Method, evaluation method and system for determining unbalance forces of a planetary gear set comprising multiple non-rigid rotors

Abstract

The application belongs to the technical field of planetary gear mechanism, and particularly discloses a method, an evaluation method and a system for determining unbalanced force of a planetary gear train containing multiple non-fixed rotary bodies, which is based on basic parameters of clutches in the non-fixed rotary body assembly, calculates eccentricity of an equivalent rotating disc of the clutch oil cylinder and total eccentricity of the equivalent rotating disc of the clutch, and further calculates eccentricity of an equivalent rotating disc of a planetary gear train frame; according to the determined eccentricity of each planetary gear train frame caused by positioning error, mass deviation, tooth side gap between the inner gear ring and the planetary gear, and support error of the frame of the planetary gear train, total eccentricity of the planetary gear train frame is calculated; according to the total eccentricity of the clutch and the total eccentricity of the planetary gear train frame, dynamic unbalanced force and dynamic unbalanced torque corresponding to each equivalent rotating disc of the planetary gear mechanism are calculated. With the technical scheme, eccentricity and deflection of each rotary body of the gear mechanism under the influence of multiple factors are fully considered, and accurate dynamic unbalanced force evaluation is realized.

Description

Technical Field

[0001] This invention belongs to the technical field of planetary transmission mechanisms, and relates to a method, evaluation method and system for determining the unbalanced force of planetary transmission including multiple non-fixed rotating bodies. Background Technology

[0002] Inhomogeneous material composition of rotating parts, assembly errors, and deformation and displacement during operation can cause misalignment between the center of mass (axis of inertia) and the center of rotation (geometric axis) of the rotating parts. This ultimately leads to additional dynamic imbalance forces generated during high-speed rotation. Dynamic imbalance forces are a significant excitation source in the rotor system of planetary transmission mechanisms (planetary gearboxes). Because the planetary gearbox frame, gears, and control components in a planetary transmission mechanism are all non-fixed rotating parts, and each rotating part inevitably has clearances, the rotating part system will inevitably have eccentricity and misalignment, generating dynamic imbalance forces during high-speed operation. These dynamic imbalance forces cause excessive vibration and stress during high-speed rotation, accelerate the wear of gears, bearings, and other components, deteriorate the working environment of the planetary gearbox, and waste energy. They also limit the maximum operating speed of the planetary gearbox.

[0003] The dynamic unbalance force of non-fixed rotating bodies in planetary transmission mechanisms is affected by a variety of factors. Currently, the calculation method for dynamic unbalance force is only for a single rotating body. For planetary transmission systems containing multiple pairs of non-fixed rotating bodies, there is no effective and reliable method for determining and evaluating dynamic unbalance force. Summary of the Invention

[0004] The purpose of this invention is to provide a method, evaluation method and system for determining the dynamic imbalance force of a planetary system containing multiple non-fixed rotating bodies, so as to achieve accurate evaluation of the dynamic imbalance force.

[0005] To achieve the above objectives, the basic solution of the present invention is: a method for determining the unbalanced force of planetary displacement involving multiple non-fixed rotating bodies, comprising the following steps:

[0006] S1, Obtain the basic parameters of the non-fixed rotating body component of the planetary gearbox transmission mechanism to be calculated;

[0007] S2, based on the basic parameters of the clutch in the non-fixed rotating body assembly, calculate the eccentricity of the equivalent turntable of the clutch cylinder caused by the eccentricity of the clutch cylinder, the eccentricity of the friction plate and the spline clearance.

[0008] S3, based on the eccentricity of the clutch equivalent turntable determined in S2, caused by cylinder eccentricity, friction plate eccentricity and spline clearance, calculate the total eccentricity of the clutch equivalent turntable.

[0009] S4. Based on the basic parameters of the planetary gear set frame in the non-fixed rotating body assembly, calculate the eccentricity of the equivalent turntable of the planetary gear set frame caused by the positioning error of the planetary gear train, mass deviation, backlash between the internal gear ring and the planetary gear teeth, and support error of the frame.

[0010] S5. Based on the eccentricity of each planetary gear set frame determined in step S4, caused by the positioning error of the planetary gear train, mass deviation, backlash between the internal gear ring and the planetary gear teeth, and support error of the frame, calculate the total eccentricity of the planetary gear set frame.

[0011] S6. Based on the total eccentricity of the clutch calculated in step S3 and the total eccentricity of the planetary gear set frame calculated in step S5, calculate the dynamic unbalance force and dynamic unbalance torque corresponding to each equivalent turntable of the planetary transmission mechanism.

[0012] The working principle and beneficial effects of this basic scheme are as follows: This technical scheme fully considers the eccentricity and skewness of each rotating body of the transmission mechanism under various influencing factors. While ensuring high calculation efficiency, it is closer to the actual situation and obtains more accurate dynamic unbalanced force and dynamic unbalanced torque, which can be used to measure the high-speed operation performance of the transmission mechanism.

[0013] Furthermore, the non-fixed rotating body assembly includes a clutch, a planetary gear set frame, planetary gears, and an internal gear ring;

[0014] The basic parameters of the non-fixed rotating body assembly include the number of teeth, module, and pitch circle pressure angle of the friction plate and inner hub in the clutch; the mass of the friction plate and inner hub; the angular velocity of the friction plate; the mass and angular velocity of the clutch; the mass and angular velocity of the planetary gear set frame; the mass and angular velocity of the internal gear ring; the center distance between the planetary gears and the sun gear; the center distance between the internal gear ring and the planetary gears; and the inner and outer ring diameters and tolerance grades of the support bearings.

