasi-zero stiffness vibration isolators are only applicable to fixed loads. After the load changes, the vibration isolator cannot meet the zero stiffness state, and the vibration isolation effect is greatly reduced.

[0005] In order to solve the problems of narrow zero stiffness range and unadjustable load of the existing quasi-zero stiffness vibration isolator, make it have a wider zero stiffness interval and have the ability to adapt to variable mass, and be more in line with engineering practice, this paper designs a variable-load quasi-zero stiffness vibration isolator based on an electromagnetic actuator, uses a disk spring after topological optimization to broaden the zero stiffness range of the vibration isolator, and eliminates the influence of mass change by controlling the active force of the electromagnetic actuator. Summary of the Invention

[0006] Technical Problems to be Solved:

[0007] To avoid the deficiencies of the prior art, the present invention provides a variable-load quasi-zero stiffness isolator based on an electromagnetic actuator, which adopts a combination of a disc spring (negative stiffness) and an annular rubber spring (positive stiffness) in parallel, and combines upper and lower electromagnetic actuators to actively adjust the output force; the force sensor is used to detect the load change in real time, and the current of the electromagnetic actuator is dynamically adjusted to form a closed-loop feedback control to ensure that the system maintains quasi-zero stiffness under variable load. The parallel structure of the present invention significantly reduces the dynamic stiffness while ensuring high static stiffness, and solves the problems of non-adjustable load and narrow zero-stiffness interval of traditional quasi-zero stiffness isolators.

[0008] The technical solution of the present invention is: a variable-load quasi-zero stiffness isolator based on an electromagnetic actuator, which includes a support structure composed of a housing and a bearing platform, and a positive stiffness element, a negative stiffness element, and an electromagnetic actuator coaxially installed in the support structure. The force sensor is used to detect the load gravity on the bearing platform in real time, and the control circuit dynamically adjusts the output force of the electromagnetic actuator according to the load change, so that the stiffness of the system is zero at the equilibrium position to complete the adaptive variable-load compensation;

[0009] The housing is a cavity with an open top, and its open top end is elastically connected to the bearing platform coaxially through a negative stiffness element; the bearing platform is used to support the equipment to be isolated;

[0010] The negative stiffness element is a disc spring, the central ring of the disc spring is coaxially connected below the bearing platform, its outer ring is coaxially installed at the open top of the housing, and its axial stiffness has a non-linear negative characteristic with displacement;

[0011] The positive stiffness element is an annular rubber spring, the top is connected to the bearing platform through an annular gland, and the bottom is connected to the inner bottom surface of the housing to provide linear positive stiffness;

[0012] The electromagnetic actuator includes an upper electromagnetic actuator and a lower electromagnetic actuator, both of which are composed of a metal cylinder wound with wires and are located on the central axis of the annular rubber spring; the upper electromagnetic actuator is rigidly connected to the bottom of the bearing platform, and the lower electromagnetic actuator is fixedly connected to the inner bottom surface of the housing.

[0013] A further technical solution of the present invention is: the control circuit includes a force sensor, an A / D conversion module, a controller, and a power conversion module connected in sequence. The force sensor is installed on the bearing platform and is used to obtain the load gravity detection value in real time; the current load gravity of the bearing platform is obtained through the force sensor and compared with the preset load at the equilibrium position, and the difference is calculated; according to the current of the electromagnetic actuator is adjusted to make the output force compensate for the load change and maintain the zero stiffness of the system at the equilibrium position.

[0014] A further technical solution of the present invention is that a plurality of ring fixators are evenly distributed along the circumferential direction on the outer edge of the bottom surface of the bearing platform, and a plurality of cylinders are evenly distributed along the circumferential direction at the open end of the outer shell. The ring fixators and the cylinder protrusions are arranged in a one-to-one correspondence coaxially. The inner diameter of the ring fixator is larger than the outer diameter of the cylinder protrusion. After the two are sleeved, they can slide relative to each other along the axial direction to form a nested connection.

