Adjustable load vibration isolator and cam profile design method thereof

Abstract

The invention discloses an adjustable load vibration isolator and a cam profile design method thereof, and belongs to the technical field of vibration isolation, the adjustable load vibration isolator comprises a cam pair composed of a cam and a swing rod, a torsion spring, a worm and gear transmission mechanism, a gear motor, a weighing sensor and a control unit, and the torsion spring is adopted in the cam pair to save space; and the multiple groups of kinematic chains form a single-degree-of-freedom Sarrus mechanism. Load is detected through the weighing sensor, the control unit drives the gear motor and the worm and gear mechanism, the pre-tightening angle of the torsion spring is automatically adjusted, the pre-tightening state is maintained through the self-locking characteristic of the worm and gear, and continuous energy consumption is not needed. According to the cam profile design method, a coordinate system and a mechanical equation are established, an equidistant line spline curve is obtained through simultaneous derivation, a cam profile is formed after offset correction, and the positive stiffness bearing-quasi-zero stiffness vibration isolation segmented function is achieved. The vibration isolator is compact in structure, small in friction loss, long in service life and capable of self-adapting to different loads, and stable and reliable vibration isolation performance is guaranteed.

Description

Technical Field

[0001] This invention belongs to the field of vibration isolation technology, and in particular relates to an adjustable load vibration isolator and its cam profile design method. Background Technology

[0002] Roadbed compaction equipment achieves road compaction through vibration, exposing operators to prolonged vibration. Insufficient vibration isolation in the accompanying vibration isolators not only reduces operator comfort but also increases the risk of occupational health problems. Therefore, there is an urgent need for vibration isolators with superior vibration isolation performance to improve the operator's experience. In existing technologies, the vibration isolation effect of various vibration isolators is directly related to their stiffness characteristics in equilibrium. The closer the equilibrium stiffness of the isolator is to zero, the better the vibration isolation effect. However, the isolator must also meet load-bearing capacity requirements, hence the widespread application of quasi-zero stiffness vibration isolators. The core working principle of this type of isolator is: it possesses high positive stiffness during the initial deformation stage, preventing loss of load-bearing capacity due to excessive deformation; when the load increases to a specific threshold, the isolator's stiffness remains near zero, thus achieving effective vibration isolation.

[0003] In the prior art, the invention disclosed in CN118462745B discloses a vibration isolator that achieves vibration isolation for a finite number of discrete loads through the multi-segment profile of a cam. However, it lacks adjustability, and for other load values, the vibration isolator exhibits significant normal stiffness in a balanced state, resulting in poor vibration isolation performance. Furthermore, friction between the cam and the moving follower is difficult to avoid.

[0004] In the prior art, the invention with publication number CN117662671A discloses a load-adaptive vibration isolation device, which uses sensors to detect load changes in real time and automatically adjusts the spring position to adapt to vibration isolation requirements under different loads. However, the load adjustment relies on the control system, which, compared with passive vibration isolation devices, has a more complex structure, more potential failure points, and lower reliability.

[0005] In the prior art, the invention with publication number CN118128851A discloses a vibration isolator with adjustable load capacity. It achieves a wide quasi-zero stiffness vibration isolation range through the inclined spring and cantilever beam mechanism, which can adapt to load conditions. However, the structure is relatively complex, the linear motion occupies a large space, and it is difficult to avoid wear between the cantilever beam and the clamp, linear bearing and guide rail and other devices.

[0006] In the prior art, utility model publication CN209309199U discloses an adjustable load metal wire mesh pad vibration isolator. By adjusting the bolts to push the wedges, the compression of the metal wire mesh pad can be changed to adapt to different loads and vibration environments. However, the metal wire mesh pad will undergo plastic deformation during use, resulting in poor stiffness consistency and potentially a short service life, affecting the overall stability of vibration isolation performance.

[0007] Furthermore, in practice, operator weight changes cannot be predicted in advance, so traditional quasi-zero stiffness vibration isolators cannot produce good results on roadbed compaction equipment. Therefore, there is an urgent need for an adjustable load vibration isolator that is compact in structure, occupies little space, has low friction loss, and has a long service life. At the same time, it can produce good vibration isolation effect for different loads through simple adjustment, in order to solve the above problems. Summary of the Invention

[0008] The purpose of this invention is to provide an adjustable load vibration isolator to overcome the shortcomings of existing vibration isolation devices, such as insufficient reliability, easy wear, or poor performance stability, so as to achieve stable and reliable load adaptive vibration isolation and ensure consistent performance during long-term service.

[0009] The second objective of this invention is to provide a method for designing the cam profile of an adjustable load vibration isolator; To solve the above-mentioned technical problems, the technical solution adopted by the present invention is an adjustable load vibration isolator, comprising: a vibration isolator body; and a cam pair composed of a cam and a rocker arm; A torsion spring, wherein the torsion spring is disposed in the cam pair; An alternating shaft helical drive mechanism, wherein the alternating shaft helical drive is connected to the torsion spring; A geared motor, wherein the geared motor is connected to the interlaced shaft helical drive; A weighing sensor is used to measure the actual load; a control unit is connected to both the weighing sensor and the geared motor via signals.

[0010] Furthermore, it also includes a preload adjustment device, which is disposed at the torsion spring and is used to adjust the pre-compression angle of the torsion spring. The preload adjustment device is any one of a lever-type pre-compression device, a cam-type pre-compression device, a wedge block / inclined surface-type pre-compression device, an eccentric wheel-type pre-compression device, or a threaded adjustment pre-compression device.

[0011] Furthermore, the worm gear transmission mechanism includes a worm and two symmetrically distributed worm wheels. The worm is fixedly connected to the output shaft of the geared motor, and the worm meshes with the two worm wheels. The vibration isolator also includes a limiting shaft. The worm wheels, torsion spring, and rocker arm are all provided with limiting holes adapted to the limiting shaft. The limiting shaft passes through the limiting holes in sequence to connect the three components in series. The worm wheel is fixedly connected to the limiting shaft, and the rocker arm is rotatably connected to the limiting shaft. The worm wheel and the rocker arm are respectively provided with eccentric through holes. The thin rods at both ends of the torsion spring are respectively inserted into the eccentric through holes to achieve linkage of the three components.