[0015] Obtain the necessary parameters for future use.

[0016] Furthermore, the method for calculating the eccentricity of the equivalent rotor of the clutch cylinder caused by the eccentricity of the clutch cylinder, the eccentricity of the friction plate, and the spline clearance in step S2 is as follows:

[0017] Based on the basic parameters of the clutch cylinder and the experimentally measured cylinder eccentricity e0, the eccentricity of the equivalent clutch cylinder turntable is e. d1 :

[0018]

[0019] Where m0 is the mass of the hydraulic cylinder; ω0 is the angular velocity of the hydraulic cylinder; M d The mass of the clutch cylinder equivalent turntable; ω d This is the angular velocity of the clutch, which is the same as the input shaft speed.

[0020] Based on the basic parameters of the inner hub and friction plate, and using geometric kinematics, the relationship between the eccentric displacement and the spline tooth backlash c is estimated. Assuming that the unilateral tooth backlash follows a Gaussian random distribution, the initial eccentricity a caused by the tooth backlash is calculated:

[0021]

[0022]

[0023]

[0024]

[0025] Among them, c i Let be the lateral clearance of the i-th tooth; normrnd represents the generation of random numbers following a Gaussian distribution; μ is the average value of the lateral clearance; σ is the dispersion of the lateral clearance, which ranges from (μ-3σ, μ+3σ); a left The eccentricity at which the left side of the tooth contacts the right side first; a right is the eccentricity distance at which the right side of the tooth contacts the left side of the tooth first; abs indicates taking the absolute value; n is the number of teeth; The pressure angle;

[0026] By analyzing the eccentricity caused by tooth flank clearance under different tooth numbers, it is found that the eccentricity generally decreases with the increase of tooth number. The larger the tooth number, the closer it is to the tooth flank clearance. Therefore, when the elastic deformation of the friction plate or spline teeth is not considered, the eccentricity caused by the clearance c of the friction plate is considered to be the same as the clearance, i.e., e. f = c;

[0027] Based on the obtained friction plate eccentricity e f Eccentricity equivalent to the clutch cylinder's equivalent turntable for:

[0028]

[0029] Where N is the number of clutch friction plates; m f ω represents the mass of the friction plate. f ω represents the angular velocity of the friction plate. d This is the angular velocity of the clutch, which is the same as the input shaft speed.

[0030] According to the involute spline clearance fit category and spline tooth tolerance grade in national standard GB / T 3478.1-2008, determine the eccentricity e caused by the spline backlash to the equivalent turntable of the clutch cylinder. d3 .

[0031] Calculate the corresponding eccentricity of the clutch for ease of use.

[0032] Furthermore, the method for calculating the total eccentricity of the clutch equivalent turntable in step S3 is as follows:

[0033] Total eccentricity e of the clutch dsum The eccentricity e calculated in step S2 is equivalent to the clutch cylinder's equivalent turntable. d1 Eccentricity equivalent to the clutch cylinder's equivalent turntable The eccentricity e caused by the spline backlash to the equivalent turntable of the clutch cylinder d3 The vector superposition, assuming that the eccentricity caused by all factors is in the same direction, calculates the maximum eccentricity of the clutch as follows: for:

[0034]

[0035] Calculate the overall eccentricity of the clutch. for:

[0036] .

[0037] The calculation is simple and easy to operate.

[0038] Furthermore, the method for calculating the eccentricity of the equivalent turntable of the planetary gear train frame caused by the positioning error, mass deviation, backlash between the internal gear ring and the planetary gear teeth, and support error of the frame in step S4 is as follows:

[0039] Calculate the equivalent eccentricity e of the planetary gear frame based on the planetary gear mass m1 and the planetary gear position error Δe. p1 for:

[0040]

[0041] Where ω1 is the angular velocity of the planetary gear; M p ω is the equivalent mass of the planetary frame; p The angular velocity of the planetary array frame;

[0042] Calculate the equivalent eccentricity e of the planetary gear set frame based on the mass deviation Δm of the planetary gears in the transmission mechanism. p2 for:

[0043]

[0044] Where r represents the distance from the center of the planetary gear hole to the center of rotation of the planetary gear frame;

[0045] The tooth flank clearance between the internal gear ring and the planetary gears is selected based on the center distance of the gear pair. The eccentricity e caused by the tooth flank clearance between the internal gear ring and the planetary gears in the planetary gear frame is considered. pr Approximately equal to the tooth flank clearance b between the internal gear ring and the planetary gears. rp The eccentricity of the planetary gear frame turntable is calculated as e.p3 :

[0046]

[0047] Where, m r ω is the mass of the internal gear ring of the planetary gear set; r ω is the angular velocity of the internal gear ring of the planetary gear set;

[0048] The support error of the planetary gear frame consists of the fit clearance between the inner ring of the support bearing and the input shaft, the fit clearance between the outer ring and the frame, and the actual radial clearance of the bearing. Based on the basic parameters of the support bearing and the national standard ISO 286.1-2010, the maximum fit clearances between the inner ring and the shaft, and between the outer ring and the frame are determined respectively. Based on the national standards GB / T 4604.1-2012 and ISO 5753.1-2009, the actual radial clearance of the bearing is determined. Finally, the maximum support error of the planetary gear frame is determined by superimposing these parameters, i.e., the maximum eccentricity e caused by the support error of the frame. p4 .

[0049] Obtain the eccentricity of the equivalent turntable of the planetary array frame for subsequent use.

