Gear gluing fault diagnosis method under friction dynamics coupling condition and related device

Abstract

The invention discloses a gear gluing fault diagnosis method under a friction dynamics coupling condition and a related device, and relates to the technical field of gear gluing fault diagnosis, and the method comprises the steps: carrying out the iterative calculation based on a dynamics model in a gear transmission process under the friction dynamics coupling condition, fluid shear stress in the gear transmission process is obtained, the contact surface heat distribution coefficient of the driving gear and the contact surface heat distribution coefficient of the driven gear are obtained through calculation based on the fluid shear stress, and the contact surface temperature of the driving gear and the contact surface temperature of the driven gear are further obtained through calculation; and further calculating to obtain a gluing safety coefficient of the driving gear and a gluing safety coefficient of the driven gear, and based on the gluing safety coefficient of the driving gear and the gluing safety coefficient of the driven gear, determining whether a gear gluing fault exists in the transmission process of the driving gear and the driven gear so as to complete gear gluing fault diagnosis. The diagnosis precision and the application range of gear gluing fault diagnosis can be improved.

Description

Technical Field

[0001] This application relates to the field of gear scuffing fault diagnosis technology, and in particular to a method and related apparatus for diagnosing gear scuffing faults under tribodynamic coupling conditions. Background Technology

[0002] Gear transmission systems operate under transient conditions such as fluctuations, heavy loads, and speed changes for extended periods, resulting in the gear teeth being in a non-steady, mixed lubrication environment characterized by frequent localized roughness peak contact and relative slippage. Under the feedback iteration of frictional excitation and lubrication conditions between the two gears, localized frictional flashing occurs on the gear teeth, potentially leading to scuffing failure. Current research on gear scuffing failure is largely applicable only to smooth surfaces, not rough ones. This is because it ignores the influence of friction, treating the gear teeth as smooth surfaces for contact surface temperature calculations and scuffing failure diagnosis, resulting in insufficient accuracy in both contact surface temperature calculations and scuffing failure diagnosis. Summary of the Invention

[0003] The purpose of this application is to provide a method and related device for diagnosing gear scuffing faults under tribodynamic coupling conditions, which can improve the diagnostic accuracy and applicability of gear scuffing fault diagnosis.

[0004] To achieve the above objectives, this application provides the following solution:

[0005] In a first aspect, this application provides a method for diagnosing gear scuffing faults under tribodynamic coupling conditions, the method comprising:

[0006] Iterative calculations are performed based on the dynamic model of gear transmission under the condition of frictional dynamic coupling to obtain the fluid shear stress during gear transmission.

[0007] Based on the fluid shear stress, the heat distribution coefficient of the contact surface of the driving gear and the heat distribution coefficient of the contact surface of the driven gear are calculated.

[0008] The contact surface temperature of the driving gear is calculated based on the heat distribution coefficient of the contact surface of the driving gear, and the contact surface temperature of the driven gear is calculated based on the heat distribution coefficient of the contact surface of the driven gear.

[0009] The adhesion safety factor of the driving gear is calculated based on the contact surface temperature of the driving gear, and the adhesion safety factor of the driven gear is calculated based on the contact surface temperature of the driven gear.

[0010] Based on the scuffing safety factor of the driving gear and the scuffing safety factor of the driven gear, it is determined whether there is a gear scuffing fault during the transmission process of the driving gear and the driven gear, and the gear scuffing fault diagnosis is completed.

[0011] Secondly, this application provides a computer device, including: a memory, a processor, and a computer program stored in the memory and capable of running on the processor, wherein the processor executes the computer program to implement the above-described method for diagnosing gear scuffing faults under tribodynamic coupling conditions.

[0012] Thirdly, this application provides a computer-readable storage medium having a computer program stored thereon, which, when executed by a processor, implements the above-described method for diagnosing gear scuffing faults under tribodynamic coupling conditions.

[0013] Fourthly, this application provides a computer program product, including a computer program that, when executed by a processor, implements the above-described method for diagnosing gear scuffing faults under tribodynamic coupling conditions.

[0014] According to the specific embodiments provided in this application, this application has the following technical effects:

[0015] This application provides a method and related apparatus for diagnosing gear scuffing faults under tribodynamic coupling conditions. Based on a dynamic model of the gear transmission process under tribodynamic coupling conditions, iterative calculations are performed to obtain the fluid shear stress during gear transmission. Based on the fluid shear stress, the contact surface heat distribution coefficients of the driving gear and driven gear are calculated. Based on the contact surface heat distribution coefficient of the driving gear, the contact surface temperature of the driving gear is calculated. Based on the contact surface heat distribution coefficient of the driven gear, the contact surface temperature of the driven gear is calculated. Based on the contact surface temperature of the driving gear, the scuffing safety factor of the driving gear is calculated. Based on the contact surface temperature of the driven gear, the scuffing safety factor of the driven gear is calculated. Based on the scuffing safety factors of the driving gear and driven gear, the presence of gear scuffing faults during the transmission process of the driving and driven gears is determined, thus completing the gear scuffing fault diagnosis. This application considers the influence of tribodynamic coupling, improving the calculation accuracy of the contact surface temperature, further improving the diagnostic accuracy of gear scuffing fault diagnosis based on contact surface temperature. Furthermore, due to the consideration of the influence of friction, it is applicable to rough surfaces, expanding the applicability of gear scuffing fault diagnosis. Attached Figure Description

[0016] To more clearly illustrate the technical solutions in the embodiments of this application or the prior art, the drawings used in the embodiments will be briefly introduced below. Obviously, the drawings described below are only some embodiments of this application. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0017] Figure 1 This is an application environment diagram for a gear scuffing fault diagnosis method under tribodynamic coupling conditions provided in Embodiment 1 of this application.

[0018] Figure 2 This is a flowchart illustrating a method for diagnosing gear scuffing faults under tribodynamic coupling conditions, as provided in Embodiment 1 of this application.

[0019] Figure 3 This is a schematic diagram of the development angle during gear transmission provided in Embodiment 1 of this application.

[0020] Figure 4 This is a schematic diagram comparing the test and simulation results of the friction coefficients of Gcr15 (high carbon chromium bearing steel) and cast aluminum bronze materials provided in Example 1 of this application; wherein, Figure 4 (a) shows a comparison of the friction coefficients of Gcr15 materials. Figure 4 (b) in the figure shows a comparison of the friction coefficients of cast aluminum bronze materials.

[0021] Figure 5 This is a schematic diagram comparing ideal working conditions with actual test working conditions provided in Embodiment 1 of this application; wherein, Figure 5 (a) in the figure shows a comparison of the slip-roll ratio between the ideal working condition and the actual test working condition. Figure 5 (b) in the figure shows the comparison between the ideal working condition and the actual test working condition load.

[0022] Figure 6 This is a schematic diagram of statistical analysis of experimental data provided in Embodiment 1 of this application; wherein, Figure 6 (a) in the figure shows the statistical data of friction coefficients for the two materials. Figure 6 (b) in the figure represents the statistical data of the deviation between the simulation and experimental results.

[0023] Figure 7 This is a schematic diagram illustrating the distribution of the sliding-rolling ratio on the surface of the gear pair under transient conditions and the verification of frictional temperature rise, provided in Embodiment 1 of this application; wherein, Figure 7 In the figure, (a) represents the slip ratio during the gear meshing cycle. Figure 7 (b) shows the flash temperature verification of the Blok formula at 1289 r / min. Figure 7 (c) in the figure represents the flash temperature verification of the Blok formula at 2089 r / min.

