Pump valve series-parallel electro-hydraulic servo system and configuration method thereof

Abstract

The invention discloses a pump valve series-parallel electro-hydraulic servo system and a configuration method thereof, and the configuration method comprises the steps: constructing a pump valve model set based on a current procureable product, and carrying out the screening and comprehensive evaluation of each pump valve combination through combining an optimization model and a multi-objective optimization algorithm; under the constraint conditions of meeting the system flow requirement, the load thrust requirement and the target tracking performance, the pump valve model combination most suitable for current engineering application is accurately matched. According to the method, the automation and optimization degree of system configuration can be remarkably improved, and the equipment type selection redundancy, the system energy consumption and the cost are effectively reduced while the tracking precision and the dynamic performance of the system are guaranteed.

Description

TECHNICAL FIELD

[0001] The present application relates to the field of hydraulic systems, in particular to a pump-valve hybrid electro-hydraulic servo system and a configuration method thereof. BACKGROUND

[0002] The pump-valve hybrid hydraulic system combines the use of variable pumps and proportional valves, servo valves and other control valves, and has high energy efficiency of pump control system and high responsiveness of valve control system, which significantly improves the comprehensive performance of the hydraulic system. In the low-speed and large-load working condition, the variable pump can directly adjust the flow, reduce the throttling loss, and realize high-efficiency energy supply; in the high-speed, small-load or fast dynamic response scene, the control valve can quickly adjust the output, improve the system control precision and response speed, and is widely used in shipbuilding and marine engineering, engineering machinery and civil industry fields.

[0003] The pump-valve hybrid system has a complex structure, involves multiple components, multiple operating modes and coupled dynamic behaviors, and the performance parameters of the pump and valve are highly discrete, resulting in a mixed multi-objective optimization problem for configuration optimization, which is difficult to solve.

[0004] The overall performance and energy efficiency of the pump-valve hybrid hydraulic system largely depend on the structural matching and dynamic coordination mechanism between the core power unit hydraulic pump and the execution control unit hydraulic valve. The ideal configuration needs to meet the dynamic demand of the load, such as fast response and accurate control, while minimizing energy consumption and control cost. However, engineering practice shows that these key objectives, dynamic performance, energy efficiency and cost, often conflict with each other. For example, to pursue faster response speed, a larger diameter valve or a higher displacement pump is usually selected, which often results in significant throttling loss and higher purchase cost; on the contrary, excessive pursuit of low cost or low energy consumption may severely restrict the dynamic ability of the system. Today, the selection of pump-valve components still generally relies on manual experience or existing engineering templates, lacking systematic and intelligent optimization design methods, which easily leads to unreasonable resource allocation, low system efficiency and high equipment cost. With the improvement of industrial automation, hydraulic systems face multiple challenges in dynamic response, energy consumption control and cost constraints. Traditional design methods based on experience or single-objective optimization are difficult to fully represent the complex trade-off relationship between multi-objectives of pump-valve systems. Therefore, it is urgent to introduce intelligent optimization algorithms to systematically configure and optimize pump-valve types, and to achieve the coordinated balance of overall performance and cost.

[0005] Chinese patent CN119353280 A discloses an electrically controlled hydraulic power steering system and a multi-objective optimization method thereof by optimizing the mechanical and control parameters of the hydraulic power steering system. Chinese patent CN107600173 A discloses an automobile hydraulic variable transmission ratio steering system and a multi-objective optimization method thereof by optimizing the key parameters of the system through an algorithm. The above patents are based on the optimization of system control parameters to inversely deduce the mechanical structure design, thereby achieving the performance target. However, for the hydraulic system, if the matching pump and valve are directly developed by optimizing the key parameters of the system, customized design is often required, which may significantly increase the system integration cost and reduce the economic efficiency and feasibility of engineering application. In traditional engineering design, the selection of existing pump and valve product types usually depends on the experience of engineers or parameter estimation based on a simplified model, which lacks systematicness and global optimality, and it is often difficult to achieve a good balance among multiple performance requirements such as response speed, control accuracy, energy efficiency and cost. SUMMARY

[0006] The purpose of the present application is to provide a pump-valve hybrid electric-hydraulic servo system with simplified pump-valve hybrid structure and organic coupling of pump and valve; another purpose of the present application is to provide a pump-valve hybrid electric-hydraulic servo system and a configuration method thereof for project working condition requirements.

[0007] Technical solution: The pump-valve hybrid electric-hydraulic servo system comprises a valve control subsystem, a pump control subsystem, a three-phase motor, an oil tank and an asymmetric hydraulic cylinder; the valve control subsystem is composed of m proportional valves and n servo valves in parallel, and the pump control subsystem is composed of h bidirectional variable pumps in parallel; the three-phase motor is connected with the fixed displacement pump, the oil inlet of the fixed displacement pump is connected with the oil tank, the three-phase motor rotates to drive the fixed displacement pump to work, the hydraulic oil output by the fixed displacement pump is distributed to each proportional valve and servo valve through a pipeline after passing through a check valve and a filter; the three-phase motor is connected with the pump control subsystem, the three-phase motor rotates to drive the bidirectional variable pump to work, the bidirectional variable pump sucks oil from the oil tank and outputs the hydraulic oil to the main oil circuit, the hydraulic oil passes through a check valve and a filter and is connected with the oil circuit of the valve control subsystem in parallel; the hydraulic cylinder is connected with the proportional valve and the servo valve, and is also connected with the main circuit of the bidirectional variable pump, the piston rod of the hydraulic cylinder is connected with a displacement sensor, and the displacement sensor is connected with a controller; the output end of the controller is connected with the signal input end of the proportional valve, the servo valve and the bidirectional variable pump.

