Equivalent stiffness and damping analysis method for vehicle-mounted motor rotor assembly

Abstract

The invention relates to the field of new energy automobile electric driving systems, in particular to an equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly. According to the method, accurate analysis and evaluation of the equivalent stiffness and damping of the rotor assembly can be realized through modal identification and frequency response analysis technologies without depending on whole vehicle and bench type test resources, a quantitative basis is provided for structural optimization and process improvement of the motor rotor assembly, the development period can be effectively shortened, and the cost can be reduced.

Description

Technical Field

[0001] The present invention relates to the field of electric drive systems for new energy vehicles, and in particular to an equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly. Background Art

[0002] As a power output source, the onboard motor rotor assembly converts electrical energy into mechanical energy, thereby providing propulsion for the vehicle. It is one of the most technologically advanced and critical components in the powertrain of new energy vehicles. The performance of the onboard motor rotor assembly not only directly impacts the vehicle's power output and range, but its stiffness and damping properties also have a crucial impact on the vibration and noise characteristics of new energy vehicles. Good stiffness and damping properties can help reduce vibration and noise, enhancing vehicle comfort and stability. Conversely, poor stiffness and damping can lead to abnormal vibration and noise during operation, impacting the user experience and overall vehicle performance.

[0003] Since there are significant differences in the on-board motor rotor assemblies of different new energy vehicle products (that is, the on-board motor rotor assemblies have diversity in structure, process and size), usually when designing the on-board motor rotor assemblies of different new energy vehicle products, it is necessary to determine the reliability of these design parameters through vehicle verification after the design parameters of the on-board motor rotor assembly are determined. In other words, the on-board motor rotor assembly manufactured using these design parameters needs to be installed on the vehicle for testing to determine its ability to suppress vibration noise. However, this method is easily interfered by various factors, resulting in extremely poor consistency in vibration and noise evaluation, long R&D cycle and high cost. The specific reasons are as follows:

[0004] ① Poor consistency in vibration and noise assessment: Since the whole vehicle test process is affected by various factors, such as the overall structure of the vehicle, the influence of other components, and different test environments, this will lead to poor consistency in the vibration and noise assessment in each test, making it difficult to accurately reflect the actual effect brought about by the adjustment of the rotor assembly design parameters.

[0005] ② Long R&D cycle: The whole vehicle test requires the rotor assembly to be installed in the complete vehicle system for testing, and this process involves multiple links such as vehicle assembly and debugging. During the vehicle assembly process, in addition to accurately installing the on-board motor rotor assembly into the motor system and ensuring that its relative position accuracy with the motor stator and other components meets strict design requirements to avoid affecting the motor performance due to assembly errors, the entire vehicle's mechanical structure must also be constructed. If the vehicle exhibits abnormal vibration, noise, and other problems during the test, it needs to be disassembled and reworked, which undoubtedly significantly increases the assembly time.

[0006] After assembly, debugging is equally complex and time-consuming. The motor's control strategy must be fine-tuned for the power system in which the vehicle's motor rotor assembly is located. This process requires the use of specialized testing equipment, repeated testing and parameter optimization under a variety of simulated operating conditions, and a comprehensive performance evaluation after each adjustment to ensure the stability and efficiency of the motor's coordinated operation with other vehicle systems. Any communication failure or signal interference can cause abnormal motor operation or even vehicle failure, and troubleshooting and resolving such issues often requires significant time and effort, extending the product development cycle.

[0007] ③ High cost: The whole vehicle test requires a lot of manpower, material and financial resources, including the purchase or modification cost of the vehicle, the use fee of the test site, the investment in test equipment, etc., which greatly increases the cost of verification.

[0008] In summary, the traditional vehicle verification method cannot meet the cycle requirements of rapid product development and production. Summary of the Invention

[0009] The purpose of the present invention is to address the corresponding deficiencies in the existing technology and provide an equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly. This method uses modal identification and frequency response analysis technology, without relying on complete vehicle and bench test resources, to accurately analyze and evaluate the equivalent stiffness and damping of the rotor assembly, providing a quantitative basis for structural optimization and process improvement of the motor rotor assembly, effectively shortening the development cycle and reducing costs.

[0010] The purpose of the present invention is to adopt the following scheme to achieve:

[0011] A method for analyzing equivalent stiffness and damping of a vehicle-mounted motor rotor assembly comprises the following steps:

[0012] 1) establishing a three-dimensional coordinate system and placing the rotor assembly in the three-dimensional coordinate system, selecting a plurality of modal points on the rotor assembly for identifying bending / torsion modes of the rotor assembly, and determining the coordinates of all modal points in the three-dimensional coordinate system;

[0013] 2) identifying the bending mode of the rotor assembly based on the vector displacement changes of the modal point determined in step 1) under all modes, and obtaining the bending modal frequencies of each order of the rotor assembly;

[0014] 3) identifying the torsional mode of the rotor assembly based on the vector displacement changes of the modal point determined in step 1) under all modes, and obtaining the torsional modal frequencies of each order of the rotor assembly;

[0015] 4) Obtain the equivalent bending stiffness of the rotor assembly based on the bending modal frequencies of each order of the rotor assembly;

[0016] 5) Obtain the equivalent torsional stiffness of the rotor assembly based on the torsional modal frequencies of each order of the rotor assembly;

[0017] 6) Calculate the equivalent axial stiffness of the rotor assembly;

[0018] 7) Calculate the equivalent damping coefficient of the rotor assembly;

[0019] 8) Based on the equivalent bending stiffness, equivalent torsional stiffness, equivalent axial stiffness, and equivalent damping coefficient of the rotor assembly, calculate the equivalent stiffness and damping coefficient of the rotor assembly, and evaluate and score the equivalent stiffness and damping coefficient of the rotor assembly.

[0020] Preferably, in step 1), the specific steps of establishing a three-dimensional coordinate system, selecting multiple modal points, and determining the coordinates of all modal points in the three-dimensional coordinate system are as follows:

[0021] 1-1) Construct a cylindrical coordinate system based on the rotor assembly: With the rotor core axis as the reference, the radial direction of the rotor core is set as the X-axis, the tangent direction of the rotor core outer edge is the Y-axis, and the axial direction of the rotating shaft is the Z-axis;

[0022] 1-2) between the first axial end face A and the second axial end face E of the rotor core, select end face B, end face C, and end face D, such that the first axial end face A, the second axial end face E, end face B, end face C, and end face D are equally spaced and parallel to each other;

[0023] 1-3) On the first shaft end surface A, the second shaft end surface E, the end surface B, the end surface C, and the end surface D, n modal points are arranged at angular intervals with the rotor core axis as the center;

[0024] 1-4) Record the spatial cylindrical coordinates of all modal points N in step 1-3) and number them in sequence.

