Damping adjusting method for suspension system

Abstract

The invention belongs to the technical field of suspension vibration reduction, and particularly relates to a suspension system damping adjusting method which comprises the following steps: taking a single-wheel suspension system as a research object, deducing a fixed point of frequency domain vibration response, and giving a suspension motion differential equation set; a function relationship between sprung mass displacement and unsprung mass displacement is deduced through Laplacian transformation, a fixed point expression of sprung mass displacement response, a frequency domain fixed point expression of unsprung mass displacement response and a frequency domain fixed point expression of suspension relative travel are respectively solved through programming, compiling of a control algorithm is carried out, and the suspension relative travel is controlled. The comparative analysis of the vibration response is carried out, and the effectiveness of the algorithm is verified.

Description

Technical Field

[0001] The invention belongs to the technical field of suspension vibration reduction, and in particular relates to a suspension system damping adjustment method. Background Art

[0002] The suspension system is usually composed of elastic elements, damping elements and guide mechanisms. Its function is to bear the weight of the vehicle body and the power transmission system; to transmit, bear and alleviate the various reaction forces and torques acting on the moving mechanism by the road surface; to reduce the impact and vibration of the vehicle body caused by uneven road excitation; and to provide a relatively stable support platform for the upper system while ensuring the off-road maneuverability of the vehicle.

[0003] With the continuous development of technology, suspension technology has changed from traditional linear characteristics to strong nonlinear characteristics, which can double the elastic buffer force and dynamic potential at the end of the stroke. Experiments have shown that the average off-road speed of the vehicle can be increased by at least 30%. For the representative of nonlinear suspension - oil-gas suspension, the off-road maneuverability of the equipment can be significantly improved even in a passive state. Therefore, with oil-gas suspension as the basic platform, further research on intelligent control has become a development necessity.

[0004] For a long time, there have been two mainstream directions in the intelligent control of suspension technology: semi-active and active control. As an active suspension, active control can adjust the actuation force in real time according to road excitation, which consumes a lot of engine power. Domestic and foreign scholars have carried out a lot of basic theoretical research and experimental verification; semi-active suspension usually includes functions such as adjustable damping, adjustable stiffness and adjustable vehicle posture, which does not consume a lot of power. It is a controllable suspension system with great development potential and is easier to enter the practical stage. Traditional semi-active suspension is mostly focused on the research of adjustable damping, but the adjustment effect is limited. Therefore, it is necessary to propose a damping characteristic frequency domain hybrid control algorithm, which can perform segmented control in the full frequency domain and determine a reasonable control algorithm in each frequency band, so as to achieve the effect of improving the driving smoothness and handling stability simultaneously. Summary of the invention

[0005] 1. Technical issues to be resolved

[0006] The technical problem to be solved by the present invention is: how to provide a suspension system damping adjustment method which is required to be able to effectively attenuate the vibration acceleration response of the vehicle body.

[0007] (II) Technical solution

[0008] In order to solve the above technical problems, the present invention provides a suspension system damping adjustment method, which solves and obtains the fixed point expression of the sprung mass displacement response according to the frequency-band hybrid control method:

[0009]

[0010] In the formula, ω 1 is the frequency of the sprung mass displacement response fixed point 1, rad / s; ω 2 is the frequency of the sprung mass displacement response fixed point 2, rad / s; ω 3 is the frequency of the sprung mass displacement response fixed point 3, rad / s; m s is the sprung mass, kg; m u is the unsprung mass; x s is the sprung mass displacement, m; x u is the unsprung mass displacement, m; k s is the suspension stiffness characteristic, N / m; k u is the wheel stiffness characteristic.

[0011] Among them, ω 2 and ω 3 and suspension stiffness k s is related to the suspension damping c s It doesn't matter, but 1 and suspension stiffness k s and suspension damping c s are irrelevant and are completely determined by the wheel stiffness k u and wheel mass m u Determine; when the excitation frequency ω is ω<ω 1 or 2 <ω<ω 3 In the range of ω, the acceleration transmission rate of the vertical vibration of the vehicle body decreases with the increase of suspension damping, that is, the inhibitory effect of increasing damping on suspension resonance is particularly significant; when the excitation frequency ω is within ω 1 <ω<ω 2 or ω>ω 3 When the vehicle body is within the specified range, the vertical vibration acceleration transmission rate increases with the increase of suspension damping, which will cause the vehicle's ride smoothness to deteriorate. This is why the passive suspension system needs to make compromises in order to improve the vehicle's ride smoothness.

[0012] Among them, the frequency domain fixed point expression for solving the unsprung mass displacement response is:

[0013]

[0014] In the formula, ω 4 is the frequency of the unsprung mass displacement response fixed point 4, rad / s; ω 5 is the frequency of the unsprung mass displacement response fixed point 5, rad / s.

[0015] Among them, ω 4 and ω 5 and suspension stiffness k sThat is, at this frequency, the wheel dynamic deformation is still affected by the suspension force; when the excitation frequency ω is within ω<ω 4 or ω>ω 5 When the excitation frequency ω is within the range of ω, the wheel dynamic deformation transmission rate decreases with the increase of suspension damping. Similarly, increasing damping is beneficial to suppress the resonance of the vehicle. 4 <ω<ω 5 When the transmission rate of wheel dynamic deformation is within the range, the wheel dynamic deformation increases with the increase of suspension damping, which will cause the vehicle's road adhesion to deteriorate, indicating that the passive suspension system cannot optimize the wheel dynamic deformation as a whole in the entire frequency band.

[0016] Among them, the frequency domain fixed point expression for solving the relative travel of the suspension is:

[0017]

[0018] Among them, ω 6 is the frequency of the fixed point 4 of the suspension relative travel, rad / s.