[0012] Furthermore, the vibration isolator also includes a fixed wall, on which a limiting hole adapted to the limiting shaft is formed. The end of the limiting shaft passes through the limiting hole and is rotatably engaged by a bearing. The cam is arranged parallel to the fixed wall, and the profile surface of the cam engages with the top end of the rocker arm. The top end of the rocker arm is rotatably fitted with a roller via a pin. The outer circumferential rolling surface of the roller is adapted to the profile surface of the cam and maintains rolling contact. The end of the profile surface of the cam is integrally formed with a protruding structure.

[0013] Furthermore, the vibration isolator body includes a base, a moving platform, and multiple sets of kinematic chains. Each kinematic chain consists of an upper connecting rod, a lower connecting rod, and a revolute joint. The multiple sets of kinematic chains are symmetrically distributed between the moving platform and the base, forming a single-degree-of-freedom Sarrus mechanism. The moving platform and the upper connecting rod, as well as the base and the lower connecting rod, are all connected by revolute joints.

[0014] Furthermore, the number of kinematic chains is four, and the four kinematic chains are symmetrically distributed around the moving platform; the base is provided with a boss, and the reduction motor is fixedly installed on the boss; the cam is fixedly connected to the bottom surface of the moving platform, and the weighing sensor is clamped between the cam and the moving platform; the seat body is fixedly installed above the moving platform, and the base is fixedly connected to the chassis of the compaction equipment; the boss is integrally formed or fixedly connected to the base.

[0015] A method for designing the cam profile of an adjustable load vibration isolator includes the following steps: S1. First, take the midpoint O of the line connecting the hinge centers of the two rocker arms as the origin, the vertical direction as the y-axis, and the horizontal direction as the x-axis, and establish a fixed coordinate system {O}. In the initial state, the equation describing the equidistant line after the right profile curve of the cam is offset outward by the roller radius r is denoted as y=f(x) in the fixed coordinate system {O}. If the cam moves downward by Δy relative to the initial state, the equation describing the equidistant line in the fixed coordinate system {O} becomes y=f(x)-Δy. S2. Determine the force relationship between the cam and the rocker arm. The cam is subjected to the resultant force F of the two rocker arms. S for:

[0016] Among them, F C The force exerted by the rocker arm on the cam is in the direction of the normal to the equidistant line, and α is the angle between the tangent of the equidistant line at the center point of the fixed hinge of the rocker arm and the x-axis. The torque balance equation at the center point of the pendulum's motion hinge is as follows:

[0017] Among them l ABθ is the distance between the centers of the two hinges on the rocker arm, k is the stiffness of the torsion spring; θ0 is the initial swing angle of the rocker arm in the initial state, φ0 is the initial pre-torsion angle of the torsion spring in the initial state; Δθ is the further rotation angle of the rocker arm relative to the initial swing angle during the downward movement of the cam, which is equal to the further torsional deformation angle of the torsion spring relative to the initial pre-torsion angle during the downward movement of the cam. S3. Mark the coordinates of the center point of the fixed hinge of the rocker arm as (x A The coordinates of the center point of the pivot hinge of the pendulum are (x, 0), B ,y B If the coordinates of the pivot hinge center are:

[0018]

[0019] And the slope of the curve of the equidistant line in the fixed coordinate system {O} :

[0020] Determine the resultant force F exerted on the cam by the two rocker arms. S for: .

[0021] Furthermore, it also includes S4: the resultant force F on the cam. S Designed as a piecewise function for displacement Δy:

[0022] Where K is the custom first-stage normal stiffness, G0 is the rated load, Δy1 is the first-stage displacement threshold of the cam, and Δy2 is the upper displacement threshold of the second-stage cam; substituting the piecewise function into the resultant force F in S3. S In the equation, the equation of the ordinate of the center point of the pendulum's motion hinge is combined with the equation of the pendulum's motion hinge. The first-stage relational equation is obtained as follows:

[0023]

[0024] And the second-stage relational equation:

[0025]

[0026] Among them, l ABφ0 is the distance between the centers of the two hinges on the rocker arm, k is the stiffness of the torsion spring; θ0 is the initial swing angle of the rocker arm in the initial state, φ0 is the initial pre-torsion angle of the torsion spring in the initial state; Δθ is the further rotation angle of the rocker arm relative to the initial swing angle during the downward movement of the cam, which is equal to the further torsional deformation angle of the torsion spring relative to the initial pre-torsion angle during the downward movement of the cam; (x A ,0) The coordinates of the center point of the fixed hinge of the rocker arm, (x B ,y B () represents the coordinates of the center point of the pivot hinge of the pendulum. Let f(x) be the slope of the curve of the equidistant line in the fixed coordinate system {O}. B ) represents the ordinate function value of the cam equidistant line.

[0027] Furthermore, it also includes S5: Numerical algorithms are used to solve the relationship equations between the two stages in S4, determining the ordinate function values ​​f(x) of the cam equidistant lines in the first and second stages. B The x-coordinate of the center point of the motion hinge of the pendulum is... B The discrete correspondence is used to fit the spline curve equation of the cam equidistant line; the spline curve equation is shifted equidistantly to the left by the roller radius r to obtain the curve equation of the right profile of the cam.

[0028] Furthermore, this also includes S6: When the initial pre-torsion angle of the torsion spring decreases to φ0', replace φ0 with φ0' in the second-stage relational equation in step S4 to determine the resultant force F on the cam in the second stage. S 'x-coordinate of the center point of the motion hinge with the pendulum rod B The relationship.