[0050] Furthermore, the method for calculating the total eccentricity of the planetary array frame in step S5 is as follows:

[0051] Total eccentricity e of the planetary array frame psum It is the equivalent eccentricity e of the planetary frame obtained in step S4. p1 The equivalent eccentricity e of the planetary frame p2 The eccentricity of the planetary gear frame turntable is e. p3 The maximum eccentricity e caused by the support error of the frame p4 Vector superposition;

[0052] Assuming that the eccentricity caused by all factors is in the same direction, the maximum eccentricity of the planetary frame is calculated as (e psum ) max for:

[0053]

[0054] Calculate the overall eccentricity of the planetary array frame according to the national standard GB / T 3748.1-2008. for:

[0055] .

[0056] It is simple to operate and easy to use.

[0057] Furthermore, the method for calculating the dynamic unbalanced force and dynamic unbalanced torque corresponding to each equivalent turntable of the planetary transmission mechanism in step S6 is as follows:

[0058] Calculate the dynamic unbalance forces of the clutch and planetary gear set frame. , for:

[0059]

[0060]

[0061] Where m is the equivalent turntable mass; e is the eccentricity between the turntable's center of mass and its rotation center; The angular velocity of the turntable; The angle between the line connecting the center of mass of the turntable and the center of rotation and the positive X-axis.

[0062] Calculate the dynamic unbalance torque of the clutch and planetary gear set frame. , for:

[0063]

[0064]

[0065] Where L is the axial width of the turntable; β is the angle of deviation between the turntable's axis of inertia and its axis of rotation.

[0066] Obtaining the final dynamic unbalance force parameters is beneficial for analyzing the planetary transmission mechanism, thereby enabling subsequent optimization of the mechanism.

[0067] The present invention also provides a method for evaluating the unbalanced forces of planetary displacement involving multiple non-fixed rotating bodies, comprising the following steps:

[0068] Determine the parameters of the optimizable planetary transmission mechanism;

[0069] Optimize the parameters of the planetary transmission mechanism, and determine the dynamic unbalance force and dynamic unbalance torque of each equivalent turntable according to the method described in this invention;

[0070] Evaluate the impact of each optimization measure on the dynamic unbalanced force and dynamic unbalanced torque, and select the optimal result.

[0071] This allows for the assessment of unbalanced forces in planetary gear shifting, the selection of optimal parameters for planetary gear shifting mechanisms, and is simple to operate and easy to use.

[0072] The present invention also provides a system for determining the dynamic unbalance force of a planetary gearbox containing multiple non-fixed rotating bodies, including a data acquisition module and a processing module. The data acquisition module is used to acquire the basic parameters of the non-fixed rotating body components of the planetary gearbox mechanism to be calculated. The output end of the data acquisition module is connected to the input end of the processing module. The processing module executes the method described in the present invention to determine the dynamic unbalance force and dynamic unbalance torque corresponding to each equivalent turntable of the planetary gearbox mechanism.

[0073] This system allows for the acquisition of planetary displacement imbalance forces, making it easy to use.

[0074] The present invention also provides a planetary displacement unbalance force assessment system comprising multiple non-fixed rotating bodies, including the planetary displacement unbalance force determination system described in the present invention, and an assessment module, wherein the assessment module assesses the degree of influence of each optimization measure on the dynamic unbalance force and dynamic unbalance torque and selects the optimal optimization measure.

[0075] This evaluation system is used to screen for optimal optimization measures to improve the performance of planetary transmission mechanisms. Attached Figure Description

[0076] Figure 1 This is a flowchart illustrating the method for determining the planetary displacement unbalance force of multiple non-fixed rotating bodies according to the present invention.

[0077] Figure 2 This is a simplified diagram of the input end structure of the double-row planetary transmission mechanism of the present invention, which includes a method for determining the unbalanced force of planetary transmission with multiple non-fixed rotating bodies.

[0078] Figure 3 This is an equivalent diagram of the dynamic imbalance force at the input end of a double-row planetary transmission mechanism, which is based on the present invention and includes a method for determining the dynamic imbalance force of a planetary transmission with multiple non-fixed rotating bodies.

[0079] Figure 4 This is a schematic diagram of the skewed structure of the CH clutch friction plate in the present invention, which includes a method for determining the unbalanced force of planetary displacement involving multiple non-fixed rotating bodies.

[0080] Figure 5 This is a schematic diagram of the eccentric structure of the CH clutch friction plate in the present invention, which includes a method for determining the unbalanced force of planetary displacement involving multiple non-fixed rotating bodies.

[0081] Figure 6 This is a schematic diagram of the CH clutch and input shaft spline backlash structure of the method for determining the planetary displacement unbalance force of multiple non-fixed rotating bodies, as described in this invention.

[0082] Figure 7 This is a schematic diagram of the spline tooth structure of the method for determining the unbalanced force of planetary displacement involving multiple non-fixed rotating bodies in this invention;

[0083] Figure 8 This is a schematic diagram of the planetary gear positioning error in the method for determining the unbalanced force of planetary displacement involving multiple non-fixed rotating bodies according to the present invention;

[0084] Figure 9 This is a schematic diagram of the planetary gear mass unevenness structure in the method for determining the unbalanced force of planetary displacement involving multiple non-fixed rotating bodies, as described in this invention.

[0085] Figure 10This is a schematic diagram of the backlash between the internal gear ring and the planetary gear teeth in the present invention, which includes a method for determining the unbalanced force of planetary displacement involving multiple non-fixed rotating bodies.