[0025] Preferably, the rotor assembly bending mode includes a first-order bending mode of the rotor assembly and a second-order bending mode of the rotor assembly. In step 2), the specific steps of obtaining the frequencies of each order bending mode of the rotor assembly according to the identified first-order bending mode of the rotor assembly and the second-order bending mode of the rotor assembly are as follows:

[0026] 2-1) Perform modal simulation on the rotor assembly and record the vector displacement changes of N modal points in all modes;

[0027] 2-2) Based on the vector displacement changes of each modal point under all modes, and in order of frequency from low to high, the first-order bending mode of the rotor assembly and the second-order bending mode of the rotor assembly are identified as follows:

[0028] 2-2-1) If the following conditions are met simultaneously, the identified mode is the first-order bending mode of the rotor assembly:

[0029] ① The displacement change directions of all modal points on the first shaft end surface A and the second shaft end surface E are consistent, and the error of the displacement change is within 5%;

[0030] ② The displacement change direction of all modal points on end surfaces B and D is consistent, and the error of the displacement change is within 5%;

[0031] ③ The displacement changes of n modal points on the end surface C are all less than 2% of the displacement change of any modal point on the end surface A of the first axis;

[0032] ④ The direction of the modal point displacement change at the center of the two end surfaces of the rotor shaft of the rotor assembly is consistent with the direction of the average displacement change of n modal points on the first shaft end surface A, and the modal point displacement change at the center of the two end surfaces of the rotor shaft of the rotor assembly is greater than 40% of the average displacement change of n modal points on the first shaft end surface A;

[0033] 2-2-2) If the following conditions are met simultaneously, the identified mode is the second-order bending mode of the rotor assembly:

[0034] ① The displacement change directions of all modal points on the first shaft end surface A and the second shaft end surface E are opposite, and the error of the displacement change is within 5%;

[0035] ② The displacement changes of all modal points on end surfaces B and D are in opposite directions, and the error of the displacement change is within 5%;

[0036] ③ The displacement changes of n modal points on end surface C are all less than 2% of the displacement change of any modal point on end surface B;

[0037] ④ The displacement changes of the modal points at the centers of the two end surfaces of the rotor shaft of the rotor assembly are in opposite directions, and the displacement changes of the modal points at the centers of the two end surfaces of the rotor shaft of the rotor assembly are both greater than 40% of the average displacement change of the n modal points on the end surface B;

[0038] 2-3) Based on the identified first-order bending mode of the rotor assembly and the second-order bending mode of the rotor assembly, the corresponding first-order bending modal frequency and second-order bending modal frequency are obtained.

[0039] Preferably, the rotor assembly torsional mode includes a first-order torsional mode of the rotor assembly and a second-order torsional mode of the rotor assembly. In step 3), the specific steps of obtaining the torsional modal frequencies of each order of the rotor assembly according to the identified first-order torsional mode of the rotor assembly and the second-order torsional mode of the rotor assembly are as follows:

[0040] 3-1) Perform modal simulation on the rotor assembly and record the vector displacement changes of all modal points in all modes;

[0041] 3-2) Based on the vector displacement changes of each modal point under all modes, and in order of frequency from low to high, the specific method for identifying the first-order torsional mode of the rotor assembly and the second-order torsional mode of the rotor assembly is as follows:

[0042] 3-2-1) If the following conditions are met at the same time, the identified mode is the first-order torsional mode of the rotor assembly:

[0043] ① The Y-axis displacement change directions of the n modal points on the first shaft end face A are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0044] ② The Y-axis displacement change directions of the n modal points on the end surface B are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0045] ③ The Y-axis displacement change directions of the n modal points on the end surface D are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0046] ④ The Y-axis displacement change directions of the n modal points on the second shaft end surface E are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0047] ⑤ The Y-axis displacement changes of n modal points on the end surface C are all less than 2% of the Y-axis displacement change of any modal point on the end surface A of the first axis;

[0048] ⑥ The Y-axis displacement change direction of the modal point on the first shaft end surface A is opposite to the Y-axis displacement change direction of the modal point on the second shaft end surface E, and the error of the Y-axis displacement change is within 5%;

[0049] ⑦ The Y-axis displacement change direction of the modal point on end surface B is opposite to the Y-axis displacement change direction of the modal point on end surface D, and the error of the Y-axis displacement change is within 5%;

[0050] ⑧ The displacement changes of the modal points at the centers of both end faces of the rotor shaft of the rotor assembly are less than 1% of the average displacement change of n modal points on the first shaft end face A;

[0051] 3-2-2) If the following conditions are met simultaneously, the identified mode is the second-order torsional mode of the rotor assembly:

[0052] ① The Y-axis displacement change directions of the n modal points on the first shaft end face A are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0053] ② The Y-axis displacement change directions of the n modal points on the end surface B are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0054] ③ The Y-axis displacement change directions of the n modal points on the end surface D are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0055] ④ The Y-axis displacement change directions of the n modal points on the second shaft end surface E are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0056] ⑤ The Y-axis displacement change directions of the n modal points on the end surface C are consistent, and the Y-axis displacement changes of the n modal points are all greater than 40% of the average Y-axis displacement change of the n modal points on the end surface A of the first axis;

[0057] ⑥ The Y-axis displacement change direction of the modal point on the end surface C is consistent with the Y-axis displacement change direction of the modal point on the end surface A of the first axis;

[0058] ⑦ The Y-axis displacement change direction of the modal point on the first shaft end surface A is consistent with the Y-axis displacement change direction of the modal point on the second shaft end surface E, and the error of the Y-axis displacement change is within 5%;

[0059] ⑧The Y-axis displacement change direction of the modal point on end surface B is consistent with the Y-axis displacement change direction of the modal point on end surface D, and the error of the Y-axis displacement change is within 5%;

[0060] ⑨ The displacement changes of the modal points at the centers of both end faces of the rotor shaft of the rotor assembly are less than 2% of the average displacement changes of the n modal points on the first shaft end face A;

[0061] ⑩ The displacement changes of all modal points on end surfaces B and D are less than 10% of the average displacement changes of n modal points on the end surface A of the first axis;

[0062] 3-3) Based on the identified first-order torsional mode of the rotor assembly and the second-order torsional mode of the rotor assembly, the corresponding first-order torsional modal frequency and second-order torsional modal frequency are obtained.

[0063] Preferably, in step 4), the formula for calculating the equivalent bending stiffness of the rotor assembly is as follows:

[0064]

[0065] Where K eq_w is the equivalent bending stiffness of the rotor assembly; F w1 is the first-order bending mode frequency; F w2 is the second-order bending modal frequency; m is the mass of the rotor assembly.

[0066] Preferably, in step 5), the formula for calculating the equivalent torsional stiffness of the rotor assembly is as follows:

[0067]

[0068] Where K eq_n is the equivalent torsional stiffness of the rotor assembly; F n1 is the first-order torsional mode frequency; F n2 is the second-order torsional modal frequency; m is the mass of the rotor assembly.

[0069] Preferably, in step 6), the specific steps of calculating the equivalent axial stiffness of the rotor assembly are as follows:

[0070] 6-1) Perform Z-axis frequency response analysis on n modal points on the first axis end surface A as follows:

[0071] 6-1-1) Apply an external force F to n modal points on the end face A of the first shaft, and perform a frequency sweep excitation with an amplitude of 1N at intervals of 1 Hz to obtain the vibration acceleration response of the corresponding modal point at each frequency;

[0072] 6-1-2) Obtain the frequency response functions of n modal points on the end surface A of the first shaft based on the vibration acceleration response signal;

[0073] 6-1-3) Take the average value of the frequency response functions of n modal points on the end face A of the first shaft to obtain the average value of the frequency response function;

[0074] 6-2) Calculate the equivalent axial stiffness of the rotor assembly based on the average value of the frequency response function and the following formula:

[0075]

[0076] Where K eq_z is the equivalent axial stiffness of the rotor assembly; H1 represents the average value of the frequency response function of n points on the first shaft end surface A when the frequency is equal to 1 Hz; H2 represents the average value of the frequency response function of n points on the first shaft end surface A when the frequency is equal to 2 Hz; H 4999 It represents the average value of the frequency response function of n points on the end face A of the first axis when the frequency is equal to 4999 Hz; H 5000 It represents the average value of the frequency response function of n points on the end surface A of the first shaft when the frequency is equal to 5000 Hz; Maxf represents the frequency at which the maximum value occurs.

