An ultra-low frequency quasi-zero stiffness vibration isolator and a debugging method

Abstract

The application discloses an ultra-low frequency quasi-zero stiffness vibration isolator and a debugging method, and the vibration isolator comprises a fixed plate, a stiffness adjusting mechanism and a dynamic mechanism; the fixed plate is connected with the stiffness adjusting mechanism through a support seat; the stiffness adjusting mechanism is connected with the dynamic mechanism through bolts; the stiffness adjusting mechanism is connected with the dynamic mechanism through first inclined rods, second inclined rods, horizontal tension spring units and vertical tension spring units to form an X-shaped full-tension spring frame body, so that the dynamic stiffness of the whole vibration isolator reaches a quasi-zero stiffness state; the horizontal tension spring units are respectively connected with the first inclined rods and the second inclined rods through first bearing pieces; the first inclined rods are hingedly connected with the dynamic mechanism through second bearing pieces; the second inclined rods are connected with the dynamic mechanism through second bearing pieces; one end of the vertical tension spring unit is connected with the horizontal tension spring unit, and the other end of the vertical tension spring unit is connected with the fixed plate; in the case of paying attention to frequency vibration noise below dozens of Hz, the application has the characteristics of high static stiffness and low dynamic stiffness in the system stiffness.

Description

Technical Field

[0001] This invention belongs to the field of vibration isolation and vibration reduction technology, and particularly relates to an ultra-low frequency quasi-zero stiffness vibration isolator and its debugging method. Background technology:

[0002] The high-precision instrument and equipment industry is developing rapidly, with increasingly higher requirements for accuracy. Equipment is also becoming more and more sensitive to environmental conditions such as micro-vibrations. Even a slight micro-vibration can reduce the output yield of the equipment or even cause it to malfunction. Therefore, isolating micro-vibrations is becoming increasingly important.

[0003] In recent years, researchers have discovered novel vibration isolation methods based on nonlinear dynamics theory—quasi-zero stiffness isolators. Micro-vibrations can reduce the quality and yield of products in many precision processes, such as photolithography, wafer inspection, and flat panel display manufacturing. In most precision applications, the primary concern is vibration noise at frequencies of 1–100 Hz or higher. These isolators exhibit high static and low dynamic stiffness characteristics, meaning that when the system is under static load, the isolator has relatively high static stiffness, providing sufficient support; when the system is under dynamic load, the isolator has low dynamic stiffness, achieving low-frequency and ultra-low-frequency vibration isolation.

[0004] However, most nonlinear vibration isolators based on quasi-zero stiffness theory can only reduce vibration of specific controlled objects and cannot meet the requirements for vibration isolation of different objects. Furthermore, existing quasi-zero stiffness vibration isolators have a narrow low-stiffness displacement range, and the stiffness increases significantly with the increase of displacement, resulting in a significant enhancement of nonlinearity. The initial isolation frequency of the vibration isolator increases, the isolation frequency band narrows, and the low-frequency isolation performance decreases. At the same time, vibration isolators designed to achieve quasi-zero stiffness generally occupy a large space, which affects the applicable range of quasi-zero stiffness vibration isolators. Summary of the Invention

[0005] To address the technical problems existing in the prior art, this invention provides an ultra-low frequency quasi-zero stiffness vibration isolator and its adjustment method. This invention employs a full-tension spring main structure to reduce system damping, while also featuring a simple structure, fewer parts, and saving space. It achieves constant-value quasi-zero stiffness characteristics, expanding the low dynamic stiffness range of the vibration isolator. For vibration and noise below tens of Hz, this invention exhibits high static and low dynamic stiffness characteristics; that is, when the system is under static load, the vibration isolator has relatively high static stiffness, providing sufficient support; when the system is under dynamic load, the vibration isolator has low dynamic stiffness, achieving low-frequency and ultra-low-frequency vibration isolation. An ultra-low frequency constant-value quasi-zero stiffness vibration isolator is composed of an isolator body; the isolator includes a fixed plate, a stiffness adjustment mechanism, and a dynamic mechanism; the fixed plate is connected to the stiffness adjustment mechanism via a support base; the stiffness adjustment mechanism is connected to the dynamic mechanism via bolts; wherein:

[0006] The stiffness adjustment mechanism, through a first inclined rod, a second inclined rod, a horizontal tension spring unit, and a vertical tension spring unit, forms an X-shaped full tension spring frame to achieve a near-zero dynamic stiffness state for the entire vibration isolator. The horizontal tension spring unit is connected to the first inclined rod and the second inclined rod respectively through a first bearing component. The first inclined rod is hinged to the dynamic mechanism through a second bearing component. The second inclined rod is connected to the dynamic mechanism through a second bearing component. One end of the vertical tension spring unit is connected to the horizontal tension spring unit, and the other end is connected to the fixed plate.

[0007] Furthermore, the horizontal tension spring unit is composed of a first horizontal tension spring and a second horizontal tension spring; the vertical tension spring unit is composed of a first vertical tension spring and a second vertical tension spring.

[0008] Furthermore, the dynamic mechanism consists of a slide rail and a slider.

[0009] The present invention is also implemented using the following technical solutions:

[0010] A method for adjusting an ultra-low frequency quasi-zero stiffness vibration isolator includes:

[0011] At the static equilibrium position, negative stiffness is generated by the first inclined rod, the second inclined rod, and the horizontal tension spring to counteract the positive stiffness of the vertical tension spring, so that the overall dynamic stiffness of the vibration isolator reaches a state of zero stiffness. When the vibration isolator deviates from the static equilibrium position, the stiffness adjustment mechanism adjusts the dynamic stiffness of the vibration isolator to a near-zero stiffness state according to the following formula:

[0012]

[0013]

[0014]

[0015] Where P1, P2, and P3 are intermediate variables. denoted as dimensionless displacement, x represents the initial state; denoted as displacement starting from point O at the top of the inclined rod, a represents the horizontal projection length of the inclined rod OO in the initial state, and h represents the vertical distance from point O at the top of the inclined rod in the initial state to the static equilibrium position. Let f be the dimensionless force, k2 be the force acting on the top of the inclined rod, k1 be the stiffness of the vertical tension spring, and α be the stiffness ratio. Dimensionless force The derivative with respect to dimensionless displacement; Let be the dimensionless value of 'a'; and let b be the horizontal projection length of the lower half of the inclined rod in the initial state. Let be the dimensionless value of b; δ is the pre-compression length of the horizontal tension spring in the initial state. Let δ be the dimensionless value; y be the displacement from the static equilibrium position, x = y + h, and the dimensionless relationship is: in: Beneficial effects

[0016] This invention employs a full-tension spring main structure, saving space in the vibration isolator and achieving a miniaturized, constant-stiffness quasi-zero-stiffness vibration isolator with fewer parts. Simultaneously, it expands the low-stiffness displacement range of the vibration isolator. When focusing on vibration noise at frequencies below tens of Hz, this invention exhibits high static and low dynamic stiffness characteristics in terms of system stiffness. That is, when the system is subjected to static loads, the vibration isolator has relatively high system stiffness, providing sufficient support; when the system is subjected to dynamic loads, the vibration isolator has low system stiffness, achieving low-frequency and ultra-low-frequency vibration isolation. Attached Figure Description

[0017] Figure 1 This is a schematic diagram of an ultra-low frequency quasi-zero stiffness vibration isolator according to the present invention;

[0018] Figure 2 This is a partial structural diagram of an ultra-low frequency quasi-zero stiffness vibration isolator according to the present invention;

[0019] Figure 3 Mechanical model and parameter labeling for quasi-zero stiffness vibration isolators.