[0015] A further technical solution of the present invention is that the top end of the bearing platform is a disc, and the part extending from the bottom of the disc into the shell body is a stepped shaft. The central ring of the disc spring is installed on the stepped surface; the end of the stepped shaft is rigidly connected to the metal cylinder of the upper electromagnetic actuator; and a first gland is coaxially fixed at the connection between the stepped shaft and the metal cylinder.

[0016] A further technical solution of the present invention is that the upper and lower ends of the circular rubber spring are respectively encapsulated with an upper ring gland and a lower ring gland, and both the upper ring gland and the lower ring gland are provided with central through holes for passing through the metal cylinders of the upper electromagnetic actuator and the lower electromagnetic actuator; the circular rubber spring, the upper ring gland and the lower ring gland are connected into one body by vulcanization.

[0017] A further technical solution of the present invention is that the stiffness calculation formula of the circular rubber spring is:

[0018]

[0019] where d1 and d2 are the inner and outer diameters of the circular rubber spring respectively, h is the axial height of the circular rubber spring, and E a =iG is the apparent elastic modulus of the compressed rubber, i = 3.6(1 + 1.65S 2 ) is the influence coefficient of the ring geometry, where

[0020] A further technical solution of the present invention is that there is a height difference between the central ring and the outer ring of the disc spring, and fifteen metal spokes are evenly distributed between the central ring and the outer ring; its stiffness calculation formula is:

[0021] k = A(1.5Y 2 - 3BY + B 2 + 1)

[0022] where E represents the elastic modulus, μ represents the Poisson's ratio, t represents the thickness of the disc spring, K1 represents the calculation coefficient, D represents the outer diameter of the disc spring, h0 represents the calculated value of the deformation amount when the disc spring is flattened, and y represents the deformation amount when the disc spring is flattened.

[0023] A further technical solution of the present invention is that the bottom of the lower electromagnetic actuator is coaxially fixed to the second gland, the second gland is coaxially fixed to the inner bottom surface of the housing, and its upper surface is in fitting connection with the lower circular gland.

[0024] A further technical solution of the present invention is that the output force expression of the electromagnetic actuator is as follows:

[0025]

[0026] Wherein, A1 is the effective area of the electromagnet pole face, μ0 is the magnetic permeability in vacuum, and B1 is the magnetic induction intensity in the working area of the electromagnet; In the formula, μ is the magnetic permeability, which is equal to the magnetic permeability of the metal core; N is the number of turns of the coil; I is the current; and L is the magnetic path length.

[0027] A control method for a variable-load quasi-zero stiffness vibration isolator based on an electromagnetic actuator specifically comprises the following steps:

[0028] Calibrate the equilibrium position of the system under a 1500N load through a static test calibration system;

[0029] Real-time detect the gravity of the equipment to be vibration-isolated and calculate the difference from the equilibrium load;

[0030] Dynamically adjust the current of the electromagnetic actuator according to the difference, so that the stiffness of the system approaches zero near the equilibrium position, and realize low-frequency broadband vibration isolation.

[0031] Beneficial effects

[0032] The beneficial effects of the present invention are as follows: A variable-load quasi-zero stiffness vibration isolator based on an electromagnetic actuator of the present invention automatically adjusts the magnitude of the current controlling the electromagnetic actuator when the load-bearing weight changes, so that the electromagnetic actuator outputs a corresponding compensating force, thereby making the stiffness of the system zero at the equilibrium position. When the vibration-isolated object makes small-amplitude vibrations near the equilibrium position, its dynamic stiffness is very small, the natural frequency of the entire system is very low, large-range frequency vibration isolation can be realized, and good low-frequency vibration isolation effect is achieved. Compared with the existing quasi-zero stiffness vibration isolators, it has the advantages of wide zero-stiffness region, adjustable load, large load-bearing capacity, small volume, compact structure, light weight, and convenient control. The specific advantage analysis is as follows:

[0033] 1. Through the parallel design of the disc spring (negative stiffness) and the circular rubber spring (positive stiffness), the present invention realizes quasi-zero stiffness near the equilibrium position, and the natural frequency of the system is greatly reduced. Experimental data shows that the acceleration transmissibility of the vibration isolation system is reduced from 1.86 of the traditional linear vibration isolator to 0.28 (reference comparison case), effectively isolating low-frequency (below 3Hz) and even ultra-low frequency vibrations, and solving the problem of poor vibration isolation effect of traditional vibration isolators in the low-frequency band.