[0086] Figure 11 This is a structural schematic diagram of the frame support error of the method for determining the unbalanced force of planetary displacement involving multiple non-fixed rotating bodies in this invention. Detailed Implementation

[0087] Embodiments of the present invention are described in detail below. Examples of these embodiments are shown in the accompanying drawings, wherein the same or similar reference numerals denote the same or similar elements or elements having the same or similar functions throughout. The embodiments described below with reference to the accompanying drawings are exemplary and are only used to explain the present invention, and should not be construed as limiting the present invention.

[0088] In the description of this invention, it should be understood that the terms "longitudinal", "lateral", "up", "down", "front", "rear", "left", "right", "vertical", "horizontal", "top", "bottom", "inner", "outer", etc., indicate the orientation or positional relationship based on the orientation or positional relationship shown in the accompanying drawings. They are only for the convenience of describing this invention and simplifying the description, and do not indicate or imply that the device or element referred to must have a specific orientation, or be constructed and operated in a specific orientation. Therefore, they should not be construed as limitations on this invention.

[0089] In the description of this invention, unless otherwise specified and limited, it should be noted that the terms "installation", "connection" and "linking" should be interpreted broadly. For example, they can refer to mechanical or electrical connections, or internal connections between two components. They can be direct connections or indirect connections through an intermediate medium. Those skilled in the art can understand the specific meaning of the above terms according to the specific circumstances.

[0090] This invention discloses a method for determining the unbalanced forces of planetary displacement involving multiple non-fixed rotating bodies, such as... Figure 1 As shown, it includes the following steps:

[0091] S1, obtain the basic parameters of the non-fixed rotating components of the planetary gearbox mechanism to be calculated; such as... Figure 2 and Figure 3 As shown, the non-fixed rotating body assembly includes a clutch, a planetary gear set, planetary gears, and an internal gear ring. The basic parameters of the non-fixed rotating body assembly include the number of teeth, module, and pitch circle pressure angle of the friction plate and inner hub in the clutch; the mass of the friction plate and inner hub; the angular velocity of the friction plate; the mass and angular velocity of the clutch; the mass and angular velocity of the planetary gear set; the mass and angular velocity of the internal gear ring; the center distance between the planetary gears and the sun gear; the center distance between the internal gear ring and the planetary gears; and the inner and outer ring diameters and tolerance grades of the support bearings.

[0092] S2, based on the basic parameters of the clutch in the non-fixed rotating body assembly, calculate the eccentricity of the equivalent turntable of the clutch cylinder caused by the eccentricity of the clutch cylinder, the eccentricity of the friction plate and the spline clearance.

[0093] S3. Based on the eccentricity of the clutch equivalent turntable determined in step S2, caused by the eccentricity of the oil cylinder, the eccentricity of the friction plate, and the spline clearance, calculate the total eccentricity of the clutch equivalent turntable.

[0094] S4. Based on the basic parameters of the planetary gear set frame in the non-fixed rotating body assembly, calculate the eccentricity of the equivalent turntable of the planetary gear set frame caused by the positioning error of the planetary gear train, mass deviation, backlash between the internal gear ring and the planetary gear teeth, and support error of the frame.

[0095] S5. Based on the eccentricity of each planetary gear set frame determined in step S4, caused by the positioning error of the planetary gear train, mass deviation, backlash between the internal gear ring and the planetary gear teeth, and support error of the frame, calculate the total eccentricity of the planetary gear set frame.

[0096] S6. Based on the total eccentricity of the clutch calculated in step S3 and the total eccentricity of the planetary gear set frame calculated in step S5, calculate the dynamic unbalance force and dynamic unbalance torque corresponding to each equivalent turntable of the planetary transmission mechanism.

[0097] This invention estimates the dynamic unbalance force and dynamic unbalance torque of a transmission mechanism, thereby estimating the magnitude of the dynamic unbalance force and dynamic unbalance torque before and after optimizing the relevant parameters of the transmission mechanism. The dynamic unbalance force and dynamic unbalance torque are indicators for measuring the high-speed operating performance of a transmission mechanism.

[0098] In a preferred embodiment of the present invention, such as Figures 4-7 As shown, the sources of clutch dynamic imbalance force are: misalignment of the CH clutch friction plate, eccentricity of the CH clutch friction plate, spline backlash between the CH clutch and the input shaft, and spline tooth profile. The method for calculating the eccentricity of the equivalent rotor of the clutch cylinder caused by clutch cylinder eccentricity, friction plate eccentricity, and spline clearance in step S2 is as follows:

[0099] Based on the basic parameters of the clutch cylinder and the experimentally measured cylinder eccentricity e0, the eccentricity of the equivalent clutch cylinder turntable is e. d1 :

[0100]

[0101] Where m0 is the mass of the hydraulic cylinder; ω0 is the angular velocity of the hydraulic cylinder; M d The mass of the clutch cylinder equivalent turntable; ω d This is the angular velocity of the clutch, which is the same as the input shaft speed.