[0029] Compared with existing technologies, the beneficial effects of this invention are as follows: Firstly, this invention innovatively employs a torsion spring as the elastic element in a cam pair composed of a cam and a rocker arm. Compared to traditional tension or compression springs, the torsion spring occupies less axial space and can provide the required restoring torque within a limited installation space, thus significantly improving the space utilization and compactness of the mechanism. Secondly, this invention constructs an adaptive adjustment system based on load feedback. This system acquires the actual load borne by the system in real time through a weighing sensor and uses this load signal as input to drive the motor to precisely control the rotation of the worm gear mechanism. The output end of the worm gear mechanism directly acts on the torsion spring, thereby automatically and accurately adjusting the preload angle of the torsion spring. This design allows key vibration isolation parameters such as the natural frequency of the vibration isolation system to be dynamically and adaptively optimized according to changes in load, ensuring that the system can provide optimal vibration isolation performance under different load conditions. Thirdly, this invention cleverly utilizes the inherent self-locking characteristics of interleaved shaft helical drives. Once the preload angle of the torsion spring is adjusted to the correct position, this self-locking characteristic prevents the torsion spring from undergoing unexpected angle changes due to external excitation or load fluctuations during operation. Therefore, after the preload angle is adjusted, the drive motor does not need to continuously provide holding torque, thus effectively avoiding continuous energy consumption caused by maintaining the preload state throughout the vibration isolation process, and achieving energy-saving operation of the system. Attached Figure Description

[0030] To more clearly illustrate the technical solutions in the embodiments of the present 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 present invention. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0031] Figure 1 This is a structural diagram showing the connection between the adjustable load vibration isolator and the seat in this embodiment; Figure 2 This is a front structural cross-sectional view of the adjustable load vibration isolator of this embodiment; Figure 3 This is an isometric view of the adjustable load vibration isolator of this embodiment; Figure 4 This is an exploded view of the swing arm connection structure of the adjustable load vibration isolator in this embodiment; Figure 5 This is a schematic diagram of the force analysis of the adjustable load isolator rocker arm when it is located in the quasi-zero stiffness profile of the cam in this embodiment. Figure 6 This is a schematic diagram of the adjustable load vibration isolator cam in its initial position according to this embodiment; Figure 7This is a schematic diagram of the adjustable load isolator rocker arm in this embodiment when it is located in the positive stiffness profile of the cam. Figure 8 This is a schematic diagram of the adjustable load isolator rocker arm in this embodiment when it is located at the quasi-zero stiffness profile of the cam. Figure 9 This is the force-displacement curve of the adjustable load vibration isolator in the vertical direction when the initial pre-torsion angle of the torsion spring decreases in this embodiment. Figure 10 This is a diagram illustrating the experimental steps of the adjustable load vibration isolator in this embodiment. Figure 11 This is a displacement transmissibility curve of the adjustable load isolator in this embodiment when the load is 20N. Figure 12 This is a displacement transmissibility curve of the adjustable load isolator in this embodiment when the load is 30N. Figure 13 This is a comparison chart of displacement transmissivity curves of the adjustable load isolator under different loads in this embodiment. In the diagram, 1. Base; 2. Lower connecting rod; 3. Upper connecting rod; 4. Boss; 5. Worm gear; 6. Gear motor; 7. Worm wheel; 8. Torsion spring; 9. Rocker arm; 10. Cam; 11. Roller; 12. Moving platform; 13. Weighing sensor; 14. Limiting shaft; 15. Fixed wall; 16. Seat body. Detailed Implementation

[0032] The technical solutions of the embodiments of the present invention will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some embodiments of the present invention, and not all embodiments. Based on the embodiments of the present invention, all other embodiments obtained by those skilled in the art without creative effort are within the scope of protection of the present invention.

[0033] This embodiment provides an adjustable load vibration isolator, specifically an adjustable load vibration isolator for roadbed compaction equipment. Figure 1 A structural diagram showing the connection between the adjustable load vibration isolator and the seat in this embodiment is shown; Figure 2 A front structural cross-sectional view of the adjustable load vibration isolator is shown. Figure 3 The overall structure of the adjustable load vibration isolator is shown. Figure 4 The rocker arm connection structure of the adjustable load isolator is shown; by Figures 1 to 3As can be seen, the adjustable load vibration isolator includes a base 1, a moving platform 12, four upper connecting rods 3, four lower connecting rods 2, a reduction motor 6, a worm gear 5, two worm wheels 7, four torsion springs 8, four rocker arms 9, two cams 10, and four load cells 13. Specifically, the main connection relationship of the above structure is that the cams 10 and the rocker arms 9 form a cam pair; the torsion springs 8 are disposed in the cam pair; the worm gear 5 and the two worm wheels 7 form an interlaced shaft helical drive, and the interlaced shaft helical drive is connected to the torsion springs 8; the reduction motor 6 is connected to the interlaced shaft helical drive; the load cells 13 are used to measure the actual load; furthermore, the adjustable load vibration isolator also includes a control unit, which is signal-connected to the load cells 13 and the reduction motor 6 respectively.

[0034] In some specific embodiments, the moving platform 12 and the base 1 are connected to the upper connecting rod 3 and the lower connecting rod 2 respectively through revolute joints; wherein, a single set of upper connecting rod 3 and lower connecting rod 2 together with three revolute joints form a kinematic chain, and four sets of the kinematic chains are symmetrically distributed around the moving platform 12. The four sets of kinematic chains combine to form a single-degree-of-freedom Sarrus mechanism, so that the vibration isolator has only a single degree of freedom of vertical movement.

[0035] In some specific embodiments, the seat body 16 is mounted above the moving platform 12. The base 1 is fixedly connected to the chassis of the compaction equipment. A boss 4 is integrally formed or fixedly provided on the base 1, and the boss 4 is used to fix the geared motor 6. The worm 5 is fixedly connected to the output shaft of the geared motor 6 by set screws. Two symmetrically distributed worm wheels 7 mesh on both sides of the worm 5. Both the worm gear 7 and the rocker arm 9 have limiting holes, the inner diameter of which matches the inner diameter of the torsion spring 8. The limiting shaft 14 passes through the limiting holes of the worm gear 7, the torsion spring 8, and the rocker arm 9 in sequence, connecting the three in series to ensure coaxiality. Furthermore, the rocker arm 9 and the limiting shaft 14 are rotatably connected by bearings, and the worm gear 7 and the limiting shaft 14 are fixedly connected by a key connection or interference fit. Even further, the rocker arm 9 and the worm gear 7 have corresponding eccentric through holes, the diameter of which matches the wire diameter of the torsion spring 8. The thin rods at both ends of the torsion spring 8 are respectively inserted into the corresponding eccentric through holes of the worm gear 7 and the rocker arm 9, realizing the linkage of the worm gear 7, the rocker arm 9, and the torsion spring 8.

