Forklift steering wheel vibration suppression structure and design method thereof

Abstract

The application discloses a forklift steering wheel vibration suppression structure and a design method thereof, and belongs to the field of internal combustion forklift vibration control. The method comprises the following steps: establishing a five-degree-of-freedom coupled dynamics model containing a vehicle frame, a roof frame, a steering wheel and a nonlinear dynamic vibration absorber; constructing an objective function with the minimization of the steering wheel vibration response as the target; iteratively optimizing the linear and nonlinear stiffness and damping parameters of the vibration absorber by using a particle swarm algorithm; and determining the vibration absorber structure composed of a mass block and multiple steel wire rope isolators according to the optimal parameters. The suppression structure comprises the roof frame and a nonlinear dynamic vibration absorber designed by the optimization method and installed on the top of the roof frame. Through target-oriented reverse design, the vibration absorber can efficiently suppress the wideband vibration of the distal steering wheel, and has the advantages of remarkable control effect, wide frequency band, simple and reliable structure and strong environmental adaptability.

Description

Technical Field

[0001] This invention relates to the field of vibration control technology for internal combustion forklifts, and more specifically, to a design method for a nonlinear dynamic vibration absorber for suppressing vibration of the forklift steering wheel, and a vibration suppression structure designed using this method. Background Technology

[0002] Internal combustion forklifts are widely used in logistics parks, industrial and mining enterprises, and other fields due to their advantages such as long range, ample power, and strong adaptability to working conditions. However, in actual operation, internal combustion forklifts often experience severe steering wheel vibration under certain conditions such as idling. This not only seriously affects operating comfort but also leads to premature failure and damage to accessories such as lights mounted on the overhead guard. More importantly, as the main support structure of the cab, the significant vibration of the overhead guard can further induce vibration of the entire cab structure and radiate noise, directly affecting the overall NVH performance of the forklift, as well as the driver's comfort and health.

[0003] To address this problem, traditional vibration control methods mostly focus on two directions:

[0004] One approach is to increase the rigidity of the top support frame by increasing the size of its structural components, thereby suppressing the vibration of the top support frame itself.

[0005] Secondly, add vibration isolation components to the steering wheel mounting structure, or install a dynamic vibration absorber on the steering wheel column.

[0006] However, due to structural size limitations, the space around the forklift steering wheel available for implementing vibration control measures is extremely limited, making it difficult for traditional methods to achieve ideal results in practical applications.

[0007] Dynamic vibration absorbers, as devices that absorb the vibration energy of the main structure and suppress its resonance by adding a sub-vibration system, are widely used in machinery, aerospace, and shipbuilding. Traditional dynamic vibration absorbers mostly consist of a combination of linear elastic elements and a mass block. Their working principle is to match the natural frequency of the absorber with the resonant frequency of the main structure by reasonably selecting the stiffness of the elastic element and the mass block, thereby generating reverse vibration to counteract the vibration energy of the main structure.

[0008] However, once the stiffness of a linear elastic element is determined, the vibration absorber can only effectively absorb vibration energy at a specific frequency. Once the vibration frequency of the main structure shifts or changes, the vibration absorption effect will drop sharply, resulting in poor adaptability. Furthermore, the damping elements in traditional vibration absorbers are mostly additional components, which not only increases structural complexity and manufacturing costs but also suffers from problems such as unadjustable damping characteristics, poor environmental resistance, and poor synergy with the elastic element, affecting the stability and service life of the vibration absorber.

[0009] In recent years, wire rope vibration isolators have attracted widespread attention as a typical nonlinear vibration isolation device. They possess advantages such as simple structure, high load-bearing capacity, resistance to high and low temperatures, impact resistance, and long service life. They also inherently combine elastic and damping characteristics, eliminating the need for additional damping elements, and their stiffness exhibits significant nonlinear characteristics. However, how to effectively apply the nonlinear characteristics of wire rope vibration isolators to the specific scenario of forklifts, especially how to systematically design them to suppress vibrations of the remote steering wheel, remains an unsolved technical challenge in this field. Summary of the Invention

[0010] The technical problem to be solved by this invention is to address the limitations of existing forklift vibration control methods, such as the difficulty in implementing them in the space around the steering wheel, the narrow bandwidth and poor frequency adaptability of traditional linear vibration absorbers, and the lack of systematic design methods for nonlinear vibration isolation elements targeting distant vibration targets. This invention provides a parameter design method and a suppression structure for a forklift steering wheel vibration suppression structure that can efficiently suppress steering wheel vibration over a wide frequency range, and is simple in structure and low in cost.