[0034] 2. The present invention uses an electromagnetic actuator combined with closed-loop feedback control. The load gravity is detected in real time through a force sensor, and the compensation force of the electromagnetic actuator is dynamically adjusted. When the load varies within the range of 500N to 2500N, the system can automatically adjust the current (within the range of ±2A), so that the vibration isolator always maintains a quasi-zero stiffness state, avoiding vibration isolation failure caused by load fluctuations. Compared with traditional quasi-zero stiffness vibration isolators (only applicable to fixed loads), the adaptability is increased by more than 300%.

[0035] 3. The optimized disc spring of the present invention adopts a topological spoke structure (15 metal spokes), and the thickness is reduced by 30% while ensuring the negative stiffness characteristics; the circular rubber spring is integrally formed by vulcanization and integrated with the electromagnetic actuator in the shell. The overall structural weight is reduced by 25% compared with similar electromagnetic vibration isolators, and the volume is reduced by 40%, which is applicable to space-limited scenarios such as vehicle-mounted and airborne.

[0036] 4. The closed-loop control algorithm of the present invention cancels the non-linear increase in stiffness caused by the increase in displacement by adjusting the electromagnetic force in real time. The simulation results show that within the displacement range of ±23mm, the dynamic stiffness of the system is always lower than 10N / mm (the stiffness of the traditional passive vibration isolator suddenly increases to 50N / mm at ±10mm), effectively avoiding amplitude jump and system instability problems, and the vibration isolation frequency band is broadened to 0.5Hz to 50Hz.

[0037] 5. Based on the force-current mapping model (the output force expression of the electromagnetic actuator), the control circuit can complete the load compensation in only 0.1 second, and the steady-state error is less than 2%. The operating current of the electromagnetic actuator near the equilibrium position is as low as 0.5A (peak current 2A), and the power consumption is reduced by 60% compared with the traditional electromagnetic vibration isolation system.

[0038] 6. This technology can be widely applied to fields such as vibration isolation of precision instruments (such as semiconductor lithography machines), vehicle-mounted medical equipment (such as mobile CT machines), and aerospace load platforms. For example, when applied in a neonatal transport vehicle, the vibration transfer rate of the baby nursery is reduced by 85%, significantly reducing the damage of mechanical vibration to sensitive equipment. Description of the Drawings

[0039] Figure 1 It is a schematic diagram of the overall structure of a variable-load quasi-zero stiffness vibration isolator based on an electromagnetic actuator in an embodiment of the present invention;

[0040] Figure 2 It is a schematic diagram of the disc spring structure in an embodiment of the present invention;

[0041] Figure 3 It is a schematic diagram of the circular rubber structure in an embodiment of the present invention;

[0042] Figure 4 It is a schematic diagram of the upper and lower electromagnetic actuator structures in an embodiment of the present invention;

[0043] Figure 5 are the force-displacement curve and stiffness curve of the disc spring of the present invention;

[0044] Figure 6 are the force-displacement curve and stiffness curve of the circular rubber ring in the embodiment of the present invention;

[0045] Figure 7 are the force-displacement curve and stiffness curve of the quasi-zero stiffness system in the embodiment of the present invention;

[0046] Figure 8 are the amplitude-frequency curves of the quasi-zero stiffness system and the linear system in the embodiment of the present invention;

[0047] Figure 9 is the structural block diagram of the electromagnetic actuator control circuit in the embodiment of the present invention;

[0048] Explanation of reference numerals: 1, bearing platform; 2, circular ring fixer; 3, cylindrical protrusion; 4, first gland; 5, housing; 6, second gland; 7, metal cylinder; 8, lower circular ring gland; 9, circular rubber spring; 10, upper circular ring gland; 11, disc spring; 12, wire; 13, wire; 14, metal cylinder. Detailed implementation manners

[0049] The embodiments described below with reference to the accompanying drawings are exemplary and are intended to explain the present invention, and should not be construed as limiting the present invention.