[0102] Based on the basic parameters of the inner hub and friction plate, and using geometric kinematics, the relationship between the eccentric displacement and the spline tooth backlash c is estimated. Assuming that the unilateral tooth backlash follows a Gaussian random distribution, the initial eccentricity a caused by the tooth backlash is calculated:

[0103]

[0104]

[0105]

[0106]

[0107] Among them, c i Let be the lateral clearance of the i-th tooth; normrnd represents the generation of random numbers following a Gaussian distribution (also known as a normal distribution); μ is the average value of the lateral clearance; σ is the dispersion (standard deviation) of the lateral clearance, which ranges between (μ-3σ, μ+3σ); a left The eccentricity at which the left side of the tooth contacts the right side first; a right The offset is the distance by which the right side of the tooth contacts the left side of the tooth first; abs indicates taking the absolute value, that is, taking the absolute value of the vector in the abs() parentheses; n is the number of teeth; The pressure angle;

[0108] By analyzing the eccentricity caused by tooth flank clearance under different tooth numbers, it is found that the eccentricity generally decreases with the increase of tooth number. The larger the tooth number, the closer it is to the tooth flank clearance. Therefore, when the elastic deformation of the friction plate or spline teeth is not considered (e.g., light load or high stiffness), the eccentricity caused by the clearance c of the friction plate is considered to be the same as the clearance, i.e., e. f = c;

[0109] Based on the obtained friction plate eccentricity e f Eccentricity equivalent to the clutch cylinder's equivalent turntable for:

[0110]

[0111] Where N is the number of clutch friction plates; m f ω represents the mass of the friction plate. f ω represents the angular velocity of the friction plate. d This is the angular velocity of the clutch, which is the same as the input shaft speed.

[0112] According to the involute spline clearance fit category and spline tooth tolerance grade in national standard GB / T 3478.1-2008, determine the eccentricity e caused by the spline backlash to the equivalent turntable of the clutch cylinder. d3 .

[0113] More preferably, the method for calculating the total eccentricity of the clutch equivalent turntable in step S3 is as follows:

[0114] Total eccentricity e of the clutch dsum The eccentricity e calculated in step S2 is equivalent to the clutch cylinder's equivalent turntable. d1 Eccentricity equivalent to the clutch cylinder's equivalent turntable The eccentricity e caused by the spline backlash to the equivalent turntable of the clutch cylinder d3 The vector superposition, assuming that the eccentricity caused by all factors is in the same direction, calculates the maximum eccentricity of the clutch as follows: for:

[0115]

[0116] Calculate the overall eccentricity of the clutch. for:

[0117] .

[0118] In a preferred embodiment of the present invention, such as Figures 8-11 As shown, the sources of imbalance force in the planetary gear train of the transmission mechanism are: planetary gear positioning error, uneven planetary gear mass, backlash between the internal gear ring and planetary gear teeth, and frame support error. The method for calculating the eccentricity of the equivalent turntable of the planetary gear train frame caused by the planetary gear train positioning error, mass deviation, backlash between the internal gear ring and planetary gear teeth, and frame support error in step S4 is as follows:

[0119] Calculate the equivalent eccentricity e of the planetary gear frame based on the planetary gear mass m1 and the planetary gear position error Δe. p1 for:

[0120]

[0121] Where ω1 is the angular velocity of the planetary gear; M p ω is the equivalent mass of the planetary frame; p The angular velocity of the planetary array frame;

[0122] Calculate the equivalent eccentricity e of the planetary gear set frame based on the mass deviation Δm of the planetary gears in the transmission mechanism. p2 for:

[0123]

[0124] Where r represents the distance from the center of the planetary gear hole to the center of rotation of the planetary gear frame;

[0125] The tooth flank clearance between the internal gear ring and the planetary gears is selected based on the center distance of the gear pair. The eccentricity e caused by the tooth flank clearance between the internal gear ring and the planetary gears in the planetary gear frame is considered. prApproximately equal to the tooth flank clearance b between the internal gear ring and the planetary gears. rp The eccentricity of the planetary gear frame turntable is calculated as e. p3 :

[0126]

[0127] Where, m r ω is the mass of the internal gear ring of the planetary gear set; r ω is the angular velocity of the internal gear ring of the planetary gear set;

[0128] The support error of the planetary gear frame consists of the fit clearance between the inner ring of the support bearing and the input shaft, the fit clearance between the outer ring and the frame, and the actual radial clearance of the bearing. Based on the basic parameters of the support bearing and the national standard ISO 286.1-2010, the maximum fit clearances between the inner ring and the shaft, and between the outer ring and the frame are determined respectively. Based on the national standards GB / T 4604.1-2012 and ISO 5753.1-2009, the actual radial clearance of the bearing is determined. Finally, the maximum support error of the planetary gear frame is determined by superimposing these parameters, i.e., the maximum eccentricity e caused by the support error of the frame. p4 .

[0129] In a preferred embodiment of the present invention, the method for calculating the total eccentricity of the planetary array frame in step S5 is as follows:

[0130] Total eccentricity e of the planetary array frame psum It is the equivalent eccentricity e of the planetary frame obtained in step S4. p1 The equivalent eccentricity e of the planetary frame p2 The eccentricity of the planetary gear frame turntable is e. p3 The maximum eccentricity e caused by the support error of the frame p4 Vector superposition;

[0131] Assuming that the eccentricity caused by all factors is in the same direction, the maximum eccentricity of the planetary frame is calculated as (e psum ) max for:

[0132]

[0133] Calculate the overall eccentricity of the planetary array frame according to the national standard GB / T 3748.1-2008. for:

[0134] .

[0135] In a preferred embodiment of the present invention, the method for calculating the dynamic unbalanced force and dynamic unbalanced torque corresponding to each equivalent turntable of the planetary transmission mechanism in step S6 is as follows:

[0136] Calculate the dynamic unbalance forces of the clutch and planetary gear set frame. , for:

[0137]

[0138]

[0139] Where m is the equivalent turntable mass; e is the eccentricity between the turntable's center of mass and its rotation center; The angular velocity of the turntable; The angle between the line connecting the center of mass of the turntable and the center of rotation and the positive X-axis.