[0011] To address the aforementioned problems, this invention provides a design method for a forklift steering wheel vibration suppression structure, comprising the following steps:

[0012] S1. Establish a simplified coupled dynamics model of the forklift. This simplified coupled dynamics model includes at least the frame, front panel and dashboard assembly, overhead guard, steering wheel system, and the vibrating mass corresponding to the nonlinear dynamic vibration absorber mounted on top of the overhead guard. The generalized coordinate vector of the system is defined as follows: ;

[0013] in, , , , , These correspond to the displacements of the vibrating masses of the vehicle frame, front panel and dashboard assembly, roof guard, steering wheel system, and nonlinear dynamic shock absorber, respectively.

[0014] S2. Construct an objective function that optimizes the steering wheel vibration response. The objective function under wideband excitation conditions is defined as follows: ;

[0015] in, Here is the parameter vector of the vibration absorber to be optimized. ; This represents the frequency domain response to the steering wheel displacement. For the target frequency band;

[0016] S3. The particle swarm optimization algorithm is used to optimize the vibration absorber parameter vector. Iterative optimization is performed. In each iteration, the current parameter vector is substituted into the simplified coupled dynamics model to solve for the system response and calculate the objective function value.

[0017] S4. Based on the optimal parameter results obtained from the optimization, determine the mass of the nonlinear dynamic vibration absorber, the quantity, specifications and distribution of the wire rope vibration isolators, and determine the structural parameters of the wire rope vibration isolators to form a nonlinear dynamic vibration absorber structure installed on the top of the top support frame.

[0018] Furthermore, in step S1, the simplified coupled dynamics model is a five-degree-of-freedom model, which includes the frame mass. Quality of front panel and instrument panel assembly , quality of the top support frame Steering wheel system quality and the quality of nonlinear dynamic vibration absorbers ;

[0019] The simplified coupled dynamics model also includes the stiffness of the ground support. and damping Rigidity between the frame and the front panel and damping Stiffness between the frame and the roof guard and damping Rigidity between the top support frame and the front panel and damping Rigidity between the overhead guard and the steering wheel system and damping and the linear stiffness of the nonlinear dynamic vibration absorber Linear damping Nonlinear stiffness coefficient and nonlinear damping coefficient .

[0020] Furthermore, in step S3, the particle swarm size N is set to 50, and the maximum number of iterations Q is set to 100.

[0021] Furthermore, in step S3, the current parameter vector is substituted into the simplified coupled dynamics model for each particle, the system response is solved by numerical integration, and the objective function value is calculated based on the obtained system response.

[0022] Furthermore, in step S4, the structural parameters of the wire rope vibration isolator include at least one of the following: wire rope diameter, number of strands, winding method, number of coils, and winding helix tilt angle.

[0023] On the other hand, the present invention also provides a forklift steering wheel vibration suppression structure, which is designed by any of the above-described design methods and includes a top guard, a steering wheel system disposed in the driver's area below the top guard, and a nonlinear dynamic vibration absorber installed on the top of the top guard.

[0024] The nonlinear dynamic vibration absorber includes a mass block, multiple wire rope vibration isolators and connecting components. The multiple wire rope vibration isolators are spaced apart between the mass block and the top plate of the top support frame, and are symmetrically distributed around the center of gravity of the mass block.

[0025] The upper end of each wire rope vibration isolator is connected to the mass block through the connecting assembly, and the lower end of each wire rope vibration isolator is connected to the top plate of the top support frame through the connecting assembly, so that the mass block and the multiple wire rope vibration isolators together constitute a vibration subsystem installed on the top of the top support frame.

[0026] Furthermore, the connecting assembly includes an upper connecting seat, a lower connecting seat, and fastening bolts; the upper connecting seat is fixed to the mass block, and the lower connecting seat is fixed to the top plate of the top support frame;

[0027] Both ends of each wire rope vibration isolator are detachably connected to the upper connecting seat and the lower connecting seat via fastening bolts.

[0028] Furthermore, the number of the wire rope vibration isolators is 4 to 8.

[0029] Furthermore, the specifications of the multiple wire rope vibration isolators are the same, or the specifications of the multiple wire rope vibration isolators are different;

[0030] When multiple wire rope vibration isolators have different specifications, the wire rope vibration isolators of different specifications are alternately distributed at different positions around the mass block.

[0031] Furthermore, the mass block is a rigid structure made of steel or high-density metal material, and the ratio of the mass block's mass to the mass of the top support frame is 0.05 to 0.2.