[0020] Figure label:

[0021] 1 is a fixed plate, 2 is a support base, 3 is the first diagonal rod, 4 is the second diagonal rod, 5 is a hinge base, 6 is a bolt, 7 is a slider, 8 is a slide rail, 9 is a support base bearing, 10 is a hinge base bearing, 11 is a vertical tension spring unit, 11a is the first vertical tension spring, 11b is the second vertical tension spring, 12 is a horizontal tension spring unit, 13 is a support base pin, and 14 is a hinge base pin. Detailed Implementation

[0022] The following is in conjunction with the appendix Figure 1-3 The present invention is described as follows:

[0023] like Figure 1 As shown, this invention provides an ultra-low frequency quasi-zero stiffness vibration isolator, which is composed of an isolator body; the isolator includes a fixed plate, a stiffness adjustment mechanism, and a dynamic mechanism; the fixed plate is connected to the stiffness adjustment mechanism via a support base; the stiffness adjustment mechanism is connected to the dynamic mechanism via bolts; the dynamic mechanism consists of a slide rail and a slider; wherein:

[0024] The fixing plate 1 can be fixedly connected to the foundation or mounting base by bolts; the support base 2 is fixedly connected to the fixing plate 1 (which has bolt holes) by bolts. The stiffness adjustment mechanism includes a first inclined rod 3, a second inclined rod 4, a horizontal tension spring unit 12, and a vertical tension spring unit 11; the horizontal tension spring unit 12 is composed of a first horizontal tension spring 12a and a second horizontal tension spring 12b; the vertical tension spring unit 11 is composed of a first vertical tension spring 11a and a second vertical tension spring 11b.

[0025] The horizontal tension spring unit 12 is connected to the first inclined rod 3 and the second inclined rod 4 respectively through a first bearing component; the first bearing component consists of a bearing seat 9 and a pin 13; the circular hole of the support seat 2 houses the support seat bearing 9, and the inner hole of the support seat bearing 9 mates with the support seat pin 13. The support seat pin 13 mates with the inner hole of the bearing 9, and the outer diameter of the bearing 9 mates with the middle hole of the first inclined rod 3, allowing the first inclined rod 3 to rotate around the support seat pin 13. The support seat pin 13 mates with the inner hole of the bearing 9, and the outer diameter of the bearing 9 mates with the middle hole of the second inclined rod 4, allowing the second inclined rod 4 to rotate around the support seat pin 13. The positional relationship between the first inclined rod 3 and the second inclined rod 4 is as follows: Figure 2 As shown.

[0026] The first inclined rod 3 is connected to the dynamic mechanism via a second bearing component; the second inclined rod 4 is also connected to the dynamic mechanism via a second bearing component; the second bearing component comprises a hinge seat bearing 10, a hinge seat pin 14, and a hinge seat 5; one side of the circular hole of the first inclined rod 3 mates with the outer diameter of the hinge seat bearing 10, the inner ring of the hinge seat bearing 10 mates with the hinge seat pin 14, and both ends of the hinge seat pin 14 are respectively connected to one end of the first horizontal tension spring 12a and the second horizontal tension spring 12b (the hooks at the ends of the tension springs are hooked onto the ends of the hinge seat pin 14). The hinge seat pin 14 mates with the inner diameter of the bearing 10, and the outer diameter of the bearing 10 mates with the circular hole of the hinge seat 5. The hinge seat 5 rotates around the hinge seat pin 14, and the first inclined rod 3 rotates around the hinge seat pin 14. The first inclined rod 3 and the hinge seat 5 are connected together via the pin 14 and the bearing 10 and can rotate around the pin. The left end of the first inclined rod 3 is located in the groove of the hinge seat 5.