[0140] Calculate the dynamic unbalance torque of the clutch and planetary gear set frame. , for:

[0141]

[0142]

[0143] Where L is the axial width of the turntable; β is the angle of deviation between the turntable's axis of inertia and its axis of rotation.

[0144] Obtaining the final dynamic unbalance force parameters is beneficial for analyzing the planetary transmission mechanism, thereby enabling subsequent optimization of the mechanism.

[0145] The present invention also provides a method for evaluating the unbalanced force of planetary transmission including multiple non-fixed rotating bodies, comprising the following steps: determining the parameters of an optimizable planetary transmission mechanism;

[0146] Optimize the parameters of the planetary transmission mechanism, and determine the dynamic unbalance force and dynamic unbalance torque of each equivalent turntable according to the determination method described in this invention;

[0147] Evaluate the impact of each optimization measure on the dynamic unbalanced force and dynamic unbalanced torque, and select the optimal result.

[0148] In a preferred embodiment of the present invention, the method for optimizing the parameters of the planetary transmission mechanism to reduce the dynamic unbalanced force and dynamic unbalanced torque corresponding to the equivalent turntable is as follows:

[0149] Reduce the clearance c between the clutch friction plates and the inner hub to reduce the eccentricity e of the clutch equivalent disc. d2 In a simulation experiment on a double-row planetary transmission mechanism, before and after appropriately reducing the clearance between the clutch friction plate and the inner hub, the equivalent eccentricity of the improved clutch was significantly reduced, and the dynamic imbalance force of the equivalent turntable of the improved clutch was reduced by 7.9% compared with that before the improvement, indicating a good optimization effect.

[0150] Reduce the tolerance grade of the clutch input shaft and spline key teeth to reduce the eccentricity e of the clutch equivalent turntable. d3 In a simulation experiment on a double-row planetary transmission mechanism, after the key tooth tolerance grade was reduced from grade 6 to grade 5, the dynamic imbalance force of the improved clutch was reduced by 20.21% compared with that before the improvement, showing a significant optimization effect.

[0151] Reduce the clearance between the inner rings of the support bearings in the planetary gear set to reduce the eccentricity e caused by the support error of the planetary gear set frame. p4 In a simulation experiment on the improved inner ring fit clearance of the support bearing of a double-row planetary gear transmission mechanism, the dynamic unbalance force of the equivalent turntable of the improved planetary gear frame was reduced by 3.0% compared with that before the improvement, and the optimization effect was slightly better.

[0152] Reduce the backlash of the planetary gear transmission pair to reduce the eccentricity e caused by the backlash between the internal gear ring and the planet gear teeth in the planetary gear carrier. p3 In a simulation experiment comparing the clearance of the planetary gear transmission pairs in a double-row planetary transmission mechanism before and after improvement, the dynamic imbalance force of the equivalent turntable of the improved first and second row planetary gear frames was reduced by 70.57% and 79.84% respectively compared with the unimproved version, demonstrating a significant optimization effect.

[0153] This invention also provides a system for determining the dynamic unbalance force of a planetary gearbox containing multiple non-fixed rotating bodies. The system includes a data acquisition module and a processing module. The data acquisition module acquires the basic parameters of the non-fixed rotating body components of the planetary gearbox mechanism to be calculated. The output of the data acquisition module is electrically connected to the input of the processing module. The processing module executes the determination method described in this invention to determine the dynamic unbalance force and dynamic unbalance torque corresponding to each equivalent turntable of the planetary gearbox mechanism. This system allows for rapid and accurate acquisition of the planetary gearbox's dynamic unbalance force, making it easy to use.

[0154] The specific data acquisition module can input basic bearing parameters through the human-computer interaction module, or it can acquire all or part of the basic bearing parameters through sensors. For example, it can acquire the number of teeth of the friction plate and inner hub in the clutch, and the diameter of the inner and outer rings of the support bearing through an image sensor; it can acquire the center distance between the planetary gear and the sun gear, and the center distance between the internal gear ring and the planetary gear through a distance sensor; and it can acquire the angular velocity of the friction plate, the clutch angular velocity, the planetary gear set frame angular velocity, and the internal gear ring angular velocity through an angular velocity sensor.

[0155] The present invention also provides a planetary displacement unbalance force assessment system comprising multiple non-fixed rotating bodies, including the planetary displacement unbalance force determination system and assessment module described in the present invention. The assessment module assesses the degree of influence of each optimization measure on the dynamic unbalance force and dynamic unbalance torque and selects the optimal optimization measure.

[0156] In the description of this specification, references to terms such as "one embodiment," "some embodiments," "example," "specific example," or "some examples," etc., indicate that a specific feature, structure, material, or characteristic described in connection with that embodiment or example is included in at least one embodiment or example of the invention. In this specification, the illustrative expressions of the above terms do not necessarily refer to the same embodiment or example. Furthermore, the specific features, structures, materials, or characteristics described may be combined in any suitable manner in one or more embodiments or examples.

[0157] Although embodiments of the invention have been shown and described, those skilled in the art will understand that various changes, modifications, substitutions and alterations can be made to these embodiments without departing from the principles and spirit of the invention, the scope of which is defined by the claims and their equivalents.