[0050] In the description of the present invention, it should be understood that the orientation or positional relationship indicated by the terms "center", "longitudinal", "transverse", "length", "width", "thickness", "upper", "lower", "front", "rear", "left", "right", "vertical", "horizontal", "top", "bottom", "inner", "outer", "clockwise", "counterclockwise", etc. are based on the orientation or positional relationship shown in the accompanying drawings, and are only for the convenience of describing the present invention and simplifying the description, rather than indicating or implying that the device or element referred to must have a specific orientation, be constructed and operated in a specific orientation, and thus should not be construed as limiting the present invention.

[0051] Based on the problems of narrow zero-stiffness range and non-adjustable load of the existing quasi-zero stiffness isolator, the present invention proposes a variable-load quasi-zero stiffness isolator based on an electromagnetic actuator, which includes a support structure composed of a housing and a bearing platform, and a positive stiffness element, a negative stiffness element, and an electromagnetic actuator coaxially installed in the support structure. The load gravity on the bearing platform is detected in real time by a force sensor, and the output force of the electromagnetic actuator is dynamically adjusted according to the load change through a control circuit, so that the stiffness of the system is zero at the equilibrium position to complete adaptive variable-load compensation;

[0052] Specifically, the housing is a cavity with an open top, and the open end of its top is elastically connected to the bearing platform coaxially through a negative stiffness element; the bearing platform is used to support the equipment to be vibration-isolated.

[0053] Specifically, the negative stiffness element is a disc spring. The central ring of the disc spring is coaxially connected below the bearing platform, and its outer ring is coaxially installed at the open top of the housing. Its axial stiffness shows a non-linear negative characteristic with displacement.

[0054] Specifically, the positive stiffness element is an annular rubber spring, which is connected to the bearing platform through an annular gland at the top and connected to the inner bottom surface of the housing at the bottom, providing linear positive stiffness.

[0055] Specifically, the electromagnetic actuator includes an upper electromagnetic actuator and a lower electromagnetic actuator, both of which are composed of a wire wound around a metal cylinder and are located on the central axis of the annular rubber spring; the upper electromagnetic actuator is rigidly connected to the bottom of the bearing platform, and the lower electromagnetic actuator is fixedly connected to the inner bottom surface of the housing.

[0056] The above technical solutions will be further described below in conjunction with examples and drawings:

[0057] In one embodiment, as shown in reference to Figure 1 An electromagnetic actuator-based variable-load quasi-zero stiffness vibration isolator in Embodiment 1 includes a bearing platform 1, a disc spring 11, an annular rubber spring 9, upper and lower electromagnetic actuators, and a housing 5. The middle of the bearing platform 1 is connected to the disc spring 11 with negative stiffness characteristics. The upper and lower parts of the axially annular rubber 9 with positive stiffness are respectively connected to the annular gland 10 and the annular gland 8. The gland 4 on the bearing platform 1 is connected to the gland 10. The wire 12 is wound around the metal cylinder 14, and the two form the upper electromagnetic actuator. The wire 13 is wound around the metal cylinder 7, and the two form the lower electromagnetic actuator. A force sensor is installed on the bearing platform 1. The input end of the control circuit is connected to the force sensor, and the output end of the control circuit is connected to the electromagnetic actuator. The disc spring 11 is connected to the upper part of the housing 5, and the gland 6 is connected to the bottom of the housing 5.

[0058] Preferably, the control circuit includes a force sensor, an A / D conversion module, a controller, and a power conversion module connected in sequence. The force sensor is installed on the bearing platform and is used to obtain the load gravity detection value in real time; obtain the current load gravity of the bearing platform through the force sensor, compare it with the load at the preset equilibrium position, and calculate the difference; adjust the current of the electromagnetic actuator according to the difference to make the output force compensate for the load change and maintain the zero stiffness of the system at the equilibrium position.

[0059] Preferably, a plurality of ring fixators are circumferentially and evenly distributed along the outer edge of the bottom surface of the bearing platform, and a plurality of cylinders are circumferentially and evenly distributed along the opening end of the outer shell. The ring fixators and the cylinder protrusions are arranged in a one-to-one correspondence coaxially. The inner diameter of the ring fixator is greater than the outer diameter of the cylinder protrusion. After the two are sleeved, they can slide relative to each other along the axial direction to form a nested connection.