[0077] Preferably, in step 7), the specific steps of calculating the equivalent damping coefficient of the rotor assembly are as follows:

[0078] 7-1) Record n ​​modal points on the end surface C of the motor rotor assembly as n excitation points, and perform frequency sweep excitation in the X-axis direction with an amplitude of 1N at an interval of 1 Hz, with a frequency change rate of 2s / 1 Hz; the frequency range is 10 Hz to 5000 Hz;

[0079] 7-2) Obtain the vibration acceleration response of each excitation point to N modal points in the motor rotor assembly under all frequency excitations, and record them as N response points;

[0080] 7-3) Obtain the frequency response function according to the following formula:

[0081] HH i_j (ω)=a i_j (ω) / F j (ω)

[0082] Where, HH i_j (ω) represents the frequency response function from the j-th excitation point to the i-th response point, j = 1, 2, 3, ..., n, i = 1, 2, 3, ..., N; a i_j (ω) represents the vibration acceleration from the jth excitation point to the ith response point; F j (ω) represents the excitation force of the jth excitation point; ω represents the frequency;

[0083] 7-4) Based on the frequency response function, the overall average response function is obtained using the following formula:

[0084]

[0085] Where Havg(ω) represents the overall average response function; M is the total number of frequency response functions; n is the number of excitation points; N is the number of response points; ω represents the frequency;

[0086] 7-5) Based on the overall average response function, the equivalent damping coefficient of the rotor assembly is obtained using the following formula:

[0087]

[0088] Where C eq is the equivalent damping coefficient of the rotor assembly; Havg(1) represents the overall average response function value at a frequency of 1 Hz; Havg(21) represents the overall average response function value at a frequency of 2 Hz; Havg(5000) represents the overall average response function value at a frequency of 5000 Hz; m is the mass of the rotor assembly.

[0089] Preferably, in step 8), according to the equivalent bending stiffness of the rotor assembly, the equivalent torsional stiffness of the rotor assembly, the equivalent axial stiffness of the rotor assembly, and the equivalent damping coefficient of the rotor assembly, the formula for obtaining the equivalent stiffness and damping coefficient of the rotor assembly is as follows:

[0090]

[0091] Where, δ eq is the equivalent stiffness and damping coefficient of the rotor assembly; K eq_wis the equivalent bending stiffness of the rotor assembly, K eq_n is the equivalent torsional stiffness of the rotor assembly, K eq_z is the equivalent axial stiffness of the rotor assembly, C eq is the equivalent damping coefficient of the rotor assembly.

[0092] Preferably, in step 8), a scoring table is used to evaluate and score the equivalent stiffness and damping coefficient of the rotor assembly.

[0093] The beneficial effects of the present invention are:

[0094] ① The present invention uses a series of analysis methods, covering the bending mode identification of the rotor assembly, the torsional mode identification of the rotor assembly, and the equivalent bending stiffness of the rotor assembly, the equivalent torsional stiffness of the rotor assembly, the equivalent axial stiffness of the rotor assembly, the equivalent damping coefficient of the rotor assembly, and the accurate calculation of the equivalent stiffness and damping coefficient of the rotor assembly, and draws conclusions by analyzing the calculation results. It can effectively provide a quantitative basis for the structural optimization and process improvement of the motor rotor assembly, and reduce costs while shortening the development cycle.

[0095] By dividing the rotor core into equally spaced axial sections and arranging the modal points on each end face at equal angular intervals, this method ensures that modal confidence is collected without blind spots along the length of the rotor core, ensuring symmetric detection of the rotor core's circumferential vibration response. By combining axial stratification with circumferential uniformity, the vibration mode characteristics corresponding to different modes can be clearly distinguished.

[0096] By setting specific conditions for identifying the bending and torsional modes of the rotor assembly, the present invention can clearly distinguish between bending and torsional characteristics of different orders. This multi-condition combined judgment mechanism efficiently and accurately filters out non-target modes. Furthermore, by ensuring the displacement consistency of each end face, local vibration interference is effectively eliminated, effectively avoiding misjudgments due to local structural differences and significantly improving the accuracy of modal identification.

[0097] ④ The present invention constructs a composite index by integrating the first-order bending modal frequency and the second-order bending rotation modal frequency, as well as the first-order torsional mode of the rotor assembly and the second-order torsional mode of the rotor assembly. This composite index can comprehensively and accurately quantify the rotor assembly's ability to resist deformation under different orders of vibration, as well as the rotor assembly's ability to resist torsional deformation under different orders of vibration. This result can provide key guidance and precise basis for the structural optimization of the motor rotor assembly and the optimization of the motor's dynamic performance, while providing a standardized evaluation basis for the comparison of multiple schemes, significantly reducing reliance on bench testing, and greatly shortening the development cycle while improving the consistency of the product's vibration and noise performance.

[0098] ⑤ This invention quantitatively analyzes the rotor assembly's deformation resistance under axial load through Z-axis frequency response analysis. This analysis provides key parameters for axial vibration control and structural optimization of automotive motors. Furthermore, by employing an efficient single-end excitation testing method, testing costs can be significantly reduced, providing a standardized evaluation method for shortening development cycles and improving the axial dynamic performance of automotive motors.

[0099] ⑥ This invention achieves precise quantification of multidimensional vibration characteristics through full-space response acquisition. Furthermore, a minimum selection mechanism is employed to accurately identify modal damping, establishing a comprehensive damping evaluation system for the rotor assembly. This provides core parameters for vibration and noise control, effectively achieving precise rotor assembly design optimization and improving the efficiency of testing costs, thereby reducing R&D costs while ensuring product performance.

[0100] ⑦ This invention constructs a multi-dimensional performance evaluation system by integrating the equivalent bending stiffness, equivalent torsional stiffness, equivalent axial stiffness, and equivalent damping coefficient of the rotor assembly. By converting complex vibration characteristics into intuitive quantitative scores, this evaluation system can effectively improve decision-making efficiency, provide precise guidance for the optimization process, and significantly shorten the product development cycle.

[0101] ⑧The present invention provides a universal method for evaluating the equivalent stiffness and damping performance of a rotor assembly, which can achieve rapid and effective evaluation regardless of the process, material, structure, etc. adopted by the rotor assembly.

[0102] Glossary:

[0103] Torsional modes: Torsional modes are the torsional deformation of the rotor about its axis. The order reflects the number of torsional vibration waves (nodes). First-order torsional vibration occurs when the entire shaft twists in the same direction (no nodes), with the two ends twisting in opposite directions (like twisting a towel). Second-order torsional vibration occurs when a torsional vibration node appears in a shaft segment, resulting in two sections twisting in opposite directions (e.g., twisting in the same direction at both ends and in the opposite direction in the middle). BRIEF DESCRIPTION OF THE DRAWINGS

[0104] Figure 1 It is a schematic diagram of the structure and end face of the rotor assembly of the present invention;

[0105] Figure 2 Schematic diagram of the modal point distribution of the rotor assembly according to an embodiment of the present invention;

[0106] Figure 3 Schematic diagram of the first-order bending mode shape of the rotor assembly according to an embodiment of the present invention;

[0107] Figure 4 Schematic diagram of the second-order bending mode shape of the rotor assembly according to an embodiment of the present invention;

[0108] Figure 5Schematic diagram of the first-order torsional mode of the rotor assembly according to an embodiment of the present invention;

[0109] Figure 6 Schematic diagram of the second-order torsional mode shape of the rotor assembly according to an embodiment of the present invention;

[0110] Figure 7 Schematic diagram of a frequency response function average value curve according to an embodiment of the present invention;

[0111] Figure 8 Schematic diagram of the average response function according to an embodiment of the present invention;

[0112] Figure 9 It is a flow chart of the present invention. DETAILED DESCRIPTION

[0113] like Figures 1 to 9 As shown, an equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly includes the following steps:

[0114] 1) Establishing a three-dimensional coordinate system and placing the rotor in the three-dimensional coordinate system, selecting multiple modal points on the rotor for identifying bending / torsion modes of the rotor assembly, and determining the coordinates of all modal points in the three-dimensional coordinate system. The specific steps are as follows;

[0115] 1-1) Construct a cylindrical coordinate system based on the rotor assembly: With the rotor core axis as the reference, the radial direction of the rotor core is set as the X-axis, the tangent direction of the rotor core outer edge is the Y-axis, and the axial direction of the rotating shaft is the Z-axis;

[0116] 1-2) between the first axial end face A and the second axial end face E of the rotor core, select end face B, end face C, and end face D, such that the first axial end face A, the second axial end face E, end face B, end face C, and end face D are equally spaced and parallel to each other;

[0117] 1-3) On the first shaft end face A, the second shaft end face E, the end face B, the end face C, and the end face D, n modal points are arranged at angular intervals with the rotor core axis as the center of the circle. In this embodiment, since the rotor has an axisymmetric structure, its bending and torsional modes have circular symmetry. Therefore, arranging four simulation points at 90° intervals can evenly cover the circumferential direction and can fully capture the bending mode and torsional mode of the rotor assembly.