Claims

1. A method of determining unbalance force of a planetary row including multiple non-rigid rotors, characterized by, The method comprises the following steps: S1, obtaining basic parameters of a non-fixed rotary body assembly of a planetary gear mechanism to be calculated; the non-fixed rotary body assembly comprises a clutch, a planetary gear frame, a planet wheel and an inner ring gear; The basic parameters of the non-fixed rotary body assembly include the number of teeth, modulus, pressure angle of the friction plate and inner hub in the clutch, the mass of the friction plate and the inner hub, the angular velocity of the friction plate, the mass and angular velocity of the clutch, the mass and angular velocity of the planetary gear frame, the mass and angular velocity of the inner ring gear, the center distance between the planet wheel and the sun gear, the center distance between the inner ring gear and the planet wheel, the diameters of the inner and outer rings of the supporting bearing and the tolerance grade; S2, based on the basic parameters of the clutch in the non-fixed rotary body assembly, calculating the eccentricity of the clutch cylinder equivalent rotating disc caused by the eccentricity of the clutch cylinder, the eccentricity of the clutch cylinder equivalent rotating disc caused by the eccentricity of the friction plate and the eccentricity of the clutch cylinder equivalent rotating disc caused by the spline gap; S3, according to the eccentricity of the clutch cylinder equivalent rotating disc caused by the eccentricity of the clutch cylinder, the eccentricity of the clutch cylinder equivalent rotating disc caused by the eccentricity of the friction plate and the eccentricity of the clutch cylinder equivalent rotating disc caused by the spline gap determined in step S2, calculating the total eccentricity of the clutch cylinder equivalent rotating disc by vector superposition; S4, based on the basic parameters of the planetary gear frame in the non-fixed rotary body assembly, calculating the eccentricity of the planetary gear frame equivalent rotating disc caused by the positioning error of the planet wheel system, the eccentricity of the planetary gear frame equivalent rotating disc caused by the mass deviation, the eccentricity of the planetary gear frame equivalent rotating disc caused by the inner ring gear and the planet wheel tooth side gap, and the eccentricity of the planetary gear frame equivalent rotating disc caused by the support error of the frame; S5, according to the eccentricity of the planetary gear frame equivalent rotating disc caused by the positioning error of the planet wheel system, the eccentricity of the planetary gear frame equivalent rotating disc caused by the mass deviation, the eccentricity of the planetary gear frame equivalent rotating disc caused by the inner ring gear and the planet wheel tooth side gap, and the eccentricity of the planetary gear frame equivalent rotating disc caused by the support error of the frame determined in step S4, calculating the total eccentricity of the planetary gear frame in the non-fixed rotary body assembly; S6, according to the total eccentricity of the clutch cylinder equivalent rotating disc calculated in step S3 and the total eccentricity of the planetary gear frame in the non-fixed rotary body assembly calculated in step S5, calculating the dynamic unbalance force and the dynamic unbalance torque corresponding to each equivalent rotating disc of the planetary gear mechanism.

2. The method for determining the unbalanced force of planetary displacement including multiple non-fixed rotating bodies as described in claim 1, characterized in that, The method for calculating the eccentricity of the clutch cylinder equivalent rotating disc caused by the eccentricity of the clutch cylinder, the eccentricity of the clutch cylinder equivalent rotating disc caused by the eccentricity of the friction plate and the eccentricity of the clutch cylinder equivalent rotating disc caused by the spline gap in step S2 is as follows: According to the basic parameters of the clutch oil cylinder and the eccentricity e0 of the oil cylinder obtained by experimental measurement, the eccentricity e of the equivalent rotating disc equivalent to the clutch oil cylinder is d1 : ， Wherein, m0 is the clutch cylinder mass; ω0 is the clutch cylinder angular velocity; M d is the equivalent rotating disc mass of the clutch cylinder; ω d is the angular velocity of the clutch, which is the same as the input shaft speed; According to the basic parameters of the inner hub and the friction plate, the relationship between the eccentric displacement and the spline tooth side gap c is estimated by using geometric kinematics, and the initial eccentricity a caused by the tooth side gap is calculated by assuming that the single-sided tooth side gap obeys Gaussian random distribution: ， ， ， ， where c i is the single flank clearance of the ith tooth; normrnd represents generating random numbers subject to Gaussian distribution; μ is the mean value of the flank clearance; σ is the dispersion degree of the flank clearance, the range of the flank clearance is between (μ-3σ, μ+3σ); a left is the eccentricity of the tooth left side contacting before the tooth right side; a right is the eccentricity of the tooth right side contacting before the tooth left side; abs represents taking absolute value; n is the number of teeth; is the pressure angle; By analyzing the eccentricity caused by the side clearance under different teeth, it is concluded that the eccentricity generally decreases with the increase of the number of teeth, and the larger the number of teeth, the closer to the side clearance. Therefore, when not considering the elastic deformation of the friction plate or the spline tooth, the eccentricity of the friction plate caused by the gap c is considered the same as the gap, i.e. f = c; According to the resulting friction plate eccentricity e f , equivalent to the clutch cylinder equivalent to the eccentricity of the rotating disc : ， Wherein, N is the number of clutch friction plate; m f is the mass of the friction plate; ω f is the angular velocity of the friction plate; ω d is the angular velocity of the clutch, which is the same as the input shaft speed; According to the spline gap class and spline tooth tolerance grade in the involute spline gap GB / T 3478.1-2008, the eccentricity e caused by the spline backlash to the clutch oil cylinder equivalent rotating disc is determined d3 .