[0024] Figure 8 This is a schematic diagram of the structure of a computer device provided in Embodiment 2 of this application. Detailed Implementation

[0025] The technical solutions of the embodiments of this application will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some embodiments of this application, and not all embodiments. Based on the embodiments of this application, all other embodiments obtained by those skilled in the art without creative effort are within the scope of protection of this application.

[0026] Example 1

[0027] The gear scuffing fault diagnosis method under tribodynamic coupling conditions provided in this application embodiment can be applied to, for example... Figure 1 The application environment shown illustrates this. The terminal communicates with the server via a network. A data storage system stores the data the server needs to process. This system can be set up independently, integrated into the server, or located in the cloud or on another server. The terminal can send diagnostic requests to the server. Upon receiving the request, the server performs iterative calculations based on a dynamic model of gear transmission under frictional dynamic coupling conditions to obtain the fluid shear stress during gear transmission. Based on this fluid shear stress, it calculates the heat distribution coefficients of the contact surfaces of the driving and driven gears. Based on these coefficients, it calculates the contact surface temperature of both gears. Based on the contact surface temperature of both gears, it calculates the scuffing safety factor of both gears. Finally, based on these scuffing safety factors, it determines whether a gear scuffing fault exists during the transmission process, thus completing the gear scuffing fault diagnosis. The server can send the diagnostic result, which indicates whether there is a gear scuffing fault, back to the terminal.

[0028] In addition, in some embodiments, the gear scuffing fault diagnosis method under tribodynamic coupling conditions can also be implemented by a server or a terminal alone. For example, the terminal can directly process the diagnosis request to be processed, or the server can obtain the diagnosis request to be processed from the data storage system and process it.

[0029] The terminals can be, but are not limited to, various desktop computers, laptops, smartphones, tablets, IoT devices, and portable wearable devices. IoT devices can include smart speakers, smart TVs, smart air conditioners, and smart in-vehicle devices, while portable wearable devices can include smartwatches, smart bracelets, and head-mounted devices. Servers can be implemented using independent servers, server clusters composed of multiple servers, or cloud servers.

[0030] In one exemplary embodiment, such as Figure 2 As shown, a method for diagnosing gear scuffing faults under tribodynamic coupling conditions is provided. This method is executed by a computer device, specifically a terminal or server, or both. In this embodiment, the method is applied to... Figure 1 The following steps are used as an example of a server in the example.

[0031] Step S1: Based on the dynamic model of the gear transmission process under the condition of frictional dynamic coupling, perform iterative calculations to obtain the fluid shear stress in the gear transmission process.

[0032] Step S2: Based on the fluid shear stress, calculate the heat distribution coefficient of the contact surface of the driving gear and the heat distribution coefficient of the contact surface of the driven gear.

[0033] Step S3: Calculate the contact surface temperature of the driving gear based on the heat distribution coefficient of the contact surface of the driving gear, and calculate the contact surface temperature of the driven gear based on the heat distribution coefficient of the contact surface of the driven gear.

[0034] Step S4: Calculate the scuffing safety factor of the driving gear based on the contact surface temperature of the driving gear, and calculate the scuffing safety factor of the driven gear based on the contact surface temperature of the driven gear.

[0035] Step S5: Based on the scuffing safety factor of the driving gear and the scuffing safety factor of the driven gear, determine whether there is a gear scuffing fault during the transmission process of the driving gear and the driven gear, and complete the gear scuffing fault diagnosis.

[0036] By implementing steps S1 to S5 above, this embodiment can take into account the influence of frictional dynamics coupling, improve the calculation accuracy of contact surface temperature, further improve the diagnostic accuracy of gear scuffing fault diagnosis based on contact surface temperature, and because it takes into account the influence of friction, it is applicable to rough surfaces, thus improving the applicability of gear scuffing fault diagnosis.

[0037] Accurately characterizing the friction and lubrication features of gears and deeply studying the vibration performance of gear transmission systems under the influence of frictional excitation are of great significance for revising the friction dynamics coupling model and predicting gear scuffing faults. Therefore, based on the established three-dimensional lubrication and vibration coupling model of the gear transmission system, the influence of factors such as the non-Newtonian fluid rheological properties of lubricating oil, thermal shear strain rate, frictional excitation behavior, and interface flash temperature theory is further coupled to develop a friction dynamics coupling model and a friction flash temperature characteristic analysis model for the gear transmission system. This enables gear friction flash temperature performance analysis and gear scuffing fault diagnosis based on the friction dynamics coupling effect. The following section further introduces the gear scuffing fault diagnosis method under the friction dynamics coupling condition used in this embodiment.

[0038] (I) Frictional Dynamics Coupled Model of Gear Transmission System

[0039] During gear transmission, the transition from a non-contact state to a meshing state inevitably involves collisions and sliding between the gear teeth, generating friction. On one hand, the transient friction coefficient, friction arm, and friction direction of gears in alternating single and double tooth states significantly affect the dynamic characteristics of the gear transmission system. On the other hand, the frictional heat generated by the friction coefficient, combined with the non-Newtonian fluid thermal shear effect, easily leads to high instantaneous interface flash temperatures at the gear meshing point, causing sudden thermal scuffing of the meshing tooth surfaces. This intensifies vibration characteristics caused by tooth profile damage and induces other forms of failure. Therefore, it is necessary to consider frictional excitation, the non-Newtonian rheological properties of lubricating oil, thermal shear strain rate, and interface flash temperature theory as key influencing factors to complete a coupled frictional dynamics model of the gear transmission system. This will effectively reduce the incidence of scuffing failures in actual gear operation and improve gear service life.

[0040] This embodiment uses an iterative calculation based on a dynamic model of the gear transmission process under frictional dynamic coupling conditions to obtain the fluid shear stress during the gear transmission process.

[0041] Iterative calculations are performed based on a dynamic model of gear transmission under frictional dynamic coupling conditions to obtain the fluid shear stress during gear transmission, specifically including:

[0042] (1) The initial fluid shear stress and initial dry shear stress are calculated based on the initial pressure. The initial total friction force is calculated based on the initial fluid shear stress and initial dry shear stress.

[0043] The initial fluid shear stress and initial dry shear stress are calculated based on the initial pressure. Based on these initial fluid shear stresses and initial dry shear stresses, the initial total frictional force is calculated, specifically including:

[0044] 1) Using the initial pressure as input, the initial fluid shear stress is calculated using the Bair-Winer viscoelastic non-Newtonian fluid model.

[0045] Within the lubrication zone, due to the high shear strain rate of the gear under load, the shear stress is not proportional to the shear rate, and the lubricating oil exhibits non-Newtonian fluid characteristics. Therefore, the Bair-Winer viscoelastic non-Newtonian fluid model is required to calculate the fluid shear stress. The formula for calculating the fluid shear stress is as follows:

[0046]

[0047] In equation (1), is the first derivative of γ, where γ is the shear strain; For τ m The first derivative, τ m For fluid shear stress; G ∞ τ is the ultimate shear modulus of the fluid. L η is the ultimate shear stress of the fluid; η is the kinematic viscosity of the lubricating oil.

[0048] Fluid limiting shear modulus G ∞ and fluid ultimate shear stress τ L Both are rheological property parameters, and both are affected by pressure p and temperature T. The calculation formula is:

[0049]

[0050] In equation (2), p is pressure and T is temperature.

[0051] τ L =0.25G ∞ (3)

[0052] Substituting the initial pressure as pressure p into equations (1)-(3), the initial fluid shear stress can be calculated.