[0008] The pump-valve hybrid electric-hydraulic servo system and the configuration method thereof, the pump-valve hybrid electric-hydraulic servo system is the above-mentioned pump-valve hybrid electric-hydraulic servo system, and the configuration method comprises the following steps:

[0009] (1) constructing a proportional valve product type set, a servo valve product type set and a bidirectional variable pump product type set according to product types, and setting performance index vectors of the bidirectional variable pump, the proportional valve and the servo valve;

[0010] (2) In the structure of the valve-pump hybrid system, the dynamics model of the electro-hydraulic servo system is constructed according to the flow superposition principle of the pump-valve subsystem, and combined with the flow continuity and mechanical balance relationship of the asymmetric hydraulic cylinder;

[0011] (3) The multi-objective optimization model of the electro-hydraulic servo system is constructed with the device cost and system throttling loss energy consumption as the objective function, and the system flow constraint, system thrust constraint, and system absolute error integral value constraint under multiple working conditions as the constraint conditions;

[0012] (4) The particle swarm optimization algorithm is used to obtain the optimal pump-valve combination scheme.

[0013] Further, the performance indicators of the bidirectional variable pump include the rated flow, rated pressure, cost, displacement gradient, variable mechanism adjustment angle, and internal and external leakage coefficient of the bidirectional variable pump, the performance indicators of the proportional valve include the flow gain, flow-pressure gain, rated pressure, rated flow, and cost of the proportional valve, and the performance indicators of the servo valve include the flow gain, flow-pressure gain, rated pressure, rated flow, and cost of the servo valve.

[0014] Further, in step (2), the output flow Q v of the valve control subsystem is

[0015]

[0016] wherein, represents the output flow of the proportional valve corresponding to the 1st, 2nd, and mth positions of the hydraulic system, respectively, and represents the proportional valve type Vb l ,Vb i ,Vb t in the 1st, 2nd, and mth positions, respectively, and K bs1 ,K bs2 ,K bsm are the flow gain and flow-pressure gain coefficients in the performance indicator vector; K b1 ,u b2 ,u bm are the 1st, 2nd, and mth position proportional valve amplifier gains; and u represents the output flow of the servo valve corresponding to the 1st, 2nd, and nth positions of the hydraulic system, respectively, and represents the servo valve type Vs z ,Vs a ,Vs d in the 1st, 2nd, and nth positions, respectively, and K ss1K ss2 K ssn K s1 K s2 K sn K L K

[0017] Q p K

[0018]

[0019] K K K o K q K y K sp1 K sp2 K sph K p1 K p2 K ph K

[0020] Q L K

[0021] Q L Q v Q p

[0022] The output flow equation of the combined electro-hydraulic servo system, the flow continuity equation of the asymmetric hydraulic cylinder and the force balance equation of the asymmetric hydraulic cylinder are as follows:

[0023]

[0024]

[0025]

[0026] wherein A is the effective action area of the hydraulic cylinder piston; V t is the sum of the volumes of the pipeline and the valve cavity; β e is the equivalent bulk modulus of the hydraulic oil, n is the ratio of the effective action areas of the rod cavity and the rodless cavity of the hydraulic cylinder, m tis the total mass of the piston and load; x p is the ideal displacement of the hydraulic cylinder piston rod; B p is the viscous damping coefficient of the piston and load; K p is the elastic stiffness coefficient; F L is the external load force acting on the piston rod; is the third order derivative of the ideal displacement of the hydraulic cylinder piston rod, is the second order derivative of the ideal displacement of the hydraulic cylinder piston rod, is the first order derivative of the ideal displacement of the hydraulic cylinder piston rod, d is the system disturbance, M, B, C, D represent the equivalent mass, damping coefficient, elastic coefficient, gain coupling coefficient of the system respectively; K1 is the first proportional input gain, K m is the mth proportional input gain, K n is the nth servo valve input gain, K h is the hth variable pump input gain, K CE is the total gain coefficient of the control system;

[0027] Take x1=x p , x1, x2 and x3 represent the displacement, velocity and acceleration in the hydraulic cylinder piston respectively; the state equation of the electro-hydraulic servo system is as follows:

[0028]

[0029] Further, in step (3), the objective function J is

[0030] J=ω1J1+ω2J2

[0031] Wherein, ω1, ω2 are optimization weights, ω1+ω2=1, J1 is the total cost of the equipment, J2 is the energy loss of the system throttling loss;

[0032] The total cost of the equipment J1 is

[0033]

[0034] Wherein, is the first position index l corresponding to the proportional valve type Vb l The cost in the performance index vector,

[0035] is the mth position index t corresponding to the proportional valve type Vb t The cost in the performance index vector, is the nth position index d corresponding to the servo valve type Vs d The cost in the performance index vector, is the hth position index y corresponding to the variable pump type P yCost in performance index vector;

[0036] System throttling loss energy loss J2 is

[0037]

[0038] E Vb = q Vb · Δp Vb

[0039] E Vs = q Vs · Δp Vs

[0040] Wherein, Vb is the proportional valve type corresponding to the 1st, mth position index l, t respectively l , Vb t Throttling loss energy generated by, Vs is the servo valve type corresponding to the 1st, nth position index z, d respectively z , Vs d Throttling loss energy generated by; E Vb E is the throttling loss energy of the proportional valve Vs E is the throttling loss energy of the servo valve; Δp b Δp is the proportional valve port pressure difference s Δp is the servo valve port pressure difference Vb q is the proportional valve output flow Vs q is the servo valve output flow.