[0019] Among them, ω 6 and suspension stiffness k s and suspension damping c s They are all unrelated, indicating that at this frequency, the suspension dynamic stroke transfer rate is not affected by the suspension force and is a constant value; in general, the suspension dynamic stroke decreases with the increase of suspension damping. Similarly, increasing damping is beneficial to suppress vehicle resonance.

[0020] The method designs a frequency divider in the following form:

[0021]

[0022] Where: is the rate of change of sprung mass acceleration; α is an intermediate variable, where α=(2πf cross )2,f cross Take the first fixed point frequency

[0023] Assume that Where f>0, then Where f is the actual input frequency, substituting it into the formula yields:

[0024] γ=((2πf) 4 -α 2 )sin 2 (2πft)

[0025] Since α=(2πf 1 ) 2 , then when 0 <f<f 1 When γ<0, and when f>f 1When γ>0, an improved frequency domain hybrid damping control method is given, which is expressed as:

[0026]

[0027] Where: c f is the adjustable damping value of the suspension; c fmax Maximum adjustable damping for the suspension.

[0028] Among them, the method establishes the comprehensive performance evaluation index of the vehicle suspension system:

[0029]

[0030] Where ζ and ξ are the weight distribution coefficients of sprung mass acceleration and wheel displacement, respectively, and ζ+ξ=1. The adaptive adjustment of the weight distribution coefficient is realized by fuzzy control method. Q RP_P and Q RP are the root mean square values ​​of the sprung vibration acceleration of the passive and controllable suspension, Q RA_P and Q RA are the root mean square values ​​of the dynamic displacements of the passive and controllable suspension wheels, respectively.

[0031] The method can perform segmented control in the entire frequency domain and determine a reasonable control algorithm in each frequency band, thereby achieving the effect of simultaneously improving driving smoothness and handling stability.

[0032] Among them, the method derives the suspension dynamics equation, finds the frequency domain fixed point that does not change with the damping coefficient, and derives it analytically. The frequency domain space is divided by the fixed point, and the law of unidirectional change of acceleration vibration response with damping is found in each frequency domain interval, so as to formulate a reasonable control algorithm to achieve the purpose of effectively suppressing the vibration influence in the whole frequency domain. Through comparison and verification with other control methods, it has a more obvious advantage in suppressing vibration response and has broad application prospects.

[0033] (III) Beneficial effects

[0034] Compared with the prior art, the present invention has the following beneficial effects:

[0035] When a vehicle is driving on a random road, the excitation frequency domain involved is very wide. Traditional algorithms such as optimal control, skylight control, and acceleration control have limited effects on improving the comprehensive driving performance of the vehicle. The method of the present invention derives the suspension dynamics equation, finds the frequency domain fixed point that does not change with the damping coefficient, and derives it analytically. The frequency domain space is divided by the fixed point, and the law of unidirectional change of acceleration vibration response with damping is found in each frequency domain interval, so as to formulate a reasonable control algorithm to achieve the purpose of effectively suppressing the influence of vibration in the entire frequency domain. By comparing and verifying with other control methods, it has a more obvious advantage in suppressing vibration response and has broad application prospects. BRIEF DESCRIPTION OF THE DRAWINGS

[0036] Figure 1 Design process for oil-gas suspension system solution;

[0037] Figure 2 Calculation process for characteristics of oil-gas suspension system;

[0038] Figure 3 Optimize the matching process for nonlinear suspension system parameters;

[0039] Figure 4 is the acceleration response of the vehicle body center of mass vibration at different natural frequencies;

[0040] Figure 5 is the acceleration response of the vehicle center of mass vibration with different damping ratios;

[0041] Figure 6 The vibration acceleration response of sprung mass for different stiffness control methods

[0042] Figure 7 Suspension relative travel for different stiffness control methods;

[0043] Figure 8 The acceleration response of unsprung mass vibration for different stiffness control methods;

[0044] Fig. 9 It is a single wheel suspension dynamics model;

[0045] Fig.10 is the amplitude-frequency characteristic of the sprung mass displacement excitation on the road surface;

[0046] Fig.11 is the amplitude-frequency characteristic of the unsprung mass displacement on the road surface excitation;

[0047] Fig.12 The amplitude-frequency characteristic of the road surface excitation caused by the relative travel of the suspension;

[0048] Fig.13 The frequency domain characteristics of sprung acceleration are compared;

[0049] Fig.14Comparison of sprung acceleration time domain characteristics. DETAILED DESCRIPTION

[0050] In order to make the purpose, content, and advantages of the present invention more clear, the specific implementation methods of the present invention are further described in detail below in conjunction with the accompanying drawings and examples.

[0051] In order to solve the above technical problems, the present invention provides a suspension system damping adjustment method, which solves and obtains the fixed point expression of the sprung mass displacement response according to the frequency-band hybrid control method:

[0052]

[0053] In the formula, ω 1 is the frequency of the sprung mass displacement response fixed point 1, rad / s; ω 2 is the frequency of the sprung mass displacement response fixed point 2, rad / s; ω 3 is the frequency of the sprung mass displacement response fixed point 3, rad / s; m s is the sprung mass, kg; m u is the unsprung mass; x s is the sprung mass displacement, m; x u is the unsprung mass displacement, m; k s is the suspension stiffness characteristic, N / m; k u is the wheel stiffness characteristic.

[0054] Among them, ω 2 and ω 3 and suspension stiffness k s is related to the suspension damping c s It doesn't matter, but 1 and suspension stiffness k s and suspension damping c s are irrelevant and are completely determined by the wheel stiffness k u and wheel mass m u Determine; when the excitation frequency ω is ω<ω 1 or 2 <ω<ω 3 In the range of ω, the acceleration transmission rate of the vertical vibration of the vehicle body decreases with the increase of suspension damping, that is, the inhibitory effect of increasing damping on suspension resonance is particularly significant; when the excitation frequency ω is within ω 1 <ω<ω 2 or ω>ω 3 When the vehicle body is within the specified range, the vertical vibration acceleration transmission rate increases with the increase of suspension damping, which will cause the vehicle's ride smoothness to deteriorate. This is why the passive suspension system needs to make compromises in order to improve the vehicle's ride smoothness.