[0118] 1-4) Record the spatial cylindrical coordinates of all modal points N obtained in step 1-3). In this embodiment, the total number of modal points is N = 5 × n + 2, where n is the number of simulation points in the same plane and 5 is the number of planes. Thus, the total number of modal points N is 22. At the same time, number all modal points in the order 1, 2, ..., 22.

[0119] 2) identifying the bending mode of the rotor assembly based on the vector displacement changes of the modal point determined in step 1) under all modes, and obtaining the bending modal frequencies of each order of the rotor assembly;

[0120] The bending modes of the rotor assembly include the first-order bending modes of the rotor assembly and the second-order bending modes of the rotor assembly. The specific steps for obtaining the frequencies of each order bending mode of the rotor assembly based on the identified first-order bending modes and second-order bending modes of the rotor assembly are as follows:

[0121] 2-1) Perform modal simulation on the rotor assembly and record the vector displacement changes of the corresponding 22 modal points under all simulated modes, including the direction and amount of displacement change;

[0122] 2-2) Based on the vector displacement changes of each modal point under all modes, and in order of frequency from low to high, the first-order bending mode of the rotor assembly and the second-order bending mode of the rotor assembly are identified as follows:

[0123] 2-2-1) If the following conditions are met simultaneously, the identified mode is the first-order bending mode of the rotor assembly:

[0124] ① The displacement change directions of all modal points on the first shaft end surface A and the second shaft end surface E are consistent, and the error of the displacement change is within 5%;

[0125] ② The displacement change direction of all modal points on end surfaces B and D is consistent, and the error of the displacement change is within 5%;

[0126] Conditions ① to ② are actually the judgment of the consistency of the end face displacement direction. The vibration mode of the first-order bending mode is a symmetrical sine curve, which is specifically manifested in that the displacement directions of the two end faces (i.e., the first axis end face A and the second axis end face E) and the middle section (i.e., the end face B and the end face D) are consistent. Therefore, the advantage of setting conditions ① and ② is that it can eliminate local vibration interference and ensure that the identified mode is the overall bending, rather than the mode caused by local deformation. At the same time, by strictly limiting the error of the displacement change, it can be ensured that each modal point is in coordinated motion, thereby verifying the structural symmetry and effectively eliminating the antisymmetric second-order bending mode.

[0127] ③ The displacement changes of n modal points on the end surface C are all less than 2% of the displacement change of any modal point on the end surface A of the first axis;

[0128] Condition ③ is actually a judgment on mid-plane displacement suppression. End surface C is located at the axial midpoint of the rotor core. Theoretically, the displacement of this section should approach zero. However, in actual operating conditions, due to the influence of many complex factors, it is difficult to achieve the ideal state of absolute zero displacement. It can only ensure that the displacement change of the modal point on end surface C is significantly smaller than the displacement change of the modal points on the two end surfaces (i.e., the first axial end surface A and the second axial end surface E). Therefore, in actual simulations, by setting this condition, the correct position of end surface C can be verified, thereby effectively eliminating high-order modes.

[0129] ④ The direction of the modal point displacement change at the center of the two end surfaces of the rotor shaft of the rotor assembly is consistent with the direction of the average displacement change of n modal points on the first shaft end surface A, and the modal point displacement change at the center of the two end surfaces of the rotor shaft of the rotor assembly is greater than 40% of the average displacement change of n modal points on the first shaft end surface A;

[0130] Condition ④ is actually a verification of shaft end displacement. The rotor shaft's two end points are constrained by bearings, and their displacement direction should align with surface A (the first-order mode is overall bending). The displacement must be large enough to account for the influence of support stiffness. This setting has the advantage of ensuring that shaft end displacement is not noise interference by setting a 40% displacement ratio.

[0131] 2-2-2) If the following conditions are met simultaneously, the identified mode is the second-order bending mode of the rotor assembly:

[0132] ① The displacement change directions of all modal points on the first shaft end surface A and the second shaft end surface E are opposite, and the error of the displacement change is within 5%;

[0133] ② The displacement changes of all modal points on end surfaces B and D are in opposite directions, and the error of the displacement change is within 5%;

[0134] Conditions ① and ② actually determine the antisymmetry of the end surface displacement directions. The second-order bending mode vibration shape is an antisymmetric sinusoidal curve, that is, the displacement directions of the two end surfaces (i.e., the first axial end surface A and the second axial end surface E) and the middle section (i.e., end surfaces B and end surfaces D) are opposite. Conditions ① and ② limit the antisymmetric deformation and can effectively exclude the symmetrical first-order bending mode.

[0135] ③ The displacement changes of n modal points on end surface C are all less than 2% of the displacement change of any modal point on end surface B;

[0136] Condition ③ actually determines the displacement suppression of the midplane. The nodes of the second-order bending mode are located at the 25% and 75% axial positions (i.e., end faces B and D). Theoretically, the displacement of end face C (50%) should be the largest. However, in actual simulations, the displacement of face C must be significantly smaller than that of face B to verify the correctness of the node positions. Therefore, by setting condition ③, higher-order modes (such as the different node positions of third-order bending) can be effectively eliminated.

[0137] ④ The displacement changes of the modal points at the centers of the two end surfaces of the rotor shaft of the rotor assembly are in opposite directions, and the displacement changes of the modal points at the centers of the two end surfaces of the rotor shaft of the rotor assembly are both greater than 40% of the average displacement change of the n modal points on the end surface B;

[0138] Condition ④ is actually a reverse determination of shaft end displacement. In the second-order bending mode, the rotor shaft's two end points are constrained by bearings, and their displacement directions should be opposite (e.g., the left end is upward, the right end is downward), and the displacement must be large enough to reflect the influence of support stiffness. Furthermore, setting a 40% displacement ratio ensures that the shaft end displacement is not affected by noise, thereby truly reflecting the modal characteristics.

[0139] 2-3) Based on the identified first-order bending mode of the rotor assembly and the second-order bending mode of the rotor assembly, the corresponding first-order bending modal frequency and second-order bending modal frequency are obtained.

[0140] According to the conditions of step 2-2-1), the first-order bending mode vibration shape of the rotor assembly is identified as follows Figure 3 As shown in FIG. 1 , under the first-order bending mode of this rotor assembly, the corresponding first-order bending mode frequency is 2458 Hz; through the conditions of the above step 2-2-2), the vibration mode of the second-order bending mode of the rotor assembly is identified as follows Figure 4 As shown, under the second-order bending mode of this rotor assembly, the corresponding second-order bending mode frequency is 3634 Hz.