[0036] In some specific embodiments, two sets of fixed walls 15 are formed on the base 1 on both sides of the boss 4 along the movement direction of the kinematic chain. The fixed walls 15 are provided with limiting holes that are adapted to the limiting shaft 14. The end of the limiting shaft 14 passes through the limiting hole of the fixed wall 15 and rotates with the limiting hole through the bearing to achieve axial and radial limiting of the limiting shaft 14.

[0037] In some specific embodiments, two cams 10 and four weighing sensors 13 are fixedly connected to the moving platform 12, and the seat body 16 is fixedly connected to the moving platform 12. Specifically, the cam 10 is arranged parallel to the fixed wall 15. The cam 10 includes a fixed surface and a contour surface arranged opposite to each other. Its fixed surface is fixedly connected to the bottom surface of the moving platform 12, and two weighing sensors 13 are sandwiched between the fixed surface of the cam 10 and the moving platform 12. The contour surface of the cam 10 is arranged downward and cooperates with the top end of the swing arm 9. Specifically, the top end of the swing arm 9 is rotatably equipped with a roller 11 through a pin. The outer circumferential rolling surface of the roller 11 is adapted to the contour surface of the cam 10 and maintains rolling contact. When the vibration generated by the compaction equipment is transmitted to the vibration isolator, the roller 11 rolls along the contour surface of the cam 10, driving the swing arm 9 to swing synchronously, causing the torsion spring 8 to generate a contraction deformation of angle Δθ. The torsion spring 8 absorbs the vibration impact force through this deformation, thereby reducing the vibration transmitted to the seat where the operator sits. The end of the profile surface of the cam 10 is integrally formed with a protruding structure. The protruding structure is used to mechanically limit the roller 11 under the condition of sudden increase in load, so as to ensure the safe operation of the vibration isolator.

[0038] Specifically, when the vibration isolation system is under static load, the geared motor 6 stops and remains stationary, the worm 5 is synchronously fixed with the output shaft of the geared motor 6, and the worm wheel 7 meshing with the worm 5 also remains fixed. At this time, one end of the torsion spring 8 becomes a fixed end due to the fixed worm wheel 7. The system's inherent weight and the external load act together on the torsion spring 8, causing it to undergo a slight contraction deformation and generate a pre-torque force. Under the synergistic effect of its own weight, the external load, and the pre-torque force of the torsion spring 8, the system maintains a stable equilibrium state. The specific force analysis is as follows: Figure 5 As shown.

[0039] When the vibration isolation system is in operation, the geared motor 6 remains stationary, the worm 5 and worm wheel 7 are fixed, and the fixed end of the torsion spring 8 remains unchanged. The vibration generated by the compaction equipment is transmitted to the vibration isolation system, causing the cam 10 to produce a quasi-static displacement. This displacement drives the rocker arm 9 to move, and the torsion spring 8 outputs a nonlinear restoring torque through the rocker arm 9, which can effectively suppress the displacement amplitude of the cam 10, thereby achieving the vibration isolation effect.

[0040] In some optional embodiments, the vibration isolation system of this embodiment further includes a preload adjustment device disposed on or near the torsion spring 8, for adjusting the pre-compression angle of the torsion spring 8 to adjust its initial stiffness. The adjustment method for the pre-compression angle of the torsion spring 8 can be any suitable mechanism known to those skilled in the art, such as, but not limited to: lever-type pre-compression device, cam-type pre-compression device, wedge / inclined surface-type pre-compression device, eccentric wheel-type pre-compression device, or threaded adjustment pre-compression device. Further, to achieve dynamic optimization of vibration isolation performance, the preload adjustment device is signal-connected to the load cell 13. The preload adjustment device can be internally or externally connected to a control unit configured to: receive a signal from the load cell 13 representing the load borne by the seat body 16; and, based on the load signal, automatically and in real-time adjust the pre-compression angle of the torsion spring 8 according to a preset control strategy or algorithm to obtain the best vibration isolation effect under different load conditions. This configuration makes the adjustment process of the pre-compression angle (system load) of the torsion spring 8 completely automated, without manual intervention.

[0041] In a specific embodiment of the present invention, the load adjustment mechanism of the vibration isolation system is as follows. When the system load changes and the vibration isolation parameters need to be adjusted, the system starts the reduction motor 6. The power output by the reduction motor 6 drives the worm gear 5 connected to it to rotate. The rotation of the worm gear 5 drives the two meshing worm wheels 7 to rotate synchronously at the same speed but in opposite directions. The rotation of the worm wheels 7 further drives the torsion spring 8 connected to it to twist, thereby changing the initial pre-torque of the torsion spring 8, and its pre-torque angle is also adjusted synchronously. Taking the case where the combined load of the moving platform and the operator decreases as an example, the adjustment process is as follows: First, the system control unit calculates the required reduction in load force based on the load change detected by the weighing sensor 13, and converts the load force change into the target angle that the worm gear 5 needs to rotate through a preset control algorithm or mapping relationship. Subsequently, the control unit sends a control command to the reduction motor 6, causing the reduction motor 6 to rotate according to the calculated target angle. Through the above mechanical transmission chain, the rotation of the reduction motor 6 is ultimately converted into a precise adaptation adjustment of the pre-torque of the torsion spring 8. In this way, the vibration isolation system of this embodiment can automatically and accurately adjust the preload of the torsion spring 8 according to the different loads it actually bears, thereby realizing the adaptive adjustment of the natural frequency of the vibration isolation system and ensuring that it can provide the best vibration isolation performance under various load conditions.