[0053] 2) Based on the initial pressure, the initial dry shear stress is calculated.

[0054] Although the boundary lubrication mechanism is highly complex and no mature model exists for accurate prediction, in engineering practice, the boundary lubrication friction coefficient only fluctuates within a small range. In the dry contact zone, using an empirical value of 0.14 for the friction coefficient, the dry shear stress is obtained by multiplying it by the corresponding pressure. The calculation formula is as follows:

[0055] τ c =0.14p(4)

[0056] In equation (4), τ c This is dry shear stress.

[0057] Substituting the initial pressure as pressure p into equation (4), the initial dry shear stress can be calculated.

[0058] 3) Based on the initial fluid shear stress, the initial fluid oil film friction force is calculated.

[0059] Fluid oil film friction force f m This can be achieved by adjusting the fluid shear stress τ m The result is obtained by integration within the computational domain, and the formula is as follows:

[0060]

[0061] In equation (5), Ω is the computational domain, which is a two-dimensional computational region determined by the user based on experience, x is the x-coordinate of a certain position, and y is the y-coordinate of a certain position.

[0062] The initial fluid shear stress is taken as the fluid shear stress τ. m Substituting into equation (5), the initial fluid oil film friction force can be calculated.

[0063] 4) The initial dry friction force is calculated based on the initial dry shear stress.

[0064] dry shear stress τ c By performing integration, the dry friction force is obtained, and the calculation formula is as follows:

[0065]

[0066] In equation (6), f c It is dry friction, which is the friction generated by the contact of micro-protrusions.

[0067] The initial dry shear stress is taken as the dry shear stress τ. c Substituting into equation (6), the initial dry friction force can be calculated.

[0068] 5) Calculate the sum of the initial fluid oil film friction and the initial dry friction to obtain the initial total friction.

[0069] Combining the fluid oil film friction and dry friction, the total friction force f is obtained, and the calculation formula is as follows:

[0070] f = f c +f m (7)

[0071] The initial fluid oil film friction force is taken as the fluid oil film friction force f. m The initial dry friction force is taken as the dry friction force f. c Substituting into equation (7), the initial total friction force can be calculated.

[0072] (2) Using the initial total friction force as input, the dynamic model of the gear transmission process under the friction dynamic coupling condition is solved to obtain the rotation angle of the driving gear and the rotation angle of the driven gear.

[0073] Based on the gear bending-torsional coupling dynamic model, the friction dynamic coupling equation of the gear transmission system under frictional excitation is introduced, and the dynamic model is:

[0074]

[0075]

[0076]

[0077]

[0078]

[0079]

[0080] The relationship between the friction lever arm H and the unfolding angle is:

[0081]

[0082]

[0083] In the above formula, m1 is the mass of the driving gear; The x-direction translational acceleration of the driving gear; x is the stiffness coefficient of the driving gear in the x direction; x1 is the translational displacement of the driving gear in the x direction. is the damping coefficient in the x-direction of the driving gear; λ is the translational velocity of the driving gear in the x-direction. f f is the coefficient of frictional direction; 1i Let m1 be the first frictional force of the i-th gear; m2 is the mass of the driven gear. Let x be the translational acceleration of the driven gear in the x-direction; x is the stiffness coefficient of the driven gear in the x-direction; x2 is the translational displacement of the driven gear in the x-direction. The damping coefficient in the x-direction of the driven gear; f is the translational velocity of the driven gear in the x-direction. 2i The second frictional force of the i-th gear; The translational acceleration in the y-direction of the driving gear; y is the stiffness coefficient in the y direction of the driving gear; y1 is the translational displacement in the y direction of the driving gear. y is the damping coefficient of the driving gear; k is the translational velocity in the y-direction of the driving gear. t R is the gear stiffness coefficient. b1 R is the base circle radius of the driving gear; θ1 is the rotation angle of the driving gear; b2 θ2 is the base circle radius of the driven gear; θ2 is the rotation angle of the driven gear; y2 is the translational displacement of the driven gear in the y direction; c t This refers to the gear damping coefficient; The rotational angular velocity of the driving gear; The angular velocity of the driven gear; Let be the translational velocity in the y-direction of the driven gear; Let y be the translational acceleration of the driven gear. y is the stiffness coefficient of the driven gear in the y direction; Iy is the damping coefficient of the driven gear in the y direction; I1 is the moment of inertia of the driving gear. k is the rotational angular acceleration of the driving gear. p θ is the torsional stiffness coefficient of the external shaft system of the driving gear. m c is the rotation angle of the motor; p The torsional damping coefficient of the external shaft system of the driving gear; H is the angular velocity of the motor. 1i Ii is the first frictional force arm of the i-th gear; I2 is the moment of inertia of the driven gear; k is the rotational angular acceleration of the driven gear. g θ is the torsional stiffness coefficient of the external shaft system of the driven gear. b c is the rotation angle of the load; g The torsional damping coefficient for the connection of the external shaft system of the driven gear; H is the rotational angular velocity of the load; 2i I is the second frictional force arm of the i-th gear; m The moment of inertia of the motor shaft system; M1 is the angular acceleration of the motor; M2 is the torque applied by the motor; I b The moment of inertia of the load shaft system; M1 is the rotational angular acceleration of the load; M2 is the torque applied by the load; H 11 (t) represents the first frictional force arm of the driving gear at time t; α A The development angle at point A during gear transmission is the initial engagement point; mod() represents the remainder operation; f m Z represents the frictional force of the fluid oil film; Z represents the number of teeth; α D H is the development angle at point D during gear transmission, where point D is the transition point from double-tooth meshing to single-tooth meshing; 12 (t) represents the first frictional force arm of the driven gear at time t; αB H is the development angle at point B during gear transmission, where point B is the transition point from single-tooth meshing to double-tooth meshing; 21 (t) represents the second frictional force arm of the driving gear at time t; β2 is the actual meshing angle; H 22 (t) represents the second frictional force arm of the driven gear at time t.

[0084] Besides inducing translational vibrations along the perpendicular line of meshing, tooth surface friction also generates a frictional torque that constrains torsional vibrations. Furthermore, the dynamic behavior of tooth surface friction excitation is complex due to time-varying effects in the friction arm, friction coefficient, and friction direction, coupled with the superimposed influence of multiple meshing gear pairs. To accurately describe the meshing position and state and obtain the directional relationship of the frictional forces, a gear development angle α is introduced. Each key point (AE) of gear meshing corresponds to a unique development angle, such as... Figure 3 As shown, Figure 3 In the middle, O p Let P, A, B, C, D, and E represent the different positions from engagement to disengagement. P is the engagement reference point, A is the engagement start point (i.e., engagement point), B is the transition point from single-tooth meshing to double-tooth meshing, C is the node, D is the transition point from double-tooth meshing to single-tooth meshing, and E is the engagement end point (i.e., disengagement point). Single-tooth meshing means that at any given time, only one pair of teeth is engaged, transmitting power and motion. Double-tooth meshing means that at any given time, two pairs of teeth are engaged simultaneously, jointly transmitting power and motion. A Let α be the development angle of point A. B Let α be the development angle of point B. C Let α be the development angle of point C. D Let α be the development angle of point D. E Let β2 be the development angle at point E, and R be the actual engagement angle. b Let λ be the base circle radius of the gear, from which the frictional force direction coefficient λ can be defined. f The formula for calculating the direction coefficient of friction is:

[0085]

[0086] In equation (18), sign() is the sign function; α C α is the development angle of point C during gear transmission; C -α A The directional coefficient is the development angle from the meshing initiation point to the node; that is, the directional coefficient changes from positive to negative after the meshing point passes the node. `mod()` performs the remainder operation. D -α A This represents the unfolding angle corresponding to a single pitch.