[0060] Preferably, the top end of the bearing platform is a disc, and the part of the disc bottom extending into the shell body is a stepped shaft. The central ring of the disc spring is installed on the stepped surface; the end of the stepped shaft is rigidly connected to the metal cylinder of the upper electromagnetic actuator; and a first gland is coaxially fixed at the connection of the stepped shaft and the metal cylinder.

[0061] Preferably, the upper and lower ends of the annular rubber spring are respectively encapsulated with an upper ring gland and a lower ring gland, and both the upper ring gland and the lower ring gland are provided with central through holes for passing through the metal cylinders of the upper electromagnetic actuator and the lower electromagnetic actuator; the annular rubber spring, the upper ring gland and the lower ring gland are connected into one body by vulcanization.

[0062] Make a detailed analysis of the technical principle of the present invention.

[0063] The mechanism for generating the quasi-zero stiffness system: When a negative stiffness elastic element is connected in parallel with a positive stiffness elastic element, and the magnitudes of the positive and negative stiffness provided by the two are the same (equilibrium position), zero stiffness can be generated. Here, the zero stiffness refers to the dynamic stiffness. When the vibration isolation equipment makes small-amplitude vibrations near the equilibrium position, its dynamic stiffness is very small, and the natural frequency of the entire system is very low, so as to achieve vibration isolation over a wide range of frequencies and have good low-frequency vibration isolation effects.

[0064] Perform topological optimization design on the disc spring. The web in the middle of the simplified disc spring becomes a metal wire. The metal wires are arranged in a circular ring in sequence and fixed to the upper and lower bearing platforms to form an elastic element similar to a disc spring. Figure 2 , as the vertical downward load is applied, the structural inclined cone surface gradually tends to the horizontal position, that is, the equilibrium position, and at this time, negative stiffness is generated. The negative stiffness element in the invention is a structure similar to a disc spring. According to the theory of elasticity, the relationship between the axial load F and the displacement deformation y of the disc spring is:

[0065]

[0066] In the formula: F——the load of the disc spring, N;

[0067] t——the thickness of the disc spring, mm;

[0068] y——the deformation amount when the disc spring is flattened, mm;

[0069] E——the elastic modulus, MPa;

[0070] μ——the Poisson's ratio;

[0071] D—the outer diameter of the disc spring, mm;

[0072] K1—calculation coefficient;

[0073] h0—the calculated deformation value when the disc spring is flattened, mm;

[0074] Derive the derivative of the displacement of formula (1) to obtain:

[0075]

[0076] Let Obtain the expression of the stiffness k of the disc spring structure:

[0077] k = A(1.5Y 2 - 3BY + B 2 + 1)(3)

[0078] It can be seen from the above formula that when the disc spring stiffness k has a negative stiffness characteristic.

[0079] The circular rubber spring provides positive stiffness. As Figure 3 shown, when the relative deformation of the rubber is less than 15%, the deformation of the rubber spring conforms to Hooke's law, and the stiffness calculation formula of the circular rubber spring is:

[0080]

[0081] where d1 and d2 are the inner and outer diameters of the circular rubber spring respectively, h is the spring height, and E a = iG is the apparent elastic modulus of the compressed rubber, and i = 3.6(1 + 1.65S 2 ) is the influence coefficient of the circular ring geometry, where

[0082] At the equilibrium position, the negative stiffness of the disc spring is the same as the positive stiffness of the rubber spring. Obtain the minimum stiffness of the disc spring through ANSYS simulation or experiment, and then determine the parameters of the rubber spring according to the calculation formula of the rubber spring. The two are combined in parallel as a quasi-zero stiffness system.

[0083] The electromagnetic actuator is a gravity compensation structure in the system. As Figure 3 shown, when the system load changes, the electromagnetic actuator will output a compensation force to restore the system to the equilibrium position. The principle of the actuator is approximately an electromagnet with a wire wound around a metal:

[0084]

[0085] Wherein, A1 is the effective area of the electromagnet pole face, μ0 is the magnetic permeability in vacuum, and B1 is the magnetic induction intensity in the working area of the electromagnet:

[0086]

[0087] Wherein, μ is the magnetic permeability, equal to the magnetic permeability of the metal core; N is the number of turns of the coil; I is the current; and L is the magnetic path length.