[0042] This embodiment also provides a method for designing the profile of the vibration isolator cam 10, the specific process of which is as follows: First, establish a fixed coordinate system {O} with the midpoint O of the line connecting the hinge centers of the two rocker arms 9 as the origin, and the vertical direction as the y-axis and the horizontal direction as the x-axis. The forces acting on the cam 10 and the rocker arms 9 are as follows: Figure 5 As shown.

[0043] In the initial state, the equation describing the equidistant line of the right profile curve of cam 10 after being offset outward by the roller radius r is denoted as y=f(x) in the fixed coordinate system {O}. If cam 10 moves downward by Δy relative to the initial state, the equation describing the equidistant line in the fixed coordinate system {O} becomes y=f(x)-Δy.

[0044] As the cam 10 moves downward, the cam 10 is subjected to the resultant force F of the two rocker arms 9. S for: (1) Among them, F C The force exerted by the rocker arm 9 on the cam 10 is in the direction of the normal to the equidistant line, and α is the angle between the tangent of the equidistant line at point B and the x-axis.

[0045] The torque balance equation of the pendulum 9 at point A is as follows: (2) Among them l AB φ is the distance between the centers of the two hinges on the rocker arm 9, and k is the stiffness of the torsion spring 8. θ0 is the initial swing angle of the rocker arm 9 in the initial state, and φ0 is the initial pre-torsion angle of the torsion spring 8 in the initial state; Δθ is the further rotation angle of the rocker arm 9 relative to the initial swing angle during the downward movement of the cam 10, and also the further torsional deformation angle of the torsion spring 8 relative to the initial pre-torsion angle during the downward movement of the cam 10.

[0046] According to formulas (1) and (2), the cam 10 is subjected to the resultant force F of the two rocker arms 9. S for: (3) Let the coordinates of point A, the center of the fixed hinge of the rocker arm 9, be (x... A (0) (the rotational connection center point between the rocker arm 9 and the limiting shaft 14), the coordinates of the center point B of the motion hinge of the rocker arm 9 are (x, 0). B ,y B (The rotational connection center point between the top of the rocker arm 9 and the roller 11). Therefore, the coordinate relationship of the hinge center of the rocker arm 9 is: (4) (5) During the downward movement of cam 10, the center point B of the motion hinge of rocker arm 9 is always located on the equidistant line, therefore: (6) The derivative of function f at point B is the slope of the curve of the equidistant line in the fixed coordinate system {O}. Therefore, it is: (7) Combining formulas (3) and (6), we can obtain: (8) In order to achieve the functions of load bearing and vibration isolation, the resultant force F on cam 10 is... S Designed as a piecewise function of Δy, where the first stage has a large positive stiffness, F S The stiffness increases significantly and linearly with the increase of Δy, thus achieving the load-bearing function; the second stage has quasi-zero stiffness, F S The piecewise function remains constant as Δy increases further, achieving quasi-zero stiffness vibration isolation. The piecewise function is defined as follows: (9) Where K is the custom first-stage positive stiffness, G0 is the rated load, Δy1 is the first-stage displacement threshold (i.e., the boundary threshold between the first and second stages of the cam 10 displacement piecewise function), and Δy2 is the second-stage displacement threshold. Both are the displacement values ​​of the cam 10 moving downward relative to the initial state.

[0047] F in the first and second stages S Substituting the expressions into formula (8), and into formulas (5) and (6) respectively, we can obtain the first-stage relational equation: (10) (11) Substituting formulas (4) and (5) into formulas (10) and (11), we can eliminate Δθ and obtain the x-coordinates of the center points of the motion hinges in the first and second stages. B The vertical coordinate function value f(x) of the equidistant line of cam 10 B The slope of the tangent line to the equidistant line at point B (the center point of the hinge of the pendulum 9) is f'(x). B The relationship is as follows: (12) (13) Based on formulas (12) and (13), the f(x) values ​​for the first and second stages can be determined using numerical algorithms. B ) and x BBased on the discrete correspondence, the spline curve equation of the equidistant line can be determined. After offsetting this curve equation equidistantly to the left by the roller radius r, the curve equation of the right profile of cam 10 can be obtained. Cam 10 made according to this profile can achieve quasi-zero stiffness vibration isolation under the following conditions: the sum of the actual operator's and the moving platform's weight is exactly the rated load G0, and the amplitude of the external excitation on the base is less than or equal to Δy2-Δy1.

[0048] For the cam 10 made according to the above profile, if the initial pre-torsion angle of the torsion spring 8 is reduced to φ0' during actual use, then φ0 in formula (12) can be replaced with φ0' to obtain the resultant force F on the cam 10 in the second stage. S 'with x B Relationship: (14) According to formula (14), when the initial pre-torsion angle of the torsion spring 8 decreases, for the same x B The resultant force on cam 10 also decreases.

[0049] Let θ0 = 20°, φ0 = 10°, x A =29mm, l AB =80mm, k=3.1Nm / rad, K=40000N / mm, G0=800N, Δy1=20mm, Δy2=50mm, according to formula (12) and formula (13), the right profile curve of cam 10 is as follows. Figure 6 As shown. Figure 7 and Figure 8 The dynamic posture of the vibration isolator is shown when the rocker arm 9 is positioned at two different characteristic contour segments of the cam 10, intuitively presenting the spatial state of the vibration isolator at different rocker arm positions. For the cam 10 made according to this contour, when the initial pre-torsion angle of the torsion spring 8 is kept constant at 10 degrees and reduced to 8 degrees, 6 degrees, 4 degrees, 2 degrees, and 0 degrees, F... S The relationship with Δy is as follows: Figure 9 As shown. By Figure 9 In the figure, the six curves from left to right are the force-displacement curves of the vibration isolator in the vertical direction when the initial pre-torsion angle of the torsion spring 8 is kept constant at 10 degrees and reduced to 8 degrees, 6 degrees, 4 degrees, 2 degrees and 0 degrees. When the initial pre-torsion angle of the torsion spring 8 decreases, the resultant force on the cam 10 also decreases, and the vibration isolator exhibits a very small positive stiffness in the y direction.