[0087] The first friction force and the second friction force are equal, both equal to the total friction force during the transmission process of the driving gear and the driven gear. That is, the first friction force of the i-th gear and the second friction force of the i-th gear are both equal to the initial total friction force.

[0088] (3) The updated pressure is calculated based on the rotation angle of the driving gear and the rotation angle of the driven gear.

[0089] Based on the rotation angles of the driving gear and the driven gear, the updated pressure is calculated, specifically including:

[0090] 1) Based on the rotation angles of the driving gear and the driven gear, the dynamic transmission error during the transmission process of the driving gear and the driven gear is calculated.

[0091] The dynamic relative displacement difference between the driving gear and the driven gear is the dynamic transmission error. The formula for calculating the dynamic transmission error is:

[0092] δ=R b1 θ1-R b2 θ2+y1-y2+e (19)

[0093] In equation (19), δ represents the dynamic transmission error; R b1 R is the base circle radius of the driving gear; θ1 is the rotation angle of the driving gear; b2 θ1 is the base circle radius of the driven gear; θ2 is the rotation angle of the driven gear; y1 is the translational displacement of the driving gear in the y direction; y2 is the translational displacement of the driven gear in the y direction; e is the gear manufacturing error, that is, the error in tooth pitch and tooth profile that exists during the manufacturing process of the gear.

[0094] The formula for calculating gear manufacturing errors is:

[0095] e(t) = e r sin(2πf mm t+φ)(20)

[0096] In equation (20), e(t) represents the gear manufacturing error at time t; e r This is a comprehensive error resulting from manufacturing errors such as gear meshing pressure angle error and pitch error; f mm φ is the meshing frequency; φ is the initial phase.

[0097] 2) Based on the dynamic transmission error, the dynamic meshing force is calculated.

[0098] The formula for calculating the dynamic meshing force during gear transmission is:

[0099] F dp =k t δ(21)

[0100] In equation (21), F dp For dynamic meshing force; k t This is the gear stiffness coefficient.

[0101] 3) The updated pressure is calculated based on the dynamic meshing force.

[0102] The updated formula for calculating pressure is:

[0103] F dp (t)=∫∫ Ω p(x,y,t)dxdy(22)

[0104] In equation (22), F dp (t) represents the dynamic meshing force at time t; p represents the pressure.

[0105] Substituting the dynamic meshing force into equation (22), p can be calculated, and p at this time is the pressure after renewal.

[0106] (4) The updated fluid shear stress and the updated dry shear stress are calculated based on the updated pressure. The updated total friction force is calculated based on the updated fluid shear stress and the updated dry shear stress.

[0107] Based on the updated pressure calculations, the updated fluid shear stress and updated dry shear stress are obtained. Based on the updated fluid shear stress and updated dry shear stress, the updated total frictional force is calculated, specifically including:

[0108] 1) Using the updated pressure as input, the updated fluid shear stress is calculated using the Bair-Winer viscoelastic non-Newtonian fluid model.

[0109] Substituting the updated pressure as pressure p into equations (1)-(3), the updated fluid shear stress can be calculated.

[0110] 2) Based on the updated pressure, the updated dry shear stress is calculated.

[0111] Substituting the updated pressure as pressure p into equation (4), the updated dry shear stress can be calculated.

[0112] 3) Based on the updated fluid shear stress, the updated fluid oil film friction force is calculated.

[0113] The updated fluid shear stress is taken as the fluid shear stress τ. m Substituting into equation (5), the frictional force of the updated fluid oil film can be calculated.

[0114] 4) Based on the updated dry shear stress, the updated dry friction force is calculated.

[0115] The updated dry shear stress is taken as the dry shear stress τ.c Substituting into equation (6), the updated dry friction force can be calculated.

[0116] 5) Calculate the sum of the updated fluid oil film friction and the updated dry friction to obtain the updated total friction.

[0117] The updated fluid oil film friction force is taken as the fluid oil film friction force f. m After the update, the dry friction force is taken as the dry friction force f. c Substituting into equation (7), the updated total friction force can be calculated.

[0118] (5) Based on the initial total friction force and the updated total friction force, determine whether to end the iteration; if yes, use the updated fluid shear stress as the fluid shear stress in the gear transmission process; if no, use the updated pressure, updated fluid shear stress, updated dry shear stress and updated total friction force as the initial pressure, initial fluid shear stress, initial dry shear stress and initial total friction force for the next iteration, and return to the step of "using the initial total friction force as input to solve the dynamic model of the gear transmission process under the friction dynamic coupling condition, and obtain the rotation angle of the driving gear and the rotation angle of the driven gear".

[0119] Based on the initial total friction force and the updated total friction force, the process determines whether to end the iteration. Specifically, this involves: calculating the absolute value of the difference between the initial and updated total friction forces; then calculating the ratio of this absolute value to the initial total friction force. If the ratio is less than or equal to a preset error, the iteration ends; if the ratio is greater than the preset error, the iteration continues. The preset error can be 10. -4 .

[0120] (II) Analysis Model of Friction Flash Temperature Characteristics of Gear Transmission System

[0121] With the advancement of current technology, the transmission power, load, and speed of gear transmission systems are continuously increasing. Under the severe friction and wear caused by dynamic loads and fluctuating speeds between meshing tooth surfaces, when the contact temperature rise between the meshing tooth surfaces exceeds the critical scuffing temperature, severe tearing and scuffing damage will occur in the surface material. Tooth surface scuffing is a very typical tooth surface damage failure phenomenon in high-speed, heavy-load gear transmission systems, characterized by rapid development and severe damage. As an important parameter for measuring the load-bearing capacity of tooth surface scuffing, the tooth surface friction flash temperature presents difficulties and challenges in calculation due to its time-varying dynamic characteristics.

[0122] This embodiment calculates the heat distribution coefficient of the contact surface of the driving gear and the driven gear based on fluid shear stress. Based on the heat distribution coefficient of the contact surface of the driving gear, the contact surface temperature of the driving gear is calculated. Based on the heat distribution coefficient of the contact surface of the driven gear, the contact surface temperature of the driven gear is calculated. Based on the contact surface temperature of the driving gear, the scuffing safety factor of the driving gear is calculated. Based on the contact surface temperature of the driven gear, the scuffing safety factor of the driven gear is calculated. Based on the scuffing safety factors of the driving gear and the driven gear, it is determined whether there is a gear scuffing fault during the transmission process of the driving gear and the driven gear, thus completing the gear scuffing fault diagnosis.

[0123] (1) Calculation of heat distribution coefficient of contact surface

[0124] Under the influence of tooth surface roughness, heat during meshing is generated by the viscous shear of the lubricating oil film and the sliding contact of micro-protrusions, and carried away by the lubricating oil or by conduction between the two contact surfaces. Therefore, based on the heat distribution model of Francis and Platt, the generated heat is divided into two parts according to the heat distribution coefficient. At the same time, the heat distribution coefficient between the two contact surfaces of the gear pair is determined according to the local velocity distribution and temperature change curve of the lubrication.