[0141] 3) identifying the torsional mode of the rotor assembly based on the vector displacement changes of the modal point determined in step 1) under all modes, and obtaining the torsional modal frequencies of each order of the rotor assembly;

[0142] The torsional modes of the rotor assembly include the first-order torsional mode of the rotor assembly and the second-order torsional mode of the rotor assembly. The specific steps for obtaining the torsional modal frequencies of each order of the rotor assembly based on the identified first-order torsional mode of the rotor assembly and the second-order torsional mode of the rotor assembly are as follows:

[0143] 3-1) Perform modal simulation on the rotor assembly and record the vector displacement changes of 22 modal points in all modes, including the direction and amount of displacement change;

[0144] 3-2) Based on the vector displacement changes of each modal point under all modes, and in order of frequency from low to high, the specific method for identifying the first-order torsional mode of the rotor assembly and the second-order torsional mode of the rotor assembly is as follows:

[0145] 3-2-1) If the following conditions are met at the same time, the identified mode is the first-order torsional mode of the rotor assembly:

[0146] ① The Y-axis displacement change directions of the n modal points on the first shaft end face A are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0147] ② The Y-axis displacement change directions of the n modal points on the end surface B are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0148] ③ The Y-axis displacement change directions of the n modal points on the end surface D are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0149] ④ The Y-axis displacement change directions of the n modal points on the second shaft end surface E are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0150] Conditions ① through ④ are essentially a single-end face displacement consistency check. In a torsional mode, modal points on the same end face should rotate synchronously about the axis, resulting in consistent Y-axis displacements and similar amplitudes. By setting conditions ① through ④, we can eliminate local vibration interference, ensure that the mode is global torsion rather than local deformation, and eliminate non-torsional modes (such as bending modes with inconsistent displacement directions).

[0151] ⑤ The Y-axis displacement changes of n modal points on the end surface C are all less than 2% of the Y-axis displacement change of any modal point on the end surface A of the first axis;

[0152] Condition ⑤ is actually a mid-plane displacement suppression check. By setting condition ⑤, we can verify the correctness of the end face C position and effectively distinguish between first-order and higher-order torsional modes. Setting a 2% threshold effectively ensures modal purity, thus avoiding confusion with other modes.

[0153] ⑥ The Y-axis displacement change direction of the modal point on the first shaft end surface A is opposite to the Y-axis displacement change direction of the modal point on the second shaft end surface E, and the error of the Y-axis displacement change is within 5%;

[0154] ⑦ The Y-axis displacement change direction of the modal point on end surface B is opposite to the Y-axis displacement change direction of the modal point on end surface D, and the error of the Y-axis displacement change is within 5%;

[0155] Conditions 6-7 actually determine the antisymmetry of the end displacements. The first-order torsional mode is an antisymmetric vibration mode, meaning that the displacement directions of the first-axis end face A and the second-axis end face E, as well as the displacement directions of the end faces B and D, are opposite. By setting conditions 6-7, we can clearly distinguish first-order torsional from other modes (for example, bending modes do not have this characteristic). Furthermore, by setting a 5% error in the Y-axis displacement variation, we can ensure the authenticity of the antisymmetry.

[0156] ⑧ The displacement changes of the modal points at the centers of both end faces of the rotor shaft of the rotor assembly are less than 1% of the average displacement change of n modal points on the first shaft end face A;

[0157] Condition ⑧ actually determines the shaft end displacement constraint. The rotor shaft of the rotor assembly is constrained by bearings at both ends. In an ideal torsional mode, the shaft end displacement should be zero. Therefore, in actual simulations, the shaft end displacement must be significantly smaller than the end face (first shaft end face A) displacement (≤1%) to verify the accuracy of the support modeling. Setting Condition ⑧ effectively ensures that the identified mode is pure torsional, rather than coupled bending and torsional.

[0158] The above conditions ① to ⑧ can effectively achieve accurate identification of the first-order torsional mode through a four-in-one screening mechanism consisting of single-end face synchronous rotation constraint, node position verification, antisymmetry matching, and shaft end displacement suppression.

[0159] 3-2-2) If the following conditions are met simultaneously, the identified mode is the second-order torsional mode of the rotor assembly:

[0160] ① The Y-axis displacement change directions of the n modal points on the first shaft end face A are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0161] ② The Y-axis displacement change directions of the n modal points on the end surface B are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0162] ③ The Y-axis displacement change directions of the n modal points on the end surface D are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0163] ④ The Y-axis displacement change directions of the n modal points on the second shaft end surface E are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%;

[0164] By setting conditions ① to ④, we can eliminate local bending interference and verify the pure torsional characteristics of the end face torsion. And by keeping the Y-axis displacement difference ≤ 3%, we can ensure the rigidity of the end face torsion.

[0165] ⑤ The Y-axis displacement change directions of the n modal points on the end surface C are consistent, and the Y-axis displacement changes of the n modal points are all greater than 40% of the average Y-axis displacement change of the n modal points on the end surface A of the first axis;

[0166] The rotor assembly's second-order torsional mode has an intermediate node (end face C), with torsional deformation on either side of the node in opposite directions. The significant displacement of end face C reflects the "double-peak" characteristic of the second-order mode. Therefore, setting condition ⑤ effectively distinguishes the rotor assembly's first-order torsional mode (with minimal intermediate displacement) from the rotor assembly's second-order torsional mode (with significant intermediate displacement).

[0167] ⑥ The Y-axis displacement change direction of the modal point on end surface C is consistent with the Y-axis displacement change direction of the modal point on end surface A of the first axis; this condition can effectively determine the node position of the second-order torsional mode (not end surface C), which is consistent with the "same direction-opposite direction-same direction" distribution of the second-order torsional vibration.

[0168] ⑦ The Y-axis displacement change direction of the modal point on the first shaft end surface A is consistent with the Y-axis displacement change direction of the modal point on the second shaft end surface E, and the error of the Y-axis displacement change is within 5%; this condition can exclude the first-order torsional mode of the rotor assembly based on the "same direction at both ends" feature of the second-order mode.

[0169] ⑧ The Y-axis displacement change direction of the modal point on end surface B is consistent with the Y-axis displacement change direction of the modal point on end surface D, and the error of the Y-axis displacement change is within 5%; this condition can verify the symmetry of the modal segment and ensure the wave-shaped distribution of the second-order torsional mode vibration shape of the rotor assembly.

[0170] ⑨ The displacement changes of the modal points at the centers of both end faces of the rotor shaft of the rotor assembly are less than 2% of the average displacement changes of the n modal points on the first shaft end face A; this condition can effectively eliminate bending mode interference (bending mode has center displacement).

[0171] ⑩ The displacement change of all modal points on end surfaces B and D is less than 10% of the average displacement change of n modal points on the end surface A of the first axis; this condition can quantify the deformation characteristics of the node area and screen the modes that conform to the node distribution law of second-order torsional vibration.

[0172] 3-3) Based on the identified first-order torsional mode of the rotor assembly and the second-order torsional mode of the rotor assembly, the corresponding first-order torsional modal frequency and second-order torsional modal frequency are obtained.

[0173] Through the conditions of step 3-2-1) above, the first-order torsional mode vibration shape of the rotor assembly is identified as follows Figure 5 As shown, under the first-order torsional mode of this rotor assembly, the corresponding first-order torsional modal frequency is 2541Hz; through the conditions of the above step 3-2-2), the second-order torsional modal vibration mode of the rotor assembly is identified as follows Figure 6 As shown, under the second-order torsional mode of this rotor assembly, the corresponding second-order torsional mode frequency is 4251 Hz.

[0174] 4) According to the bending modal frequencies of each order of the rotor assembly, the equivalent bending stiffness of the rotor assembly is obtained. The formula for calculating the equivalent bending stiffness of the rotor assembly is as follows:

[0175]

[0176] Where K eq_w is the equivalent bending stiffness of the rotor assembly; F w1 is the first-order bending mode frequency; F w2 is the second-order bending modal frequency; m is the mass of the rotor assembly.

