Bench load simulation compensation control method for vehicle electric drive system in anechoic chamber environment

Abstract

The present invention relates to a load simulation compensation control method for a vehicle electric drive system bench in an anechoic chamber environment, comprising: equating a rigid-flexible coupling transmission shaft of a vehicle electric drive system bench to a multi-degree-of-freedom inertia-damping-rigidity concentrated mass model; establishing its dynamic model according to the inertia-damping-rigidity concentrated mass model; establishing an electric drive system bench dynamic model that takes into account the characteristics of the rigid-flexible coupling transmission shaft according to the established rigid-flexible coupling transmission shaft dynamic model, combined with a target vehicle longitudinal dynamic model and a load simulation device model; and performing load simulation compensation control on the vehicle electric drive system bench based on the electric drive system bench dynamic model. Compared with the prior art, this method can effectively improve the dynamic response and loading accuracy of load simulation of an electric drive system bench with a rigid-flexible coupling transmission shaft in an anechoic chamber environment, and greatly improve the accuracy and effectiveness of the test evaluation results of the vehicle electric drive system.

Description

Technical Field

[0001] The present invention relates to the technical field of vehicle electric drive system performance testing, and in particular to a vehicle electric drive system bench load simulation compensation control method in an anechoic chamber environment. Background Art

[0002] The vehicle electric drive system test bench based on the anechoic chamber is widely used in the field of powertrain NVH performance testing. The load simulation device (dynamometer) is often placed outside the anechoic chamber to reduce the impact of its noise on the performance test results of the DUT, and a rigid-flexible coupling transmission shaft is usually used to connect the DUT and the load simulation device. However, the parameter uncertainty and unmodeled error of the rigid-flexible coupling transmission shaft will seriously affect the robustness of the load simulation control system; at the same time, the difference between the mechanical inertia of the rigid-flexible coupling transmission shaft and the vehicle inertia will also cause a large static and dynamic error between the actual loading torque of the DUT and the target value, making it difficult to accurately achieve the consistency of the load characteristic test of the electric drive system and the whole vehicle. Therefore, for this type of electric drive test bench system, how to make the DUT achieve the same loading torque as the actual vehicle on the electric drive system test bench is the key issue of load simulation compensation control.

[0003] In order to solve the problem of poor stability of speed closed-loop tracking control of load simulation system in anechoic chamber environment, speed feedforward compensator is usually used to improve its performance. However, traditional speed feedforward compensator is usually affected by the perturbation of transmission shaft parameters, which makes the anti-interference ability of speed closed-loop tracking performance worse. At the same time, the speed feedforward compensator cannot well achieve the loading torque accuracy of load simulation device. In the traditional compensation control method, how to reduce the dynamic error of loading torque and improve its load simulation accuracy is not considered.

[0004] Chinese patent CN114221832A discloses a vehicle bench dynamic load simulation device and control method, the device includes a simulation calculation device, a vehicle composite brake system, a dynamic load simulation device, a CAN communication network and a coupling; the vehicle composite brake system and the dynamic load simulation device are mechanically connected through a coupling; the simulation calculation device, the vehicle composite brake system and the dynamic load simulation device exchange data through the CAN communication network; wherein the simulation calculation device calculates the reference torque of the vehicle composite brake system and the reference speed of the load motor of the dynamic load simulation device based on the driver's operating instructions and the running vehicle dynamics model; the vehicle composite brake system brakes the vehicle based on the reference torque; the dynamic load simulation device dynamically loads the vehicle composite brake system based on the reference speed of the load motor. The load simulation method of the vehicle bench dynamic load simulation device adopts the first-order and third-order Taylor expansions for delay compensation, does not consider the influence of the measured system and the load simulation device on the design of the delay compensation link, and also cannot effectively solve the dynamic error of the loading torque caused by the uncertainty of the transmission system parameters. The applicability of the load simulation compensator designed based on this theory has certain limitations. Summary of the invention

[0005] The purpose of the present invention is to overcome the defects of the above-mentioned prior art and to provide a vehicle electric drive system bench load simulation compensation control method in an anechoic chamber environment with good stability, strong anti-interference ability and high precision.

[0006] The purpose of the present invention can be achieved by the following technical solutions:

[0007] A load simulation compensation control method for a vehicle electric drive system bench in an anechoic chamber environment, the load simulation compensation control method comprising:

[0008] Step 1: Equivalent the rigid-flexible coupled transmission shaft of the vehicle electric drive system test bench to a multi-degree-of-freedom inertia-damping-stiffness concentrated mass model;

[0009] Step 2: Establish its dynamic model based on the inertia-damping-stiffness concentrated mass model;

[0010] Step 3: Based on the established rigid-flexible coupling transmission shaft dynamics model, combined with the target vehicle longitudinal dynamics model and the load simulation device model, establish the electric drive system bench dynamics model considering the rigid-flexible coupling transmission shaft characteristics;

[0011] Step 4: Based on the electric drive system bench dynamics model, load simulation compensation control is performed on the vehicle electric drive system bench, specifically:

[0012] Step 4-1: Based on the dynamic model of the electric drive system bench, a speed closed-loop tracking control sub-method of the load simulation device is constructed;

[0013] Step 4-2: construct a speed feedforward correction compensator for the speed closed-loop tracking control of the load simulation device based on the speed closed-loop tracking control sub-method of the load simulation device;

[0014] Step 4-3: Based on the speed feedforward correction compensator of the load simulation device speed closed-loop tracking control, a load simulation device loading torque feedforward correction compensator considering the dynamic compensation gain is constructed;

[0015] Step 4-4: Use the feedforward correction compensator constructed in step 4-3 to control the vehicle electric drive system test bench. According to the control result of the load simulation device loading torque feedforward correction compensator, determine whether the load simulation device meets the stability of the speed closed-loop tracking and the loading torque accuracy requirements. If the preset conditions are met, the control result is directly output; if the conditions are not met, the dynamic compensation gain is adjusted and step 4-3 is repeated.

[0016] As a preferred technical solution, the vehicle electric drive system test bench in the anechoic chamber environment includes a dynamometer, a soundproof wall, a test piece, an electric drive system and a rigid-flexible coupled transmission shaft; the rigid-flexible coupled transmission shaft includes a vehicle half-axle, a diaphragm spring coupling, a left transmission long shaft and a right transmission long shaft.

