Vehicle nonlinear inerter double-air-chamber hydro-pneumatic suspension structure and parameter determination method

Abstract

The invention discloses a vehicle nonlinear inerter double-air-chamber hydro-pneumatic suspension structure and a parameter determination method. The method comprises the following working steps that 1, a nonlinear fluid inerter model is constructed; 2, constructing a 1 / 4 nonlinear inerter hydro-pneumatic suspension model; 3, parameter optimization based on a non-dominated sorting genetic algorithm II; and 4, performing dynamic performance simulation analysis. The method has the beneficial effects that the nonlinear fluid inerter and the double-air-chamber hydro-pneumatic suspension are combined, and the inerter hydro-pneumatic suspension model is constructed; a non-dominated sorting genetic algorithm II is used for optimizing target parameters, the algorithm is an optimization algorithm used for solving a plurality of conflict targets, and finally an optimal solution set of the Pareto graph is found; the overall performance of the suspension system can be effectively improved; and an obvious optimization effect is achieved in the aspect of vehicle vibration.

Description

Technical Field

[0001] The invention relates to the technical field of vehicle suspension systems, and in particular to a vehicle nonlinear inertia double-chamber oil-gas suspension structure and a parameter determination method. Background Art

[0002] In the field of vehicle suspension systems, with the increasing requirements for vehicle ride smoothness, handling stability and ride comfort, traditional suspension systems have been unable to meet these advanced requirements; oil-gas suspension, with its superior nonlinear elastic characteristics and good vibration reduction performance, has gradually become the preferred suspension system for high-end automobiles, engineering vehicles and military special vehicles; however, traditional oil-gas suspension still has certain limitations under complex working conditions, such as insufficient dynamic response and insufficient vibration control; in order to further improve the performance of oil-gas suspension, researchers began to explore new technical paths that combine fluid inertia vessels with oil-gas suspension; as a device that can store and release kinetic energy, the unique inertial characteristics of fluid inertia vessels can effectively improve the dynamic response and vibration control effects of the suspension system; combining fluid inertia vessels with oil-gas suspensions can fully utilize the advantages of both to form a more efficient and stable suspension system;

[0003] In view of the above situation, it is particularly important to construct a new inertial oil-pneumatic suspension model and establish its parameter method. The present invention intends to develop a more advanced suspension system design to meet higher suspension performance requirements. Summary of the invention

[0004] In order to solve the above technical problems, the present invention provides a vehicle nonlinear inertia dual-chamber oil-gas suspension structure and parameter determination method, comprising the following working steps:

[0005] Step 1: Construction of nonlinear fluid inertia container model;

[0006] Step 2: 1 / 4 nonlinear inertia oil-pneumatic suspension (FI-HPS) model construction;

[0007] Step 3: Parameter optimization based on non-dominated sorting genetic algorithm II (NSGA-II);

[0008] Step 4: Dynamic performance simulation analysis.

[0009] To further supplement the technical solution, the nonlinear fluid inertia container model in step 1 is constructed, and its specific implementation method is as follows:

[0010] The fluid inertia container is composed of a piston rod, a piston, a spiral channel, a hydraulic cylinder, etc. Its composition structure is similar to that of the oil-gas suspension. Therefore, this paper conducts an integrated design of the oil-gas suspension and the fluid inertia container based on this structure. The fluid inertia container uses the rapid flow of fluid in the spiral pipe to form a "fluid flywheel" to provide inertial force.

[0011] The force expression between the two end points of an ideal fluid inertia container is:

[0012]

[0013] During the flow process, the viscosity effect is generated due to the influence of factors such as the length, radius, and temperature of the pipeline. The damping force is the main nonlinear characteristic of the fluid inertia container. When the present invention studies the double-chamber oil-gas inertia container suspension, the piston friction force is within the controllable range, so the influence of friction force is not considered.

[0014] The force expression between the two ends of the nonlinear fluid inertia container is:

[0015] F=f+f.

[0016]

[0017] In the formula, Z2, Z1 are the displacements of the two ends of the inertia container, fb is the inertial force, fc is the nonlinear damping force, b is the inertia coefficient, C1, C2 are the damping coefficients, which can be expressed by the following formula:

[0018]

[0019] Where ρ is the oil density, l is the length of the spiral pipe, h is the pitch of the spiral channel, r4 is the spiral radius of the spiral pipe, S A is the area of ​​the hydraulic cylinder, S3 is the cross-sectional area of ​​the spiral pipe, r3 is the radius of the spiral pipe, and μ is the viscosity of the liquid. From the above formula, we can see that C2 and b are both related to the length of the spiral pipe, and the formula can be obtained:

[0020]

[0021] It can be seen from the above formula that the change of h and r4 will synchronously change b and C2 to obtain different inertia coefficients and damping coefficients, thereby establishing different nonlinear fluid inertia container models.