[0177] In this embodiment, the mass of the rotor assembly is m=12.0 kg, and the equivalent bending stiffness K of the rotor assembly is obtained by calculation. eq_w 13967 N·m 2 .

[0178] 5) According to the torsional modal frequencies of each order of the rotor assembly, the equivalent torsional stiffness of the rotor assembly is obtained. The formula for calculating the equivalent torsional stiffness of the rotor assembly is as follows:

[0179]

[0180] Where K eq_n is the equivalent torsional stiffness of the rotor assembly; F n1 is the first-order torsional mode frequency; F n2 is the second-order torsional modal frequency; m is the mass of the rotor assembly.

[0181] In this embodiment, the equivalent torsional stiffness K of the rotor assembly is obtained by calculation. eq_n 16165 N·m 2 .

[0182] 6) Calculate the equivalent axial stiffness of the rotor assembly;

[0183] 6-1) Perform Z-axis frequency response analysis on the four modal points on the first axis end surface A as follows:

[0184] 6-1-1) Apply an external force F to the four modal points on the end surface A of the first shaft. In this embodiment, F = 10N. Perform a frequency sweep excitation with an amplitude of 1N at 1Hz intervals (i.e., 1Hz, 2Hz, ..., 5000Hz) to obtain the vibration acceleration response of the corresponding modal point at each frequency.

[0185] 6-1-2) Based on the vibration acceleration response signal, obtain the frequency response functions of the four modal points on the first shaft end surface A;

[0186] 6-1-3) Take the average value of the frequency response function of the four modal points on the first shaft end face A to obtain the average value of the frequency response function. The average value curve of the frequency response function of the four modal points is as follows: Figure 7 As shown;

[0187] 6-2) Calculate the equivalent axial stiffness of the rotor assembly based on the average value of the frequency response function and the following formula:

[0188]

[0189] Where K eq_z is the equivalent axial stiffness of the rotor assembly; H1 represents the average value of the frequency response function of n points on the first shaft end surface A when the frequency is equal to 1 Hz; H2 represents the average value of the frequency response function of n points on the first shaft end surface A when the frequency is equal to 2 Hz; H 4999 It represents the average value of the frequency response function of n points on the end face A of the first axis when the frequency is equal to 4999 Hz; H 5000 It represents the average value of the frequency response function of n points on the end surface A of the first shaft when the frequency is equal to 5000 Hz; Maxf represents the frequency at which the maximum value occurs.

[0190] In this embodiment, the frequency of the maximum value is 4139 Hz, that is, the equivalent axial stiffness K of the rotor assembly is obtained by calculation. eq_z 14524 N·m 2 .

[0191] 7) Calculate the equivalent damping coefficient of the rotor assembly;

[0192] 7-1) Mark the four modal points on the end surface C of the motor rotor assembly as the four excitation points, and perform a frequency sweep excitation with an amplitude of 1N in the X-axis direction at an interval of 1 Hz, with a frequency change rate of 2s / 1 Hz; the frequency range is 10 Hz to 5000 Hz;

[0193] 7-2) Obtain the vibration acceleration response of each excitation point to 22 modal points in the motor rotor assembly under all frequency excitations and record them as 22 response points;

[0194] 7-3) Obtain the frequency response function according to the following formula:

[0195] HH i_j (ω)=a i_j (ω) / F j (ω)

[0196] Where, HH i_j (ω) represents the frequency response function from the jth excitation point to the ith response point, j = 1, 2, 3, 4, i = 1, 2, 3, ..., 22; a i_j(ω) represents the vibration acceleration from the jth excitation point to the ith response point; F j (ω) represents the excitation force of the jth excitation point; ω represents the frequency;

[0197] 7-4) Based on the frequency response function, the overall average response function is obtained using the following formula:

[0198]

[0199] Where Havg(ω) represents the overall average response function; M is the total number of frequency response functions; n is the number of excitation points; N is the number of response points; ω represents the frequency;

[0200] Specifically, one excitation point can excite 22 response points of the rotor assembly, and one response point is divided into three directions. In this embodiment, there are four excitation points on the end surface C. Therefore, the total number of frequency response functions is M = 22 × 3 × 4 = 264, that is:

[0201]

[0202] In this embodiment, the average response function obtained by calculation is shown in Figure 8 .

[0203] 7-5) Based on the overall average response function, the equivalent damping coefficient of the rotor assembly is obtained using the following formula:

[0204]

[0205] Where C eq is the equivalent damping coefficient of the rotor assembly; Havg(1) represents the overall average response function value at a frequency of 1 Hz; Havg(21) represents the overall average response function value at a frequency of 2 Hz; Havg(5000) represents the overall average response function value at a frequency of 5000 Hz; m is the mass of the rotor assembly.

[0206] In this embodiment, the equivalent damping coefficient C of the rotor assembly is obtained by calculation. eq It is 3.9.

[0207] 8) Based on the equivalent bending stiffness, equivalent torsional stiffness, equivalent axial stiffness, and equivalent damping coefficient of the rotor assembly, calculate the equivalent stiffness and damping coefficient of the rotor assembly, and evaluate and score the equivalent stiffness and damping coefficient of the rotor assembly.

[0208] According to the equivalent bending stiffness of the rotor assembly, the equivalent torsional stiffness of the rotor assembly, the equivalent axial stiffness of the rotor assembly and the equivalent damping coefficient of the rotor assembly, the formulas for the equivalent stiffness and damping coefficient of the rotor assembly are as follows:

[0209]

[0210] Where, δ eq is the equivalent stiffness and damping coefficient of the rotor assembly; K eq_w is the equivalent bending stiffness of the rotor assembly, K eq_n is the equivalent torsional stiffness of the rotor assembly, K eq_z is the equivalent axial stiffness of the rotor assembly, C eq is the equivalent damping coefficient of the rotor assembly.

[0211] In this embodiment, the equivalent stiffness and damping coefficient δ of the rotor assembly are obtained by calculation. eq It is 18.8.

[0212] A scoring table is constructed to evaluate and score the equivalent stiffness and damping coefficient of the rotor assembly. The scoring table is constructed through a large number of early part experiments and vehicle experiments. That is, the equivalent stiffness and damping coefficient δ of the rotor assembly are obtained through a large number of part experiments. eq The performance parameters of the vehicle are combined with the vehicle test to analyze its impact on the vibration and noise performance of the vehicle, thus establishing the δ eq The relationship between the actual performance.

[0213] Table 1: Score Sheet

[0214] Score 1 point 2 points 3 points 4 points 5 points 6 points 7 points 8 points 9 points 10 points <![CDATA[δ eq ]]> >50 45～50 40～45 35～40 30～35 25～30 20～25 15～20 10～15 <10 Rating Perfect Excellent excellent Very good good average generally Poor Difference Extremely poor

[0215] The equivalent stiffness and damping coefficient of the rotor assembly are evaluated and scored using the scoring table (see Table 1 for details). In this embodiment, δ eq The score is 18.8, which is in the range of 15 to 20, so the score is 8, which is a poor rating. Through evaluation and scoring, complex multi-physics performance parameters are converted into intuitive engineering decision signals, allowing R&D personnel to determine whether the rotor assembly should be optimized based on the final score and rating, and based on the rating results.

[0216] The rating results are a systematic assessment of the stiffness and damping characteristics of the rotor assembly components. A high rating indicates excellent stiffness and damping performance, effectively reducing vibration and noise generated by the motor assembly. It also demonstrates the rotor assembly's excellent resistance to deformation, ensuring that critical performance will not rapidly decline over time, even with increasing mileage.