[0017] As a preferred technical solution, the method for constructing the inertia-damping-stiffness concentrated mass model in step 1 is:

[0018] According to the different masses of the rigid-flexible coupling transmission shaft assembly, it is equivalent to a concentrated mass body connected by several rigid elements and flexible elements; the mass of the rigid-flexible coupling transmission shaft assembly is determined by the actual structural characteristics of the electric drive test bench, and different types of electric drive test benches have different numbers of equivalent concentrated mass degrees of freedom.

[0019] As a preferred technical solution, the inertia-damping-stiffness concentrated mass model is specifically:

[0020]

[0021] Among them, J is the moment of inertia parameter; c is the damping parameter; k is the stiffness parameter of the flexible element; ω is the angular velocity of the rigid element; T e is the torque at the output end of the vehicle electric drive system; T l is the loading torque of the load simulation device; n is the number of degrees of freedom of the concentrated mass model;

[0022] The moment of inertia parameter J and the flexible element stiffness parameter k are calculated by finite element analysis; the damping parameter c is obtained by parameter identification method; the angular velocity ω of each rigid element is obtained by the angular velocity of the load simulation device using the Taylor series expansion of the n-order inertia link, satisfying the following relationship:

[0023]

[0024] Among them, τ ω is the angular velocity lag time of the rigid element; s is the Laplace operator; ω t is the actual angular velocity of the load simulation device.

[0025] As a preferred technical solution, the electric drive system bench dynamics model in step 3 includes a target vehicle longitudinal dynamics model, a load simulation device model, and a rigid-flexible coupled transmission shaft multi-degree-of-freedom inertia-damping-rigidity concentrated mass model.

[0026] As a preferred technical solution, the target vehicle longitudinal dynamics model is specifically:

[0027]

[0028] Among them, F t is the vehicle driving force; m is the vehicle mass; f is the rolling friction coefficient; g is the acceleration of gravity; θ is the road slope; ρ is the air density; C D is the air resistance coefficient; v veh is the vehicle speed; A is the vehicle frontal area;

[0029] The load simulation device adopts a three-phase asynchronous motor, and its load simulation device model is expressed as:

[0030] T c =-N p 1.5m sr 2 ·L r ·(I rd I sq -I rq I sd )

[0031] Among them, T c The torque is calculated by closed-loop speed tracking of the load simulation device; N p is the pole pair number; m sr is the rotor flux; L r is the rotor inductance; I rd ,I rq are the motor rotor dq axis currents respectively; I sd ,I sq They are the motor rotor dq axis currents respectively;

[0032] The vehicle driving force is calculated by using a proportional-differential controller to track the target vehicle speed through the actual angular velocity signal of the load simulation device. The controller is specifically:

[0033]

[0034] in, is the proportional link coefficient; is the differential link coefficient; K T is the analytical proportional coefficient of the required torque;

[0035] The load simulation device adopts a speed closed-loop control method, with the wheel end target angular velocity and the actual angular velocity of the load simulation device as input, and the load simulation device controller is designed as follows:

[0036]

[0037] In the formula, is the coefficient of the integration link; is the proportional link coefficient; is the differential link coefficient.

[0038] As a preferred technical solution, the step 4-1 is specifically as follows:

[0039] Step 4-1-1: According to the constructed vehicle driving force proportional-differential controller, the actual angular velocity of the load simulation device is fed back to calculate the output torque of the vehicle electric drive system. The transfer function relationship is expressed as:

[0040] T e (s) = ω t (s)G d (s)

[0041] Step 4-1-2: Based on the calculated output torque of the vehicle electric drive system, use it as the input variable of the target vehicle longitudinal dynamics model to obtain the wheel end target angular velocity ω em , the transfer function is expressed as:

[0042] ω em (s)=[T e (s)-T d (s)]G em (s)

[0043] Among them, G em (s) is the open-loop transfer function of the vehicle model; T d is the sliding resistance torque of the target vehicle;

[0044] Step 4-1-3: Use the calculated output torque of the vehicle electric drive system as the input variable of the rigid-flexible coupling transmission shaft model to obtain the actual angular velocity ω of the load simulation device t , the transfer function is expressed as:

[0045] ω t(s)=[T e (s)-T l (s)]G t (s)

[0046] Step 4-1-4: Take the target angular velocity ω at the wheel end em is the expected angular velocity, and the actual angular velocity ω of the load simulation device t As the control variable, the angular velocity error is calculated and used as the input of the load simulation controller to calculate the loading torque of the load simulation device. The expression is:

[0047] T l (s) = [ω t (s)-ω em (s)]G(s)G c (s)

[0048] Combining the above expressions, the closed-loop tracking control sub-method of the electric drive system bench load simulation speed is obtained, which can be expressed as:

[0049]

[0050] Among them, G ω (s) is the speed closed-loop tracking transfer function of the load simulation device; ω em is the target angular velocity of the wheel end; G(s) is the electromagnetic torque calculation transfer function of the load simulation device; G c (s) is the load simulation device controller; G d (s) is the required torque analysis link; G t (s) is the transfer function of the rigid-flexible coupled transmission shaft.

[0051] As a preferred technical solution, the step 4-2 is specifically as follows:

[0052] Step 4-2-1: According to the electric drive system bench load simulation speed closed-loop tracking control sub-method, construct the load simulation compensation control relationship:

[0053]

[0054] Step 4-2-2: Superimpose the wheel end target angular velocity ω in the load simulation speed closed-loop tracking control loop em The input vehicle pseudo-dynamics model 1 / G em (s), and the speed feedforward correction compensator is obtained:

[0055]

[0056] Among them, G comp (s) is the speed feedforward correction compensator.