[0022] To further supplement the technical solution, the 1 / 4 nonlinear inertia oil-pneumatic suspension (FI-HPS) model in step 2 is constructed, and its specific implementation method is as follows:

[0023] In the dual-chamber oil-gas suspension, the B and C chambers are connected to the damping hole through a one-way valve, and the oil flow formula is as follows:

[0024]

[0025] Where: is the speed of the piston moving up and down; Q1 is the oil flow between chambers B and C; S C is the C cavity area.

[0026] When the oil-gas spring is in compression stroke, the oil flows through the one-way valve and the damping hole, and the flow formula is:

[0027]

[0028] Where: A1 is the equivalent area of ​​the damping hole, A2 is the equivalent area of ​​the one-way valve.

[0029] When the oil-gas spring is in the extension stroke, the oil flows through the hole frequently, and the flow formula is:

[0030]

[0031] From the above formula, we can get the pressure loss caused by chambers B and C as follows:

[0032]

[0033] The damping force caused by the pressure loss in chambers B and C can be expressed as:

[0034]

[0035] Where: sign is the sign function. When the oil-gas spring is in the extension stroke, When the oil-gas spring is in the compression stroke,

[0036] In the dual-chamber oil-gas suspension proposed by the present invention, there is also pressure loss when the A chamber is connected to the R chamber of the accumulator by an oil pipe. The formula is as follows:

[0037]

[0038] In the formula, A3 is the cross-sectional area of ​​the pipeline, η1 and η2 are the oil pressure loss coefficients between the oil pipe and the accumulator. When the oil flows into the accumulator R, η1 = 1, and when the oil flows out of the accumulator R, η1 = 0.5; η2 is the opposite of η1. Q2 is the flow formula between chamber A and accumulator E, S A is the area of ​​cavity A.

[0039] The damping force generated by the pressure loss caused by the A and R chambers can be expressed as:

[0040]

[0041] Therefore, the total damping force expression of the dual-chamber oil-gas suspension proposed in the present invention is:

[0042] F C =F1+F2

[0043] The damping coefficient of the hydro-pneumatic suspension is:

[0044]

[0045] To further supplement the technical solution, the 1 / 4 nonlinear inertia oil-pneumatic suspension (FI-HPS) model in step 2 is constructed, and its specific implementation method is as follows:

[0046] The stiffness characteristic of the dual-chamber inertial oil-gas suspension in the present invention is the relationship between the elastic force on the main piston and the relative displacement of the piston without considering the damping force and ignoring the effect of oil compression on the gas volume. In the dual-chamber oil-gas suspension, when the influence of the damping characteristic is not considered, the pressure P of chamber A is A , and the pressure P in the accumulator R R Equal in size, P A =P R , similarly, the pressure of chamber B is P B 、C chamber pressure P C , D cavity pressure P D Equal in size, P B =P C =P D .

[0047] The formula for the elastic force of the piston at any time is:

[0048] F S =P C S C -P A S A

[0049] Where: F S is the total spring force of the suspension, S C is the area of ​​the annular cavity C, S A is the area of ​​chamber A.

[0050] The gas state equation of the accumulator R chamber is:

[0051]

[0052] Where: V R0 、V RS and V R are the gas volumes of the accumulator R chamber at the initial state, equilibrium position and transient state respectively; P R0 , P RS and P R are the gas pressures in the initial state, equilibrium position and instantaneous state of the R chamber of the accumulator respectively; n is the gas polytropic index.

[0053] The gas state equation in chamber D is:

[0054]

[0055] Where: V D0 、V DS and V D are the gas volumes in the initial state, equilibrium position and instantaneous state of cavity D respectively; P D0 , P DS and P D are the gas pressures in the initial state, equilibrium position and instantaneous state of the D chamber of the accumulator respectively; n is the gas polytropic index.

[0056] When the piston rod is in the equilibrium position, let the downward movement be x. The change of the gas volume in the R and D chambers of the accumulator in the instantaneous state is as follows:

[0057] V R =V RS +S A ·x

[0058] V D =V DS -S C ·x

[0059] The mass on the suspension spring is M. When the oil-gas spring moves from the initial state to the equilibrium state, the pressure in each chamber is the same, and the process is regarded as an isothermal state. Take n = 1, and the pressure and volume of the air cavity are as follows:

[0060]

[0061] By combining the above formulas, we can get:

[0062]

[0063] For the instantaneous elastic force F S The corresponding stiffness coefficient can be obtained by differentiating the displacement x:

[0064]

[0065] To further supplement the technical solution, the 1 / 4 nonlinear inertia oil-pneumatic suspension (FI-HPS) model in step 2 is constructed, and its specific implementation method is as follows:

[0066] Construct the dynamic equation of the 1 / 4 vehicle suspension structure:

[0067]

[0068] In the formula, m s The sprung mass is set to 400 kg, m u The unsprung mass is set to 40 kg, K t The tire stiffness is set to 200 kN·m -1 , F is the force between the suspensions. s , Z uare the vertical displacements of the sprung and unsprung masses, Z r is the vertical displacement of the road surface input, and its formula is:

[0069]

[0070] In the formula, z r (t) is the vertical input displacement, G q (n0) is the road roughness, v is the vehicle speed, and when the car is driving on a Class C road, the road roughness coefficient is 2.56×10 -5 , W t is Gaussian white noise with a mean of 0.