3. The method of claim 2, wherein the planet row is a multiple non- fixed-point gyroid. The method for calculating the total eccentricity of the clutch cylinder equivalent rotating disc in step S3 is as follows: ​ Total eccentricity e of the clutch dsum The eccentricity e calculated in step S2 is equivalent to the clutch cylinder's equivalent turntable. d1 Eccentricity equivalent to the clutch cylinder's equivalent turntable The eccentricity e caused by the spline backlash to the equivalent turntable of the clutch cylinder d3 The vector superposition, assuming that the eccentricity caused by all factors is in the same direction, calculates the maximum eccentricity of the clutch as follows: for: ， Calculating the integrated eccentricity of a clutch is: 。 4. The method of claim 1, wherein the planetary row includes multiple non-rigid rotors. The method for calculating the eccentricity of the equivalent rotating disc of the planetary gear frame caused by the positioning error of the planetary gear train in step S4 is as follows: According to the planetary gear mass m1 and the planetary gear position error Δe, the equivalent eccentricity e of the planetary row frame is calculated p1 is: ， where ω1is the angular velocity of the planetary gear; M p is the equivalent mass of the planetary row frame; ω p is the angular velocity of the planetary row frame; According to the mass deviation Δm of the planetary gear of the planetary gear train, the equivalent eccentricity e of the planetary frame is calculated p2 is: ， Wherein, r represents the distance from the center of the planetary gear hole to the center of the planetary gear frame rotation; According to the gear pair center distance, the inner ring and the planetary gear tooth side clearance are selected, and the eccentricity e caused by the inner ring and the planetary gear tooth side clearance in the planetary gear frame pr is approximately equal to the inner ring and the planetary gear tooth side clearance b rp , and the eccentricity of the planetary gear frame rotating disc is calculated as e p3 : ， wherein m r is the mass of the planetary row ring gear; ω r is the angular velocity of the planetary row ring gear; The support error of the planetary wheel frame is composed of the matching gap between the inner ring of the support bearing and the input shaft, the matching gap between the outer ring and the frame, and the actual radial play of the bearing. Based on the basic parameters of the support bearing and the national standard ISO 286.1-2010, the maximum matching gaps between the inner ring and the shaft and between the outer ring and the frame are determined, respectively. Based on the national standard GB / T 4604.1-2012 and the national standard ISO 5753.1-2009, the actual radial play of the bearing is determined. Finally, the maximum support error of the planetary wheel frame, i.e. the maximum eccentricity e caused by the support error of the frame, is determined by superimposed calculation p4 .

5. The method of claim 4, wherein the planet row is a multiple non- rigidly connected gyrators. The method for calculating the total eccentricity of the planetary gear frame in the non-fixed rotating body assembly in step S5 is as follows: the total eccentricity e of the planetary gear frame psum is the equivalent eccentricity e of the planetary gear frame obtained in step S4 p1 , the equivalent eccentricity e of the planetary gear frame p2 , the eccentricity e of the planetary gear frame turntable p3 , the maximum eccentricity e caused by the support error of the frame p4 vector superposition Assuming that the eccentricity directions caused by each factor are in the same direction, the maximum eccentricity of the planetary gear train frame is calculated as (e psum ) max is: ， Referring to GB / T 3748.1-2008, the integrated eccentricity of the planetary gear frame is calculated is: 。 6. The method of claim 1, wherein the planetary row includes multiple non-rigid rotors. The method for calculating the dynamic unbalance force and the dynamic unbalance torque corresponding to each equivalent rotating disc of the planetary gear mechanism in step S6 is as follows: Computing dynamic unbalance forces of a clutch, a planetary row frame , : ， ， Wherein, m is clutch oil cylinder equivalent rotating disc mass; e is eccentricity of clutch oil cylinder equivalent rotating disc centroid and rotation center; is clutch oil cylinder equivalent rotating disc angular velocity; is clutch oil cylinder equivalent rotating disc centroid and rotation center connecting line distance X axis positive direction angle; Computing the dynamic unbalance moment of a clutch, a planetary row frame , is: ， ， Wherein, L is the axial width of the clutch oil cylinder equivalent rotating disc; β is the deflection angle between the inertia axis of the clutch oil cylinder equivalent rotating disc and the rotating shaft.

7. A method of evaluating unbalance forces of a planetary row comprising multiple non-rigid rotors, characterized in that, The method comprises the following steps: Determining the parameters of the planetary gear mechanism which can be optimized; Optimizing the parameters of the planetary gear mechanism, and determining the dynamic unbalance force and the dynamic unbalance torque of each equivalent rotating disc according to the method in any one of claims 1-6; Evaluating the influence degree of each optimization measure on the dynamic unbalance force and the dynamic unbalance torque and screening the optimal result.

8. A system for determining unbalance forces of a planetary row comprising multiple non-rigid rotors, characterized in that The system comprises a data acquisition module and a processing module, the data acquisition module is used to acquire the basic parameters of the non-fixed rotating body assembly of the planetary gear mechanism to be calculated, the output end of the data acquisition module is connected with the input end of the processing module, and the processing module executes the method in any one of claims 1-6 to determine the dynamic unbalance force and the dynamic unbalance torque corresponding to each equivalent rotating disc of the planetary gear mechanism.

9. A system for evaluating unbalance forces of a planetary row comprising multiple non-rigid rotors, characterized in that, The system comprises the planetary gear dynamic unbalance force determination system in claim 8, and an evaluation module, the evaluation module is used to evaluate the influence degree of each optimization measure on the dynamic unbalance force and the dynamic unbalance torque and screen the optimal optimization measure.

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