[0125] The formulas for calculating the heat distribution coefficient of the contact surface of the driving gear and the heat distribution coefficient of the contact surface of the driven gear are as follows:

[0126]

[0127] In equation (23), q(x,y) is the heat distribution coefficient at (x,y), and (x,y) are the x and y coordinates of a certain position; τ(x,y) is the fluid shear stress at (x,y); U2 is the tangential velocity of the driven gear; U1 is the tangential velocity of the driving gear; q A (x,y) is the heat distribution coefficient of the contact surface of the driving gear at (x,y); K f The thermal conductivity of the lubricating oil; The transient temperature of the driven gear; q represents the transient temperature of the driving gear. B (x,y) is the heat distribution coefficient of the contact surface of the driven gear at (x,y).

[0128] (2) Calculation of contact surface temperature

[0129] The calculation of flash temperature is based on the theory of moving heat sources on a semi-infinite body. Based on this theory, a calculation model for the transient temperature of the contact surface of a gear pair is established, and the second type of Volterra integral equation is solved using an iterative method. The formulas for calculating the contact surface temperatures of the driving gear and the driven gear are as follows:

[0130]

[0131] In the above formula, T1(ξ) is the contact surface temperature of the driving gear, and ξ is the position of the calculation point, which can be represented by coordinates; T b1 ρ is the initial surface temperature of the driving gear; ρ1 is the density of the driving gear; C1 is the specific heat capacity of the driving gear; k1 is the thermal conductivity of the driving gear; x is the initial position of the contact area; k f λ is the thermal conductivity; h is the equivalent thickness of the gear contact surface; T2(λ) is the instantaneous contact temperature of the driven gear, λ is the integral variable; T1(λ) is the instantaneous contact temperature of the driving gear; q A (λ) is the heat distribution coefficient of the contact surface of the driving gear; T2(ξ) is the contact surface temperature of the driven gear; T b2 ρ is the initial surface temperature of the driven gear; ρ2 is the density of the driven gear; C2 is the specific heat capacity of the driven gear; k2 is the thermal conductivity of the driven gear; q B (λ) is the heat distribution coefficient of the contact surface of the driven gear.

[0132] (3) Calculation of the safety factor for bonding

[0133] The formula for calculating the bonding safety factor is:

[0134]

[0135] In equation (26), S sint θ is the safety factor for bonding. sint The limiting temperature; θ sin To calculate the temperature, when the calculated temperature is the contact surface temperature of the driving gear, the scuffing safety factor of the driving gear is calculated; when the calculated temperature is the contact surface temperature of the driven gear, the scuffing safety factor of the driven gear is calculated.

[0136] To assess the degree of frictional flash failure during gear meshing, the limiting temperature for tooth surface scuffing failure is calculated. The formula for calculating the limiting temperature is as follows:

[0137] θ sint = (0.85 + 1.4X) w )(ρη) -0.05 S FZG 2 (27)

[0138] In equation (27), X w Let X be the welding coefficient of the gear material. For the gear under study, X w =1.15; ρ is the density of the gear material; η is the kinematic viscosity of the lubricating oil; S FZG For the oil bonding load rating, S is the specific rating for the model under study.FZG =8.

[0139] (4) Gear scuffing fault diagnosis

[0140] Based on the scuffing safety factors of the driving gear and the driven gear, the system determines whether scuffing faults exist during the transmission process of the driving and driven gears, thus completing the gear scuffing fault diagnosis. Specifically, this includes: if the scuffing safety factor of the driving gear is less than a first preset value, or the scuffing safety factor of the driven gear is less than a first preset value, then a high scuffing safety risk is determined to exist during the transmission process of the driving and driven gears; if the scuffing safety factor of the driving gear is greater than or equal to the first preset value and less than or equal to a second preset value, or the scuffing safety factor of the driven gear is greater than or equal to the first preset value and less than or equal to the second preset value, then a medium scuffing safety risk is determined to exist during the transmission process of the driving and driven gears; if the scuffing safety factor of the driving gear is greater than the second preset value, or the scuffing safety factor of the driven gear is greater than the second preset value, then a low scuffing safety risk is determined to exist during the transmission process of the driving and driven gears.

[0141] Where the first preset value is 1 and the second preset value is 2, at this time, S sint <1 indicates a high risk of adhesive bonding safety, 1≤S sint ≤2 represents a medium-strength bonding safety risk, S sint A score greater than 2 indicates a low risk of adhesive bonding.

[0142] In the calculation process of tribodynamic coupling and flash temperature of gear transmission systems, frictional excitation plays a crucial role in connecting the vibration characteristics of the system with the mechanical features of the interface. While gear interface friction affects the vibration characteristics of the transmission shaft, the shaft vibration velocity and vibration load also have a feedback effect on the interface friction, thus forming a tribodynamic coupling chain. Based on this, the focus of this embodiment is to construct a tribodynamic coupling model of the gear transmission system with dynamic response, oil film state, and frictional excitation as coupling factors. Based on the three-dimensional lubrication state and vibration characteristic coupling analysis model of the gear transmission system, and further considering the influence of interface friction-heat effects, when calculating the friction coefficient and interface contact flash temperature during gear meshing, firstly, considering the influence of the non-Newtonian fluid effect of the lubricating oil film during gear meshing, the Bair-Winer rheological model and the modified Dowson model are used to calculate the oil film shear stress and shear modulus during meshing. Secondly, considering the influence of fluctuating loads and instantaneous rotational speed in the dynamic response, the contact shear stress of the tooth surface micro-protrusions during meshing is solved. The frictional force in the contact domain during gear meshing is obtained by integrating the oil film shear stress and the micro-protrusion shear stress, and this frictional force is fed back as a frictional excitation to the vibration characteristic model of the gear system, further refining the coupled analysis of micro-interface mechanics and macro-transmission system vibration. Subsequently, the heat generated in the tooth surface contact domain during meshing is solved by combining the oil film shear stress, the micro-protrusion contact shear stress, and the relative slip between the two interfaces. Finally, based on the theory of rapidly moving heat sources, the obtained heat is substituted into the second type of Volterra integral equation, and the interface flash temperature is calculated using an iterative method.

[0143] The following experiment will verify this.

[0144] (I) Verification by Equivalent Simulation Test of Friction Excitation in Gear Transmission System

[0145] To verify the accuracy of the frictional excitation caused by the friction coefficient in the coupled friction dynamics model of the gear transmission system, a verification test of the simulated gear line contact friction coefficient was carried out by using the equivalent method of gear slip-roll ratio and load during operation. The input parameters of lubricating oil calculation and material property parameters during the verification process are shown in Table 1 and Table 2. Since the stress calculation is related to the elastic modulus of the material, the stress range is different in the equivalent process of different materials during the test. The specific equivalent working condition correspondence is shown in Table 3.

[0146] Table 1 Main parameters of lubricant

[0147] name numerical values <![CDATA[Density (kg·m -3 )]]> 890 <![CDATA[Specific heat capacity (J·(kg·℃) -1 )]]> 2000 <![CDATA[Thermal conductivity (W·(m·℃) -1 )]]> 0.14 Dynamic viscosity (Pa·s) 0.075 <![CDATA[Barus coefficient of piezoviscosity (Pa -1 )]]> <![CDATA[2.2e -8 ]]>

[0148] Table 2. Property parameters of the two materials used in the experiment.