[0041] Further, in step (3), the proportional valve port pressure difference Δp b is

[0042]

[0043] The servo valve port pressure difference Δp s is

[0044]

[0045] Wherein, K qs is the servo valve flow gain coefficient sc K is the servo valve flow-pressure gain coefficient s u is the servo valve input electric signal.

[0046] Further, in step (3), the system flow constraint is

[0047]

[0048] Wherein, Q is the rated flow required by the system work, Variable pump model P corresponding to the 1th position index o, respectively o y Rated flow in the performance index vector, Proportional valve model Vb corresponding to the 1th position index l, respectively l t Rated flow in the performance index vector, Servo valve model Vs corresponding to the 1th position index z, respectively z d Rated flow in the performance index vector.

[0049] Further, in step (3), the system thrust constraint is

[0050]

[0051] Where, A is the piston effective area, P L is the load pressure, P rated is the system rated pressure, and λ is the safety margin coefficient; Variable pump model P corresponding to the 1th position index o, respectively o Rated pressure in the performance index vector, Proportional valve model Vb corresponding to the 1th position index l, respectively l Rated pressure in the performance index vector, Servo valve model Vs corresponding to the 1th position index z, respectively z Rated pressure in the performance index vector.

[0052] Further, in step (3), the system absolute error integral value constraint under multiple working conditions is

[0053]

[0054] Where, IAE i is the system absolute error integral value under the currently selected working condition, i = 1, 2, 3, δ i (t) is the working condition indicator function, m(t) ∈ {1, 2, 3} is the currently selected working condition, and x d is the actual tracking displacement of the hydraulic cylinder piston rod.

[0055] Further, a penalty term is introduced into the target function J to obtain the fitness function J penalized , as follows:

[0056] J penalized = J + Penalty

[0057]

[0058] ​​​Wherein, alpha1, alpha2, alpha3 are penalty coefficients.

[0059] Advantages: Compared with the prior art, the present application has the following advantages: 1. The present application fully utilizes the performance advantages of the pump-valve hybrid hydraulic system, reasonably selects the pump and valve and optimally configures the system structure, improves the control accuracy, energy consumption level and equipment cost, and the stability and service life of the system; 2. The present application fuses the pump-valve flow superposition principle and the dynamic characteristics of the asymmetric hydraulic cylinder, establishes a unified modeling system suitable for various driving combinations, systematically identifies the key structural parameters of the pump and valve that significantly affect the dynamic performance of the system, and accordingly establishes a multi-objective optimization model with system energy consumption and equipment cost as the optimization objectives; 3. By constructing a pump-valve type set based on currently available products, and combining the optimization model and multi-objective optimization algorithm to screen and comprehensively evaluate each pump-valve combination, the most suitable pump-valve type combination for current engineering applications is accurately matched under the constraint conditions of meeting the system flow demand, load thrust requirement and absolute error integral value under different working conditions. This method can significantly improve the automation and optimization level of system configuration, while ensuring the tracking accuracy and dynamic performance of the system, effectively reducing equipment selection redundancy, system energy consumption and cost. BRIEF DESCRIPTION OF DRAWINGS

[0060] Figure 1 Fig. 1 is a structural schematic diagram of a pump-valve hybrid drive electro-hydraulic servo system of the present application;

[0061] Figure 2 Fig. 4 is a flow chart of a pump-valve hybrid configuration optimization method for a multi-objective optimization electro-hydraulic servo system of the present application;

[0062] Figure 3 Fig. 5 is a flow chart of a pump-valve hybrid configuration based on a particle swarm optimization algorithm of the present application. DETAILED DESCRIPTION

[0063] The present application will be further described below in conjunction with the drawings.

[0064] As Figure 1As shown, the pump-valve hybrid electric-hydraulic servo system of the present application comprises a valve control subsystem, a pump control subsystem, a constant-displacement pump 1, a three-phase motor 2, a one-way valve 3, a filter 4, a proportional valve 5, a servo valve 6, an asymmetric hydraulic cylinder 7, a displacement sensor 8, an overflow valve 9, a bidirectional variable pump 10, a controller 11, and a switch valve 12. The valve control subsystem is composed of m proportional valves and n servo valves in parallel, and the pump control subsystem is composed of h bidirectional variable pumps in parallel. The three-phase motor is connected to the constant-displacement pump, the oil inlet of the constant-displacement pump is connected to an oil tank, the three-phase motor rotates to drive the constant-displacement pump to work, the hydraulic oil output by the constant-displacement pump is distributed to each proportional valve and servo valve through a pipeline after passing through the one-way valve and the filter. The three-phase motor is connected to the pump control subsystem, the three-phase motor rotates to drive the bidirectional variable pump to work, the bidirectional variable pump sucks oil from the oil tank and outputs the hydraulic oil to the main oil circuit, and the hydraulic oil is connected in parallel with the oil circuit of the valve control subsystem after passing through the one-way valve and the filter. The hydraulic cylinder is connected to the proportional valve and the servo valve, and is also connected to the main circuit of the bidirectional variable pump, the piston rod of the hydraulic cylinder is connected to the displacement sensor, and the displacement sensor is connected to the controller. The signal input ends of the proportional valve, the servo valve, and the bidirectional variable pump are respectively connected to the output end of the controller. The m output ends of the controller are connected to the signal input ends of the proportional valves Vb1…Vbm, the n output ends are connected to the signal input ends of the servo valves Vs1…Vsn, and the h output ends are connected to the signal input ends of the bidirectional variable pumps P1…Ph.