[0217] From the perspective of the operating principle of the vehicle power system, the higher the power of the vehicle power system, the greater the speed and torque of the rotor assembly during operation, which is more likely to cause significant vibration and noise. If the vibration and noise performance of the entire vehicle is to reach the high-end level corresponding to the power level, the equivalent stiffness and damping coefficient δ of the rotor assembly should be eqThe most stringent criterion must be met, namely δ eq >50, corresponding to 1 point in the scoring table, is rated as "perfect".

[0218] In actual application, if only the vehicle power system power dimension is considered: when the power of the vehicle power system is above 60kW, it is necessary to require δ eq >50 (i.e., a score of 1 point, a rating of "perfect"); when the power of the vehicle's power system is between 40kW and 60kW, the score is at least 2 to 4 points; when the power of the vehicle's power system is between 20kW and 40kW, the score is at least 5 to 6 points; when the power of the vehicle's power system is below 20kW, the score is at least 7 to 8 points; and once the score is 9 to 10 points, the rotor assembly must be optimized.

[0219] If the durability of the rotor assembly components is also taken into consideration, to ensure that the rotor assembly does not experience durability failures during long-term use, the score must be at least 1 to 4. In other words, a score of 1 to 4 means that the rotor assembly has excellent stiffness and damping performance, strong deformation resistance, and can effectively ensure long-term performance stability; a score of 5 to 7 means that the durability risk is low, but there may be potential after-sales problems; and a score of 9 to 10 means that the rotor assembly cannot meet both vibration and noise and durability requirements and is judged as unqualified.

[0220] In summary, by adopting the above method and using modal identification and frequency response analysis technology, it is possible to accurately analyze and evaluate the equivalent stiffness and damping of the rotor assembly without relying on complete vehicle and bench test resources, thereby providing a quantitative basis for the structural optimization and process improvement of the motor rotor assembly. It can effectively avoid the problem of poor consistency in vibration and noise evaluation (i.e., NVH performance evaluation), shorten the development cycle, reduce development costs, and effectively improve the production qualification rate of products.

[0221] The above description is only a preferred embodiment of the present invention and is not intended to limit the present invention. Any modifications made to the present invention by those skilled in the art without departing from the spirit of the present invention shall fall within the scope of protection of the present invention.

Claims

1. A method for analyzing the equivalent stiffness and damping of a vehicle-mounted motor rotor assembly, characterized in that: The following steps are involved: 1) establishing a three-dimensional coordinate system and placing the rotor assembly in the three-dimensional coordinate system, selecting a plurality of modal points on the rotor assembly for identifying bending / torsion modes of the rotor assembly, and determining the coordinates of all modal points in the three-dimensional coordinate system; 2) identifying the bending mode of the rotor assembly based on the vector displacement changes of the modal point determined in step 1) under all modes, and obtaining the bending modal frequencies of each order of the rotor assembly; 3) identifying the torsional mode of the rotor assembly based on the vector displacement changes of the modal point determined in step 1) under all modes, and obtaining the torsional modal frequencies of each order of the rotor assembly; 4) Obtain the equivalent bending stiffness of the rotor assembly based on the bending modal frequencies of each order of the rotor assembly; 5) Obtain the equivalent torsional stiffness of the rotor assembly based on the torsional modal frequencies of each order of the rotor assembly; 6) Calculate the equivalent axial stiffness of the rotor assembly; 7) Calculate the equivalent damping coefficient of the rotor assembly; 8) Based on the equivalent bending stiffness, equivalent torsional stiffness, equivalent axial stiffness, and equivalent damping coefficient of the rotor assembly, calculate the equivalent stiffness and damping coefficient of the rotor assembly, and evaluate and score the equivalent stiffness and damping coefficient of the rotor assembly.

2. The equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly according to claim 1 is characterized in that: In step 1), the specific steps of establishing a three-dimensional coordinate system, selecting multiple modal points, and determining the coordinates of all modal points in the three-dimensional coordinate system are as follows: 1-1) Construct a cylindrical coordinate system based on the rotor assembly: With the rotor core axis as the reference, the radial direction of the rotor core is set as the X-axis, the tangent direction of the rotor core outer edge is the Y-axis, and the axial direction of the rotating shaft is the Z-axis; 1-2) between the first axial end face A and the second axial end face E of the rotor core, select end face B, end face C, and end face D, such that the first axial end face A, the second axial end face E, end face B, end face C, and end face D are equally spaced and parallel to each other; 1-3) On the first shaft end surface A, the second shaft end surface E, the end surface B, the end surface C, and the end surface D, n modal points are arranged at angular intervals with the rotor core axis as the center; 1-4) Record the spatial cylindrical coordinates of all modal points N in step 1-3) and number them in sequence.

3. The equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly according to claim 2, characterized in that: The rotor assembly bending mode includes the first-order bending mode of the rotor assembly and the second-order bending mode of the rotor assembly. In step 2), the specific steps of obtaining the bending modal frequencies of each order of the rotor assembly based on the identified first-order bending mode of the rotor assembly and the second-order bending mode of the rotor assembly are as follows: 2-1) Perform modal simulation on the rotor assembly and record the vector displacement changes of N modal points in all modes; 2-2) Based on the vector displacement changes of each modal point under all modes, and in order of frequency from low to high, the first-order bending mode of the rotor assembly and the second-order bending mode of the rotor assembly are identified as follows: 2-2-1) If the following conditions are met simultaneously, the identified mode is the first-order bending mode of the rotor assembly: ① The displacement change directions of all modal points on the first shaft end surface A and the second shaft end surface E are consistent, and the error of the displacement change is within 5%; ② The displacement change direction of all modal points on end surfaces B and D is consistent, and the error of the displacement change is within 5%; ③ The displacement changes of n modal points on the end surface C are all less than 2% of the displacement change of any modal point on the end surface A of the first axis; ④ The direction of the modal point displacement change at the center of the two end surfaces of the rotor shaft of the rotor assembly is consistent with the direction of the average displacement change of n modal points on the first shaft end surface A, and the modal point displacement change at the center of the two end surfaces of the rotor shaft of the rotor assembly is greater than 40% of the average displacement change of n modal points on the first shaft end surface A; 2-2-2) If the following conditions are met simultaneously, the identified mode is the second-order bending mode of the rotor assembly: ① The displacement change directions of all modal points on the first shaft end surface A and the second shaft end surface E are opposite, and the error of the displacement change is within 5%; ② The displacement changes of all modal points on end surfaces B and D are in opposite directions, and the error of the displacement change is within 5%; ③ The displacement changes of n modal points on end surface C are all less than 2% of the displacement change of any modal point on end surface B; ④ The displacement changes of the modal points at the centers of the two end surfaces of the rotor shaft of the rotor assembly are in opposite directions, and the displacement changes of the modal points at the centers of the two end surfaces of the rotor shaft of the rotor assembly are both greater than 40% of the average displacement change of the n modal points on the end surface B; 2-3) Based on the identified first-order bending mode of the rotor assembly and the second-order bending mode of the rotor assembly, the corresponding first-order bending modal frequency and second-order bending modal frequency are obtained.