[0055] Among them, the frequency domain fixed point expression for solving the unsprung mass displacement response is:

[0056]

[0057] In the formula, ω 4 is the frequency of the unsprung mass displacement response fixed point 4, rad / s; ω 5 is the frequency of the unsprung mass displacement response fixed point 5, rad / s.

[0058] Among them, ω 4 and ω 5 and suspension stiffness k s That is, at this frequency, the wheel dynamic deformation is still affected by the suspension force; when the excitation frequency ω is within ω<ω 4 or ω>ω 5 When the excitation frequency ω is within the range of ω, the wheel dynamic deformation transmission rate decreases with the increase of suspension damping. Similarly, increasing damping is beneficial to suppress the resonance of the vehicle. 4 <ω<ω 5 When the transmission rate of wheel dynamic deformation is within the range, the wheel dynamic deformation increases with the increase of suspension damping, which will cause the vehicle's road adhesion to deteriorate, indicating that the passive suspension system cannot optimize the wheel dynamic deformation as a whole in the entire frequency band.

[0059] Among them, the frequency domain fixed point expression for solving the relative travel of the suspension is:

[0060]

[0061] Among them, ω 6 is the frequency of the fixed point 4 of the suspension relative travel, rad / s.

[0062] Among them, ω 6 and suspension stiffness k s and suspension damping c s They are all unrelated, indicating that at this frequency, the suspension dynamic stroke transfer rate is not affected by the suspension force and is a constant value; in general, the suspension dynamic stroke decreases with the increase of suspension damping. Similarly, increasing damping is beneficial to suppress vehicle resonance.

[0063] The method designs a frequency divider in the following form:

[0064]

[0065] Where: is the rate of change of sprung mass acceleration; α is an intermediate variable, where α=(2πf cross ) 2 , f cross Take the first fixed point frequency

[0066] Assume that Where f>0, then Where f is the actual input frequency, substituting it into the formula yields:

[0067] γ=((2πf) 4 -α 2 )sin 2 (2πft)

[0068] Since α=(2πf 1 ) 2 , then when 0 <f<f 1 When γ<0, and when f>f 1 When γ>0, an improved frequency domain hybrid damping control method is given, which is expressed as:

[0069]

[0070] Where: c f is the adjustable damping value of the suspension; c fmax Maximum adjustable damping for the suspension.

[0071] Among them, the method establishes the comprehensive performance evaluation index of the vehicle suspension system:

[0072]

[0073] Where ζ and ζ are the weight distribution coefficients of sprung mass acceleration and wheel displacement, respectively, and ζ+ξ=1. The adaptive adjustment of the weight distribution coefficient is realized by fuzzy control method. Q PR_P and Q RP are the root mean square values ​​of the sprung vibration acceleration of the passive and controllable suspension, Q RP_P and Q RA are the root mean square values ​​of the dynamic displacements of the passive and controllable suspension wheels, respectively.

[0074] The method can perform segmented control in the entire frequency domain and determine a reasonable control algorithm in each frequency band, thereby achieving the effect of simultaneously improving driving smoothness and handling stability.

[0075] Among them, the method derives the suspension dynamics equation, finds the frequency domain fixed point that does not change with the damping coefficient, and derives it analytically. The frequency domain space is divided by the fixed point, and the law of unidirectional change of acceleration vibration response with damping is found in each frequency domain interval, so as to formulate a reasonable control algorithm to achieve the purpose of effectively suppressing the vibration influence in the whole frequency domain. Through comparison and verification with other control methods, it has a more obvious advantage in suppressing vibration response and has broad application prospects.

[0076] Example 1

[0077] In this embodiment, the suspension system not only needs to be designed with a highly compact integrated structure, but also needs to carry out special dynamic modeling research to clarify the influence of parameters such as wheel rubber deformation, positioning parameters and hard point coordinates of various structural parts, and cylinder layout on the smooth transmission of suspension force, so as to avoid the problem of excessive lateral force and eccentric wear of the piston and hinge point, so as to achieve smooth and efficient transmission of ground impact excitation.

[0078] On the basis of effectively solving the stress problem of suspension system structural parts, it is also necessary to sort out the complete suspension system design process and method, and formulate a systematic evaluation index system, so as to ensure the reliability and stability of the actuator while achieving efficient isolation of ground impact excitation. This requires the reasonable matching of elasticity and damping characteristics to effectively suppress the vibration energy from the outside world. The function of the elastic element is mainly as an energy storage device, usually through the deformation of the spring or accumulator to achieve the purpose of absorbing part of the vibration energy and realize the buffering function, but it cannot dissipate the energy and needs to be released in the next vibration cycle. The damping element throttles the oil at high speed through the damping valve system, which can convert the vibration energy into heat energy for dissipation. The elasticity and damping characteristics of the passive suspension have been determined at the beginning of the design. Usually, a compromise design is made based on various working conditions, striving to meet the use requirements under different driving conditions. However, once the characteristic parameters are determined at the beginning of the design, they cannot be adjusted, and it is impossible to achieve the best use state on various road surfaces. Therefore, on the basis of formulating the suspension system design methods and processes, it is necessary to further carry out the optimization matching and coordinated control research of the suspension characteristics, which is also the key to fully realizing the buffering and vibration isolation performance. It is also necessary to clarify the best matching scheme of suspension elasticity and damping characteristics under different road conditions and vehicle speed conditions, so as to achieve the purpose of significantly reducing the vehicle vibration response.