[0065] The pump-valve hybrid electric-hydraulic servo system of the present application and its configuration method, the pump-valve hybrid electric-hydraulic servo system is the above-mentioned pump-valve hybrid electric-hydraulic servo system, and the configuration method comprises the following steps:

[0066] (1) Establishing a current selectable proportional valve product model set S Vb ={Vb1, Vb2, …, Vb w}, a servo valve product model set S Vs ={Vs1, Vs2, …, Vs r}, and a variable pump product model set S P ={P1, P2, …, P s}, wherein w, r, and s are the total number of indexes of each product, and the variable pump, the proportional valve, and the servo valve are set as performance index vectors when the indexes are different. P o is the variable pump model corresponding to the oth index in the model set S P , Vb l is the proportional valve model corresponding to the lth index in the model set S Vb , and Vs z is the servo valve model corresponding to the zth index in the model set S Vs .

[0067] Further, the variable pump, the constant-displacement pump, the proportional valve, and the servo valve corresponding index model performance index vectors are represented as follows:

[0068] where, is the variable pump rated flow, is the variable pump rated pressure, is the variable pump cost, is the variable pump displacement gradient, γ o is the variable mechanism adjustment swing angle, is the variable pump internal and external leakage coefficient, is the proportional valve flow gain, is the proportional valve flow pressure gain, is the proportional valve rated pressure, is the proportional valve rated flow, is the proportional valve cost, is the servo valve flow gain, is the servo valve flow pressure gain, is the servo valve rated pressure, is the servo valve rated flow, is the servo valve cost.

[0069] Further, similarly, P q , P y can represent the variable pump model corresponding to the q,y index in the model set S P , Vb i , Vb t represents the proportional valve model corresponding to the i,t index in the model set S Vb , Vs a , Vs d represents the servo valve model corresponding to the a,d index in the model set S Vs , and their performance index vectors can be represented as

[0070] (2) Establish the system dynamics model, through the pump, valve sub-model and non-symmetrical hydraulic cylinder model, the complete system dynamics model can be constructed, and the key model is represented as follows:

[0071] (21) Servo valve model

[0072] The linear flow equation of a single servo valve can be represented as:

[0073] q s = K qs x s - K sc P L

[0074] where, q s represents the output flow of the servo valve, K qs and K scx s x L p

[0075] x

[0076] x s = K ss u s

[0077] x ss = K s u

[0078] q

[0079] q s = K qs K ss u s -K sc P L

[0080] (22) Proportional valve model

[0081] q

[0082] q b = K qb x b -K bc P L

[0083] q b = K qb x bc -K b P L

[0084] x

[0085] x b = K bs u b

[0086] x bs = K b u

[0087] q

[0088] ​q b = K qb K bs u b - K bc P L

[0089] (23) Asymmetric cylinder model

[0090] The flow continuity equation of the asymmetric cylinder is:

[0091]

[0092] where A is the effective area of the cylinder piston; V t is the sum of the volumes of the pipeline, valve chamber, etc.; β e is the equivalent bulk modulus of the hydraulic oil; n is the ratio of the actual effective areas of the rod chamber and the rodless chamber of the hydraulic cylinder; C t is the total leakage coefficient of the hydraulic cylinder, where C t = C i + C e / 2 + C tc , C i and C e are the internal and external leakage coefficients of the hydraulic cylinder, and n is the ratio of the effective areas of the rod chamber and the rodless chamber of the hydraulic cylinder.

[0093] The force balance equation of the asymmetric cylinder is:

[0094]

[0095] where m t is the total mass of the piston and the load; x p is the displacement of the hydraulic cylinder piston rod; B p is the viscous damping coefficient of the piston and the load; K p is the elastic stiffness coefficient; and F is the external load force acting on the piston.

[0096] (24) Variable pump model

[0097] The output flow equation of the variable pump can be expressed as:

[0098] q P = K d ω p γ - C tp P L

[0099] where K d is the displacement gradient of the variable pump, ω p is the rotational speed of the variable pump; γ is the adjustment swing angle of the variable mechanism; and C tp = C ip + C epwhere C ip and C ep are the inner and outer leakage coefficients of the pump, respectively.

[0100] The relationship between the pump speed and the input signal u p is:

[0101] ω p = K sp u p

[0102] where K sp is the pump amplifier gain coefficient.

[0103] Therefore, the variable pump output flow equation is:

[0104] q P = K d K sp γu p -C tp P L

[0105] The structure of the pump-valve hybrid hydraulic system can be composed of a valve control subsystem obtained by connecting m proportional valves and n servo valves in parallel, and a pump control subsystem composed of h variable pumps. The output flow Q v of the valve control subsystem can be represented as:

[0106]

[0107] where, represents the output flow of the proportional valve corresponding to the first, second, and mth positions of the hydraulic system, respectively, with indexes l, i, and t, respectively. represents the type of proportional valve corresponding to the first, second, and mth positions, respectively, with indexes l, i, and t, respectively. l Vb i Vb t are the flow gain and flow-pressure gain coefficients in the performance index vector; K bs1 K bs2 K bsm are the amplifier gains of the first, second, and mth proportional valves; u b1 u b2 u bm are the input electrical signals of the first, second, and mth proportional valves, respectively. represents the output flow of the servo valve corresponding to the first, second, and nth positions of the hydraulic system, respectively, with indexes z, a, and d, respectively. represents the type of servo valve corresponding to the first, second, and nth positions, respectively, with indexes z, a, and d, respectively. z Vs a Vs d are the flow gain and flow-pressure gain coefficients in the performance index vector; Kss1 ,K ss2 ,K ssm is the 1st, 2nd, nth position servo valve amplifier gain; u s1 ,u s2 ,u sn is the 1st, 2nd, nth position servo valve input electrical signal; p L is the load pressure.