[0050] If the sum of the actual weight of the operator and the moving platform is less than the rated load G0, the sum of the actual weight G of the operator and the moving platform can be measured by four load cells mounted on the upper plane of the two cams 10. Under static equilibrium, according to the force balance relationship, we can obtain: (15) Let Δy = (Δy2 - Δy1) / 2 be the displacement of cam 10 in the y direction under static equilibrium. Substituting it into formula (6), we can obtain the x-coordinate of the center B of the movable hinge under static equilibrium. B The obtained x B Substituting into formula (15), the initial pre-torsion angle φ0' corresponding to G can be obtained.

[0051] Then, the worm gear 5 is driven to rotate by the reduction motor 6, causing the left worm wheel to rotate counterclockwise by φ0-φ0' and the right worm wheel to rotate clockwise by φ0-φ0', thereby adjusting the initial pre-torsion angle values ​​of the two torsion springs 8 to the expected value φ0'. At this time, for the load G, the vibration isolator has a very small positive stiffness when the displacement of the moving platform is between Δy1 and Δy2, thus still having a good vibration isolation effect.

[0052] To investigate the vibration isolation performance of the vibration isolator in this embodiment, a series of vibration tests were conducted under two different load conditions, approximately 20N and 30N respectively. The vibration isolation test platform includes the following components: (1) Vibration System: The core equipment of the vibration testing system includes a computer equipped with a vibration control program developed by Yiheng Technology Co., Ltd., a VSA-H102A power amplifier, a VT-900X vibration controller, and a VE-5150 small electric vibration table. Before starting the experiment, the control program on the computer is activated, a suitable excitation method is selected, and relevant parameters are set. Once the settings are completed, these parameters are sent to the output unit of the vibration controller to provide analog signals to the power amplifier. Then the power amplifier is activated, and the gain is adjusted to about 2 / 3 of its position to drive the vibration table.

[0053] (2) Quasi-zero stiffness vibration isolation system based on cam mechanism: This system consists of a quasi-zero stiffness vibration isolation mechanism and a fixed connecting plate. The fixed connecting plate connects the vibration table and the vibration isolation mechanism and is fastened to the vibration table with bolts, ensuring structural stability. The vibration isolation mechanism is fixed on the connecting plate, ensuring structural reliability and effectively transmitting the vibration of the vibration table. The design principle of the vibration isolation mechanism is that it utilizes a special mechanical structure to generate near-zero dynamic stiffness within a small vibration range, thereby achieving a vertical vibration isolation effect. The main function of the quasi-zero stiffness vibration isolation mechanism is to reduce or eliminate the vertical vibration of the vibration table. By transmitting energy through the vibration of internal positive and negative stiffness units, the energy is dissipated and eliminated within the mechanism itself, maintaining the acceleration of the device and structure on the vibration isolator below the vibration acceleration caused by the vibration table, thus protecting precision instruments and equipment.

[0054] (3) Data Acquisition System: Composed of an accelerometer computer, a vibration controller input section, and vibration analysis software. It enables comprehensive monitoring of the vibration environment and analysis of the performance of the vibration isolation mechanism. Accelerometers are fixed to the vibration table surface and the quasi-zero stiffness vibration isolation mechanism, respectively, to detect vibration excitation signals and real-time system response signals. They are the system's "ears," accurately measuring and recording vibration acceleration. The signals acquired by the sensors are sent to the vibration control input unit for vibration control. The vibration control input unit functions as a signal conditioning and AD conversion unit, ensuring the stability and accuracy of signal transmission. It converts the analog signals transmitted from the sensors into digital signals for input calculation.

[0055] Vibration isolation experiment flowchart as follows Figure 10 As shown, in the dynamic experiment of the quasi-zero stiffness vibration isolation mechanism based on the cam mechanism, the processed parts are assembled to form the quasi-zero stiffness vibration isolation mechanism based on the cam mechanism. The vibration isolation mechanism is then fixedly connected to the vibration table through a fixed connecting plate to achieve the stability requirement of transmitting the vibration of the vibration table to the quasi-zero stiffness vibration isolation mechanism based on the cam mechanism.

[0056] To accurately acquire measurement data of excitation and response signals, this study uses screws to mount accelerometer 1 on the surface of the vibration table to monitor the excitation signal. Accelerometer 2 is placed on the isolation platform of the quasi-zero stiffness isolation mechanism to collect the dynamic response characteristics of the isolation system in real time. Each sensor is connected to the vibration controller via appropriate wiring, and a computer equipped with dedicated vibration experiment analysis software performs real-time signal acquisition, processing, and analysis. Before starting the experiment, this implementation starts the control program on the computer, selects the sinusoidal excitation mode, and sets the parameters. In this implementation, the gain of the power amplifier is adjusted to approximately 2 / 3 to ensure appropriate vibration intensity. Once preparation is complete, pressing the start button on the control program will activate the vibration table according to the set mode.

[0057] After the experiment, the experimental data was imported into the analysis software. The response signal of accelerometer 2 was divided by the corresponding value of the excitation signal of accelerometer 1 at each frequency to calculate the displacement transmissibility of the vibration isolation mechanism at each frequency band. This displacement transmissibility serves as an indicator for evaluating the vibration isolation effect, reflecting the vibration isolation performance based on the frequency characteristics of the vibration isolation mechanism. Based on the results of the displacement transmissibility analysis, this embodiment will comprehensively evaluate the vibration isolation performance of the near-zero stiffness mechanism and use this as the basis for design improvement.

[0058] Since the minimum excitation frequency of the vibration platform is 5Hz, under the premise of meeting the platform parameter settings, the excitation amplitude is 3mm, and the excitation frequency is extended from 5Hz to 40Hz, adjusted every 1Hz. After the system response stabilizes, the output signal data of accelerometer 2 is recorded. The displacement transmissibility curves obtained under different load conditions are shown below. Figure 11 and Figure 12 As shown.

[0059] Depend on Figure 11 and Figure 12 As can be seen from the example graph, the horizontal axis represents the excitation frequency f, and the vertical axis represents the displacement transmissibility Ta. Within the excitation frequency range, the displacement transmissibility at the corresponding frequency can be accurately calculated. The graph shows that, regardless of whether the load is 20N or 30N, the displacement transmissibility gradually decreases with increasing excitation frequency.