[0071] In order to analyze the influence of three models of traditional double-chamber oil-gas suspension, ideal inertia oil-gas suspension and nonlinear inertia oil-gas suspension on vehicle suspension performance, this paper provides three different suspension structures, S0 is the traditional double-chamber oil-gas suspension, S1 is the ideal inertia oil-gas suspension, S2 is the nonlinear inertia oil-gas suspension, and the corresponding suspension forces F 0、 F1 and F2 are as follows.

[0072]

[0073] To further supplement the technical solution, the parameter optimization based on the non-dominated sorting genetic algorithm II (NSGA-II) in step 3 is specifically implemented as follows:

[0074] The present invention adopts NSGA-II, which is the full name of non-dominated sorting genetic algorithm, to solve the optimization algorithm of multiple conflicting objectives, and finally find the optimal solution set of Pareto graph, which has the characteristics of fast algorithm operation speed and good convergence. The present invention takes the parameters of traditional double-chamber oil-gas suspension as reference, ensures that other parameters remain unchanged, and selects the initial air pressure P of the air cavity D in the inertia oil-gas suspension. D0 , the optimization range is [15,35]bar; volume V D0 , the optimization range is [0.2,1.3]L; the initial pressure P of the accumulator R chamber R0 , the optimization range is [15,35]bar, volume V R0 , the optimization range is [0.2,1.5]L, the nonlinear damping coefficient C2, the optimization range is [458.3,3000]N s / m; the inertia coefficient b optimization range is [0,1000]kg, which is the optimization object.

[0075] The vehicle selected in this invention is a passenger car, and in actual use, the emphasis is more on riding comfort. Therefore, the body acceleration should be as small as possible, and the suspension travel and tire dynamic load should also be within the normal range to ensure the normal working condition of the vehicle suspension system. There are mutual constraints for the three evaluation indicators, so the optimization objective function is as follows:

[0076]

[0077] Where: J1 is the overall evaluation index reflecting the vehicle's ride comfort and stability, while J2 is the evaluation index of the vehicle's road friendliness. This formula can directly compare the performance of the inertia container-based hydro-pneumatic suspension and the traditional hydro-pneumatic suspension. pas -RMS, SWS pas -RMS and DTL pas -RMS are the root mean square values ​​of the body acceleration, suspension working space and dynamic tire load of the traditional hydro-pneumatic suspension. BA_RMS, SWS_RMS and DTL_RMS are the root mean square values ​​of the body acceleration, suspension working space and dynamic tire load of the hydro-pneumatic suspension based on inertia vessel. W1 and W2 are the optimized weight coefficients, which are set to 0.6 and 0.4 respectively.

[0078] As a further supplement to the technical solution, the dynamic performance simulation analysis in step 4 is specifically implemented as follows: in the simulation analysis, the vehicle body acceleration, suspension dynamic deflection, and tire dynamic load are used as indicators, and Matlab and Simulink are used to perform road excitation analysis on the above three indicators at a vehicle speed of 20m / s; the present invention fully considers the nonlinear parasitic damping force factor of the fluid inertia container, establishes three structural models of traditional dual-chamber oil-gas suspension S0, ideal inertia oil-gas suspension S1, and nonlinear inertia oil-gas suspension S2, and performs dynamic simulation analysis.

[0079] The beneficial effects are as follows: the present invention combines a nonlinear fluid inertia container with a dual-chamber oil-gas suspension to construct an inertia container oil-gas suspension model; utilizes a non-dominated sorting genetic algorithm II (NSGA-II) to optimize target parameters, and the algorithm is used to solve the optimization algorithm of multiple conflicting targets, and finally finds the optimal solution set of the Pareto graph; compares the nonlinear inertia container oil-gas suspension with an ideal inertia container oil-gas suspension and a traditional dual-chamber oil-gas suspension, and concludes that the inertia force and parasitic damping force generated by the fluid inertia container are combined with the nonlinear characteristics of the oil-gas suspension itself to effectively improve the overall performance of the suspension system; the nonlinear inertia container dual-chamber oil-gas suspension structure of the present invention has an obvious optimization effect in terms of vehicle vibration. BRIEF DESCRIPTION OF THE DRAWINGS

[0080] Figure 1 It is a flow chart of a vehicle nonlinear inertia double-chamber oil-gas suspension structure and parameter determination method;