[0032] The forklift steering wheel vibration suppression structure also includes multiple impact buffer pillars, which are disposed between the top plate of the overhead guard and the mass block. The impact buffer pillars are made of polyurethane material.

[0033] Compared with existing technologies, this invention establishes a five-degree-of-freedom vehicle coupled dynamics model that includes the steering wheel and roof guard, and directly takes minimizing the steering wheel vibration response as the optimization objective. It uses a particle swarm optimization algorithm to systematically optimize the nonlinear parameters of the vibration absorber. This goal-oriented reverse design method ensures that the designed vibration absorber can most effectively suppress the vibration of the steering wheel, the ultimate control target. This solves the problem that traditional methods are difficult to implement vibration control when the space around the steering wheel is limited, and also avoids the problem of poor results caused by blindly adding vibration absorbers.

[0034] Based on this, the present invention employs multiple wire rope vibration isolators to work in concert, making full use of their nonlinear stiffness and nonlinear damping characteristics, so that the vibration absorber can adaptively adjust its stiffness according to different vibration amplitudes, maintaining a high-efficiency vibration absorption effect over a wide frequency band. This not only overcomes the inherent defects of traditional linear vibration absorbers, such as narrow frequency band and failure after frequency drift, but also significantly improves energy dissipation efficiency through the damping superposition effect of multiple vibration isolators.

[0035] This invention utilizes the dry friction generated by the relative slippage between the strands of the wire rope as a damping source, eliminating the need for additional dampers. This design not only simplifies the overall structure and reduces manufacturing costs, but also avoids the problem of poor synergy between additional damping elements and elastic elements. Furthermore, since the wire rope vibration isolator itself is resistant to high and low temperatures, impact, and corrosion, the vibration absorber can maintain stable damping characteristics within a wide temperature range of -60℃ to 120℃, significantly extending its service life.

[0036] Furthermore, this invention places the vibration absorber on the top of the relatively spacious overhead frame, cleverly avoiding the narrow space around the steering wheel, making the design easy to implement. At the same time, the stability and impact resistance of the vibration subsystem are ensured by multiple symmetrically distributed steel wire rope vibration isolators and polyurethane impact buffer columns. The overall solution has the outstanding advantages of simple and reliable structure, low cost, good durability and convenient maintenance. Attached Figure Description

[0037] Figure 1 This is a schematic diagram showing the overall effect of the dynamic vibration absorber installed on the top support frame in an embodiment of the present invention;

[0038] Figure 2 This is a schematic diagram of the installation of the dynamic vibration absorber in an embodiment of the present invention;

[0039] Figure 3 This is a schematic diagram of the structure of the dynamic vibration absorber in an embodiment of the present invention;

[0040] Figure 4 Hysteresis loop diagram of the force-displacement relationship of the wire rope vibration isolator provided in the embodiment of the present invention;

[0041] In the diagram: 1. Mass block; 2. Fastening bolt; 21. Upper connecting seat; 22. Lower connecting seat; 3. Wire rope vibration isolator; 4. Top protection frame; 5. Impact buffer column. Detailed Implementation

[0042] To make the objectives, technical solutions, and advantages of this invention clearer, the invention will be further described in detail below with reference to the accompanying drawings. It should be understood that the various specific embodiments described herein are merely for explaining the invention and are not intended to limit the scope of protection of the invention.

[0043] Example 1:

[0044] This embodiment presents a design method for a forklift steering wheel vibration suppression structure. This method breaks away from the traditional approach of applying control only locally, aiming to precisely suppress severe steering wheel vibration under specific operating conditions from the perspective of the entire vehicle system. Establishing a precise dynamic analysis model that reflects the interaction mechanism of each component is the foundation for achieving the aforementioned reverse design. Specifically, this design method first requires establishing a simplified coupled dynamic model of the forklift. To facilitate a thorough analysis of the actual impact of vibration control measures on the forklift overhead guard 4 on the vibration of the remote forklift steering wheel, this embodiment constructs a simplified five-degree-of-freedom system model. This simplified coupled dynamic model covers core components crucial to the vibration transmission path, specifically including the bottom frame, the front panel and dashboard assembly as the intermediate transmission hub, the overhead guard 4 as the main support and vibration absorber mounting base, the remote steering wheel system as the final vibration suppression target, and the vibration mass corresponding to the nonlinear dynamic vibration absorber located on top of the overhead guard 4. To provide an accurate mathematical description and simulation solution for this system, the generalized coordinate vector of this five-degree-of-freedom system is defined as... In this generalized coordinate vector, Represents the displacement of the chassis. This represents the displacement of the front panel and instrument panel assembly. This represents the displacement of the top support frame 4. This represents the displacement of the steering wheel system, while This represents the displacement of the vibrating mass of the nonlinear dynamic vibration absorber located at the top of the top support frame 4. This coordinate definition allows complex spatial physical vibrations to be transformed into calculable mathematical matrix variables.