4. The equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly according to claim 2, characterized in that: The rotor assembly torsional mode includes the first-order torsional mode of the rotor assembly and the second-order torsional mode of the rotor assembly. In step 3), the specific steps of obtaining the torsional modal frequencies of each order of the rotor assembly according to the identified first-order torsional mode of the rotor assembly and the second-order torsional mode of the rotor assembly are as follows: 3-1) Perform modal simulation on the rotor assembly and record the vector displacement changes of all modal points in all modes; 3-2) Based on the vector displacement changes of each modal point under all modes, and in order of frequency from low to high, the specific method for identifying the first-order torsional mode of the rotor assembly and the second-order torsional mode of the rotor assembly is as follows: 3-2-1) If the following conditions are met at the same time, the identified mode is the first-order torsional mode of the rotor assembly: ① The Y-axis displacement change directions of the n modal points on the first shaft end face A are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%; ② The Y-axis displacement change directions of the n modal points on the end surface B are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%; ③ The Y-axis displacement change directions of the n modal points on the end surface D are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%; ④ The Y-axis displacement change directions of the n modal points on the second shaft end surface E are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%; ⑤ The Y-axis displacement changes of n modal points on the end surface C are all less than 2% of the Y-axis displacement change of any modal point on the end surface A of the first axis; ⑥ The Y-axis displacement change direction of the modal point on the first shaft end surface A is opposite to the Y-axis displacement change direction of the modal point on the second shaft end surface E, and the error of the Y-axis displacement change is within 5%; ⑦ The Y-axis displacement change direction of the modal point on end surface B is opposite to the Y-axis displacement change direction of the modal point on end surface D, and the error of the Y-axis displacement change is within 5%; ⑧ The displacement changes of the modal points at the centers of both end faces of the rotor shaft of the rotor assembly are less than 1% of the average displacement change of n modal points on the first shaft end face A; 3-2-2) If the following conditions are met simultaneously, the identified mode is the second-order torsional mode of the rotor assembly: ① The Y-axis displacement change directions of the n modal points on the first shaft end face A are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%; ② The Y-axis displacement change directions of the n modal points on the end surface B are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%; ③ The Y-axis displacement change directions of the n modal points on the end surface D are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%; ④ The Y-axis displacement change directions of the n modal points on the second shaft end surface E are consistent, and the error of the Y-axis displacement change between the n modal points is within 3%; ⑤ The Y-axis displacement change directions of the n modal points on the end surface C are consistent, and the Y-axis displacement changes of the n modal points are all greater than 40% of the average Y-axis displacement change of the n modal points on the end surface A of the first axis; ⑥ The Y-axis displacement change direction of the modal point on the end surface C is consistent with the Y-axis displacement change direction of the modal point on the end surface A of the first axis; ⑦ The Y-axis displacement change direction of the modal point on the first shaft end surface A is consistent with the Y-axis displacement change direction of the modal point on the second shaft end surface E, and the error of the Y-axis displacement change is within 5%; ⑧The Y-axis displacement change direction of the modal point on end surface B is consistent with the Y-axis displacement change direction of the modal point on end surface D, and the error of the Y-axis displacement change is within 5%; ⑨ The displacement changes of the modal points at the centers of both end faces of the rotor shaft of the rotor assembly are less than 2% of the average displacement changes of the n modal points on the first shaft end face A; ⑩ The displacement changes of all modal points on end surfaces B and D are less than 10% of the average displacement changes of n modal points on the end surface A of the first axis; 3-3) Based on the identified first-order torsional mode of the rotor assembly and the second-order torsional mode of the rotor assembly, the corresponding first-order torsional modal frequency and second-order torsional modal frequency are obtained.

5. The equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly according to claim 3 is characterized in that: In step 4), the formula for calculating the equivalent bending stiffness of the rotor assembly is as follows: Where K eq_w is the equivalent bending stiffness of the rotor assembly; F w1 is the first-order bending mode frequency; F w2 is the second-order bending modal frequency; m is the mass of the rotor assembly.

6. The equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly according to claim 4, characterized in that: In step 5), the formula for calculating the equivalent torsional stiffness of the rotor assembly is as follows: Where K eq_n is the equivalent torsional stiffness of the rotor assembly; F n1 is the first-order torsional mode frequency; F n2 is the second-order torsional modal frequency; m is the mass of the rotor assembly.

7. The equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly according to claim 2, characterized in that: In step 6), the specific steps for calculating the equivalent axial stiffness of the rotor assembly are as follows: 6-1) Perform Z-axis frequency response analysis on n modal points on the first axis end surface A as follows: 6-1-1) Apply an external force F to n modal points on the end face A of the first shaft, and perform a frequency sweep excitation with an amplitude of 1N at intervals of 1 Hz to obtain the vibration acceleration response of the corresponding modal point at each frequency; 6-1-2) Obtain the frequency response functions of n modal points on the end surface A of the first shaft based on the vibration acceleration response signal; 6-1-3) Take the average value of the frequency response functions of n modal points on the end face A of the first shaft to obtain the average value of the frequency response function; 6-2) Calculate the equivalent axial stiffness of the rotor assembly based on the average value of the frequency response function and the following formula: Where K eq_z is the equivalent axial stiffness of the rotor assembly; H1 represents the average value of the frequency response function of n points on the first shaft end surface A when the frequency is equal to 1 Hz; H2 represents the average value of the frequency response function of n points on the first shaft end surface A when the frequency is equal to 2 Hz; H 4999 It represents the average value of the frequency response function of n points on the end face A of the first axis when the frequency is equal to 4999 Hz; H 5000 It represents the average value of the frequency response function of n points on the end surface A of the first shaft when the frequency is equal to 5000 Hz; Maxf represents the frequency at which the maximum value occurs.

8. The equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly according to claim 2, characterized in that: In step 7), the specific steps for calculating the equivalent damping coefficient of the rotor assembly are as follows: 7-1) Record n ​​modal points on the end surface C of the motor rotor assembly as n excitation points, and perform frequency sweep excitation in the X-axis direction with an amplitude of 1N at an interval of 1 Hz, with a frequency change rate of 2s / 1 Hz; the frequency range is 10 Hz to 5000 Hz; 7-2) Obtain the vibration acceleration response of each excitation point to N modal points in the motor rotor assembly under all frequency excitations, and record them as N response points; 7-3) Obtain the frequency response function according to the following formula: HH i_j (ω)=a i_j (ω) / F j (oh) Where, HH i_j (ω) represents the frequency response function from the j-th excitation point to the i-th response point, j = 1, 2, 3, ..., n, i = 1, 2, 3, ..., N; a i_j (ω) represents the vibration acceleration from the jth excitation point to the ith response point; F j (ω) represents the excitation force of the jth excitation point; ω represents the frequency; 7-4) Based on the frequency response function, the overall average response function is obtained using the following formula: Where Havg(ω) represents the overall average response function; M is the total number of frequency response functions; n is the number of excitation points; N is the number of response points; ω represents the frequency; 7-5) Based on the overall average response function, the equivalent damping coefficient of the rotor assembly is obtained using the following formula: Where C eq is the equivalent damping coefficient of the rotor assembly; Havg(1) represents the overall average response function value at a frequency of 1 Hz; Havg(21) represents the overall average response function value at a frequency of 2 Hz; Havg(5000) represents the overall average response function value at a frequency of 5000 Hz; m is the mass of the rotor assembly.

9. The equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly according to claim 8, characterized in that: In step 8), based on the equivalent bending stiffness of the rotor assembly, the equivalent torsional stiffness of the rotor assembly, the equivalent axial stiffness of the rotor assembly, and the equivalent damping coefficient of the rotor assembly, the formula for the equivalent stiffness and damping coefficient of the rotor assembly is as follows: Where, δ eq is the equivalent stiffness and damping coefficient of the rotor assembly; K eq_w is the equivalent bending stiffness of the rotor assembly, K eq_n is the equivalent torsional stiffness of the rotor assembly, K eq_z is the equivalent axial stiffness of the rotor assembly, C eq is the equivalent damping coefficient of the rotor assembly.

10. The equivalent stiffness and damping analysis method for a vehicle-mounted motor rotor assembly according to claim 9, characterized in that: In step 8), the equivalent stiffness and damping coefficient of the rotor assembly are evaluated and scored using a scoring table.

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