[0057] As a preferred technical solution, the step 4-3 is specifically as follows:

[0058] Step 4-3-1: According to the target vehicle wheel end angular velocity ω em (s) and introduce 1 / G v (s) The loading torque is observed to obtain the transmission torque of the transmission shaft system and perform a first-order low-pass filter on it, that is, the transmission torque of the rigid-flexible coupling transmission shaft is obtained:

[0059]

[0060] Among them, τ d is the lag time; T el (s) is the rigid-flexible coupling transmission torque; G v (s) is the transfer function of the load simulation device;

[0061] Step 4-3-2: Obtain the approximate torque T at the output of the load simulation device through a torque sensor or observer n-1 , and obtain the reference compensation amount of loading torque:

[0062] △T el (s) = T e (s)-T n-1 (s)

[0063] Step 4-3-3: Observe the system state variables of the rigid-flexible coupling transmission shaft and express them as:

[0064]

[0065]

[0066] Among them, x is the system state parameter; u is the system control quantity to be determined; d is the external interference input; A n is the first coefficient matrix; B n is the second coefficient matrix; C n is the third coefficient matrix; ω1 is the equivalent angular velocity of the first concentrated mass; T1 is the equivalent transmitted torque of the first elastic element; ω2 is the equivalent angular velocity of the second concentrated mass; T2 is the equivalent transmitted torque of the second elastic element; ω n-1 is the equivalent angular velocity of the n-1th concentrated mass; ω n is the equivalent angular velocity of the load simulation device;

[0067] Step 4-3-4: In view of the dynamic error of the loading torque of the load simulation device caused by the difference in the moment of inertia of the rigid-flexible coupling transmission shaft, an adjustment gain is introduced to compensate for it, namely:

[0068]

[0069] Among them, Δ is the actual adjustment compensation amount; ΔG(s) is the compensation adjustment gain; K f is the reference compensation gain; ΔK f To dynamically adjust the gain;

[0070] Step 4-3-5: Obtain the loading torque T according to the above load simulation compensation relationship l With input torque T e and T d The relationship between the load simulation device and the actual loading torque after load simulation compensation is:

[0071]

[0072] As a preferred technical solution, the actual loading torque of the load simulation device after the load simulation compensation is adjusted by dynamically adjusting the gain so as to meet the stability of the speed closed-loop tracking and the loading torque accuracy requirements.

[0073] Compared with the prior art, the present invention has the following beneficial effects:

[0074] 1. Good stability: The vehicle electric drive system bench load simulation compensation control method in the anechoic chamber environment of the present invention can eliminate the problem of poor stability caused by the inverse dynamics model of the traditional speed closed-loop tracking system by considering the required torque analysis link of the electric drive system under test in the load simulation speed closed-loop tracking control system, thereby effectively improving its stability.

[0075] 2. Strong anti-interference ability: The vehicle electric drive system test bench load simulation compensation control method in the anechoic chamber environment of the present invention can eliminate the influence of the rigid-flexible coupling transmission shaft model parameter disturbance by constructing a speed feedforward compensator that considers the required torque decision and the vehicle simulation object, thereby improving the anti-interference ability of the load simulation device.

[0076] 3. High precision: The vehicle electric drive system test bench load simulation compensation control method in the anechoic chamber environment of the present invention can reduce the influence of the inertia error between the actual vehicle and the test bench on the load simulation loading accuracy by constructing a torque feedforward correction compensator that takes into account the compensation adjustment gain. BRIEF DESCRIPTION OF THE DRAWINGS

[0077] Figure 1 1 is a flow chart of a vehicle electric drive system bench load simulation compensation control method suitable for an anechoic chamber environment in an embodiment of the present invention;

[0078] Figure 2 A schematic diagram of a test bench structure of a vehicle electric drive system based on an anechoic chamber environment in an embodiment of the present invention;

[0079] Figure 3 It is a 4-DOF rigid-flexible coupled concentrated mass model of the gantry transmission shaft in the embodiment of the present invention;

[0080] Figure 4 Schematic diagram of a vehicle electric drive system bench load simulation control system suitable for an anechoic chamber environment in an embodiment of the present invention;

[0081] Figure 5 This is a schematic diagram of the load simulation speed closed-loop tracking control principle of the electric drive system test bench in an embodiment of the present invention;

[0082] Figure 6 It is an algorithm structure diagram of a vehicle electric drive system bench load simulation compensation control method suitable for an anechoic chamber environment in an embodiment of the present invention. DETAILED DESCRIPTION

[0083] The following will be combined with the drawings in the embodiments of the present invention to clearly and completely describe the technical solutions in the embodiments of the present invention. Obviously, the described embodiments are part of the embodiments of the present invention, not all of the embodiments. Based on the embodiments of the present invention, all other embodiments obtained by ordinary technicians in this field without creative work should fall within the scope of protection of the present invention.

[0084] The term "one embodiment" or "embodiment" as used herein refers to a specific feature, structure or characteristic that may be included in at least one implementation of the present application. In the description of the present application, it should be understood that the orientation or positional relationship indicated by the terms "upper", "lower", "left", "right", "top", "bottom", etc. is based on the orientation or positional relationship shown in the accompanying drawings, which is only for the convenience of describing the present application and simplifying the description, rather than indicating or implying that the device or element referred to must have a specific orientation, be constructed and operated in a specific orientation, and therefore cannot be understood as a limitation of the present application. In addition, the terms "first" and "second" are used only for descriptive purposes, and cannot be understood as indicating or implying relative importance or implicitly indicating the number of technical features indicated. Thus, the features defined as "first" and "second" may include one or more of the features explicitly or implicitly. Moreover, the terms "first", "second", etc. are used to distinguish similar objects, and are not necessarily used to describe a specific order or sequence. It should be understood that the data used in this way can be interchangeable where appropriate, so that the embodiments of the present application described here can be implemented in an order other than those illustrated or described here.

[0085] Example 1

[0086] Figure 1This is a flow chart of a method for simulating load compensation control of a vehicle electric drive system in an anechoic chamber environment provided in an embodiment of the present application. This application provides method operation steps as described in the embodiment or flow chart, but based on routine or non-creative labor, more or fewer operation steps may be included. The order of steps listed in the embodiment is only one of many step execution orders and does not represent the only execution order. The method should be implementable by software and / or hardware. Please refer to Figure 1 , the method may include:

[0087] Step 101: Equivalently converting the rigid-flexible coupling transmission shaft of the vehicle electric drive system test bench into a multi-degree-of-freedom inertia-damping-stiffness concentrated mass model;

[0088] Step 102: Establishing its dynamic model according to the inertia-damping-stiffness concentrated mass model;

[0089] Step 103: Based on the established rigid-flexible coupling transmission shaft dynamics model, combined with the target vehicle longitudinal dynamics model and the load simulation device model, an electric drive system bench dynamics model considering the rigid-flexible coupling transmission shaft characteristics is established;