[0081] Figure 2 (a) is a schematic diagram of a fluid inertia container; (b) is an equivalent structure diagram of an ideal fluid inertia container; (c) is an equivalent structure diagram of a nonlinear fluid inertia container;

[0082] Figure 3 (a) is the structure diagram of the traditional dual-chamber oil-gas suspension; (b) is the structure diagram of the inertial oil-gas suspension;

[0083] Figure 4 This is the 1 / 4 suspension structure diagram;

[0084] Figure 5 These are diagrams of three different suspension structures;

[0085] Figure 6 It is the parameter optimization flow chart of the non-dominated sorting genetic algorithm NSGA-II;

[0086] Figure 7 (a) is the time domain diagram of vehicle body acceleration; (b) is the time domain diagram of suspension dynamic travel; (c) is the time domain diagram of tire dynamic load;

[0087] In the formula of the present invention, ρ is the density of the oil, l is the length of the spiral pipe, h is the pitch of the spiral channel, r4 is the spiral radius of the spiral pipe, S A is the area of ​​the hydraulic cylinder, S3 is the cross-sectional area of ​​the spiral pipe, r3 is the radius of the spiral pipe, μ is the viscosity of the liquid, is the speed of the piston moving up and down; Q1 is the oil flow between chambers B and C, S C is the area of ​​C cavity, A1 is the equivalent area of ​​damping hole, A2 is the equivalent area of ​​check valve, A3 is the cross-sectional area of ​​pipeline, η1 and η2 are the oil pressure loss coefficients between oil pipe and accumulator, F S is the total spring force of the suspension, S C is the area of ​​the annular cavity C, S A A is the chamber area, V R0 、V RS and V R are the gas volumes of the accumulator R chamber at the initial state, equilibrium position and instantaneous state, respectively, P R0 , P RS and P R are the gas pressures in the initial state, equilibrium position and instantaneous state of the accumulator R chamber, n is the gas polynomial index, m s is the sprung mass, m u is the unsprung mass, K t The stiffness of the tire, F is the force between the suspension. s , Z u are the vertical displacements of the sprung and unsprung masses, Z ris the vertical displacement input by the road surface, S0 is the traditional dual-chamber oil-gas suspension, S1 is the ideal inertia oil-gas suspension, and S2 is the nonlinear inertia oil-gas suspension. DETAILED DESCRIPTION

[0088] In order to make the technical solution more clear to those skilled in the art, Figure 1-7 The technical solution of the present invention is described in detail:

[0089] like Figure 1 As shown, a vehicle nonlinear inertia double-chamber oil-gas suspension structure and parameter determination method includes the following working steps:

[0090] Step 1: Construction of nonlinear fluid inertia container model;

[0091] Step 2: 1 / 4 nonlinear inertia oil-pneumatic suspension (FI-HPS) model construction;

[0092] Step 3: Parameter optimization based on non-dominated sorting genetic algorithm II (NSGA-II);

[0093] Step 4: Dynamic performance simulation analysis;

[0094] Among them, the specific implementation method of constructing the nonlinear fluid inertia container model in step 1 is:

[0095] Depend on Figure 2 (a) It can be seen that the fluid inertia container is composed of a piston rod, a piston, a spiral channel, a hydraulic cylinder, etc., and its composition structure is similar to that of the oil-gas suspension. Therefore, the present invention integrates the oil-gas suspension and the fluid inertia container based on this structure. The fluid inertia container uses the rapid flow of oil in the spiral pipe to form a "fluid flywheel" to provide inertial force.

[0096] Depend on Figure 2 (b) It can be seen that the force expression between the two end points of an ideal fluid inertia container is:

[0097]

[0098] During the flow process, the viscosity effect is generated due to the influence of factors such as the length, radius, and temperature of the pipeline. The damping force is the main nonlinear characteristic of the fluid inertia container. When the present invention studies the double-chamber oil-gas inertia container suspension, the piston friction force is within the controllable range, so the influence of friction force is not considered.

[0099] Depend on Figure 2 (c) It can be seen that the force expression between the two end points of the nonlinear fluid inertia container is:

[0100] F==f+fc

[0101]

[0102] In the formula, Z2, Z1 are the displacements of the two ends of the inertia container, fb is the inertial force, fc is the nonlinear damping force, b is the inertia coefficient, C1, C2 are the damping coefficients, which can be expressed by the following formula:

[0103]

[0104] Among them, the 1 / 4 nonlinear inertia oil-pneumatic suspension (FI-HPS) model in step 2 is constructed, and its specific implementation method is as follows:

[0105] Depend on Figure 3 It can be seen that in the dual-chamber oil-gas suspension, the B and C chambers are connected to the damping hole through a one-way valve, and the oil flow formula is as follows:

[0106]

[0107] Where: is the speed of the piston moving up and down; Q1 is the oil flow between chambers B and C; S C is the C cavity area.