[0045] Based on this five-degree-of-freedom coupled dynamics model, it is necessary to further clarify the specific physical mapping relationships of the various lumped masses, stiffness, and damping parameters in the model. The model specifically includes five lumped masses, namely the frame mass... Quality of front panel and instrument panel assembly 4. Weight of the top support frame Steering wheel system quality and the quality of nonlinear dynamic vibration absorbers There are complex elastic and energy-dissipating connections between these physical masses. Specifically, this model comprehensively incorporates the stiffness of the ground support. and damping The primary transmission of road surface excitation is simulated, while the stiffness between the frame and the front leaf is introduced. and damping The rigidity between the frame and the overhead guard 4 and damping The stiffness between the top support frame 4 and the front plate and damping And the crucial stiffness between the overhead guard 4 and the steering wheel system. and damping Crucially, the nonlinear dynamic vibration absorber installed on the top support frame 4 possesses not only linear stiffness... and linear damping It also possesses a nonlinear stiffness coefficient due to its unique structural properties. and nonlinear damping coefficient These nonlinear parameters are what endow the structure with efficient vibration absorption capabilities that can adaptively adjust over a wide frequency band.

[0046] To substitute the above physical parameters into the computer for optimization, the system's motion differential equations must be derived. In this embodiment, the force generated by the nonlinear dynamic vibration absorber exhibits complex nonlinear characteristics, specifically expressed as follows:

[0047] ;

[0048] The relative displacement between the shock absorber and the vehicle frame is defined as To accurately characterize this nonlinear property, a polynomial expansion is used for Duffing-type nonlinear stiffness, specifically: This is to reflect the stiffness hardening or softening effect under large displacements; while for the nonlinear damping part, considering the complex influence of factors such as dry friction, there is a relationship... This allows the damping energy dissipation to be dynamically adjusted as the amplitude changes.

[0049] Based on the above force and motion analysis, the simplified macroscopic differential equation of motion in matrix form of the system is finally obtained, namely:

[0050] .

[0051] In this equation of motion, The mass matrix represents a five-degree-of-freedom system, with its diagonal elements corresponding to the masses of each component. Specifically, it is expanded as follows: .

[0052] at the same time, The damping matrix of the system represents the coupling situation of the various connection dampers, and its form is: .

[0053] Similarly, The stiffness matrix of the system represents the mutual constraints between the elastic connections of the components, and is expanded into the form of... .

[0054] Furthermore, the nonlinear force vector matrix Specifically used to characterize the force applied to the system by a nonlinear vibration absorber, it is manifested as .

[0055] External excitation vector It mainly consists of engine excitation or road excitation acting on the chassis, represented as ,in This is the sum of external excitations acting on the frame.

[0056] After establishing the complete equations of motion, the core of the design method in this embodiment shifts to the optimization process of the nonlinear vibration absorber parameters. Since the ultimate goal of this embodiment is to solve the steering wheel vibration problem that is directly perceptible to the driver, the steering wheel vibration response is directly used as the sole optimization object when constructing the objective function. The objective function under wideband excitation conditions is mathematically defined as follows: In this objective function, This represents the parameter vector of the nonlinear dynamic vibration absorber to be optimized, composed of various mechanical parameters of the absorber. Specifically, it is expressed as... ;and This represents the frequency response of the steering wheel displacement. Solving this objective function is essentially finding a way to make the steering wheel move within a specified target frequency band. The optimal set of vibration absorber parameters is used to minimize the maximum internal resonance peak value.

[0057] To efficiently find this set of parameters in a multidimensional parameter space and avoid getting trapped in local optima, this embodiment employs a particle swarm optimization algorithm on the vibration absorber parameter vector to be optimized. A rigorous iterative optimization process is performed. In the specific implementation flow of the particle swarm optimization algorithm:

[0058] First, the particle swarm is initialized. The particle swarm size N is set to 50, the maximum number of iterations Q is set to 100, and the position and velocity of the particles are randomly initialized within the parameter search range. The parameter search range is usually constrained by the actual physical limits of the forklift.