[0090] Step 104: Based on the electric drive system test bench dynamics model, load simulation compensation control is performed on the vehicle electric drive system test bench, specifically:

[0091] Step 104-1: constructing a speed closed-loop tracking control sub-method of a load simulation device according to the dynamics model of the electric drive system bench;

[0092] Step 104-2: constructing a speed feedforward correction compensator for a speed closed-loop tracking control of a load simulation device based on a speed closed-loop tracking control sub-method of a speed of the load simulation device;

[0093] Step 104-3: constructing a load simulation device loading torque feedforward correction compensator taking into account a dynamic compensation gain based on the speed feedforward correction compensator of the speed closed-loop tracking control of the load simulation device;

[0094] Step 104-4: Use the feedforward correction compensator constructed in step 104-3 to control the vehicle electric drive system test bench, and judge whether the load simulation device meets the stability of the speed closed-loop tracking and the loading torque accuracy requirements based on the control result of the load simulation device loading torque feedforward correction compensator. If the preset conditions are met, the control result is directly output; if the conditions are not met, the dynamic compensation gain is adjusted and step 104-3 is repeated.

[0095] Specifically, the vehicle electric drive system test bench in the anechoic chamber environment includes a dynamometer, a soundproof wall, a test piece, an electric drive system, and a rigid-flexible coupling transmission shaft. The rigid-flexible coupling transmission shaft includes a vehicle half shaft, a diaphragm spring coupling, a left transmission shaft, and a right transmission shaft.

[0096] Specifically, the method for constructing the inertia-damping-stiffness concentrated mass model in step 101 is:

[0097] According to the different masses of the rigid-flexible coupling transmission shaft assembly, it is equivalent to a concentrated mass body connected by several rigid elements and flexible elements; the mass of the rigid-flexible coupling transmission shaft assembly is determined by the actual structural characteristics of the electric drive test bench, and different types of electric drive test benches have different numbers of equivalent concentrated mass degrees of freedom.

[0098] Optionally, the inertia-damping-stiffness concentrated mass model is specified as:

[0099]

[0100] Among them, J is the moment of inertia parameter; c is the damping parameter; k is the stiffness parameter of the flexible element; ω is the angular velocity of the rigid element; T e is the torque at the output end of the vehicle electric drive system; T l is the loading torque of the load simulation device; n is the number of degrees of freedom of the concentrated mass model;

[0101] The moment of inertia parameter J and the flexible element stiffness parameter k are calculated by finite element analysis; the damping parameter c is obtained by parameter identification method; the angular velocity ω of each rigid element is obtained by the angular velocity of the load simulation device using the Taylor series expansion of the n-order inertia link, satisfying the following relationship:

[0102]

[0103] Among them, τ ω is the angular velocity lag time of the rigid element; s is the Laplace operator; ω t is the actual angular velocity of the load simulation device.

[0104] Specifically, the electric drive system bench dynamics model in step 103 includes a target vehicle longitudinal dynamics model, a load simulation device model, and a rigid-flexible coupled transmission shaft multi-degree-of-freedom inertia-damping-rigidity concentrated mass model.

[0105] Optionally, the target vehicle longitudinal dynamics model is specifically:

[0106]

[0107] Among them, F tis the vehicle driving force; m is the vehicle mass; f is the rolling friction coefficient; g is the acceleration of gravity; θ is the road slope; ρ is the air density; C D is the air resistance coefficient; v veh is the vehicle speed; A is the vehicle frontal area;

[0108] The load simulation device uses a three-phase asynchronous motor, and its load simulation device model is expressed as:

[0109] T c =-N p 1.5m sr 2 ·L r ·(I rd I sq -I rq I sd )

[0110] Among them, T c The torque is calculated by closed-loop speed tracking of the load simulation device; N p is the pole pair number; m sr is the rotor flux; L r is the rotor inductance; I rd ,I rq are the motor rotor dq axis currents respectively; I sd ,I sq They are the motor rotor dq axis currents respectively;

[0111] The vehicle driving force is calculated by using the proportional-differential controller to track the target driving speed through the actual angular velocity signal of the load simulation device. The specific controller is:

[0112]

[0113] in, is the proportional link coefficient; is the differential link coefficient; K T is the analytical proportional coefficient of the required torque;

[0114] The load simulation device adopts a speed closed-loop control method, with the target angular velocity of the wheel end and the actual angular velocity of the load simulation device as input. The load simulation device controller is designed as follows:

[0115]

[0116] In the formula, is the coefficient of the integration link; is the proportional link coefficient; is the differential link coefficient.

[0117] Specifically, step 104-1 is as follows:

[0118] Step 104-1-1: According to the constructed vehicle driving force proportional-differential controller, the actual angular velocity of the load simulation device is fed back to calculate the output torque of the vehicle electric drive system. The transfer function relationship is expressed as:

[0119] T e (s) = ω t (s)G d (s)

[0120] Step 104-1-2: Based on the calculated output torque of the vehicle electric drive system, use it as the input variable of the target vehicle longitudinal dynamics model to obtain the wheel end target angular velocity ω em , the transfer function is expressed as:

[0121] ω em (s)=[T e (s)-T d (s)]G em (s)

[0122] Among them, G em (s) is the open-loop transfer function of the vehicle model; T d is the sliding resistance torque of the target vehicle;

[0123] Step 104-1-3: Use the calculated output torque of the vehicle electric drive system as the input variable of the rigid-flexible coupling transmission shaft model to obtain the actual angular velocity ω of the load simulation device t , the transfer function is expressed as:

[0124] ω t (s)=[T e (s)-T l (s)]G t (s)

[0125] Step 104-1-4: Take the wheel end target angular velocity ω em is the expected angular velocity, and the actual angular velocity ω of the load simulation device t As the control variable, the angular velocity error is calculated and used as the input of the load simulation controller to calculate the loading torque of the load simulation device. The expression is:

[0126] T l (s) = [ω t (s)-ω em (s)]G(s)G c (s)

[0127] Combining the above expressions, the closed-loop tracking control sub-method of the electric drive system bench load simulation speed is obtained, which can be expressed as:

[0128]

[0129] Among them, G ω (s) is the speed closed-loop tracking transfer function of the load simulation device; ω em is the target angular velocity of the wheel end; G(s) is the electromagnetic torque calculation transfer function of the load simulation device; G c (s) is the load simulation device controller; G d (s) is the required torque analysis link; G t (s) is the transfer function of the rigid-flexible coupled transmission shaft.