[0149] name Gcr15 Cast aluminum bronze Elastic modulus (GPa) 206 103 Poisson's ratio 0.3 0.3 <![CDATA[Density (kg·m -3 )]]> 7800 8200 <![CDATA[Specific heat capacity (J·(kg·℃) -1 )]]> 470 420 <![CDATA[Thermal conductivity (W·(m·℃) -1 )]]> 46 56

[0150] Table 3 Equivalent Test Conditions

[0151]

[0152]

[0153] This embodiment provides a testing machine for testing the tribological properties of sliding-rolling contact. The machine mainly consists of a drive shaft motion system, a test shaft motion system, a test force loading system, a measurement system, a lubrication system, an electrical control box, and a measurement and control system. The lower part of the machine is the base, with the lower spindle and its drive system mounted on the upper left side. A hydraulic cylinder and piston are located in the middle of the base. The upper spindle and its drive system are mounted on the upper right rear side of the base, and the high-voltage electrical control system is housed inside the base. The main hydraulic oil source and lubricating oil source are located at the rear of the main machine, connected to the main testing machine via hydraulic circuits and control cables. The electrical control box performs signal acquisition, transformation, and transmission tasks, and is connected to the industrial control computer measurement and control system via a serial cable. The testing principle of this testing machine is that when the test rings rotate and there is a certain pressure between them, a frictional torque will be generated between them. This frictional torque acts on the speed and torque sensor through the main shaft and coupling. The control system then processes the collected frictional torque, calculates the frictional force through the relationship between the frictional torque and the radius of the test piece, and obtains the coefficient of friction by the quotient of the frictional force and the normal force.

[0154] The geometric dimensions of the upper and lower test pieces in this experiment conform to the standard "Metallic Materials Rolling Contact Fatigue Test Method". The upper test piece is mounted on the shaft of the upper test piece, and the lower test piece is mounted similarly. Both the upper and lower test pieces are 60mm diameter rings. The equivalent simulation of line contact during gear meshing is achieved by the contact and grinding of the upper and lower test pieces made of the same material. Figure 4 and Figure 5 The figures show the deviations in the friction coefficient tests of Gcr15 and cast aluminum bronze (CuAl9Fe4) specimens using polishing processes, as well as the dynamic deviations of the sliding-rolling ratio and load conditions between the actual operation of the testing machine and the ideal conditions shown in Table 3. Figure 6 The friction coefficient and deviation data of the two friction characteristics test are statistically described.

[0155] Depend on Figure 4 and Figure 5It can be seen that the gear line contact friction coefficient calculated by the proposed method has a maximum deviation of only 16.37% compared with the friction coefficient measured by the sliding-rolling contact fatigue tribological characteristic test. This maximum deviation is due to the unavoidable speed fluctuations in machine control, the difference between the actual sliding-rolling ratio and the ideal working condition, and the fact that the actual applied load is not completely consistent with the equivalent load of gear meshing under the ideal condition. Furthermore, it cannot be guaranteed that the surface morphology environment during the test is the same as that in the simulation process. Figure 6 Statistical analysis of the experimental data shows that, considering the friction coefficients of the two materials, Gcr15 has a higher friction coefficient than cast aluminum bronze, and the dispersion of Gcr15 is significantly better than that of cast aluminum bronze, resulting in more stable experimental performance. In conclusion, the reliability of the friction excitation applied to the friction dynamics coupling model of the gear transmission system has been verified.

[0156] (II) Verification of interface temperature rise characteristics under gear tribodynamic coupling

[0157] The accuracy of the flash temperature prediction model of the gear transmission system after frictional dynamic coupling analysis was verified based on Blok flash temperature theory. According to the Blok flash temperature criterion, the maximum temperature of the tooth surface contact is the sum of the body temperature and flash temperature of the gear in the tooth surface friction contact area during the transmission process.

[0158] θ B =θ M +θ fla (28)

[0159] In equation (28), θ B θ is the contact temperature of the tooth surface. M θ represents the temperature of the gear body. fla This refers to the flash temperature.

[0160] According to Blok's flash temperature theory, the flash temperature θ fla It is possible to obtain:

[0161]

[0162] In equation (29), u is the temperature rise coefficient; f μ f is the coefficient of friction; e denoted as , where is the normal load on the tooth surface per unit tooth width; b is the contact half-width between the driving gear and the driven gear.

[0163] The flash temperature θ caused by friction between gear pairs during transmission can be obtained. fla The expression for how it changes with time t:

[0164]

[0165] During gear transmission, the tooth surface contact temperature θ B The function with respect to time t can be expressed as:

[0166] θ B (t)=θ M +θ fla (t)(31)

[0167] Based on the above explanation of the theoretical calculation method for tooth surface contact flash temperature, and given that the body temperature of the meshing gear remains basically constant after the gear pair enters a stable operating state, the tooth surface contact temperature θ B The trend of (t) should be related to the flash temperature θ fla The trend of change of (t) is the same. When the meshing point of the gears in the ideal meshing state moves to the pitch point, the relative sliding velocity between the tooth profiles of the driving and driven gears is zero, so the flash temperature at the pitch point at this time is... It can be represented as:

[0168]

[0169] Considering the two speed conditions during gear meshing at the oil pump shaft in a gear transmission system, these conditions are substituted into the Blok fitting formula and the friction flash temperature calculation model for a gear pair with a purely smooth surface, and then... Figure 7 The two calculation results shown are compared and analyzed. Figure 7 As shown in (a), compared to the ideal meshing state where the relative sliding velocity at the gear node is zero, during actual meshing, due to the speed fluctuations caused by the dynamic response of the gear transmission system, the velocities between the two contact surfaces at the node position within the gear meshing cycle have a very small difference but are not completely equal. During actual meshing, the slip-roll ratio is slightly greater than 0, specifically 0.015, indicating that there is basically no relative sliding between the two tooth surfaces. However, during the gear engagement and disengagement stages, the slip-roll ratio is large, the velocity difference between the two contact tooth surfaces is large, and relative sliding occurs.

[0170] Depend on Figure 7 (b) and Figure 7 As shown in (c), the flash temperature on the smooth gear surface exhibits an approximately "V"-shaped distribution during the complete gear meshing cycle. The flash temperature reaches its maximum at the gear engagement and disengagement points, with the maximum value occurring at a relatively low speed of 1289 r / min. Simultaneously, at both speeds, the flash temperature reaches its minimum at the contact point, and this minimum value is slightly greater than 0. This is because the flash temperature distribution on the gear surface during the meshing cycle under a smooth surface is related to the velocity state of the two surfaces, i.e., the slip-roll ratio distribution trend. At the contact point, the two tooth surfaces basically do not slide relative to each other, resulting in low oil film shear force and low flash temperature within the contact area. However, during the gear engagement and disengagement stages, the relative sliding between the two surfaces increases the shear stress of the lubricating oil film, leading to an increase in flash temperature. (Comprehensive analysis) Figure 7 It can be seen that in the process of comparing and verifying the numerical calculation results of smooth surfaces with the Blok flash temperature empirical formula, the numerical trends are basically consistent. Although the Blok flash temperature empirical formula cannot take into account the influence of roughness factors, it can prove that the flash temperature characteristic analysis model under the friction dynamic coupling of gear transmission system is reliable in engineering applications.

[0171] This application also provides an application scenario in which the above-described gear scuffing fault diagnosis method under tribodynamic coupling conditions is applied. Specifically, the gear scuffing fault diagnosis method under tribodynamic coupling conditions provided in this embodiment can be applied in a gear fault diagnosis scenario. The gear fault diagnosis scenario includes a diagnosis stage and a display stage. The diagnosis stage is used to complete the gear scuffing fault diagnosis, and the display stage is used to show the gear scuffing fault diagnosis results to the user. The gear scuffing fault diagnosis method under tribodynamic coupling conditions provided in this embodiment belongs to the diagnosis stage.