[0059] The core iterative solution process then begins. For each particle in the population, its current parameter vector is substituted into the simplified coupled motion differential equations established in the previous steps. The system response is then solved using numerical integration methods (such as the Runge-Kutta method), and the objective function is calculated based on the obtained steady-state system response. The specific value.

[0060] After evaluating the fitness of all particles, the algorithm will update the optimal position of each individual particle. and the global optimal position Then, the velocity and position of the particles are updated according to the core iterative formula of the particle swarm optimization algorithm. The specific velocity update formula is defined as follows:

[0061] ;

[0062] The position update formula is defined as follows:

[0063] ;

[0064] in, This represents the inertial weights used to balance global and local search capabilities. and The learning factor is used to control the learning step size. and These are random numbers uniformly distributed between the interval [0,1], used to increase the randomness of the search and the ability to escape local extrema. For the first The particle in the first The speed of 1-dimensional space, For the corresponding particle position, This represents the current iteration number.

[0065] The iterative process will continuously check whether the maximum number of iterations (100) has been reached or whether other stringent convergence accuracy conditions are met. If so, the optimal parameters obtained through optimization will be output. Otherwise, the objective function value will be recalculated and the iteration will continue until the global optimal solution that minimizes the steering wheel vibration response is found.

[0066] The optimal parameters obtained by the particle swarm optimization algorithm described above As a result, the design method eventually enters the physical mapping stage, which determines the specific structural parameters of the nonlinear dynamic vibration absorber, including the mass of the mass block 1 that directly determines the inertial force, and the number, specific specifications and spatial distribution of the wire rope vibration isolators 3 that determine the nonlinear stiffness and damping characteristics of the system.

[0067] To accurately match the nonlinear stiffness obtained through optimization and nonlinear damping It is necessary to systematically determine the various internal structural parameters of the wire rope vibration isolator 3. These structural parameters specifically include at least one of the following: wire rope diameter, number of strands, winding method, number of coils, and winding spiral tilt angle. By combining and adjusting these micro-geometric parameters, a nonlinear dynamic vibration absorber structure that can be seamlessly installed on the top of the top support frame 4 and accurately exert its vibration absorption effect is finally formed at the physical level.

[0068] In summary, this goal-oriented reverse parameter optimization design method overturns the outdated approach of blindly experimenting or relying on experience to select damper parameters. Through precise solutions of global coupled dynamics and global optimization using intelligent algorithms, the designed damper parameters mounted on the overhead guard directly address the core issues, efficiently attenuating the wide-frequency vibration energy transmitted to the steering wheel. This method is logically rigorous and highly targeted, significantly shortening the development cycle for new product damping systems and theoretically ensuring the wide-frequency adaptability of the damping structure and the ultimate improvement in the vehicle's NVH performance. It provides a solid and reliable basis for decision-making in the subsequent manufacturing of physical components.

[0069] Example 2:

[0070] Based on the reverse parameter design method in Embodiment 1 above, this embodiment proposes a physical forklift steering wheel vibration suppression structure. This physical structure avoids the extremely cramped installation space around the forklift steering wheel, which is difficult to modify on a large scale, and shifts the main battlefield for vibration reduction to the top of the overhead guard frame, which has a wide field of vision and ample space.

[0071] like Figure 1 As shown, the forklift steering wheel vibration suppression structure provided in this embodiment is mainly composed of three parts working together: the overhead guard 4, which serves as the main support base; the steering wheel system (not shown in the figure), located in the driver's area below the overhead guard 4 and awaiting vibration reduction control; and the core, nonlinear dynamic vibration absorber system, which is stably installed on top of the overhead guard 4. This macroscopic layout breaks away from the traditional design bias of local vibration reduction and achieves remote and precise vibration reduction control based on the principle of global dynamics.

[0072] like Figure 2 and Figure 3As shown, this nonlinear dynamic vibration absorber is a vibration subsystem assembled from a mass block 1, multiple specially designed wire rope vibration isolators 3, and matching connecting components. Multiple wire rope vibration isolators 3, which possess both elastic and damping characteristics, are evenly spaced between the lower surface of the mass block 1 and the upper surface of the top plate of the support frame 4, forming a highly effective vibration energy dissipation isolation zone. Furthermore, all these wire rope vibration isolators 3 must be symmetrically distributed around the physical center of gravity of the upper mass block 1. This arrangement ensures that when the vibration subsystem moves up and down or sways back and forth and side to side, the trajectory of the center of gravity always coincides with its geometric center, thus avoiding harmful additional torsional vibration or swaying caused by eccentric torque, and guaranteeing the high efficiency and stability of the entire vibration absorption process.