[0130] Specifically, step 104-2 is as follows:

[0131] Step 104-2-1: According to the electric drive system bench load simulation speed closed-loop tracking control sub-method, construct a load simulation compensation control relationship:

[0132]

[0133] Step 104-2-2: Superimpose the wheel end target angular velocity ω in the load simulation speed closed-loop tracking control loop em The input vehicle pseudo-dynamics model 1 / G em (s), and the speed feedforward correction compensator is obtained:

[0134]

[0135] Among them, G comp (s) is the speed feedforward correction compensator.

[0136] Specifically, step 104-3 is as follows:

[0137] Step 104-3-1: Based on the target vehicle wheel end angular velocity ω em (s) and introduce 1 / G v (s) The loading torque is observed to obtain the transmission torque of the transmission shaft system and perform a first-order low-pass filter on it, that is, the transmission torque of the rigid-flexible coupling transmission shaft is obtained:

[0138]

[0139] Among them, τ d is the lag time; T el (s) is the rigid-flexible coupling transmission torque; G v (s) is the transfer function of the load simulation device;

[0140] Step 104-3-2: Obtain the approximate torque T at the output end of the load simulation device through a torque sensor or observer n-1, and obtain the reference compensation amount of loading torque:

[0141] △T el (s) = T e (s)-T n-1 (s)

[0142] Step 104-3-3: Observe the system state variables of the rigid-flexible coupling transmission shaft and express them as:

[0143]

[0144]

[0145] Among them, x is the system state parameter; u is the system control quantity to be determined; d is the external interference input; A n is the first coefficient matrix; B n is the second coefficient matrix; C n is the third coefficient matrix; ω1 is the equivalent angular velocity of the first concentrated mass; T1 is the equivalent transmitted torque of the first elastic element; ω2 is the equivalent angular velocity of the second concentrated mass; T2 is the equivalent transmitted torque of the second elastic element; ω n-1 is the equivalent angular velocity of the n-1th concentrated mass; ω n is the equivalent angular velocity of the load simulation device;

[0146] Step 104-3-4: In view of the dynamic error of the loading torque of the load simulation device caused by the difference in the moment of inertia of the rigid-flexible coupling transmission shaft, an adjustment gain is introduced to compensate for it, namely:

[0147]

[0148] Among them, Δ is the actual adjustment compensation amount; ΔG(s) is the compensation adjustment gain; K f is the reference compensation gain; ΔK f To dynamically adjust the gain;

[0149] Step 104-3-5: Obtain the loading torque T according to the above load simulation compensation relationship l With input torque T e and T d The relationship between the load simulation device and the actual loading torque after load simulation compensation is:

[0150]

[0151] Optionally, the actual loading torque of the load simulation device after load simulation compensation is adjusted by dynamically adjusting the gain so as to meet the stability of the speed closed-loop tracking and the loading torque accuracy requirements.

[0152] Example 2

[0153] This embodiment provides a more specific vehicle electric drive system bench load simulation compensation control method in an anechoic chamber environment, including:

[0154] Step 1: Based on the mechanical structure characteristics of the rigid-flexible coupling transmission shaft between the electric drive system in the anechoic chamber and the dynamometer outside the anechoic chamber, the vehicle electric drive system test bench based on the anechoic chamber environment is as follows: Figure 2 As shown, it includes left / right dynamometers, soundproof walls, a test piece, an electric drive system and a rigid-flexible coupled transmission shaft; the rigid-flexible coupled transmission shaft includes a vehicle half-axle, a diaphragm spring coupling and a left transmission long shaft and a right transmission long shaft; the left transmission long shaft and the right transmission long shaft are respectively connected to the left dynamometer and the right dynamometer through the soundproof wall. Based on the above system structure, the centralized mass method is used to equate the rigid-flexible coupled transmission shaft to a multi-degree-of-freedom "inertia-damping-stiffness" centralized mass model. The "inertia-damping-stiffness" centralized mass model can be equivalent to a centralized mass body connected by multiple rigid elements and flexible elements according to the different masses of the rigid-flexible coupled transmission shaft assembly; the mass of the rigid-flexible coupled transmission shaft assembly is determined by the actual structural characteristics of the electric drive test bench, and different types of electric drive test benches have different numbers of equivalent centralized mass degrees of freedom. In this embodiment, Figure 3 As shown in the figure, the rigid-flexible coupled transmission shaft is equivalent to a 4-DOF “inertia-damping-stiffness” concentrated mass model.

[0155] Step 2: In order to express the dynamic performance of the established 4-DOF "inertia-damping-stiffness" concentrated mass model as much as possible, the established 4-DOF "inertia-damping-stiffness" dynamic model is expressed as:

[0156]

[0157] In the formula, J a_eq is the first concentrated mass inertia; ω a_eq is the equivalent angular velocity of the first concentrated mass; c a_eq The first concentrated mass damping; J L_eq is the second concentrated mass inertia; ω L_eq is the equivalent angular velocity of the second concentrated mass; c L_eq is the second concentrated mass damping; J b_eq is the third concentrated mass inertia; ω b_eq is the equivalent angular velocity of the third concentrated mass; c b_eq is the third concentrated mass damping; J v_eq is the mass inertia of the fourth set; ω v_eq is the equivalent angular velocity of the fourth concentrated mass; c v_eq is the fourth concentrated mass damping; T e is the torque at the output end of the vehicle electric drive system; T lApply torque to the load simulator.

[0158] The moment of inertia parameter J and the flexible body stiffness parameter k in the above formula can be calculated by establishing a three-dimensional model of the rigid-flexible coupling transmission shaft and using the finite element analysis method; at the same time, the damping parameter c is obtained by parameter identification method through the bench modal analysis test. The angular velocity ω of each rigid element is obtained by the angular velocity of the load simulation device using the Taylor series expansion of the n-order inertia link, satisfying the following relationship:

[0159]

[0160] Among them, τ ω is the angular velocity lag time of the rigid element; s is the Laplace operator.