[0172] Example 2

[0173] In one exemplary embodiment, a computer device is provided, which may be a server or a terminal, and its internal structure diagram may be as follows. Figure 8 As shown, this computer device includes a processor, memory, input / output (I / O) interfaces, and a communication interface. The processor, memory, and I / O interfaces are connected via a system bus, and the communication interface is also connected to the system bus via the I / O interfaces. The processor provides computational and control capabilities. The memory includes non-volatile storage media and internal memory. The non-volatile storage media stores the operating system, computer programs, and a database. The internal memory provides the environment for the operation of the operating system and computer programs stored in the non-volatile storage media. The database stores data. The I / O interfaces are used for exchanging information between the processor and external devices. The communication interface is used for communication with external terminals via a network connection. When executed by the processor, the computer program implements a method for diagnosing gear scuffing faults under tribodynamic coupling conditions.

[0174] Those skilled in the art will understand that Figure 8 The structure shown is merely a block diagram of a portion of the structure related to the present application and does not constitute a limitation on the computer device to which the present application is applied. Specific computer devices may include more or fewer components than those shown in the figure, or combine certain components, or have different component arrangements.

[0175] In one exemplary embodiment, a computer device is provided, including a memory and a processor, wherein the memory stores a computer program, and the processor executes the computer program to implement the gear scuffing fault diagnosis method under tribodynamic coupling conditions of Embodiment 1.

[0176] Example 3

[0177] In one exemplary embodiment, a computer-readable storage medium is provided storing a computer program that, when executed by a processor, implements the gear scuffing fault diagnosis method under tribodynamic coupling conditions of Embodiment 1.

[0178] Example 4

[0179] In one exemplary embodiment, a computer program product is provided, including a computer program that, when executed by a processor, implements the gear scuffing fault diagnosis method under tribodynamic coupling conditions of Embodiment 1.

[0180] It should be noted that the user information (including but not limited to user device information, user personal information, etc.) and data (including but not limited to data used for analysis, data stored, data displayed, etc.) involved in this application are all information and data authorized by the user or fully authorized by all parties, and the collection, use and processing of the relevant data must comply with relevant regulations.

[0181] The technical features of the above embodiments can be combined in any way. For the sake of brevity, not all possible combinations of the technical features in the above embodiments are described. However, as long as there is no contradiction in the combination of these technical features, they should be considered to be within the scope of this specification.

[0182] This document uses specific examples to illustrate the principles and implementation methods of this application. The descriptions of the above embodiments are only for the purpose of helping to understand the methods and core ideas of this application. Furthermore, those skilled in the art will recognize that, based on the ideas of this application, there will be changes in the specific implementation methods and application scope. Therefore, the content of this specification should not be construed as a limitation of this application.

Claims

1. A method for diagnosing gear scuffing faults under tribodynamic coupling conditions, characterized in that, The method for diagnosing gear scuffing faults under tribodynamic coupling conditions includes: Iterative calculations are performed based on the dynamic model of gear transmission under the condition of frictional dynamic coupling to obtain the fluid shear stress during gear transmission. Based on the fluid shear stress, the heat distribution coefficient of the contact surface of the driving gear and the heat distribution coefficient of the contact surface of the driven gear are calculated. The contact surface temperature of the driving gear is calculated based on the heat distribution coefficient of the contact surface of the driving gear, and the contact surface temperature of the driven gear is calculated based on the heat distribution coefficient of the contact surface of the driven gear. The adhesion safety factor of the driving gear is calculated based on the contact surface temperature of the driving gear, and the adhesion safety factor of the driven gear is calculated based on the contact surface temperature of the driven gear. Based on the scuffing safety factor of the driving gear and the scuffing safety factor of the driven gear, it is determined whether there is a gear scuffing fault during the transmission process of the driving gear and the driven gear, and the gear scuffing fault diagnosis is completed.

2. The method for diagnosing gear scuffing faults under tribodynamic coupling conditions according to claim 1, characterized in that, Iterative calculations are performed based on a dynamic model of gear transmission under frictional dynamic coupling conditions to obtain the fluid shear stress during gear transmission, specifically including: The initial fluid shear stress and initial dry shear stress are calculated based on the initial pressure, and the initial total friction force is calculated based on the initial fluid shear stress and the initial dry shear stress. Using the initial total friction force as input, the dynamic model of the gear transmission process under the condition of friction dynamic coupling is solved to obtain the rotation angle of the driving gear and the rotation angle of the driven gear. The updated pressure is calculated based on the rotation angles of the driving gear and the driven gear. Based on the updated pressure, the updated fluid shear stress and the updated dry shear stress are calculated. Based on the updated fluid shear stress and the updated dry shear stress, the updated total friction force is calculated. Based on the initial total friction force and the updated total friction force, determine whether to end the iteration; if yes, use the updated fluid shear stress as the fluid shear stress in the gear transmission process; if no, use the updated pressure, the updated fluid shear stress, the updated dry shear stress, and the updated total friction force as the initial pressure, initial fluid shear stress, initial dry shear stress, and initial total friction force for the next iteration, and return to the step of "using the initial total friction force as input to solve the dynamic model of the gear transmission process under the friction dynamic coupling condition, and obtain the rotation angle of the driving gear and the rotation angle of the driven gear".

3. The method for diagnosing gear scuffing faults under tribodynamic coupling conditions according to claim 2, characterized in that, The initial fluid shear stress and initial dry shear stress are calculated based on the initial pressure. The initial total friction force is then calculated based on these initial fluid shear stresses and the initial dry shear stresses. Specifically, this includes: using the initial pressure as input, calculating the initial fluid shear stress using the Bair-Winer viscoelastic non-Newtonian fluid model; calculating the initial dry shear stress based on the initial pressure; calculating the initial fluid oil film friction force based on the initial fluid shear stress; calculating the initial dry friction force based on the initial dry shear stress; and calculating the sum of the initial fluid oil film friction force and the initial dry friction force to obtain the initial total friction force. Based on the updated pressure, the updated fluid shear stress and updated dry shear stress are calculated. Based on the updated fluid shear stress and the updated dry shear stress, the updated total friction force is calculated, specifically including: using the updated pressure as input, calculating the updated fluid shear stress using the Bair-Winer viscoelastic non-Newtonian fluid model; calculating the updated dry shear stress based on the updated pressure; calculating the updated fluid oil film friction force based on the updated fluid shear stress; calculating the updated dry friction force based on the updated dry shear stress; and calculating the sum of the updated fluid oil film friction force and the updated dry friction force to obtain the updated total friction force.