[0060] Under a working load of 20N, by Figure 11 It can be seen that the displacement transmissibility decreases sharply in the excitation frequency range of 5Hz-10Hz, indicating that the vibration isolation system is playing a vibration isolation role, and the effect is becoming more and more significant. When the excitation frequency is 10Hz, the displacement transmissibility has decreased to about 0.3. In the excitation frequency range of 10Hz-40Hz, although the displacement transmissibility gradually decreases with the increase of the excitation frequency, the rate of decrease tends to be gradual. Compared with the displacement transmissibility in the previous low frequency range, although the rate has decreased, the displacement transmissibility is mostly below 0.15, indicating that the vibration isolation system has played a superior vibration isolation role in this excitation frequency range.

[0061] Under a working load of 30N, by Figure 12 It can be seen that within the excitation frequency range of 5Hz-10Hz, the displacement transmissibility decreases sharply, similar to the working state with a load of 20N. This indicates that the vibration isolation system consistently performs vibration isolation within the appropriate load range, and the effect becomes increasingly significant. At an excitation frequency of 10Hz, the displacement transmissibility has decreased to approximately 0.2, a 33% reduction compared to the displacement transmissibility with a load of 20N. In the excitation frequency range of 10Hz-40Hz, the displacement transmissibility also gradually decreases with increasing excitation frequency, but the rate of decrease becomes more gradual. Compared to the displacement transmissibility in the previous low-frequency range, although the rate of decrease is somewhat reduced, the displacement transmissibility is mostly below 0.1, indicating that within this excitation frequency range, the vibration isolation system exhibits superior vibration isolation performance.

[0062] Depend on Figure 13As the load increases, it is evident that the displacement transmissibility curve at a load of 30N is consistently lower than that at a load of 20N, especially in the low-frequency range of 5-15Hz. As shown in the force-displacement curves under different loads in Chapter 3, when the load is 30N, the vibration isolation system is in a quasi-zero stiffness state, exhibiting significant vibration isolation performance. When the load is 20N, the system is in a positive stiffness state, indicating that the quasi-zero stiffness vibration isolation system is significantly more effective than the positive stiffness system in this frequency range. However, as the excitation frequency gradually increases, the displacement response curves of both systems converge. Once the frequency exceeds 30Hz, their dynamic characteristics become essentially the same. Measured data shows that under any load condition, the vibration isolation device exhibits stable displacement attenuation performance, and its transmissibility curve maintains a monotonically decreasing characteristic, consistent with previous theoretical derivations. In-depth observation reveals that the displacement transmissibility in the theoretical model exhibits typical nonlinear characteristics. Its curve shows a rapid upward trend in the low-frequency region, forming a prominent peak at the critical frequency before decaying. The experimental test used 5Hz as the initial excitation frequency, effectively avoiding the system's inherent resonance region. The measured curve only captures the gradual decay process to the right of the peak, which precisely verifies the theoretical model's prediction regarding the system's resonance avoidance mechanism. The displacement transmissibility curve of the quasi-zero stiffness isolation mechanism during the vibration isolation experiment is roughly the same as that in the dynamic analysis, verifying the calculated analysis results and the accuracy of the experimental results. From the above simulation and experimental data, it can be seen that the displacement transmissibility of the quasi-zero stiffness mechanism conforms to the calculated results, and the vibration isolation effect meets the design requirements.

[0063] The various embodiments in this specification are described in a related manner. Similar or identical parts between embodiments can be referred to mutually. Each embodiment focuses on describing the differences from other embodiments. In particular, the system embodiments are basically similar to the method embodiments, so the description is relatively simple; relevant parts can be referred to the descriptions of the method embodiments.

[0064] The above description is merely a preferred embodiment of the present invention and is not intended to limit the scope of protection of the present invention. Any modifications, equivalent substitutions, improvements, etc., made within the spirit and principles of the present invention are included within the scope of protection of the present invention.

Claims

1. An adjustable load vibration isolator, characterized in that, include: Vibration isolator body; A cam pair consisting of a cam (10) and a rocker arm (9); A torsion spring (8) is disposed in the cam pair; An alternating shaft helical drive mechanism, wherein the alternating shaft helical drive is connected to the torsion spring (8); A geared motor (6) is connected to the staggered shaft helical drive; A load cell (13) is used to measure the actual load; The control unit is connected to the weighing sensor (13) and the geared motor (6) respectively.

2. The adjustable load vibration isolator according to claim 1, characterized in that, It also includes a preload adjustment device, which is located at the torsion spring (8) and is used to adjust the pre-compression angle of the torsion spring (8). The preload adjustment device is any one of a lever-type pre-compression device, a cam-type pre-compression device, a wedge-block / inclined-plane type pre-compression device, an eccentric wheel type pre-compression device, or a threaded adjustment type pre-compression device.

3. The adjustable load vibration isolator according to claim 1, characterized in that, The worm gear transmission mechanism includes a worm (5) and two symmetrically distributed worm wheels (7). The worm (5) is fixedly connected to the output shaft of the geared motor (6), and the worm (5) meshes with the two worm wheels (7). The vibration isolator also includes a limiting shaft (14). The worm wheel (7), the torsion spring (8), and the rocker arm (9) are all provided with limiting holes adapted to the limiting shaft (14). The limiting shaft (14) passes through the limiting holes in sequence to connect the three in series. The worm wheel (7) is fixedly connected to the limiting shaft (14), and the rocker arm (9) is rotatably connected to the limiting shaft (14). The worm wheel (7) and the rocker arm (9) are respectively provided with eccentric through holes. The thin rods at both ends of the torsion spring (8) are respectively inserted into the eccentric through holes to realize the linkage of the three.