[0108] When the oil-gas spring is in compression stroke, the oil flows through the one-way valve and the damping hole, and the flow formula is:

[0109]

[0110] Where: A1 is the equivalent area of ​​the damping hole, A2 is the equivalent area of ​​the one-way valve.

[0111] When the oil-gas spring is in the extension stroke, the oil flows through the hole frequently, and the flow formula is:

[0112]

[0113] From the above formula, we can get the pressure loss caused by chambers B and C as follows:

[0114]

[0115] The damping force caused by the pressure loss in chambers B and C can be expressed as:

[0116]

[0117] Where: sign is the sign function. When the oil-gas spring is in the extension stroke, When the oil-gas spring is in the compression stroke,

[0118] The double air chamber oil-gas suspension proposed by the present invention, Figure 3 In the example, there is also pressure loss when the A chamber is connected to the R chamber of the accumulator by an oil pipe. The formula is as follows:

[0119]

[0120] In the formula, A3 is the cross-sectional area of ​​the pipeline, η1 and η2 are the oil pressure loss coefficients between the oil pipe and the accumulator. When the oil flows into the accumulator R, η1 = 1, and when the oil flows out of the accumulator R, η1 = 0.5; η2 is the opposite of η1. Q2 is the flow formula between chamber A and accumulator E, S A is the area of ​​cavity A.

[0121] The damping force generated by the pressure loss caused by the A and R chambers can be expressed as:

[0122]

[0123] Therefore, the total damping force expression of the dual-chamber oil-gas suspension proposed in the present invention is:

[0124] F C =F1+F2

[0125] The damping coefficient of the hydro-pneumatic suspension is:

[0126]

[0127] The stiffness characteristic of the dual-chamber inertial oil-gas suspension in the present invention is the relationship between the elastic force on the main piston and the relative displacement of the piston without considering the damping force and ignoring the effect of oil compression on the gas volume. Figure 3 In the dual-chamber oil-gas suspension, without considering the influence of damping characteristics, the pressure P of chamber A is A , and the pressure P in the accumulator R R Equal in size, P A =P R , similarly, the pressure of chamber B is P B 、C chamber pressure P C , D cavity pressure P D Equal in size, P B =P C =P D .

[0128] The formula for the elastic force of the piston at any time is:

[0129] F S =P C S C -P A S A

[0130] Where: F S is the total spring force of the suspension, S C is the area of ​​the annular cavity C, S A is the area of ​​chamber A.

[0131] The gas state equation of the accumulator R chamber is:

[0132]

[0133] Where: V R0 、V RS and V R are the gas volumes of the accumulator R chamber at the initial state, equilibrium position and transient state respectively; P R0 , P RS and P R are the gas pressures in the initial state, equilibrium position and instantaneous state of the R chamber of the accumulator respectively; n is the gas polytropic index.

[0134] The gas state equation in chamber D is:

[0135]

[0136] Where: V D0 、V DS and V D are the gas volumes in the initial state, equilibrium position and instantaneous state of cavity D respectively; P D0 , P DS and P D are the gas pressures in the initial state, equilibrium position and instantaneous state of the D chamber of the accumulator respectively; n is the gas polytropic index.

[0137] When the piston rod is in the equilibrium position, let the downward movement be x. The change of the gas volume in the R and D chambers of the accumulator in the instantaneous state is as follows:

[0138] V R =V RS +S A ·x

[0139] V D =V DS -S C ·x

[0140] The mass on the suspension spring is M. When the oil-gas spring moves from the initial state to the equilibrium state, the pressure in each chamber is the same, and the process is regarded as an isothermal state. Take n = 1, and the pressure and volume of the air cavity are as follows:

[0141]

[0142] By combining the above formulas, we can get:

[0143]

[0144] For the instantaneous elastic force F S The corresponding stiffness coefficient can be obtained by differentiating the displacement x:

[0145]

[0146] 1 / 4 vehicle suspension diagram Figure 4 As shown, Figure 5 S0: traditional dual-chamber oil-gas suspension, S1: ideal inertia oil-gas suspension, S2: nonlinear inertia oil-gas suspension. The dynamic equations are established based on different structures:

[0147]

[0148] In the formula, m s The sprung mass is set to 400 kg, m u The unsprung mass is set to 40 kg, K t The tire stiffness is set to 200 kN·m -1 , F is the inter-suspension force. Zs and Zu are the vertical displacements of the sprung mass and the unsprung mass respectively, and Zr is the vertical displacement of the road input, and the formula is:

[0149]

[0150] In the formula, z r (t) is the vertical input displacement, G q (n0) is the road roughness, v is the vehicle speed, and when the car is driving on a Class C road, the road roughness coefficient is 2.56×10 -5 , W t is Gaussian white noise with a mean of 0.

[0151] In order to analyze the influence of three models of traditional double-chamber oil-gas suspension, ideal inertia oil-gas suspension and nonlinear inertia oil-gas suspension on vehicle suspension performance, this paper provides three different suspension structures, S0 is the traditional double-chamber oil-gas suspension, S1 is the ideal inertia oil-gas suspension, S2 is the nonlinear inertia oil-gas suspension, and the corresponding suspension forces F 0、 F1 and F2 are as follows.