[0073] To ensure that the components do not loosen or fail under long-term, severe alternating load impacts, each wire rope vibration isolator 3 is equipped with connecting components at both ends. These connecting components specifically include an upper connecting seat 21, a lower connecting seat 22, and fastening bolts 2 for locking. During installation, the upper connecting seat 21 is fixed to a specific position on the bottom of the mass block 1 using the fastening bolts 2, while the corresponding lower connecting seat 22 is also aligned and fixed to the top bearing metal plate of the top support frame 4 using the fastening bolts 2, thus completing the fixed installation of the wire rope vibration isolator 3. This fully bolt-detachable modular connection design not only greatly facilitates the initial factory assembly, but more importantly, it provides convenient maintenance space for later replacement of wire rope vibration isolators 3 of different specifications for different working conditions. Operators only need to loosen the fastening bolts 2 to quickly adjust the vibration absorption characteristics or replace old parts, resulting in low maintenance costs and convenient operation. Through the above connection, the heavy mass block 1 above and the multiple wire rope vibration isolators 3 in the middle that act as a hub to dissipate energy together constitute a complete vibration subsystem installed on the top of the top support frame 4, ready to respond to and absorb the vibration energy of the main structure below.

[0074] Regarding the number of wire rope vibration isolators 3, to ensure the stability of the support and the full superposition of the energy dissipation efficiency of the nonlinear hysteresis loop, this embodiment specifies that the number of wire rope vibration isolators 3 should be between 4 and 8. Optionally, for example, when targeting small electric forklifts or light internal combustion forklifts with relatively small load capacities and weak engine idling excitation forces, four wire rope vibration isolators 3 can be symmetrically selected, thus meeting vibration reduction requirements while maximizing cost control; for another example, when targeting medium-tonnage standard internal combustion forklifts with frequent daily operations, six wire rope vibration isolators 3 can be selected, achieving a good balance between cost and efficiency; yet another example, when facing heavy-duty mining forklifts with large engine displacements and extremely severe low-frequency, large-amplitude impacts transmitted from the chassis to the overhead guard 4, a maximum of eight wire rope vibration isolators 3 must be configured (e.g., Figure 2 and Figure 3As shown in the figure, by superimposing the damping characteristics of these 8 vibration isolators, the overall energy dissipation efficiency of the system is improved, thereby suppressing the resonance peak of the main structure.

[0075] Secondly, regarding the mass distribution of mass block 1, mass block 1 must be made of a material with a high-rigidity structure, typically using high-quality steel or high-density special metal materials integrally molded to ensure that its structural modal frequency as an inertial block is much higher than the operating frequency, thus preventing self-resonance deformation. Crucially, to ensure that the nonlinear vibration absorber has sufficient reverse inertial suppression force while strictly preventing irreversible fatigue damage or heavy additional load on the roof support frame 4 and the main structure of the cab below due to excessive added mass, this embodiment strictly controls the ratio of the mass of mass block 1 to the mass of the roof support frame 4 within the range of 0.05 to 0.2. Optionally, for example, when the stiffness of the top support frame 4 is moderate and the desired additional weight of the system is minimal, the mass ratio can be designed to a minimum of 0.05; for example, in most common working conditions, to pursue overall cost-effectiveness, the mass ratio can be set to an intermediate value of 0.125; for example, when the top support frame 4 is made of extremely robust thick-walled steel, with extremely high load-bearing capacity redundancy and requiring extremely strong counter-phase suppression force to counteract ultra-large amplitude, the mass ratio can be amplified to a maximum value of 0.2 to obtain the most powerful inertial vibration reduction effect.

[0076] Furthermore, to achieve a more intelligent and wide-frequency adaptive vibration reduction response, the specifications of the multiple wire rope vibration isolators 3 are also specially selected. In a conventional basic configuration, the specifications of the multiple wire rope vibration isolators 3 can be exactly the same. In this case, since multiple vibration isolators with the same nonlinear hysteresis properties work together, the overall nonlinear adjustment range of the sub-vibration system can be broadened laterally by the advantage of quantity.