[0027] The right-side circular hole of the second inclined rod 4 mates with the outer diameter of the hinge seat bearing 10. The inner ring of the hinge seat bearing 10 mates with the hinge seat pin 14. Both ends of the hinge seat pin 14 are connected to the other ends of the first horizontal tension spring 12a and the second horizontal tension spring 12b, respectively (the hooks at the ends of the tension springs are hooked onto the ends of the hinge seat pin 14). The hinge seat pin 14 mates with the inner diameter of the bearing 10, and the outer diameter of the bearing 10 mates with the circular hole of the hinge seat 5. The hinge seat 5 rotates around the hinge seat pin 14, and the second inclined rod 4 rotates around the hinge seat pin 14. The second inclined rod 4 and the hinge seat 5 are connected together by the pin 14 and the bearing 10 and can rotate around the pin. The right end of the second inclined rod 4 is located in the groove of the hinge seat 5. The threaded countersunk hole of the hinge seat 5 is threadedly connected to one end of the double-ended bolt 6, and the other end of the double-ended bolt 6 is connected to the threaded hole of the slider 7. The slider 7 and the slide rail 8 are slidably connected.

[0028] The vertical tension spring unit 11 consists of a first vertical tension spring 11a and a second vertical tension spring 11b. The bottom ends of the first vertical tension spring 11a and the second vertical tension spring 11b are hooked onto the upper ends of the bolts on the fixing plate 1 (the upper ends of the bolts have small holes drilled in them). The upper end of the first vertical tension spring 11a is hooked into the small hole at the right end of the first diagonal rod 3, and the upper end of the second vertical tension spring 11b is hooked into the small hole at the left end of the second side diagonal rod 4. The vertical tension spring unit 11 provides positive stiffness to bear the vibration isolation mass.

[0029] After a suitable vibration isolation mass is placed on the slide rail 8, it reaches a static equilibrium position (the first inclined rod 3 and the second inclined rod 4 are simultaneously in a horizontal state). In the static equilibrium position, the first inclined rod 3, the second inclined rod 4, and the horizontal tension spring unit 12 generate negative stiffness, which cancels out the positive stiffness of the vertical tension spring unit 11, making the overall dynamic stiffness of the vibration isolator reach a state of zero stiffness. When excited, when deviating from the static equilibrium position, the overall dynamic stiffness is in a quasi-zero stiffness state. At the static equilibrium position, the load is entirely borne by the vertical tension spring unit 11. This vibration isolator is a quasi-zero stiffness vibration isolator, which has both high static stiffness and low dynamic stiffness. The low dynamic stiffness is used for low-frequency and ultra-low-frequency vibration isolation, while the high static stiffness is used to bear higher vibration isolation mass.

[0030] The quasi-zero stiffness vibration isolator proposed in this invention and its parameter specifications are as follows: Figure 3 As shown, the force-displacement and stiffness-displacement expressions can be obtained from the force analysis.

[0031]

[0032]

[0033]

[0034] P1, P2, and P3 are intermediate variables. The displacement is dimensionless, and x is the initial state. Figure 2 (As shown) The displacement starting from point O at the top of the inclined rod, a is the projection length of the inclined rod OO' in the horizontal direction in the initial state, and h is the vertical distance from point O at the top of the inclined rod in the initial state to the static equilibrium position. Let f be the dimensionless force, k2 be the stiffness of the vertical tension spring unit 11, k1 be the stiffness of the horizontal tension spring unit 12, and α be the stiffness ratio. Dimensionless force The derivative with respect to dimensionless displacement.

[0035] Let be the dimensionless value of 'a'. Let b be the horizontal projection length of the lower half of the inclined rod in the initial state. Let be the dimensionless value of b. δ is the pre-compression length of the horizontal tension spring in its initial state. Let be the dimensionless value of δ. y is the displacement from the static equilibrium position, and x = y + h, with the dimensionless relation as follows: in

[0036] At the static equilibrium position, where both diagonal rods are horizontal, and with two quasi-zero stiffness conditions—stiffness equal to 0 and the second derivative of stiffness equal to 0—the following parametric relationship is obtained.