[0152]

[0153] Among them, step three is based on parameter optimization of non-dominated sorting genetic algorithm II (NSGA-II), and the specific implementation method is as follows:

[0154] NSGA-II is the full name of non-dominated sorting genetic algorithm, which is used to solve the optimization algorithm of multiple conflicting objectives and finally find the optimal solution set of Pareto graph. It has the characteristics of fast algorithm running speed and good convergence. Its optimization process is as follows Figure 6 shown.

[0155] The present invention takes the parameters of the traditional dual-chamber oil-gas suspension as a reference, and selects the initial air pressure P of the air chamber D in the inertial oil-gas suspension while ensuring that other parameters remain unchanged. D0 , the optimization range is [15,35]bar; volume V D0 , the optimization range is [0.2,1.3]L; the initial pressure P of the accumulator R chamber R0 , the optimization range is [15,35]bar, volume V R0 , the optimization range is [0.2,1.5]L, the nonlinear damping coefficient C2, the optimization range is [458.3,3000]N s / m; the inertia coefficient b optimization range is [0,1000]kg, which is the optimization object.

[0156] The vehicle selected in this invention is a passenger car, and in actual use, the emphasis is more on riding comfort. Therefore, the body acceleration should be as small as possible, and the suspension travel and tire dynamic load should also be within the normal range to ensure the normal working condition of the vehicle suspension system. There are mutual constraints for the three evaluation indicators, so the optimization objective function is as follows:

[0157]

[0158] Where: J1 is the overall evaluation index reflecting the vehicle's ride comfort and stability, while J2 is the evaluation index of the vehicle's road friendliness. This formula can directly compare the performance of the inertia container-based hydro-pneumatic suspension and the traditional hydro-pneumatic suspension. pas -RMS, SWS pas -RMS and DTL pas -RMS are the root mean square values ​​of the body acceleration, suspension working space and dynamic tire load of the traditional hydro-pneumatic suspension. BA_RMS, SWS_RMS and DTL_RMS are the root mean square values ​​of the body acceleration, suspension working space and dynamic tire load of the hydro-pneumatic suspension based on inertia vessel. W1 and W2 are the optimized weight coefficients, which are set to 0.6 and 0.4 respectively.

[0159] The parameters of the conventional dual-chamber oil-gas suspension selected by the present invention are as follows:

[0160] Table 1 Suspension parameters

[0161]

[0162] In the parameters of the inertia oil-gas suspension, the spiral pipe radius r3 is set to 0.007m, and the C1 calculation formula shows that C1 = 1621.3. Based on the NSGA-II algorithm, in the optimal solution set obtained under the reasonable range of optimization parameters, considering the overall vibration isolation performance of the suspension system, the inertia oil-gas suspension parameters that can improve the ride comfort of the vehicle as much as possible are selected, as shown in Table 2, and the performance comparison of the three different structures S0, S1 and S2 is shown in Table 3.

[0163] Table 2 Optimization parameters

[0164]

[0165] Table 3 Optimization performance comparison

[0166]

[0167] The present invention performs time domain simulation analysis on three different suspension structures. The results are shown in Table 3. Figure 7 As shown in Table 3, compared with S0, the root mean square value of the body acceleration of S1 is reduced by 9.7%. On the contrary, the root mean square value of the suspension dynamic stroke is increased by 9.1%, and the root mean square value of the tire dynamic load is also increased by 11.0%. The S1 structure improves the smoothness of the vehicle to a certain extent, but it does not play a significant positive role in improving the overall performance of the suspension system. Compared with the S0 structure, the S2 structure reduces the three evaluation indicators to varying degrees. Among them, the root mean square value of the body acceleration is reduced by 20.5%, the root mean square value of the suspension dynamic stroke is reduced by 12.7%, and the root mean square value of the tire dynamic load is reduced by 8.0%, which greatly improves the ride comfort and driving stability.

[0168] In summary, from the perspective of time domain analysis, the S2 structure not only significantly improves the vehicle's ride comfort, but also improves the vehicle's handling performance and road friendliness to a certain extent. The S1 structure only improves the ride comfort, and other performance indicators have a negative impact.

[0169] The above technical solutions only reflect the preferred technical solutions of the technical solutions of the present invention. Some changes that may be made to certain parts thereof by technicians in this technical field all reflect the principles of the present invention and fall within the protection scope of the present invention.

Claims

1. A vehicle nonlinear inertia dual-chamber oil-gas suspension structure and parameter determination method, characterized in that: The following steps are involved: Step 1: Construction of nonlinear fluid inertia container model; Step 2: 1 / 4 nonlinear inertia oil-pneumatic suspension model construction; Step 3: Parameter optimization based on non-dominated sorting genetic algorithm II; Step 4: Dynamic performance simulation analysis.