[0108] Pump control subsystem output flow model Q p can be expressed as:

[0109]

[0110] where, represents the variable pump output flow corresponding to the 1st, 2nd, hth position index o, q, y of the hydraulic system; respectively, represents the variable pump model corresponding to the 1st, 2nd, hth position index o, q, y of the hydraulic system o ,P q ,P y is the displacement gradient, mechanism adjustment swing angle and leakage coefficient in the performance index vector; K sp1 ,K sp2 ,K sph is the 1st, 2nd, hth position variable pump amplifier gain coefficient; u p1 ,u p2 ,u ph is the 1st, 2nd, hth position variable pump input electrical signal.

[0111] Therefore, the system output flow is the sum of the valve control subsystem and pump control subsystem output flow, which can be expressed as:

[0112] Q L = Q v + Q p

[0113] The joint system output flow equation, the flow continuity equation of the asymmetric hydraulic cylinder and the force balance equation of the asymmetric hydraulic cylinder can be obtained:

[0114]

[0115] where,

[0116] where, A is the effective action area of the hydraulic cylinder piston; V t is the sum of the volumes of the pipeline, valve cavity, etc.; β e is the equivalent volume elastic modulus of the hydraulic oil, n is the ratio of the effective action area of the rod cavity to the rodless cavity of the hydraulic cylinder, m t is the total mass of the piston and the load; xp is the displacement of the hydraulic cylinder piston rod; B p is the viscous damping coefficient of the piston and load; K p is the elastic stiffness coefficient; F L is the external load force acting on the piston rod.

[0117] Take the state of the controlled system x1=x p , x1, x2 and x3 represent the displacement, velocity and acceleration in the hydraulic cylinder piston respectively. Therefore, the state equation of the controlled system is expressed as follows:

[0118]

[0119] Starting from the state space and transfer function model of the controlled system, the embedding method of the structural parameters of each type of pump and each type of valve in the dynamic model is analyzed, the mapping relationship of the structural parameters of the pump and the valve in the model is analyzed, and the influence mechanism of the structural parameters on the input-output characteristics and the internal state evolution process of the system is clarified. It can be observed that the structural parameters of the valve that mainly affect the system are: the flow gain K qb of the proportional valve, the flow pressure gain K bc of the proportional valve, the flow gain K qs of the servo valve, the flow pressure gain K sc of the servo valve; the structural parameters of the variable pump that mainly affect the system are: the displacement gradient K d of the variable pump, the adjusting swing angle γ of the variable mechanism, and the internal and external leakage coefficients C tp of the pump. These parameters jointly determine the dynamic response characteristics and steady-state performance of the system through coupling.

[0120] (3) The performance index vectors of the variable pump, the proportional valve and the servo valve are used as optimization configuration variables to construct an optimization objective function or to impose constraint conditions.

[0121] (4) An optimization model is established, and the objective functions of the equipment cost and the energy loss of the system throttling loss are constructed, which are expressed as follows:

[0122] J=ω1J1+ω2J2

[0123] Where ω1, ω2 are the optimization weights of the objective function, which can be valued according to the actual situation, and they satisfy the condition ω1+ω2=1.

[0124] The total cost of the equipment J1 can be expressed as follows:

[0125]

[0126] Where, ​Vb for the proportional valve model corresponding to the 1st position index l l Cost in the performance index vector, Vb for the proportional valve model corresponding to the mth position index t t Cost in the performance index vector, Vs for the servo valve model corresponding to the nth position index d d Cost in the performance index vector, P for the variable pump model corresponding to the hth position index y y Cost in the performance index vector.

[0127] The J2 is the throttling loss energy loss of the system, which is represented as follows:

[0128]

[0129] Where, Vb for the proportional valve model corresponding to the 1st, mth position index l, t l Vb t Throttling loss energy loss generated, Vs for the servo valve model corresponding to the 1st, nth position index z, d z Vs d Throttling loss energy loss generated.

[0130] Throttling loss energy loss of the proportional valve and servo valve E Vb E Vs Can be represented as follows:

[0131] E Vb = q Vb · Δp Vb

[0132] E Vs = q Vs · Δp Vs

[0133] Where, Δp b is the proportional valve port pressure difference, which can be represented as:

[0134]

[0135] Similarly, Δp s is the servo valve port pressure difference, which can be represented as:

[0136]

[0137] Further, the system flow constraint can be represented as:

[0138]

[0139] where Q is the required flow rate of the system, are the variable pump model numbers corresponding to the 1th, hth position index respectively o y the required flow rate in the performance index vector, are the proportional valve model numbers corresponding to the 1th, mth position index respectively l, t l t the required flow rate in the performance index vector, are the servo valve model numbers corresponding to the 1th, nth position index respectively z, d z d the required flow rate in the performance index vector.