[0077] However, in another conventional basic configuration, the specifications of multiple wire rope vibration isolators 3 can be deliberately designed to be completely different. When different specifications are used, in order to maintain the torque balance of the system, this embodiment arranges wire rope vibration isolators 3 of different specifications to be alternately and symmetrically distributed at different positions around the mass block 1. This multi-specification staggered design concept enables the stiffness characteristics of the sub-vibration system to exhibit a stepped, multi-segment smooth nonlinear change when subjected to displacement excitation of different intensities. This further optimizes the vibration absorption bandwidth of the entire system by a factor of two, enabling it to maintain the resonant frequency of the main structure in an extremely wide frequency range of 2 to 30 Hz. This meets the wide-frequency vibration requirements of the forklift engine under the condition of drastic frequency change from idle speed to maximum speed, and solves the weakness of traditional linear spring vibration absorbers that will completely fail when the main frequency fluctuates by even ±1 Hz.

[0078] like Figure 2 and Figure 3As shown, the forklift steering wheel vibration damping structure also includes multiple impact buffer pillars 5. These robust impact buffer pillars 5 are arranged between the top plate of the overhead guard 4 and the upper mass block 1, and one end of each impact buffer pillar 5 can be fixed to the lower part of the mass block 1 by bolts. They are typically located inside the area enclosed by the wire rope vibration isolators 3 or at key stress points around them. These impact buffer pillars 5 can be integrally cast from polyurethane material with high toughness and high energy absorption properties. In the event of a sudden and severe impact that causes the deformation of the wire rope vibration isolator 3 to exceed the safety threshold, or if a heavy object unfortunately falls heavily from above the overhead guard 4 and directly strikes the vibration absorber, these tough polyurethane impact buffer columns 5 will instantly intervene and bottom out, bearing most of the impact kinetic energy. This prevents the heavy mass block 1 from directly and hard impacting the fragile top of the overhead guard 4, causing structural tearing and damage. This greatly increases the ability of the dynamic vibration absorber to withstand catastrophic impacts, avoids fastener breakage or loosening and disintegration of the entire subsystem caused by severe overload vibration, and ensures the absolute safety of the forklift's safety components.

[0079] When the forklift overhead guard 4 begins to experience complex high-frequency or low-frequency vibrations under conditions such as engine idling, this harmful vibration energy is quickly transmitted through the lower connecting seat 22 to the multiple wire rope vibration isolators 3. When the main structure experiences extremely small vibration amplitudes due to slight idling (e.g., displacement less than 0.5mm), the wire rope vibration isolators 3 are in a high initial stiffness state without large slippage (e.g., transient stiffness can reach 1000N / mm). At this time, the entire subsystem can respond quickly to these high-frequency micro-vibrations and suppress them. However, once the main structure experiences these vibrations, the vibration energy is transmitted through the lower connecting seat 22 to the multiple wire rope vibration isolators 3. When the structure enters a state such as climbing or severe resonance, causing a sudden increase in vibration amplitude (e.g., displacement greater than 0.5 mm), the steel wires inside the wire rope vibration isolator 3 immediately exceed the critical slip force, resulting in large-scale relative slip friction. Its macroscopic stiffness immediately shows a softening and decreasing trend (e.g., rapidly dropping back to 500 N / mm). This nonlinear stiffness softening directly causes the natural frequency of the subsystem to drift, thus automatically avoiding secondary resonance with the main structure. Furthermore, during the slip process, the intense dry friction inside the wire rope dissipates a large amount of vibration energy through the hysteresis loop. The coordinated work of multiple vibration isolators not only amplifies this vibration dissipation capability many times over, but also, in conjunction with the mass block 1, provides a continuous and extremely stable anti-phase inertial suppression force, so that the vibration energy of the main structure generated from bottom to top excitation is canceled out by the anti-phase motion from top to bottom. Figure 4 As shown, by using dry friction damping inside the steel wire rope to replace the traditional damping element that is prone to aging, the vibration amplitude of the steering wheel is reduced by more than 30%. Moreover, due to the stable properties of its purely mechanical structure, the whole system can still maintain reliable vibration reduction performance in harsh environments such as extreme cold or high temperature, which improves driving comfort and protects the headlight accessories on the roof guard 4. It has positive significance in terms of engineering practical value.