2. The vehicle nonlinear inertia dual-chamber oil-gas suspension structure and parameter determination method according to claim 1, characterized in that: The nonlinear fluid inertia container model in step 1 is constructed in the following specific implementation method: The force expression between the two end points of an ideal fluid inertia container is: During the flow process, the viscosity effect is generated due to the influence of the length, radius and temperature of the pipeline. The damping force is the main nonlinear characteristic of the fluid inertia container. When the present invention studies the double-chamber oil-gas inertia container suspension, the piston friction force is within the controllable range, so the influence of the friction force is not considered. The force expression between the two ends of the nonlinear fluid inertia container is: F=f b +f c Where Z2 and Z1 are the displacements of the two ends of the inertia container, and f b is the inertial force, f c is the nonlinear damping force; where b is the inertia coefficient, C1 and C2 are the damping coefficients, which can be expressed by the following formula: Where ρ is the oil density, l is the length of the spiral pipe, h is the pitch of the spiral channel, r4 is the spiral radius of the spiral pipe, S A is the area of ​​the hydraulic cylinder, S3 is the cross-sectional area of ​​the spiral pipe, r3 is the radius of the spiral pipe, and μ is the viscosity of the liquid; from the above formula, it can be seen that C2 and b are both related to the length of the spiral pipe, and the formula can be obtained: It can be seen from the above formula that the change of h and r4 will synchronously change b and C2 to obtain different inertia coefficients and damping coefficients, thereby establishing different nonlinear fluid inertia container models.

3. The vehicle nonlinear inertia dual-chamber oil-gas suspension structure and parameter determination method according to claim 1, characterized in that: The specific implementation method of the 1 / 4 nonlinear inertia oil-gas suspension model construction in step 2 is as follows: In the dual-chamber oil-gas suspension, the B and C chambers are connected to the damping hole through a one-way valve, and the oil flow formula is as follows: Where: is the speed of the piston moving up and down; Q1 is the oil flow between chambers B and C; S C is the C cavity area; When the oil-gas spring is in compression stroke, the oil flows through the one-way valve and the damping hole, and the flow formula is: Where: A1 is the equivalent area of ​​the damping hole, A2 is the equivalent area of ​​the one-way valve; When the oil-gas spring is in the extension stroke, the oil flows through the hole frequently, and the flow formula is: From the above formula, we can get the pressure loss caused by chambers B and C as follows: The damping force caused by the pressure loss in chambers B and C can be expressed as: Where: sign is the sign function. When the oil-gas spring is in the extension stroke, When the oil-gas spring is in the compression stroke, In the dual-chamber oil-gas suspension proposed by the present invention, there is also pressure loss when the A chamber is connected to the R chamber of the accumulator by an oil pipe. The formula is as follows: Where A3 is the cross-sectional area of ​​the pipeline, η1 and η2 are the oil pressure loss coefficients between the oil pipe and the accumulator. When the oil flows into the accumulator R, η1 = 1, and when the oil flows out of the accumulator R, η1 = 0.5; η2 is the opposite of η1; Q2 is the flow formula between chamber A and accumulator E, S A is the area of ​​cavity A; The damping force generated by the pressure loss caused by the A and R chambers can be expressed as: Therefore, the total damping force expression of the dual-chamber oil-gas suspension proposed in the present invention is: F C =F1+F2 The damping coefficient of the hydro-pneumatic suspension is:

4. The vehicle nonlinear inertia dual-chamber oil-gas suspension structure and parameter determination method according to claim 1, characterized in that: The specific implementation method of the 1 / 4 nonlinear inertia oil-gas suspension model construction in step 2 is as follows: The stiffness characteristic of the dual-chamber inertial oil-gas suspension in the present invention is the relationship between the elastic force on the main piston and the relative displacement of the piston when the damping force is not considered and the effect of oil compression on the gas volume is ignored; in the dual-chamber oil-gas suspension, when the influence of the damping characteristic is not considered, the pressure P of the A chamber is A The pressure P in the accumulator R R Equal in size, P A =P R , similarly, the pressure of chamber B is P B 、C chamber pressure P C , D cavity pressure P D Equal in size, P B =P C =P D ; The formula for the elastic force of the piston at any time is: F S =P C S C -P A S A Where: F S is the total spring force of the suspension, S C is the area of ​​the annular cavity C, S A is the area of ​​chamber A; The gas state equation of the accumulator R chamber is: Where: V R0 、V RS and V R are the gas volumes of the accumulator R chamber at the initial state, equilibrium position and transient state respectively; P R0 , P RS and P R are the gas pressures in the initial state, equilibrium position and instantaneous state of the R chamber of the accumulator respectively; n is the gas polytropic index; The gas state equation in chamber D is: Where: V D0 、V DS and V D are the gas volumes in the initial state, equilibrium position and instantaneous state of cavity D respectively; P D0 , P DS and P D are the gas pressures in the initial state, equilibrium position and instantaneous state of the D chamber of the accumulator respectively; n is the gas polytropic index. When the piston rod is in the equilibrium position, let the downward movement be x. The change of the gas volume in the R and D chambers of the accumulator in the instantaneous state is as follows: V R =V RS +S A ·x V D =V DS -S C ·x The mass on the suspension spring is M. When the oil-gas spring moves from the initial state to the equilibrium state, the pressure in each chamber is the same, and the process is regarded as an isothermal state. Take n = 1, and the pressure and volume of the air cavity are as follows: By combining the above formulas, we can get: For the instantaneous elastic force F S The corresponding stiffness coefficient can be obtained by differentiating the displacement x:

5. The vehicle nonlinear inertia dual-chamber oil-gas suspension structure and parameter determination method according to claim 1, characterized in that: In the step 2, a 1 / 4 nonlinear inertia oil-gas suspension model is constructed, and the specific implementation method is as follows: Construct the dynamic equation of the 1 / 4 vehicle suspension structure: In the formula, m s The sprung mass is set to 400 kg, m u The unsprung mass is set to 40 kg, K t The tire stiffness is set to 200 kN·m -1 , F is the force between the suspensions; Z s , Z u are the vertical displacements of the sprung and unsprung masses, Z r is the vertical displacement of the road surface input, and its formula is: In the formula, Z r (t) is the vertical input displacement, G q (n0) is the road roughness, v is the vehicle speed, and when the car is driving on a Class C road, the road roughness coefficient is 2.56×10 -5 , W t is a Gaussian white noise with a mean of 0; In order to analyze the influence of three models of traditional double-chamber oil-gas suspension, ideal inertia oil-gas suspension and nonlinear inertia oil-gas suspension on vehicle suspension performance, this paper provides three different suspension structures, S0 is the traditional double-chamber oil-gas suspension, S1 is the ideal inertia oil-gas suspension, S2 is the nonlinear inertia oil-gas suspension, and the corresponding suspension forces F 0、 F1 and F2 are as follows; 6. The vehicle nonlinear inertia dual-chamber oil-gas suspension structure and parameter determination method according to claim 1, characterized in that: The specific implementation method of the parameter optimization based on the non-dominated sorting genetic algorithm II in step 3 is as follows: The present invention takes the parameters of the traditional dual-chamber oil-gas suspension as a reference, and selects the initial air pressure P of the air chamber D in the inertial oil-gas suspension while ensuring that other parameters remain unchanged. D0 , the optimization range is [15,35]bar; volume V D0 , the optimization range is [0.2,1.3]L; the initial pressure P of the accumulator R chamber R0 , the optimization range is [15,35]bar, volume V R0 , the optimization range is [0.2,1.5]L, the nonlinear damping coefficient C2, the optimization range is [458.3,3000]N s / m; the inertia coefficient b optimization range is [0,1000]kg, which is the optimization object; The vehicle selected in the present invention is a passenger car, and in actual use, more emphasis is placed on riding comfort. Therefore, the vehicle body acceleration should be as small as possible, and the suspension dynamic stroke and tire dynamic load should also be within the normal range to ensure the normal working condition of the vehicle suspension system; there are mutual constraints on the three evaluation indicators, so the optimization objective function is as follows: Where: J1 is the overall evaluation index reflecting the vehicle's ride comfort and stability, while J2 is the evaluation index of the vehicle's road friendliness; this formula can directly compare the performance of the inertia container-based hydro-pneumatic suspension and the traditional hydro-pneumatic suspension; where BA pas -RMS, SWS pas -RMS and DTL pas -RMS are the root mean square values ​​of the vehicle acceleration, suspension working space and dynamic tire load of the traditional hydro-pneumatic suspension; BA_RMS, SWS_RMS and DTL_RMS are the root mean square values ​​of the vehicle acceleration, suspension working space and dynamic tire load of the hydro-pneumatic suspension based on inertia vessel; W1 and W2 are the optimized weight coefficients, which are set to 0.6 and 0.4 respectively.

7. The vehicle nonlinear inertia dual-chamber oil-gas suspension structure and parameter determination method according to claim 1, characterized in that: The dynamic performance simulation analysis in step 4 is specifically implemented as follows: in the simulation analysis, the vehicle body acceleration, suspension dynamic deflection, and tire dynamic load are used as indicators, and the above three indicators are subjected to road excitation analysis at a vehicle speed of 20 m / s using Matlab and Simulink; the present invention fully considers the nonlinear parasitic damping force factor of the fluid inertia container, establishes three structural models of the traditional double-chamber oil-gas suspension S0, the ideal inertia container oil-gas suspension S1, and the nonlinear inertia container oil-gas suspension S2, and performs dynamic simulation analysis.

Images