[0037]

[0038]

[0039] When the parameters satisfy formulas (4) and (5), formula (1) is a constant force line and formula (2) is a constant zero stiffness; when the parameters only satisfy formula (4), formula (2) is a nonlinear quasi-zero stiffness characteristic; when the parameters only satisfy formula (5), formula (2) is a constant stiffness characteristic.

[0040] To achieve better vibration isolation at low and ultra-low frequencies, the parameters should satisfy formula (5) and simultaneously satisfy the following conditions: (indicating that the value of α is less than) And the α value is close to ).

[0041] Although the present invention has been described above, the present invention is not limited to the specific embodiments described above. The specific embodiments described above are merely illustrative and not restrictive. Those skilled in the art can make many modifications under the guidance of the present invention without departing from the spirit of the present invention, and these modifications are all within the protection scope of the present invention.

Claims

1. An ultra-low frequency quasi-zero stiffness vibration isolator, comprising an isolator body; characterized in that: The vibration isolator includes a fixed plate, a stiffness adjustment mechanism, and a dynamic mechanism; the fixed plate is connected to the stiffness adjustment mechanism via a support base; the stiffness adjustment mechanism is connected to the dynamic mechanism via bolts; wherein: The stiffness adjustment mechanism, through a first inclined rod, a second inclined rod, a horizontal tension spring unit, and a vertical tension spring unit, forms an X-shaped full tension spring frame to achieve a state of zero dynamic stiffness for the entire vibration isolator. The horizontal tension spring unit is connected to the first inclined rod and the second inclined rod respectively through a first bearing component. The first inclined rod is hinged to the dynamic mechanism through a second bearing component. The second inclined rod is connected to the dynamic mechanism through a second bearing component. One end of the vertical tension spring unit is connected to the horizontal tension spring unit, and the other end is connected to the fixed plate.

2. The ultra-low frequency quasi-zero stiffness vibration isolator according to claim 1, characterized in that: The horizontal tension spring unit is composed of a first horizontal tension spring and a second horizontal tension spring; the vertical tension spring unit is composed of a first vertical tension spring and a second vertical tension spring.

3. The ultra-low frequency quasi-zero stiffness vibration isolator according to claim 1, characterized in that: The dynamic mechanism consists of a slide rail and a slider.

4. A method for adjusting an ultra-low frequency quasi-zero stiffness vibration isolator, characterized in that: The process of adjusting the quasi-zero stiffness at ultra-low frequencies using the vibration isolators of claims 1-3 includes: At the static equilibrium position, i.e., when the first and second inclined rods are horizontal, negative stiffness is generated through the first and second inclined rods and the horizontal tension spring, which counteracts the positive stiffness of the vertical tension spring, bringing the overall dynamic stiffness of the vibration isolator to a state of zero stiffness. When deviating from the static equilibrium position, the dynamic stiffness of the vibration isolator is in a quasi-zero stiffness state. The stiffness adjustment mechanism is adjusted according to the following formula: Where P1, P2, and P3 are intermediate variables. For dimensionless displacement, x is the displacement from the initial state; for displacement starting from point O at the top of the inclined rod, a is the horizontal projection length of the inclined rod OO′ in the initial state, and h is the vertical distance from point O at the top of the inclined rod in the initial state to the static equilibrium position. Let f be the dimensionless force, k2 be the force acting on the top of the inclined rod, k1 be the stiffness of the vertical tension spring, and α be the stiffness ratio. Dimensionless force The derivative with respect to dimensionless displacement; Let be the dimensionless value of 'a'; and let b be the horizontal projection length of the lower half of the inclined rod O′O″ in the initial state. Let be the dimensionless value of b; δ be the pre-tension length of the horizontal tension spring in the initial state. Let δ be the dimensionless value; y be the displacement from the static equilibrium position, x = y + h, and the dimensionless relationship is: in:

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