[0079] Suspension system optimization matching flow chart as follows Figure 1As shown in the figure, first, according to the overall layout requirements, the center of gravity of the vehicle load is determined, the action system is preliminarily laid out, and the wheel system layout and the suspension guiding mechanism layout are carried out separately. In terms of the wheel system layout, the position of the wheel system is preliminarily arranged according to the principle that the center of gravity of the load and the elastic center should coincide as much as possible. Usually, it is grouped by front and rear. Off-road vehicles are configured according to the front and rear wheel load distribution ratio of 6:4 to ensure the ground contact ability of the front wheels during high-speed off-road driving. To facilitate the setting of characteristic parameters, first, a random road surface excitation is generated, and an equivalent linear dynamic model of the vehicle is established. The road surface model is used as the input of the equivalent linear dynamic model to carry out the suspension characteristic check and analysis. The suspension equilibrium position stiffness and damping characteristic parameters are input into the calculation program in the form of arrays respectively, and an array combination is carried out to carry out the dynamic matching calculation. The expected maximum driving speed is input, and the key vibration response conditions such as the mass above the spring, the mass below the spring, the relative suspension travel, and the vehicle body stability are monitored. The time-domain and frequency-domain response data are calculated respectively, and the root mean square value and the maximum vibration response data are used as the basis for evaluation to judge whether the vibration threshold is met. According to experience, it is set that the root mean square value of the vibration response of the mass above the spring acceleration should be less than 0.3g, the maximum impact acceleration vibration response should be less than 3.5g, the suspension vibration response travel should not be greater than the full suspension travel to avoid hitting the stopper, and the dynamic load of the mass below the spring should be less than the mass below the spring itself to avoid the driving safety problem of the wheel leaving the ground. A thermodynamics model of the suspension system is established, and at the same time, the maximum driving speed and the road surface model are input into the thermodynamics model to realize the joint operation of the vehicle equivalent linear dynamic model and the thermal model. The thermal equilibrium temperature T < T0 = 150°C is used as the evaluation threshold to carry out the optimization and matching of the stiffness and damping parameters to ensure that the suspension temperature will not be too high and cause seal failure when the vehicle is driving at high speed on a complex road surface. Taking the suspension thermal equilibrium threshold as the threshold for the optimization and matching of characteristic parameters is proposed for the first time and has played an important guiding role in the project development process. Through the comprehensive evaluation of the vehicle vibration threshold and the suspension thermal equilibrium temperature threshold, after the model is repeatedly iterated and calculated to meet the threshold requirements, the suspension damping ratio can be preliminarily determined.

[0080] In terms of the suspension guiding mechanism layout, according to the layout of the action system, the 1. pull arm length, 2. balance elbow length, 3. distance from the balance elbow axis to the cylinder center line, 4. pull arm installation angle, and 5. link length are preliminarily determined, and a suspension leverage ratio calculation module is generated, and finally a mathematical model of the oil-gas suspension is generated. The determined suspension damping ratio is input into the mathematical model of the oil-gas suspension to guide the design of the damping valve and generate a nonlinear damping characteristic. At the same time, the calculated axle loads of each wheel are input into the mathematical model of the oil-gas suspension to carry out the generation calculation of the nonlinear elastic characteristic. The calculated nonlinear elastic characteristic and damping characteristic curves are input into the suspension thermal model to complete the calculation of the thermal equilibrium temperature rise characteristic curve.

[0081] The nonlinear elastic characteristics and damping characteristics, as well as the generated road surface model, are input into the nonlinear dynamic model of the whole vehicle for further vibration response verification calculation. The obtained vibration response results can be used as the load of key components of the action system, for reliability analysis of key parts, and to guide the engineering design of the actuator. Comprehensively evaluate whether the vehicle body vibration response characteristics, suspension thermal balance temperature rise characteristic curve, and key parts reliability analysis results meet the design index requirements. If they do, the suspension matching optimization process ends. If they do not meet the requirements, the program returns to the first step of the action system layout and re-carries out iterative optimization calculations until the index requirements are met and the best solution is matched.

[0082] To further optimize the suspension system structure and design an integrated solution, we first need to conduct a qualitative analysis of the main components of the suspension. Starting from the overall layout of the vehicle, we can determine that the main design elements of the suspension system include the balance elbow, rocker arm, connecting rod, oil and gas spring and other components. Then, combined with the suspension system design theory, we will conduct an engineering design, considering the rigid constraint connection between the balance elbow and the vehicle body of the integrated suspension system and the setting of the slewing pair. The design process is as follows: Figure 2 shown.

[0083] According to the overall layout requirements and the load distribution of each wheel position, the requirements of the dynamic and static travel, static balance position, etc. of the suspension vibration isolation system are first refined, and the vertical stiffness of the static balance position of the nonlinear suspension is checked and calculated by load and static travel, and the natural frequency is set between 1-1.3Hz. The reason for setting it in this range is mainly to take into account the particularity of the nonlinear characteristics of the suspension. It can not only ensure that the suspension has low stiffness and good ride comfort when moving near the balance position, but also when the movement travel is increased, with the nonlinear increase of the elastic force value, it can also generate sufficient dynamic specific potential energy at the end of the travel to prevent the phenomenon of rigid collision with the limiter.