[0161] Step 3: Based on the established rigid-flexible coupling transmission shaft dynamics model, combined with the target vehicle longitudinal dynamics model and the load simulation device model, an electric drive system bench model considering the rigid-flexible coupling transmission shaft dynamics characteristics is established. The load simulation control system is as follows: Figure 4 shown.

[0162] The target vehicle longitudinal dynamics model is considered in the bench load simulation control method of the electric drive system and is expressed as:

[0163]

[0164] Among them, F t is the vehicle driving force; m is the vehicle mass; f is the rolling friction coefficient; g is the acceleration of gravity; θ is the road slope; ρ is the air density; C D is the air resistance coefficient; v veh is the vehicle speed; A is the vehicle’s frontal area.

[0165] like Figure 4 As shown, by taking the target vehicle longitudinal dynamics model into account in the load simulation system, we can obtain:

[0166]

[0167] Among them, J em is the equivalent inertia of the whole vehicle; c em is the equivalent damping of the whole vehicle.

[0168] At the same time, the dynamic model of the rigid-flexible coupling transmission shaft is established:

[0169] T e -T l = a_eq (J a_eq s+c a_eq )+ω L_eq (JL_eq s+c L_eq )+ω b_eq (J b_eq s+c b_eq )+ω v_eq (J v_eq s+c v_eq )

[0170] If the load simulation device uses a three-phase asynchronous motor, according to the relationship between the motor electromagnetic torque and current, the current I m is the cross product of the dq axis current of the stator and rotor windings, then the electromagnetic torque relationship of the speed closed-loop tracking feedback can be expressed as:

[0171]

[0172] Among them, T c The torque is calculated by closed-loop speed tracking of the load simulation device; N p is the pole pair number; m sr is the rotor flux; L r is the rotor inductance; I rd ,I rq are the motor rotor dq axis currents respectively; I sd ,I sq They are the motor rotor dq axis current respectively; K s It is the speed closed-loop tracking proportional coefficient.

[0173] The bench dynamics model of the electric drive system includes the target vehicle longitudinal dynamics model, the load simulation device model, and the rigid-flexible coupled transmission shaft multi-degree-of-freedom inertia-damping-rigidity concentrated mass model; the vehicle driving force is calculated by using the proportional-differential controller to track the target driving speed through the actual angular velocity signal of the load simulation device. The controller is designed as follows:

[0174]

[0175] in, is the proportional link coefficient; is the differential link coefficient; K T is the analytical proportional coefficient of the required torque.

[0176] The load simulation device adopts a speed closed-loop control method, with the target angular velocity of the wheel end and the actual angular velocity of the load simulation device as input. The load simulation device controller is designed as follows:

[0177]

[0178] In the formula, is the coefficient of the integration link; is the proportional link coefficient; is the differential link coefficient.

[0179] Step 4: Based on the established electric drive system bench model, develop a vehicle electric drive system bench load simulation composite correction compensation control method, which includes the following three parts:

[0180] Step 4-1: In the established load simulation device model, design a load simulation speed closed-loop tracking control method that takes into account the required torque analysis link, such as Figure 5 shown.

[0181] Step 4-1-1: According to the constructed vehicle driving force proportional-differential controller, the actual angular velocity of the load simulation device is fed back to calculate the output torque of the vehicle electric drive system. The transfer function relationship is expressed as:

[0182] T e (s) = ω t (s)G d (s)

[0183] Step 4-1-2: Based on the calculated output torque of the vehicle electric drive system, use it as the input variable of the target vehicle longitudinal dynamics model to obtain the wheel end target angular velocity ω em , the transfer function is expressed as:

[0184] ω em (s)=[T e (s)-T d (s)]G em (s)

[0185] Among them, G em (s) is the open-loop transfer function of the vehicle model; T d is the sliding resistance torque of the target vehicle.

[0186] Step 4-1-3: Use the calculated output torque of the vehicle electric drive system as the input variable of the rigid-flexible coupling transmission shaft model to obtain the actual angular velocity ω of the load simulation device t , the transfer function is expressed as:

[0187] ω t (s)=[T e (s)-T l (s)]G t (s)

[0188] Step 4-1-4: Take the target angular velocity ω at the wheel end em is the expected angular velocity, and the actual angular velocity ω of the load simulation device t As the control variable, the angular velocity error is calculated and used as the input of the load simulation controller to calculate the loading torque of the load simulation device. The expression is:

[0189] T l (s) = [ω t (s)-ω em (s)]G(s)G c (s)

[0190] Combining the above relationship, the closed-loop tracking control method of the electric drive system bench load simulation speed can be obtained, which can be expressed as:

[0191]

[0192] in

[0193]

[0194] Combining the above relationships, the speed closed-loop tracking relationship of the load simulation system of the electric drive system test bench can be obtained as follows:

[0195]

[0196] in

[0197]

[0198] Step 4-2: In order to overcome the influence of parameter perturbation of the rigid-flexible coupled transmission shaft of the load simulation system on the speed closed-loop tracking performance, a speed feedforward correction compensator is introduced into the speed closed-loop tracking system to achieve the stability of the load simulation system and improve its anti-interference ability. The design steps of the speed feedforward correction compensator are as follows:

[0199] Step 4-2-1: According to the electric drive system bench load simulation speed closed-loop tracking control sub-method, construct the load simulation compensation control relationship:

[0200]

[0201] Step 4-2-2: Superimpose the wheel end target angular velocity ω in the load simulation speed closed-loop tracking control loop em The input vehicle pseudo-dynamics model 1 / G em (s), and the speed feedforward correction compensator is obtained:

[0202]

[0203] in

[0204]

[0205] In order to eliminate the influence of parameter perturbation of the rigid-flexible coupling transmission shaft on the system, the speed feedforward correction compensator can be designed as:

[0206]

[0207] Step 4-3: Based on the speed closed-loop tracking speed feedforward correction compensator, design a load simulation device loading torque feedforward correction compensator considering the dynamic compensation gain, such as Figure 6 The design process shown is as follows:

[0208] Step 4-3-1: According to the simulated vehicle object angular velocity ω em (s) and introduce 1 / G v (s) The loading torque is observed to obtain the transmission torque of the transmission shaft system and perform a first-order low-pass filter on it, that is, the transmission torque of the rigid-flexible coupling transmission system is obtained:

[0209]

[0210] Among them, τ d is the lag time; T el (s) is the rigid-flexible coupling transmission torque; G v (s) is the transfer function of the load simulation device.