[0140] Further, the system thrust constraint can be expressed as:

[0141]

[0142] where A is the piston effective area, P L is the load pressure, P rated is the system rated pressure, and λ is the safety margin coefficient. denotes the variable pump model number corresponding to the index o o the required pressure in the performance index vector, denotes the proportional valve model number corresponding to the index l l the required pressure in the performance index vector, denotes the servo valve model number corresponding to the index z z the required pressure in the performance index vector.

[0143] Combined with the state space expression in step (2), a PID controller is added, which calculates the control output signal according to the error value, and its expression can be expressed as:

[0144]

[0145] where u(t) is the control output signal; e(t) is the displacement tracking error; K p is the proportional gain, used to quickly respond to errors; K i is the integral gain, used to eliminate steady-state errors; K d is the differential gain, used to suppress system oscillation.

[0146] The proportional valve input signal u b in the system can be expressed as:

[0147]

[0148] The servo valve input signal u​​​s , which can be expressed as:

[0149]

[0150] The input signal u of the variable pump in the system p , which can be expressed as:

[0151]

[0152] The actual displacement tracking error e of the hydraulic cylinder piston rod is:

[0153] e = x p - x d

[0154] wherein x p is the ideal tracking displacement, and x d is the actual tracking displacement.

[0155] Further, the absolute error integral value constraint under different working conditions in step (4) can be expressed as follows:

[0156]

[0157] wherein δ i (t) is a working condition indication function, which can be expressed as:

[0158]

[0159] wherein m(t) ∈ {1, 2, 3} is the currently selected working condition. It can be divided into three working conditions, and is specifically expressed as:

[0160] i = 1, x p (t) = A sin(ωt + φ) + B

[0161]

[0162] At this time, the IAE i value of the system under different working conditions is obtained, and the IAE i values of other working conditions are 0.

[0163] Further, the following constraints are set according to different working condition requirements:

[0164] The system tracks a sine signal:

[0165]

[0166] The system tracks a ramp signal:

[0167]

[0168] The system tracks a trapezoidal signal:

[0169]

[0170] where ε1, ε2, ε3 are empirical set thresholds.

[0171] (5) To solve the constructed multi-objective optimization model, the particle swarm optimization (PSO) algorithm is introduced to efficiently explore the pump-valve configuration space with its high global search ability, so as to realize the optimal matching and performance improvement of the pump-valve combination scheme. The main steps of the optimization algorithm are as follows:

[0172] (51) Particle coding: To adapt to the discrete structure optimization characteristics of this problem, integer coding and fixed-length vector coding are used to construct the particle structure. The dimension of each particle is composed of the following two types of variables:

[0173] Pump-valve number variables: the number of proportional valves m, the number of servo valves n, and the number of variable pumps h, and set the maximum number of proportional valves M max , the maximum number of servo valves N max , and the maximum number of variable pumps H max ;

[0174] The type of the pump-valve at the i-th position of the system: the index type is selected from the given type set S P , S Vb , S Vs .

[0175] Further, the position vector coding of each particle is expressed as:

[0176] X = [m, n, h, P1 o , … Ph y , …, PH max , Vb1 l , … Vbm t , …, VbMmax, Vs1 z , …, Vsn d , …, VsN max ]

[0177] Where m ∈ [0, M max ], n ∈ [0, N max ], h ∈ [0, H max ] are integers.

[0178] P1 o represents the variable pump type P o corresponding to the first position index o, and Ph y represents the variable pump type P o corresponding to the h-th position index y., PHmax represents the maximum number of positions configurable by the variable pump, Vb1 l represents the proportional valve type Vb corresponding to the 1st position index l l , Vbm t represents the proportional valve type Vb corresponding to the mth position index t t , VbMmax represents the maximum number of positions configurable by the proportional valve, Vs1 z represents the servo valve type Vs corresponding to the 1st position index z z , Vsn d represents the servo valve type Vs corresponding to the n-th position index d d , VsNmax represents the maximum number of positions configurable by the servo valve.

[0179] If h < Hmax, h+1 ~ Hmax, the variable pump positions in this part are not enabled, and the proportional valve and servo valve are consistent.

[0180] (52) Initialize the particle swarm: set the number of particles to N s , and the maximum number of iterations to k max . When the number of iterations k = 1, for each particle α: the number of proportional valves m = randint(0, M max ), the number of servo valves n = randint(0, N max ), the number of variable pumps h = randint(0, H max ), initialize the velocity vector v α,k+1 , and initialize the position vector x α,k+1 .

[0181] (53) Fitness function: based on the objective function J, combined with hydraulic system performance constraints such as flow matching, load capacity, tracking error, and penalty term to construct the overall fitness function J penalized , which is as follows:

[0182] J penalized = J + Penalty

[0183]

[0184] Where α1, α2, α3 are penalty coefficients used to balance the influence of each constraint condition on the optimization target.

[0185] This structure ensures that not only the target performance function value is considered in the optimization process, but also the configuration scheme that does not meet the working condition requirements is appropriately penalized, improving the feasibility and practicality of the converged solution.

[0186] (54) Particle swarm update process: according to the update rule of the particle swarm algorithm, the particle velocity update can be represented as:

[0187] v α,k+1 = w · v α,k + c1 · r1 · (p α,best - x α,k ) + c2 · r2 · (g best - x α,k )

[0188] where v α,k and x α,k are the velocity and position of particle α at the kth generation, p α,best and g best represent the personal best position and global best position, respectively.

[0189] The particle position update formula can be expressed as:

[0190] x α,k+1 = round(x α,k + v α,k+1 )

[0191] The updated real position is rounded to ensure that the variable remains an integer.