[0084] Then, the oil-gas spring characteristics matching and the hard point coordinates of the guide mechanism are designed respectively. Kinematic simulation analysis is carried out with the layout space and suspension geometric kinematics as constraints, and the hard point design of the suspension vibration isolation system is carried out. The constructed dynamic model is used for iterative optimization. The optimal hard point coordinates are determined with the minimum lateral force on the piston as the evaluation criterion. Then the guide mechanism is designed according to the hard point coordinates, and the topological optimization of the guide mechanism is carried out through finite element analysis. At the same time, the oil-gas spring characteristics are matched, and the key parameters such as the oil-gas spring cylinder diameter, piston stroke, and air chamber volume are determined according to the design requirements such as bearing capacity and frequency deviation. The corresponding nonlinear stiffness model and adjustable damping valve system damping characteristic model are constructed in combination with the suspension thermodynamic model, and the structural parameters of the key components of the oil-gas spring are determined. Finally, the structural design of the suspension system is carried out in combination with the design scheme of the oil-gas spring and the guide mechanism. The ultimate impact load is input for finite element analysis, and the strength of its key components is checked. The process of iterative hard point coordinates, structural parameters, and strength analysis is finally formed. A highly compact nonlinear suspension system solution is finally formed.

[0085] On the basis of forming a specific implementation plan for the suspension system, further detailed characteristic calculations of the suspension system are carried out. The process is as follows: Figure 3 First, input the vehicle weight G, vehicle bottom height from the ground H, balance elbow axis distance from the vehicle bottom E, balance elbow length R, road wheel diameter D, number of road wheels n, dynamic stroke f d 、Static stroke f j , coordinates of the upper hinge point (x1, y1), length of the pull arm r, angle between the pull arm and the balance elbow β, piston area A h , initial air column length h 0 , calculate the angle α between the balance elbow and the horizontal line, and further calculate the oil-gas spring piston stroke s and lever ratio i in sequence, according to the initial gas column length h 0 , and use the gas state equation to derive the gas pressure p of the oil-gas spring at any position, and use the calculated gas pressure p and piston area A h Multiply them to get the nonlinear elastic characteristics of the oil-gas spring, and then multiply them with the lever ratio i to further get the elastic characteristics of the suspension system. The generation of nonlinear elastic characteristic curves can be achieved through independent programming. The natural frequency of the sprung mass at the static equilibrium position and the dynamic potential energy at the end of the suspension system stroke are calculated respectively, and the vehicle dynamics model is established to check and calculate the ride comfort and driving stability of the vehicle. The vibration response results of the vehicle dynamics simulation are used to make a judgment. If the acceleration vibration response is greater than the set threshold, the initial inflation pressure p needs to be increased. 0 , the characteristics are repeatedly calculated and input into the vehicle dynamics model for analysis again until the threshold setting requirements are met. In addition, if the relative suspension travel is greater than the set threshold, it means that the suspension buffering capacity is insufficient, and the initial inflation pressure p needs to be reduced. 0, and iteratively calculate the characteristics and input them into the vehicle dynamics model for analysis until the threshold setting requirements are met.

[0086] In order to make the vehicle have good driving performance on various complex roads, it is necessary to further study the control method of suspension stiffness and damping characteristics. After adopting the technical solution of oil-gas suspension, stiffness adjustment becomes possible by adjusting the high-pressure gas volume.

[0087] The adjustment of stiffness characteristics is mainly reflected in two aspects: static adjustment and dynamic adjustment during driving. In terms of static adjustment, the internal working pressure of the oil-gas suspension is first collected through sensors, the wheel load state is calculated, and the suspension accumulator is charged and discharged through the hydraulic pneumatic system. The vehicle body frequency deviation is adjusted to a reasonable range (set to 1Hz-1.3Hz) using the gas state equation. After the gas volume adjustment is completed, it is necessary to consider that the vehicle body posture will also change accordingly. Therefore, it is necessary to formulate a control algorithm to drive the vehicle posture adjustment hydraulic valve group to charge and discharge the suspension cylinder, so as to achieve the target position of the vehicle bottom height from the ground to meet the driving requirements. The static adjustment of stiffness characteristics can effectively solve the problem of vehicle driving performance changes caused by the large difference between empty and full load conditions, and meet the use requirements under various load conditions.

[0088] In terms of dynamic adjustment of stiffness characteristics during driving, a suspension system dynamics model is established to carry out simulation analysis of the optimal matching of stiffness and damping characteristics, such as Figure 4-Figure 5 As shown in the figure, as the natural frequency, i.e. the suspension stiffness, decreases, the root mean square of the body vibration acceleration also decreases, and the sensitivity of the change is significantly greater than the effect of the damping ratio on it, but it will also increase the relative travel of the suspension. Figure 4 It can be seen that the root mean square acceleration of the vehicle body vibration will gradually decrease with the increase of the damping ratio. When the damping ratio reaches about 0.5, the root mean square acceleration reaches the minimum. As the damping ratio continues to increase, the root mean square acceleration gradually shows an increasing trend, that is, there is a nonlinear trend, which means that under specific road conditions and vehicle speeds, the optimal suspension damping can usually be selected. Figure 5 It can be seen that the root mean square acceleration of the vehicle body vibration will increase with the increase of the natural frequency, that is, the suspension stiffness, and basically presents a simple linear relationship, which means that when the relative travel of the suspension allows, reducing the stiffness can achieve the effect of suppressing the vehicle body vibration.

[0089] Based on the above matching analysis results, the vibration response of the suspension system in the time domain was further studied. The suspension stiffness was increased by 2 times, i.e., the high stiffness state, and the suspension stiffness was reduced to 0.5 times, i.e., the low stiffness state. The suspension stiffness was adjusted according to the stroke change, i.e., the stroke-varied stiffness state, and the acceleration-varied stiffness state, which further considered the body acceleration response, was determined. The control method of the stroke-varied stiffness is as follows:

[0090]

[0091] Where k is the target suspension stiffness, k b is the passive suspension stiffness, s is the suspension response travel, s 0 is the set suspension range travel.

[0092] The control method of acceleration variable stiffness is as follows:

[0093]

[0094] Where: is the sprung mass acceleration; Preset threshold value for acceleration.