[0088] Referring to Figure 2 as shown, the disc spring is subjected to simulation or static experiment, the force-displacement curve of the disc spring is analyzed, and the stiffness curve of the structure is obtained by differentiation as Figure 5 , it can be seen that the disc spring has the negative stiffness characteristic, and the minimum stiffness is about -245 N / mm. According to the minimum stiffness of the disc spring and formula (4), the size of the annular rubber spring is calculated, as Figure 3 shown. And the force-displacement curve and stiffness of the rubber spring are verified according to the static experiment or simulation as Figure 6 . The disc spring and the annular rubber spring are connected in parallel to form a quasi-zero stiffness structure, and the force-displacement curve and stiffness curve of the quasi-zero stiffness system are analyzed and obtained, as Figure 7 , it can be seen that at the equilibrium position, the stiffness of the system is zero, which indicates that the structure has the quasi-zero stiffness characteristic.

[0089] In order to keep the quasi-zero stiffness system always in the equilibrium position, the electromagnetic actuator shown in the figure is used for displacement compensation. When the weight of the vibration-isolated object changes, the gravity sensor will measure the load gravity and compare it with the force at the equilibrium position. The control module of the system, such as Figure 9 , will adjust the magnitude and direction of the current according to the change of gravity to compensate for the changed gravity. The relationship between the current and the compensation force is given by formula (5) and formula (6). When the object is in the equilibrium state, the amplitude-frequency curves of the quasi-zero stiffness system and the linear system are as Figure 8 , it can be seen that the resonance peak of the quasi-zero stiffness system moves forward compared with the linear system. It can be said that the invention can effectively broaden the vibration isolation frequency band and realize low-frequency vibration isolation.

[0090] Although the embodiments of the present invention have been shown and described above, it can be understood that the above embodiments are exemplary and should not be construed as limiting the present invention. Those of ordinary skill in the art can make changes, modifications, substitutions, and variations to the above embodiments within the scope of the present invention without departing from the principle and purpose of the present invention.

Claims

1. A variable-load quasi-zero stiffness isolator based on an electromagnetic actuator, characterized in that: It includes a support structure composed of a shell and a load-bearing platform, and a positive stiffness element, a negative stiffness element, and an electromagnetic actuator coaxially installed in the support structure. The load gravity on the load-bearing platform is detected in real time by a force sensor, and the output force of the electromagnetic actuator is dynamically adjusted according to the load change by a control circuit, so that the stiffness of the system is zero at the equilibrium position to complete adaptive variable load compensation; The housing is a cavity with an opening at the top, and the opening end at the top is coaxially elastically connected to the bearing platform through a negative stiffness element; the bearing platform is used to support the vibration-isolated equipment; The negative stiffness element is a disc spring, the central ring of the disc spring is coaxially connected to the bottom of the bearing platform, and the outer ring is coaxially installed at the top opening of the shell, and the axial stiffness thereof presents a nonlinear negative characteristic with displacement; The positive stiffness element is a circular rubber spring, the top of which is connected to the bearing platform through a circular gland, and the bottom is connected to the inner bottom surface of the shell, providing linear positive stiffness; The electromagnetic actuator comprises an upper electromagnetic actuator and a lower electromagnetic actuator, both of which are composed of wires wrapped around a metal cylinder and are located on the central axis of the annular rubber spring; the upper electromagnetic actuator is rigidly connected to the bottom of the bearing platform, and the lower electromagnetic actuator is fixedly connected to the inner bottom surface of the shell.

2. The variable load quasi-zero stiffness vibration isolator based on an electromagnetic actuator according to claim 1, wherein: The control circuit includes a force sensor, an A / D conversion module, a controller and a power conversion module connected in sequence. The force sensor is installed on the load-bearing platform and is used to obtain the load gravity detection value in real time; the current load gravity of the load-bearing platform is obtained through the force sensor, and compared with the preset balance position load to calculate the difference; according to the current of the electromagnetic actuator, the output force compensates for the load change and maintains the zero stiffness of the system at the balance position.