[0192] The particle boundary processing formula can be expressed as:

[0193] x α,k+1 = min(x max , max(x min , x α,k+1 ))

[0194] The integer value is limited within the feasible range to prevent index out-of-bounds or invalid device quantities.

[0195] (55) Convergence criterion and output strategy: Set the maximum number of iterations N max or the convergence threshold δ of the objective function as the algorithm termination condition.

[0196] IF k < k max , J penalized < δ, let k = k + 1, then jump to (53) to start the next round of loop.

[0197] IF η = η max or J penalized ≥ δ, then the experiment is stopped and the optimal combination X best is output.

[0198] The output of the current global optimal particle corresponding to the pump valve type combination can be expressed as:

[0199] X = [m, n, h, P1 o * , …, Ph y * , …, PHmax Vb1 l * Vbm t * VbMmax, Vs1 z * Vsn d * VsN max ]

[0200] The combination is the optimal configuration scheme that meets the engineering constraints and optimizes the target value in the current design space.

[0201] Through the proposed particle swarm optimization algorithm, the optimal pump valve type combination that meets multiple engineering constraint conditions can be efficiently screened, realizing the collaborative optimization between the performance indicators of the hydraulic system, such as flow matching, output pressure, tracking accuracy, and the economic target of the system, as well as the cost of equipment and operating energy consumption, and constructing an efficient and intelligent configuration scheme for practical engineering applications.

Claims

1. A pump-valve hybrid electro-hydraulic servo system, characterized by, The pump-valve hybrid electro-hydraulic servo system comprises a valve control subsystem, a pump control subsystem, a three-phase motor, an oil tank and an asymmetric hydraulic cylinder; the valve control subsystem is composed of m proportional valves and n servo valves in parallel, the pump control subsystem is composed of h bidirectional variable pumps in parallel; the three-phase motor is connected with the quantitative pump, the oil inlet of the quantitative pump is connected with the oil tank, the three-phase motor rotates to drive the quantitative pump to work, the hydraulic oil output by the quantitative pump is distributed to each proportional valve and servo valve through a pipeline after passing through a check valve and a filter; the three-phase motor is connected with the pump control subsystem, the three-phase motor rotates to drive the bidirectional variable pump to work, the bidirectional variable pump sucks oil from the oil tank and outputs the hydraulic oil to a main oil circuit, the hydraulic oil passes through a check valve and a filter and is connected with the oil circuit of the valve control subsystem in parallel; the hydraulic cylinder is connected with the proportional valve and the servo valve and is also connected with the main circuit of the bidirectional variable pump, the piston rod of the hydraulic cylinder is connected with a displacement sensor, and the displacement sensor is connected with a controller; the signal input ends of the proportional valve, the servo valve and the bidirectional variable pump are respectively connected with the output end of the controller.

2. A pump-valve hybrid electro-hydraulic servo system and a configuration method thereof, characterized by, The configuration method of the pump-valve hybrid electro-hydraulic servo system comprises the following steps: (1) constructing proportional valve product type set, servo valve product type set and bidirectional variable pump product type set according to product type, and setting performance index vectors of the bidirectional variable pump, the proportional valve and the servo valve; (2) under the structure of the pump-valve hybrid system, constructing an electro-hydraulic servo system dynamics model according to pump-valve subsystem flow superposition principle, combining the flow continuity and mechanical balance relationship of the asymmetric hydraulic cylinder; (3) taking device cost and system throttling loss energy consumption as objective functions, taking system flow constraint, system thrust constraint and system absolute error integral value constraint under multiple working conditions as constraint conditions, and constructing a multi-objective optimization model of the electro-hydraulic servo system; (4) solving the optimal pump-valve combination scheme by using a particle swarm optimization algorithm.

3. The pump-valve hybrid electro-hydraulic servo system and its configuration method according to claim 2, characterized in that, The performance indexes of the bidirectional variable pump include rated flow, rated pressure, cost, displacement gradient, variable mechanism adjustment angle and internal and external leakage coefficients, the performance indexes of the proportional valve include flow gain, flow pressure gain, rated pressure, rated flow and cost, and the performance indexes of the servo valve include flow gain, flow pressure gain, rated pressure, rated flow and cost.