[0211] Step 4-3-2: Obtain approximate torque T through torque sensor or observer s , and obtain the reference compensation amount of loading torque:

[0212] △T el (s) = T e (s)-T s (s)

[0213] Step 4-3-3: If the state observer is used to s (s) can be calculated and the system state variables of the rigid-flexible coupling transmission shaft can be observed, which can be expressed as:

[0214]

[0215] in

[0216]

[0217]

[0218] Among them, x is the system state parameter; u is the system control quantity to be determined; d is the external interference input; A n is the first coefficient matrix; B n is the second coefficient matrix; C n is the third coefficient matrix; T k is the equivalent transmission torque of the first elastic element; T L is the equivalent transmission torque of the second elastic element; T s k is the equivalent transmitted torque of the third elastic element; a_eqk is the elastic coefficient of the first elastic element; L_eq k is the elastic coefficient of the second elastic element; s_eq is the elastic coefficient of the third elastic element.

[0219] Step 4-3-4: For the dynamic error of loading torque caused by the difference in the moment of inertia of the rigid-flexible coupling transmission shaft, an adjustment gain can be introduced to compensate for it, thereby reducing the steady-state error between the loading torque and the target torque, that is:

[0220]

[0221] Among them, Δ is the actual adjustment compensation amount; ΔG(s) is the compensation adjustment gain; K f is the reference compensation gain; ΔK f To dynamically adjust the gain.

[0222] Step 4-3-5: According to the above load simulation relationship, the loading torque T l With input torque T e and T d The actual loading torque after load simulation compensation is:

[0223]

[0224] Step 4-4: According to the control result of the load simulation device loading torque composite correction compensator, determine whether the load simulation device meets the stability of the speed closed-loop tracking and the loading torque accuracy requirements. If the conditions are met, directly output the result; if the conditions are not met, adjust the dynamic compensation gain and repeat step 4-3.

[0225] The above is only a specific embodiment of the present invention, but the protection scope of the present invention is not limited thereto. Any technician familiar with the technical field can easily think of various equivalent modifications or replacements within the technical scope disclosed by the present invention, and these modifications or replacements should be included in the protection scope of the present invention. Therefore, the protection scope of the present invention shall be based on the protection scope of the claims.

Claims

1. A vehicle electric drive system bench load simulation compensation control method in an anechoic chamber environment, characterized in that: The load simulation compensation control method comprises: Step 1: Equivalent the rigid-flexible coupled transmission shaft of the vehicle electric drive system test bench to a multi-degree-of-freedom inertia-damping-stiffness concentrated mass model; Step 2: Establish its dynamic model based on the inertia-damping-stiffness concentrated mass model; Step 3: Based on the established rigid-flexible coupling transmission shaft dynamics model, combined with the target vehicle longitudinal dynamics model and the load simulation device model, establish the electric drive system bench dynamics model considering the rigid-flexible coupling transmission shaft characteristics; Step 4: Based on the electric drive system bench dynamics model, load simulation compensation control is performed on the vehicle electric drive system bench, specifically: Step 4-1: Based on the dynamic model of the electric drive system bench, a speed closed-loop tracking control sub-method of the load simulation device is constructed; Step 4-2: construct a speed feedforward correction compensator for the speed closed-loop tracking control of the load simulation device based on the speed closed-loop tracking control sub-method of the load simulation device; Step 4-3: Based on the speed feedforward correction compensator of the load simulation device speed closed-loop tracking control, a load simulation device loading torque feedforward correction compensator considering the dynamic compensation gain is constructed; Step 4-4: Use the feedforward correction compensator constructed in step 4-3 to control the vehicle electric drive system test bench. According to the control result of the load simulation device loading torque feedforward correction compensator, determine whether the load simulation device meets the stability of the speed closed-loop tracking and the loading torque accuracy requirements. If the preset conditions are met, the control result is directly output; if the conditions are not met, the dynamic compensation gain is adjusted and step 4-3 is repeated.

2. The method for simulating load compensation and control of a vehicle electric drive system in an anechoic chamber environment according to claim 1, characterized in that: The vehicle electric drive system test bench in the anechoic chamber environment includes a dynamometer, a soundproof wall, a test piece, an electric drive system and a rigid-flexible coupling transmission shaft; the rigid-flexible coupling transmission shaft includes a vehicle half shaft, a diaphragm spring coupling, a left transmission long shaft and a right transmission long shaft.

3. The method for simulating load compensation and control of a vehicle electric drive system in an anechoic chamber environment according to claim 1, characterized in that: The method for constructing the inertia-damping-stiffness concentrated mass model in step 1 is: According to the different masses of the rigid-flexible coupling transmission shaft assembly, it is equivalent to a concentrated mass body connected by several rigid elements and flexible elements; the mass of the rigid-flexible coupling transmission shaft assembly is determined by the actual structural characteristics of the electric drive test bench, and different types of electric drive test benches have different numbers of equivalent concentrated mass degrees of freedom.

4. The method for simulating load compensation of a vehicle electric drive system in an anechoic chamber environment according to claim 1, characterized in that: The inertia-damping-stiffness concentrated mass model is specifically: Among them, J is the moment of inertia parameter; c is the damping parameter; k is the stiffness parameter of the flexible element; ω is the angular velocity of the rigid element; T e is the torque at the output end of the vehicle electric drive system; T l is the loading torque of the load simulation device; n is the number of degrees of freedom of the concentrated mass model; The moment of inertia parameter J and the flexible element stiffness parameter k are calculated by finite element analysis; the damping parameter c is obtained by parameter identification method; the angular velocity ω of each rigid element is obtained by the angular velocity of the load simulation device using the Taylor series expansion of the n-order inertia link, satisfying the following relationship: Among them, τ ω is the angular velocity lag time of the rigid element; s is the Laplace operator; ω t is the actual angular velocity of the load simulation device.

5. The method for simulating load compensation and controlling a vehicle electric drive system in an anechoic chamber environment according to claim 1, characterized in that: The electric drive system bench dynamics model in step 3 includes a target vehicle longitudinal dynamics model, a load simulation device model, and a rigid-flexible coupled transmission shaft multi-degree-of-freedom inertia-damping-stiffness concentrated mass model.