[0095] A comparative analysis of vibration responses based on vehicle dynamics was conducted, and the results are as follows: Figure 6 As shown in the figure, it can be seen that with the increase of suspension stiffness, the body vibration acceleration also shows a significant increasing trend, and when the stiffness decreases, it shows a significant downward trend. Compared with the passive suspension, the travel variable stiffness and acceleration variable stiffness control methods have obvious advantages in suppressing the body vibration acceleration, and are comparable in suppressing the relative travel of the suspension and the acceleration of the unsprung mass, such as Figure 7-Figure 8 Compared with the traditional passive suspension, the realization of the adjustable stiffness function significantly expands the dimension of suspension characteristic adjustment and the amplitude of vibration response suppression, laying the foundation for the research and development of intelligent suspension vibration reduction technology.

[0096] It should be noted that in the process of dynamically adjusting the suspension stiffness characteristics during driving, it is necessary to accurately monitor the filling and releasing volume of the high-pressure gas in the suspension air chamber, and at the same time compensate the amount of oil in the oil-gas suspension oil chamber through the vehicle posture hydraulic adjustment system. That is, the vehicle posture hydraulic adjustment system needs to drain oil when the air chamber is inflated, and needs to fill oil when the air chamber is deflated, so as to ensure the stability of the vehicle body posture.

[0097] In view of the fact that the stiffness adjustment system is relatively complex and the fact that the function is achieved by adjusting the high-pressure air volume, it has typical nonlinear and time-lag characteristics, which is not conducive to real-time adjustment. Therefore, it is further proposed to first suppress the acceleration vibration response of the vehicle body by adjusting the suspension stiffness in steps, and then combine it with the damping characteristic control method to achieve the purpose of significantly improving the off-road maneuverability of the vehicle. The realization of stiffness adjustment can effectively solve the use requirements of large changes in sprung mass and special equipment operations.

[0098] Aiming at the adaptive adjustment of suspension damping characteristics, the present invention proposes a frequency-band hybrid control method, such as Fig. 9 As shown in the figure, taking the single-wheel suspension system as the research object, the fixed point of the frequency domain vibration response is derived, and the suspension motion differential equations are first given:

[0099]

[0100] In the formula, m s is the sprung mass, m u is the unsprung mass, is the sprung mass acceleration, is the unsprung mass acceleration, x s is the sprung mass displacement, x u is the unsprung mass displacement, is the sprung mass velocity, is the unsprung mass velocity, x r is the road roughness, k s is the suspension stiffness characteristic, k u is the wheel stiffness characteristic, c s is the suspension damping characteristic.

[0101] Through Laplace transform, we get:

[0102]

[0103] The functional relationship between the sprung mass and the unsprung mass displacement is derived as follows:

[0104]

[0105] And the functional relationship between the sprung mass and the ground excitation elevation is summarized as:

[0106]

[0107] The fixed point expression of the sprung mass displacement response is obtained by programming:

[0108]

[0109] Depend on Fig.10 As shown, by programming and analyzing the amplitude-frequency characteristics, it can be seen that: ①ω 2 and ω 3 and suspension stiffness k s is related to the suspension damping c s It doesn't matter, but 1 and suspension stiffness k s and suspension damping c s are irrelevant and are completely determined by the wheel stiffness k u and unsprung mass m u Determine; ② When the excitation frequency ω is ω<ω 1 or 2 <ω<ω 3 When the excitation frequency ω is within the range of ω, the vertical vibration acceleration transmission rate of the vehicle body decreases with the increase of suspension damping, that is, the inhibitory effect of increasing damping on suspension resonance (low-frequency resonance area and high-frequency resonance area) is particularly significant; ③ When the excitation frequency ω is within ω1 <ω<ω 2 or ω>ω 3 When the vehicle body is within the specified range, the vertical vibration acceleration transmission rate increases with the increase of suspension damping, which will cause the vehicle's ride smoothness to deteriorate. This is why the passive suspension system needs to make compromises in order to improve the vehicle's ride smoothness.

[0110] Furthermore, the frequency domain fixed point expression of the unsprung mass displacement response is derived, and the functional relationship between the unsprung mass displacement and the road elevation excitation is summarized as follows:

[0111]

[0112] set up:

[0113]

[0114] Solving the above equation is equivalent to the following form:

[0115]

[0116] Further sorting:

[0117] {k u (-m s ω 2 +k s )[k u ω-(m u +m s )ω 3 ]} 2 -{(k u ω)[m u m s ω 4 -(k s m s +k u m s +m u k s )ω 2 +k u k s ]} 2 =0 The fixed point expression of the frequency domain of the unsprung mass displacement response is solved by programming:

[0118]

[0119] The amplitude-frequency characteristics of the unsprung mass displacement to the road surface excitation are analyzed by programming, such as Fig.11 As shown, it can be seen that: ①ω 4 and ω 5 and suspension stiffness k sThat is, at this frequency, the wheel dynamic deformation is still affected by the suspension force; ② When the excitation frequency ω is within ω<ω 4 or ω>ω 5 When the excitation frequency ω is within the range of ω, the wheel dynamic deformation transmission rate decreases with the increase of suspension damping. Similarly, increasing damping is beneficial to suppress the resonance of the vehicle. 4 <ω<ω 5 When the transmission rate of wheel dynamic deformation is within the range, the wheel dynamic deformation increases with the increase of suspension damping, which will cause the vehicle's road adhesion to deteriorate, indicating that the passive suspension system cannot optimize the wheel dynamic deformation as a whole in the entire frequency band.