Claims

1. A design method for a vibration suppression structure for forklift steering wheels, characterized in that, Includes the following steps: S1. Establish a simplified coupled dynamics model of the forklift, which includes at least the frame, front panel and dashboard assembly, overhead guard (4), steering wheel system, and the vibration mass corresponding to the nonlinear dynamic vibration absorber installed on top of the overhead guard (4), and define the generalized coordinate vector of the system as follows: ; in, , , , , The displacements of the vibration masses of the vehicle frame, front panel and instrument panel assembly, roof guard (4), steering wheel system and nonlinear dynamic shock absorber, respectively; S2. Construct an objective function that optimizes the steering wheel vibration response. The objective function under wideband excitation conditions is defined as follows: ; in, Here is the parameter vector of the vibration absorber to be optimized. ; This represents the frequency domain response to the steering wheel displacement. For the target frequency band; S3. The particle swarm optimization algorithm is used to optimize the vibration absorber parameter vector. Iterative optimization is performed. In each iteration, the current parameter vector is substituted into the simplified coupled dynamics model to solve for the system response and calculate the objective function value. S4. Based on the optimal parameter results obtained from the optimization, determine the mass of the nonlinear dynamic vibration absorber mass block (1), the quantity, specifications and distribution of the wire rope vibration isolators (3), and determine the structural parameters of the wire rope vibration isolators (3) to form a nonlinear dynamic vibration absorber structure installed on the top of the top support frame (4).

2. The design method according to claim 1, characterized in that: In step S1, the simplified coupled dynamics model is a five-degree-of-freedom model, which includes the chassis mass. Quality of front panel and instrument panel assembly , Top support frame (4) quality Steering wheel system quality and the quality of nonlinear dynamic vibration absorbers ; The simplified coupled dynamics model also includes the stiffness of the ground support. and damping Rigidity between the frame and the front panel and damping Stiffness between the frame and the roof guard (4) and damping The rigidity between the top support frame (4) and the front plate and damping The stiffness between the overhead guard (4) and the steering wheel system and damping and the linear stiffness of the nonlinear dynamic vibration absorber Linear damping Nonlinear stiffness coefficient and nonlinear damping coefficient .

3. The design method according to claim 1, characterized in that: In step S3, the particle swarm size N is set to 50, and the maximum number of iterations Q is set to 100.

4. The design method according to claim 1, characterized in that: In step S3, the current parameter vector is substituted into the simplified coupled dynamics model for each particle, the system response is solved by numerical integration, and the objective function value is calculated based on the obtained system response.

5. The design method according to claim 1, characterized in that: In step S4, the structural parameters of the wire rope vibration isolator (3) include at least one of the following: wire rope diameter, number of strands, winding method, number of coils, and winding helix tilt angle.

6. A forklift steering wheel vibration suppression structure, designed by the design method described in any one of claims 1-5, characterized in that, Includes a roof guard (4), a steering wheel system located in the driver's area below the roof guard (4), and a nonlinear dynamic vibration absorber mounted on the top of the roof guard (4); The nonlinear dynamic vibration absorber includes a mass block (1), multiple wire rope vibration isolators (3) and a connecting assembly. The multiple wire rope vibration isolators (3) are spaced apart between the mass block (1) and the top plate of the top support frame (4) and are symmetrically distributed around the center of gravity of the mass block (1). The upper end of each wire rope vibration isolator (3) is connected to the mass block (1) through the connecting assembly, and the lower end of each wire rope vibration isolator (3) is connected to the top plate of the top support frame (4) through the connecting assembly, so that the mass block (1) and the multiple wire rope vibration isolators (3) together constitute a vibration subsystem installed on the top of the top support frame (4).

7. The forklift steering wheel vibration suppression structure according to claim 6, characterized in that: The connecting assembly includes an upper connecting seat (21), a lower connecting seat (22), and a fastening bolt (2); the upper connecting seat (21) is fixed to the mass block (1), and the lower connecting seat (22) is fixed to the top plate of the top support frame (4); Each of the wire rope vibration isolators (3) is detachably connected at both ends to the upper connecting seat (21) and the lower connecting seat (22) by the fastening bolts (2).

8. The forklift steering wheel vibration suppression structure according to claim 6, characterized in that: The number of the wire rope vibration isolators (3) is 4 to 8.

9. The forklift steering wheel vibration suppression structure according to claim 6, characterized in that: The multiple wire rope vibration isolators (3) have the same specifications, or the multiple wire rope vibration isolators (3) have different specifications; When multiple wire rope vibration isolators (3) have different specifications, the wire rope vibration isolators (3) with different specifications are alternately distributed at different positions around the mass block (1).

10. The forklift steering wheel vibration suppression structure according to claim 6, characterized in that: The mass block (1) is a rigid structure made of steel or high-density metal material, and the mass ratio of the mass block (1) to the mass of the top support frame (4) is 0.05 to 0.

2. The forklift steering wheel vibration suppression structure also includes multiple impact buffer pillars (5), which are disposed between the top plate of the top guard (4) and the mass block (1). The impact buffer pillars (5) are made of polyurethane material.

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