6. The method for simulating load compensation and controlling a vehicle electric drive system in an anechoic chamber environment according to claim 5, characterized in that: The target vehicle longitudinal dynamics model is specifically: Among them, F t is the vehicle driving force; m is the vehicle mass; f is the rolling friction coefficient; g is the acceleration of gravity; θ is the road slope; ρ is the air density; C D is the air resistance coefficient; v veh is the vehicle speed; A is the vehicle frontal area; The load simulation device adopts a three-phase asynchronous motor, and its load simulation device model is expressed as: T c =-N p ·1.5m sr 2 ·L r ·(I rd ·I sq -I rq ·I sd ) Among them, T c The torque is calculated by closed-loop speed tracking of the load simulation device; N p is the pole pair number; m sr is the rotor flux; L r is the rotor inductance; I rd ,I rq are the motor rotor dq axis currents respectively; I sd ,I sq They are the motor rotor dq axis currents respectively; The vehicle driving force is calculated by using a proportional-differential controller to track the target vehicle speed through the actual angular velocity signal of the load simulation device. The controller is specifically: in, is the proportional link coefficient; is the differential link coefficient; K T is the analytical proportional coefficient of the required torque; The load simulation device adopts a speed closed-loop control method, with the wheel end target angular velocity and the actual angular velocity of the load simulation device as inputs, and the load simulation device controller is designed as follows: In the formula, is the coefficient of the integration link; is the proportional link coefficient; is the differential link coefficient.

7. The method for simulating load compensation and controlling a vehicle electric drive system in an anechoic chamber environment according to claim 1, characterized in that: The step 4-1 is specifically as follows: Step 4-1-1: According to the constructed vehicle driving force proportional-differential controller, the actual angular velocity of the load simulation device is fed back to calculate the output torque of the vehicle electric drive system. The transfer function relationship is expressed as: T e (s)=ω t (s)G d (s) Step 4-1-2: Based on the calculated output torque of the vehicle electric drive system, use it as the input variable of the target vehicle longitudinal dynamics model to obtain the wheel end target angular velocity ω em , the transfer function is expressed as: ω em (s)=[T e (s)-T d (s)]G em (s) Among them, G em (s) is the open-loop transfer function of the vehicle model; T d is the sliding resistance torque of the target vehicle; Step 4-1-3: Use the calculated output torque of the vehicle electric drive system as the input variable of the rigid-flexible coupling transmission shaft model to obtain the actual angular velocity ω of the load simulation device t , the transfer function is expressed as: ω t (s)=[T e (s)-T l (s)]G t (s) Step 4-1-4: Take the target angular velocity ω at the wheel end em is the expected angular velocity, and the actual angular velocity ω of the load simulation device t As the control variable, the angular velocity error is calculated and used as the input of the load simulation controller to calculate the loading torque of the load simulation device. The expression is: T l (s)=[ω t (s)-ω em (s)]G(s)G c (s) Combining the above expressions, the closed-loop tracking control sub-method of the electric drive system bench load simulation speed is obtained, which can be expressed as: Among them, G ω (s) is the speed closed-loop tracking transfer function of the load simulation device; ω em is the target angular velocity of the wheel end; G(s) is the electromagnetic torque calculation transfer function of the load simulation device; G c (s) is the load simulation device controller; G d (s) is the required torque analysis link; G t (s) is the transfer function of the rigid-flexible coupled transmission shaft.

8. The method for simulating load compensation and controlling a vehicle electric drive system in an anechoic chamber environment according to claim 1, characterized in that: The step 4-2 is specifically as follows: Step 4-2-1: According to the electric drive system bench load simulation speed closed-loop tracking control sub-method, construct the load simulation compensation control relationship: Step 4-2-2: Superimpose the wheel end target angular velocity ω in the load simulation speed closed-loop tracking control loop em The input vehicle pseudo-dynamics model 1 / G em (s), and the speed feedforward correction compensator is obtained: Among them, G comp (s) is the speed feedforward correction compensator.

9. The method for simulating load compensation and controlling a vehicle electric drive system in an anechoic chamber environment according to claim 1, characterized in that: The steps 4-3 are specifically as follows: Step 4-3-1: According to the target vehicle wheel end angular velocity ω em (s) and introduce 1 / G v (s) The loading torque is observed to obtain the transmission torque of the transmission shaft system and perform a first-order low-pass filter on it, that is, the transmission torque of the rigid-flexible coupling transmission shaft is obtained: Among them, τ d is the lag time; T el (s) is the rigid-flexible coupling transmission torque; G v (s) is the transfer function of the load simulation device; Step 4-3-2: Obtain the approximate torque T at the output of the load simulation device through a torque sensor or observer n-1 , and obtain the reference compensation amount of loading torque: ΔT el (s)=T e (s)-T n-1 (s) Step 4-3-3: Observe the system state variables of the rigid-flexible coupling transmission shaft and express them as: Among them, x is the system state parameter; u is the system control quantity to be determined; d is the external interference input; A n is the first coefficient matrix; B n is the second coefficient matrix; C n is the third coefficient matrix; ω1 is the equivalent angular velocity of the first concentrated mass; T1 is the equivalent transmitted torque of the first elastic element; ω2 is the equivalent angular velocity of the second concentrated mass; T2 is the equivalent transmitted torque of the second elastic element; ω n-1 is the equivalent angular velocity of the n-1th concentrated mass; ω n is the equivalent angular velocity of the load simulation device; Step 4-3-4: In view of the dynamic error of the loading torque of the load simulation device caused by the difference in the moment of inertia of the rigid-flexible coupling transmission shaft, an adjustment gain is introduced to compensate for it, namely: Among them, Δ is the actual adjustment compensation amount; ΔG(s) is the compensation adjustment gain; K f is the reference compensation gain; ΔK f To dynamically adjust the gain; Step 4-3-5: Obtain the loading torque T according to the above load simulation compensation relationship l With input torque T e and T d The relationship between the load simulation device and the actual loading torque after load simulation compensation is:

10. The method for simulating load compensation and controlling a vehicle electric drive system in an anechoic chamber environment according to claim 9, characterized in that: The actual loading torque of the load simulation device after the load simulation compensation is adjusted by dynamically adjusting the gain so as to meet the stability of the speed closed-loop tracking and the loading torque accuracy requirements.

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