[0120] The relationship between the relative travel of the suspension and the road surface excitation elevation is further derived:

[0121]

[0122] set up:

[0123]

[0124] Arrange the transformation into the form of a system of equations:

[0125]

[0126] The frequency domain fixed point expression for solving the relative travel of the suspension is:

[0127]

[0128] Through programming, the amplitude-frequency characteristics of the relative travel of the suspension to the road excitation are analyzed, such as Fig.12 As shown, it can be seen that: ①ω 6 and suspension stiffness k s and suspension damping c s ① They are all unrelated, indicating that at this frequency, the suspension dynamic stroke transmission rate is not affected by the suspension force and is a constant value; ② In general, the suspension dynamic stroke decreases with the increase of suspension damping. Similarly, increasing damping is beneficial to suppress vehicle resonance.

[0129] Based on the above analysis of the six fixed points of vehicle body vertical vibration acceleration, wheel dynamic deformation and suspension dynamic travel, it can be seen that from the perspective of frequency, there is a certain consistency in the influence of suspension damping on the transfer characteristics of vehicle body vertical vibration acceleration, wheel dynamic deformation and suspension dynamic travel: ① In both the low-frequency resonance zone and the high-frequency resonance zone, the suspension system needs to provide greater damping to suppress resonance; ② In both the medium-frequency zone and the high-frequency zone, the suspension system needs to provide smaller damping to reduce and improve the vehicle's driving smoothness and road adhesion.

[0130] Although the suspension system usually has two vibration control goals - improving vehicle ride smoothness and road adhesion (while also taking into account the vehicle driving safety goal), and there is only one control quantity, so no matter what control method is used, these two performance indicators cannot be simultaneously optimized. However, from the above analysis of the fixed point of the transmissibility, it can be seen that the transmission characteristics of the suspension damping to the three indicators in the frequency domain are consistent to a certain extent. As long as the frequency intervals are appropriately divided and different controls are applied to different frequency intervals, the overall performance of the suspension can be approximately optimal.

[0131] According to the above analysis, the frequency divider γ is designed as follows:

[0132]

[0133] Where: is the rate of change of sprung mass acceleration; α is an intermediate variable, where α=(2πf cross ) 2 , f cross Take the first fixed point frequency

[0134] Assume that Where f>0, then Where f is the actual input frequency, substituting it into the formula yields:

[0135] γ=((2πf) 4 -α 2 )sin 2 (2πft)

[0136] Since α=(2πf 1 ) 2 , then when 0 <f<f 1 When γ<0, and when f>f 1 When γ>0, an improved frequency domain hybrid damping control method is given, which can be expressed as:

[0137]

[0138] Where: c f is the adjustable damping value of the suspension; c fmax Maximum adjustable damping for the suspension.

[0139] The evaluation index and vibration control target of the vehicle suspension system are based on the root mean square value of the sprung vibration acceleration, the root mean square value of the wheel dynamic displacement and the comprehensive performance of the suspension system formed by them. The convex function is selected to establish the comprehensive performance evaluation index of the vehicle suspension system, which can be expressed as:

[0140]

[0141] Where ζ and ξ are the weight distribution coefficients of sprung mass acceleration and wheel displacement, respectively, and ζ+ξ=1. The adaptive adjustment of the weight distribution coefficient is realized by fuzzy control method. Q RP_P and Q RP are the root mean square values ​​of the sprung vibration acceleration of the passive and controllable suspension, Q RA_P and Q RA are the root mean square values ​​of the dynamic displacements of the passive and controllable suspension wheels, respectively.

[0142] Through programming, the vibration response comparison analysis of different control methods is carried out, and the frequency response characteristics of the vehicle suspension system are analyzed using the sine excitation method. The reference frequency range is 0.011m, referring to the random road surface. -1 <n<2.83m -1 , taking the F-class standard random road surface as the excitation input, the time-frequency domain comparison curves and the root mean square value of the sprung vibration acceleration of the passive suspension, skyhook damping control, acceleration damping control and frequency domain hybrid control suspension systems are generated, such as Figure 13-14 As shown in the figure, it can be seen that (1) the ceiling damping has a good inhibitory effect on the sprung vibration acceleration in the low-frequency resonance area, but has an obvious amplification trend in the vibration response in the high-frequency area, and the root mean square value of the sprung acceleration response in the full-frequency domain has increased; (2) ADD damping is not as good as ceiling damping and frequency domain hybrid control in suppressing the sprung vibration acceleration in the low-frequency resonance area, but has a very obvious inhibitory effect on the sprung vibration acceleration in the frequency domain between low-frequency resonance and high-frequency resonance, which is equivalent to frequency domain hybrid control. Therefore, the root mean square value of the sprung acceleration response in the full-frequency domain is attenuated by 1.3% compared with the passive suspension. 14.3%; (3) Compared with ceiling damping and ADD damping control, hybrid control tracks the response characteristics of ceiling damping to sprung vibration acceleration in the low-frequency range, and tracks the response characteristics of acceleration damping control to sprung vibration acceleration in the medium and high-frequency range, indicating that the crossover designed by the improved algorithm is effective; (4) Compared with passive suspension, frequency domain hybrid control can reduce sprung vibration acceleration in most frequency bands, with a RMS attenuation of 16.2% compared with passive suspension, which is more obvious than ceiling damping and ADD damping, thus verifying the effectiveness of the algorithm.

[0143] The present invention belongs to the technical field of suspension vibration reduction, and specifically relates to a suspension system damping adjustment method. The method proposes a frequency-band hybrid control method, takes a single-wheel suspension system as a research object, derives the fixed point of the frequency-domain vibration response, gives a suspension motion differential equation group, derives the functional relationship between the sprung mass and the unsprung mass displacement by Laplace transform, and solves the fixed point expression of the sprung mass displacement response, the frequency-domain fixed point expression of the unsprung mass displacement response, and the frequency-domain fixed point expression of the suspension relative stroke by programming, respectively. The control algorithm is written, and a comparative analysis of the vibration response is performed to verify the effectiveness of the algorithm.