4. The adjustable load vibration isolator according to claim 1 or 3, characterized in that, The vibration isolator also includes a fixed wall (15), on which a limiting hole adapted to the limiting shaft (14) is provided. The end of the limiting shaft (14) passes through the limiting hole and is rotated and engaged by a bearing. The cam (10) is arranged parallel to the fixed wall (15), and the profile surface of the cam (10) is engaged with the top end of the rocker arm (9). The top end of the rocker arm (9) is rotatably fitted with a roller (11) through a pin. The outer circumferential rolling surface of the roller (11) is adapted to the profile surface of the cam (10) and maintains rolling contact. The end of the profile surface of the cam (10) is integrally formed with a protruding structure.

5. The adjustable load vibration isolator according to claim 1, characterized in that, The vibration isolator body includes a base (1), a moving platform (12) and multiple sets of kinematic chains. The kinematic chains consist of an upper connecting rod (3), a lower connecting rod (2) and a revolute joint. The multiple sets of kinematic chains are symmetrically distributed between the moving platform (12) and the base (1) to form a single-degree-of-freedom Sarrus mechanism. The moving platform (12) and the upper connecting rod (3), and the base (1) and the lower connecting rod (2) are all connected by revolute joints.

6. The adjustable load vibration isolator according to claim 1 or 5, characterized in that, The number of kinematic chains is four, and the four kinematic chains are symmetrically distributed around the moving platform (12); the base (1) is provided with a boss (4), and the geared motor (6) is fixedly installed on the boss (4); the cam (10) is fixedly connected to the bottom surface of the moving platform (12), and the weighing sensor (13) is sandwiched between the cam (10) and the moving platform (12); the seat body (16) is fixedly installed above the moving platform (12), and the base (1) is fixedly connected to the chassis of the compaction equipment; the boss (4) and the base (1) are integrally formed or fixedly connected.

7. A method for designing the cam profile of an adjustable load vibration isolator, characterized in that, Includes the following steps: S1. First, take the midpoint O of the line connecting the hinge centers of the two pendulum rods (9) as the origin, the vertical direction as the y-axis and the horizontal direction as the x-axis, and establish a fixed coordinate system {O}. In the initial state, the equation describing the equidistant line of the right profile curve of the cam (10) after being offset outward by the roller radius r is y=f(x) in the fixed coordinate system {O}. If the cam (10) moves downward by Δy relative to the initial state, the equation describing the equidistant line in the fixed coordinate system {O} becomes y=f(x)-Δy. S2. Determine the force relationship between the cam (10) and the rocker arm (9). The cam (10) is subjected to the resultant force F of the two rocker arms (9). S for: ； Among them, F C The force exerted by the rocker arm (9) on the cam (10) is in the direction of the normal to the equidistant line, and α is the angle between the tangent of the equidistant line at the center point of the fixed hinge of the rocker arm (9) and the x-axis. The torque balance equation at the center point of the motion hinge of the pendulum (9) is as follows: ； Among them l AB θ0 is the distance between the centers of the two hinges on the rocker arm (9), k is the stiffness of the torsion spring (8); θ0 is the initial swing angle of the rocker arm (9) in the initial state, φ0 is the initial pre-torsion angle of the torsion spring (8) in the initial state; Δθ is the swing angle of the rocker arm (9) relative to the initial swing angle during the downward movement of the cam (10), which is equal to the further torsional deformation angle of the torsion spring (8) relative to the initial pre-torsion angle during the downward movement of the cam (10). S3. The coordinates of the fixed hinge center point of the rocker arm (9) are marked as (x A The coordinates of the center point of the motion hinge of the pendulum (9) are (x, 0), and (x, 0). B ,y B If the coordinates of the hinge center of the pendulum (9) are: ； ； And the slope of the curve of the equidistant line in the fixed coordinate system {O} : ； The cam (10) is subjected to the resultant force F from the two rocker arms (9). S for: 。 8. The vibration isolator cam profile design method according to claim 7, characterized in that, It also includes S4: the resultant force F on the cam (10) S Designed as a piecewise function for displacement Δy: ； Where K is the custom first-stage positive stiffness, G0 is the rated load, Δy1 is the first-stage displacement threshold of cam (10), and Δy2 is the second-stage displacement upper limit threshold of cam (10); Substitute the piecewise function into the resultant force F in S3. S In the equation, the equation of the ordinate of the center point of the motion hinge of the pendulum (9) is combined with the equation of the motion hinge center point. The first-stage relational equation is obtained as follows: ； ； And the second-stage relational equation: ； ； Among them, l AB Let θ0 be the distance between the centers of the two hinges on the rocker arm (9), k be the stiffness of the torsion spring (8); θ0 be the initial swing angle of the rocker arm (9) in the initial state, φ0 be the initial pre-torsion angle of the torsion spring (8) in the initial state; Δθ be the further rotation angle of the rocker arm (9) relative to the initial swing angle during the downward movement of the cam (10), which is equal to the further torsional deformation angle of the torsion spring (8) relative to the initial pre-torsion angle during the downward movement of the cam (10); (x A ,0) The coordinates of the center point of the fixed hinge of the pendulum (9), (x B ,y B ) represents the coordinates of the center point of the motion hinge of the pendulum (9). Let f(x) be the slope of the curve of the equidistant line in the fixed coordinate system {O}. B ) is the ordinate function value of the equidistant line of cam (10).

9. The vibration isolator cam profile design method according to claim 8, characterized in that, Also includes S5: Numerical algorithms are used to solve the relationship equations between the two stages in S4, and to determine the ordinate function values ​​f(x) of the equidistant lines of the cam (10) in the first and second stages. B The x-coordinate of the center point of the motion hinge of the pendulum (9) and the rocker arm (9) B The discrete correspondence is used to fit the spline curve equation of the equidistant line of the cam (10); the spline curve equation is shifted equidistantly to the left by the roller radius r to obtain the curve equation of the right profile of the cam (10).

10. The vibration isolator cam profile design method according to claim 8, characterized in that, Also includes S6: When the initial pre-torsion angle of the torsion spring (8) decreases to φ0', replace φ0 with φ0' in the second-stage relational equation in step S4 to determine the resultant force F on the cam (10) in the second stage. S The x-coordinate of the center point of the motion hinge with the pendulum (9) is x B The relationship.

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