3. The variable load quasi-zero stiffness vibration isolator based on an electromagnetic actuator according to claim 1, characterized in that: The outer edge of the bottom surface of the supporting platform is evenly distributed with multiple circular ring fixers along the circumferential direction, and the open end of the outer shell is evenly distributed with multiple cylinders along the circumferential direction. The circular ring fixers and cylindrical protrusions are coaxially arranged in a one-to-one correspondence. The inner diameter of the circular ring fixer is larger than the outer diameter of the cylindrical protrusion. After the two are assembled, they can slide relative to each other in the axial direction to form a nested connection.

4. The variable load quasi-zero stiffness vibration isolator based on an electromagnetic actuator according to claim 3, characterized in that: The top of the bearing platform is a disc, and the part extending into the shell from the bottom of the disc is a stepped shaft, and a central ring of a disc spring is installed on the stepped surface; the end of the stepped shaft is rigidly connected to the metal cylinder of the upper electromagnetic actuator; and a first pressure cover is coaxially fixed at the connection between the stepped shaft and the metal cylinder.

5. The variable load quasi-zero stiffness vibration isolator based on an electromagnetic actuator according to claim 1, characterized in that: The upper and lower ends of the annular rubber spring are respectively encapsulated with an upper annular pressure cover and a lower annular pressure cover, and both the upper annular pressure cover and the lower annular pressure cover are provided with a central through hole for passing through the metal cylinders of the upper electromagnetic actuator and the lower electromagnetic actuator; the annular rubber spring, the upper annular pressure cover and the lower annular pressure cover are connected as a whole by vulcanization.

6. The variable load quasi-zero stiffness vibration isolator based on an electromagnetic actuator according to claim 5, characterized in that: The bottom of the lower electromagnetic actuator is coaxially fixed to the second gland, and the second gland is coaxially fixed to the inner bottom surface of the housing, and the upper surface of the second gland is fitted and connected to the lower annular gland.

7. The variable load quasi-zero stiffness vibration isolator based on an electromagnetic actuator according to claim 5, characterized in that: The calculation formula of the annular rubber spring stiffness is: where d1 and d2 are the inner and outer diameters of the annular rubber spring respectively, h is the axial height of the annular rubber spring, and E a = iG is the apparent elastic modulus of the compressed rubber, and i = 3.6(1 + 1.65S 2 ) is the influence coefficient of the annular geometry, where 8. The variable load quasi-zero stiffness vibration isolator based on an electromagnetic actuator according to claim 1, characterized in that: There is a height difference between the central ring and the outer ring of the disc spring, and fifteen metal spokes are evenly distributed between the central ring and the outer ring; the stiffness calculation formula is: k = A(1.5Y 2 - 3BY + B 2 + 1) Among them, E represents the modulus of elasticity, μ represents the Poisson's ratio, t represents the thickness of the disc spring, K1 represents the calculation coefficient, D represents the outer diameter of the disc spring, h0 represents the calculated value of the deformation amount when the disc spring is flattened, and y represents the deformation amount when the disc spring is flattened.

9. The variable-load quasi-zero stiffness vibration isolator based on an electromagnetic actuator according to claim 1, characterized in that: The output force expression of the electromagnetic actuator is as follows: Among them, A1 is the effective area of the electromagnet pole face, μ0 is the magnetic permeability in vacuum, and B1 is the magnetic induction intensity in the working area of the electromagnet; In the formula, μ is the magnetic permeability, which is equal to the magnetic permeability of the metal core; N is the number of turns of the coil; I is the current; and L is the magnetic path length.

10. A control method for a quasi-zero stiffness vibration isolator with adjustable load-bearing based on an electromagnetic actuator according to any one of claims 1-9, characterized in that The specific steps are as follows: The equilibrium position of the system under a load of 1500N is calibrated through a static test; Detect the gravity of the vibration-isolated equipment in real time and calculate the difference from the balanced load; Dynamically adjust the current of the electromagnetic actuator according to the difference, so that the system stiffness approaches zero near the equilibrium position, realizing low-frequency wide-band vibration isolation.

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