[0144] The above is only a preferred embodiment of the present invention. It should be pointed out that for ordinary technicians in this technical field, several improvements and modifications can be made without departing from the technical principles of the present invention. These improvements and modifications should also be regarded as the scope of protection of the present invention.

Claims

1. A suspension system damping adjustment method, characterized in that: The method solves the fixed point expression of the sprung mass displacement response based on the frequency-band hybrid control method: Where ω1 is the frequency of the sprung mass displacement response fixed point 1, rad / s; ω2 is the frequency of the sprung mass displacement response fixed point 2, rad / s; ω3 is the frequency of the sprung mass displacement response fixed point 3, rad / s; m s is the sprung mass, kg; m u is the unsprung mass; x s is the sprung mass displacement, m; x u is the unsprung mass displacement, m; k s is the suspension stiffness characteristic, N / m; k u is the wheel stiffness characteristic.

2. The method for adjusting the damping of a suspension system according to claim 1, characterized in that: in, ω2 and ω3 and suspension stiffness k s is related to the suspension damping c s is independent of, while ω1 is related to the suspension stiffness k s and suspension damping c s are irrelevant and are completely determined by the wheel stiffness k u and wheel mass m u Decide; When the excitation frequency ω is in the range of ω<ω1 or ω2<ω<ω3, the vertical vibration acceleration transmission rate of the vehicle body decreases with the increase of suspension damping, that is, the inhibitory effect of increasing damping on suspension resonance is particularly significant; when the excitation frequency ω is in the range of ω1<ω<ω2 or ω>ω3, the vertical vibration acceleration transmission rate of the vehicle body increases with the increase of suspension damping, which will cause the vehicle's ride smoothness to deteriorate. This is why the passive suspension system needs to make compromises in order to improve the vehicle's ride smoothness.

3. The method for adjusting the damping of a suspension system according to claim 1, characterized in that: The frequency domain fixed point expression for solving the displacement response of the unsprung mass is: Where ω4 is the frequency of the unsprung mass displacement response fixed point 4, rad / s; ω5 is the frequency of the unsprung mass displacement response fixed point 5, rad / s.

4. The method for adjusting the damping of a suspension system according to claim 3, characterized in that: in, ω4 and ω5 and suspension stiffness k s That is, at this frequency the wheel dynamic deformation is still affected by the suspension force; when the excitation frequency ω is in the range of ω<ω4 or ω>ω5, the wheel dynamic deformation transmission rate decreases with the increase of suspension damping. Similarly, increasing the damping is beneficial to suppressing the resonance of the vehicle; when the excitation frequency ω is in the range of ω4<ω<ω5, the wheel dynamic deformation transmission rate increases with the increase of suspension damping, which will cause the vehicle's road adhesion to deteriorate, indicating that the passive suspension system cannot optimize the wheel dynamic deformation as a whole in the entire frequency band.

5. The method for adjusting the damping of a suspension system according to claim 3, characterized in that: The frequency domain fixed point expression for solving the relative travel of the suspension is: Wherein, ω6 is the frequency of the fixed point 4 of the suspension relative travel, rad / s.

6. The method for adjusting the damping of a suspension system according to claim 5, characterized in that: in, ω6 and suspension stiffness k s and suspension damping c s They are all unrelated, indicating that at this frequency, the suspension dynamic stroke transfer rate is not affected by the suspension force and is a constant value; in general, the suspension dynamic stroke decreases with the increase of suspension damping. Similarly, increasing damping is beneficial to suppress vehicle resonance.

7. The method for adjusting the damping of a suspension system according to claim 5, characterized in that: The method designs a frequency divider of the following form: Where: is the rate of change of sprung mass acceleration; α is an intermediate variable, where α=(2πf cross ) 2 , f cross Take the first fixed point frequency Assume that there is Where f>0, then Where f is the actual input frequency, substituting it into the formula yields: γ=((2πf) 4 -a 2 )sin 2 (2πft) Since there is α = (2πf1) 2 , when 0 < f < f1, γ < 0; when f > f1, γ > 0. Based on this, an improved frequency-domain hybrid damping control method is given, expressed as: Where: c f is the adjustable damping value of the suspension; cfma x Maximum adjustable damping for the suspension.

8. The method for adjusting the damping of a suspension system according to claim 7, characterized in that: The method establishes comprehensive performance evaluation index of vehicle suspension system: Where ζ and ξ are the weight distribution coefficients of sprung mass acceleration and wheel displacement, respectively, and ζ+ξ=1. The adaptive adjustment of the weight distribution coefficient is realized by fuzzy control method. Q RP_P and Q RP are the root mean square values ​​of the sprung vibration acceleration of the passive and controllable suspension, Q RA_P and Q RA are the root mean square values ​​of the dynamic displacements of the passive and controllable suspension wheels, respectively.

9. The method for adjusting the damping of a suspension system according to claim 8, characterized in that: The method can perform segmented control within the entire frequency domain and determine a reasonable control algorithm in each frequency band, thereby achieving the effect of simultaneously improving driving smoothness and handling stability.

10. The method for adjusting the damping of a suspension system according to claim 9, characterized in that: The method derives the suspension dynamics equation, finds the frequency domain fixed point that does not change with the damping coefficient, and derives the analytical formula. The frequency domain space is divided by the fixed point, and the law of unidirectional change of acceleration vibration response with damping is found in each frequency domain interval, so as to formulate a reasonable control algorithm to achieve the purpose of effectively suppressing the vibration influence in the whole frequency domain. Through comparison and verification with other control methods, it has obvious advantages in suppressing vibration response and has broad application prospects.

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