4. The method for diagnosing gear scuffing faults under tribodynamic coupling conditions according to claim 2, characterized in that, The dynamic model is as follows: Where m1 is the mass of the driving gear; The x-direction translational acceleration of the driving gear; x is the stiffness coefficient of the driving gear in the x direction; x1 is the translational displacement of the driving gear in the x direction. is the damping coefficient in the x-direction of the driving gear; λ is the translational velocity of the driving gear in the x-direction. f f is the coefficient of frictional direction; 1i Let m1 be the first frictional force of the i-th gear; m2 is the mass of the driven gear. Let x be the translational acceleration of the driven gear in the x-direction; x is the stiffness coefficient of the driven gear in the x-direction; x2 is the translational displacement of the driven gear in the x-direction. The damping coefficient in the x-direction of the driven gear; f is the translational velocity of the driven gear in the x-direction. 2i The second frictional force of the i-th gear; The translational acceleration in the y-direction of the driving gear; y is the stiffness coefficient in the y direction of the driving gear; y1 is the translational displacement in the y direction of the driving gear. y is the damping coefficient of the driving gear; k is the translational velocity in the y-direction of the driving gear. t R is the gear stiffness coefficient. b1 R is the base circle radius of the driving gear; θ1 is the rotation angle of the driving gear; b2 θ2 is the base circle radius of the driven gear; θ2 is the rotation angle of the driven gear; y2 is the translational displacement of the driven gear in the y direction; c t This refers to the gear damping coefficient; The rotational angular velocity of the driving gear; The angular velocity of the driven gear; Let be the translational velocity in the y-direction of the driven gear; Let be the translational acceleration in the y-direction of the driven gear; y is the stiffness coefficient of the driven gear in the y direction; Iy is the damping coefficient of the driven gear in the y direction; I1 is the moment of inertia of the driving gear. k is the rotational angular acceleration of the driving gear. p θ is the torsional stiffness coefficient of the external shaft system of the driving gear. m c is the rotation angle of the motor; p The torsional damping coefficient of the external shaft system of the driving gear; H is the angular velocity of the motor. 1i Ii is the first frictional force arm of the i-th gear; I2 is the moment of inertia of the driven gear; k is the rotational angular acceleration of the driven gear. g θ is the torsional stiffness coefficient of the external shaft system of the driven gear. b c is the rotation angle of the load; g The torsional damping coefficient for the connection of the external shaft system of the driven gear; H is the rotational angular velocity of the load; 2i I is the second frictional force arm of the i-th gear; m The moment of inertia of the motor shaft system; M1 is the angular acceleration of the motor; M2 is the torque applied by the motor; I b The moment of inertia of the load shaft system; M1 is the rotational angular acceleration of the load; M2 is the torque applied by the load; H 11 (t) represents the first frictional force arm of the driving gear at time t; α A The development angle at point A during gear transmission is the initial engagement point; mod() represents the remainder operation; f m Z represents the frictional force of the fluid oil film; Z represents the number of teeth; α D H is the development angle at point D during gear transmission, where point D is the transition point from double-tooth meshing to single-tooth meshing; 12 (t) represents the first frictional force arm of the driven gear at time t; α B H is the development angle at point B during gear transmission, where point B is the transition point from single-tooth meshing to double-tooth meshing; 21 (t) represents the second frictional force arm of the driving gear at time t; β2 is the actual meshing angle; H 22 (t) represents the second frictional force arm of the driven gear at time t; The first frictional force of the i-th gear and the second frictional force of the i-th gear are both equal to the initial total frictional force.

5. The method for diagnosing gear scuffing faults under tribodynamic coupling conditions according to claim 2, characterized in that, Based on the rotation angles of the driving gear and the driven gear, the updated pressure is calculated, specifically including: Based on the rotation angles of the driving gear and the driven gear, the dynamic transmission error during the transmission process of the driving gear and the driven gear is calculated. Based on the aforementioned dynamic transmission error, the dynamic meshing force is calculated. Based on the dynamic meshing force, the updated pressure is calculated; The formula for calculating the dynamic transmission error is as follows: δ=R b1 θ1-R b2 θ2+y1-y2+e; Where δ represents the dynamic transmission error; R b1 R is the base circle radius of the driving gear; θ1 is the rotation angle of the driving gear; b2 θ1 is the base circle radius of the driven gear; θ2 is the rotation angle of the driven gear; y1 is the translational displacement of the driving gear in the y direction; y2 is the translational displacement of the driven gear in the y direction; e is the gear manufacturing error. The formula for calculating the dynamic meshing force is: F dp =k t d; Among them, F dp For dynamic meshing force; k t This is the gear stiffness coefficient.

6. The method for diagnosing gear scuffing faults under tribodynamic coupling conditions according to claim 1, characterized in that, The formulas for calculating the heat distribution coefficient of the contact surface of the driving gear and the heat distribution coefficient of the contact surface of the driven gear are as follows: Where q(x,y) is the heat distribution coefficient at (x,y); τ(x,y) is the fluid shear stress at (x,y); U2 is the tangential velocity of the driven gear; U1 is the tangential velocity of the driving gear; q A (x,y) is the heat distribution coefficient of the contact surface of the driving gear at (x,y); K f The thermal conductivity of the lubricating oil; The transient temperature of the driven gear; q represents the transient temperature of the driving gear. B (x,y) is the heat distribution coefficient of the contact surface of the driven gear at (x,y); The formulas for calculating the contact surface temperature of the driving gear and the contact surface temperature of the driven gear are as follows: Where T1(ξ) is the contact surface temperature of the driving gear, and ξ is the position of the calculation point; T b1 ρ is the initial surface temperature of the driving gear; ρ1 is the density of the driving gear; C1 is the specific heat capacity of the driving gear; k1 is the thermal conductivity of the driving gear; x is the initial position of the contact area; k f λ is the thermal conductivity; h is the equivalent thickness of the gear contact surface; T2(λ) is the instantaneous contact temperature of the driven gear, λ is the integral variable; T1(λ) is the instantaneous contact temperature of the driving gear; q A (λ) is the heat distribution coefficient of the contact surface of the driving gear; T2(ξ) is the contact surface temperature of the driven gear; T b2 ρ is the initial surface temperature of the driven gear; ρ2 is the density of the driven gear; C2 is the specific heat capacity of the driven gear; k2 is the thermal conductivity of the driven gear; q B (λ) is the heat distribution coefficient of the contact surface of the driven gear.

7. The method for diagnosing gear scuffing faults under tribodynamic coupling conditions according to claim 1, characterized in that, The formula for calculating the bonding safety factor is: Among them, S sint θ is the safety factor for bonding. sint The limiting temperature; θ sin To calculate the temperature, when the calculated temperature is the contact surface temperature of the driving gear, the scuffing safety factor of the driving gear is calculated; when the calculated temperature is the contact surface temperature of the driven gear, the scuffing safety factor of the driven gear is calculated. Based on the scuffing safety factors of the driving gear and the driven gear, the system determines whether scuffing faults exist during the transmission process of the driving and driven gears, thus completing the gear scuffing fault diagnosis. Specifically, this includes: if the scuffing safety factor of the driving gear is less than a first preset value, or the scuffing safety factor of the driven gear is less than a first preset value, then a high scuffing safety risk is determined to exist during the transmission process of the driving and driven gears; if the scuffing safety factor of the driving gear is greater than or equal to the first preset value and less than or equal to a second preset value, or the scuffing safety factor of the driven gear is greater than or equal to the first preset value and less than or equal to the second preset value, then a medium scuffing safety risk is determined to exist during the transmission process of the driving and driven gears; if the scuffing safety factor of the driving gear is greater than the second preset value, or the scuffing safety factor of the driven gear is greater than the second preset value, then a low scuffing safety risk is determined to exist during the transmission process of the driving and driven gears.

8. A computer device, comprising: A memory, a processor, and a computer program stored in the memory and capable of running on the processor, characterized in that the processor executes the computer program to implement the gear scuffing fault diagnosis method under tribodynamic coupling conditions as described in any one of claims 1-7.

9. A computer-readable storage medium having a computer program stored thereon, characterized in that, When executed by a processor, the computer program implements the gear scuffing fault diagnosis method under the tribodynamic coupling conditions described in any one of claims 1-7.

10. A computer program product, comprising a computer program, characterized in that, When executed by a processor, the computer program implements the gear scuffing fault diagnosis method under the tribodynamic coupling conditions described in any one of claims 1-7.

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