4. The pump-valve hybrid electro-hydraulic servo system and its configuration method according to claim 2, characterized in that, In step (2), the output flow rate Q of the valve control subsystem v is wherein, Ql,i,t represents the output flow rate of the proportional valve corresponding to the first, second, and mth positions with indexes l, i, and t, respectively, Vbl,i,t represents the type of the proportional valve corresponding to the first, second, and mth positions with indexes l, i, and t, respectively l Vbl,i,t i Vbl,i,t t K represents the flow gain and flow-pressure gain coefficients in the performance index vector bs1 K bs2 K bsm ul,i,t represents the amplifier gain of the proportional valve corresponding to the first, second, and mth positions b1 ul,i,t b2 ul,i,t bm ul,i,t represents the input electrical signal of the proportional valve corresponding to the first, second, and mth positions Qz,a,d represents the output flow rate of the servo valve corresponding to the first, second, and nth positions with indexes z, a, and d, respectively Vsz,a,d represents the type of the servo valve corresponding to the first, second, and nth positions with indexes z, a, and d, respectively z Vsz,a,d a Vsz,a,d d K represents the flow gain and flow-pressure gain coefficients in the performance index vector ss1 K ss2 K ssn uz,a,d represents the amplifier gain of the servo valve corresponding to the first, second, and nth positions s1 uz,a,d s2 uz,a,d sn uz,a,d represents the input electrical signal of the servo valve corresponding to the first, second, and nth positions L P represents the load pressure Output flow rate Q of the pump control subsystem p To wherein, represents the variable pump output flow corresponding to the 1st, 2nd, hth position index o, q, y of the hydraulic system; represents the variable pump model P corresponding to the 1st, 2nd, hth position index o, q, y of the hydraulic system, respectively o ,P q ,P y the displacement gradient, the mechanism adjustment swing angle and the leakage coefficient in the performance index vector; K sp1 ,K sp2 ,K sph represents the variable pump amplifier gain coefficient of the 1st, 2nd, hth position; u p1 ,u p2 ,u ph represents the variable pump input electric signal of the 1st, 2nd, hth position; The output flow rate Q of the electro-hydraulic servo system L To Q L = Q v + Q p The output flow equation of the combined electro-hydraulic servo system, the flow continuity equation of the asymmetric hydraulic cylinder and the force balance equation of the asymmetric hydraulic cylinder are as follows: Wherein, A is the effective area of hydraulic cylinder piston; V t The sum of the volume of pipeline and valve cavity; β e The effective volume modulus of hydraulic oil, n is the ratio of effective area of hydraulic cylinder rod cavity and rodless cavity, m t The total mass of piston and load; x p The ideal displacement of hydraulic cylinder piston rod; B p The viscous damping coefficient of piston and load; K p The elastic stiffness coefficient; F L The external load force acting on the piston rod; The third derivative of the ideal displacement of hydraulic cylinder piston rod, The second derivative of the ideal displacement of hydraulic cylinder piston rod, The first derivative of the ideal displacement of hydraulic cylinder piston rod, d is the system disturbance, M, B, C, D respectively represent the equivalent mass, damping coefficient, elastic coefficient, gain coupling coefficient of the system; K1 is the input gain of the first proportional valve, K m The input gain of the mth proportional valve, K n The input gain of the nth servo valve, K h The input gain of the hth variable pump, K CE The total gain coefficient of the control system; Take x1 = x p , x1, x2 and x3 represent displacement, velocity and acceleration in hydraulic cylinder piston respectively; the state equation of electro-hydraulic servo system is as follows:

5. The hybrid electro-hydraulic servo system of claim 4, wherein, In step (3), the objective function J is J = ω1J1 + ω2J2 Wherein, ω1, ω2 are optimization weights, ω1 + ω2 = 1, J1 is the total cost of the device, and J2 is the system throttling loss energy consumption; The total cost of the device J1 is wherein, Vb is the proportional valve model corresponding to the lth position index l cost in the performance index vector, Vb is the proportional valve model corresponding to the mth position index t cost in the performance index vector, Vs is the servo valve model corresponding to the nth position index d cost in the performance index vector, P is the variable pump model corresponding to the hth position index y cost in the performance index vector; The system throttling loss energy consumption J2 is E Vb = q Vb · Δp Vb E Vs = q Vs · Δp Vs wherein, Vb is the proportional valve model corresponding to the 1st, mth position index l, t in the system l Vb t the throttling loss energy consumption generated, Vs is the servo valve model corresponding to the 1st, nth position index z, d in the system z Vs d the throttling loss energy consumption generated, Vb E is the throttling loss energy consumption of the proportional valve, Vs Δp is the throttling loss energy consumption of the servo valve; b Δp is the proportional valve port pressure difference, s q is the servo valve output flow rate, Vb q is the proportional valve output flow rate, Vs q is the servo valve output flow rate.

6. The pump-valve hybrid electro-hydraulic servo system and its configuration method according to claim 5, characterized in that, In step (3), the proportional valve port differential pressure Δp b is Servo valve port differential pressure Δp s To where K qs is the servo valve flow gain coefficient, K sc is the servo valve flow-pressure gain coefficient, u s is the servo valve input electrical signal.

7. The hybrid electro-hydraulic servo system of claim 6, wherein, In step (3), the system flow constraint is wherein Q is a rated flow required for system operation, are the variable pump models P corresponding to the 1st, hth position indices o, h respectively o ,P y rated flow in the performance index vector, are the proportional valve models Vb corresponding to the 1st, mth position indices l, t respectively l ,Vb t rated flow in the performance index vector, are the servo valve models Vs corresponding to the 1st, nth position indices z, d respectively z ,Vs d rated flow in the performance index vector.

8. The pump-valve hybrid electro-hydraulic servo system and its configuration method according to claim 7, characterized in that, In step (3), the system thrust constraint is where A is the piston effective area, P L is the load pressure, P rated is the system rated pressure, and λ is the safety margin coefficient. denotes the variable pump model P o rated pressure in the performance index vector, denotes the proportional valve model Vb l rated pressure in the performance index vector, denotes the servo valve model Vs z rated pressure in the performance index vector.

9. The pump-valve hybrid electro-hydraulic servo system and its configuration method according to claim 8, characterized in that, In step (3), the system absolute error integral value constraint under multiple working conditions is Where, IAE i is the absolute error integral value of the system under the current selected working condition, i = 1, 2, 3, δ i (t) is a working condition indicator function, m(t) ∈ {1, 2, 3} is the current selected working condition, x d is the actual tracking displacement of the hydraulic cylinder piston rod.

10. The pump-valve hybrid electro-hydraulic servo system and its configuration method according to claim 9, characterized in that, A penalty term is introduced into the objective function J to obtain the fitness function J penalized As follows: J penalized = J + Penalty Wherein, α1, α2, α3 are penalty coefficients.

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