Bevel gear pair friction-flash temperature characteristic calculation method under real machining surface condition

Abstract

The invention relates to the field of bevel gear pair friction-flash temperature characteristic calculation, and particularly discloses a bevel gear pair friction-flash temperature characteristic calculation method under real machining surface conditions. Comprising the following steps: calculating oil film shear stress, shear modulus and tooth surface micro-bulge contact shear stress in a helical gear meshing process by adopting a non-Newtonian fluid model and a micro-bulge contact model; summing the shearing stress of the oil film and the shearing stress of the micro-bulge, and performing integration to solve the friction force of the contact domain in the meshing process of the bevel gear and the friction coefficient in the single-tooth meshing period of the bevel gear; coupling oil film shear stress, micro-convex body contact shear stress and relative slippage of two interfaces, and solving heat generated by a tooth surface contact domain in the meshing process through a heat distribution model; substituting the obtained heat into a second kind of Volterra integral equation based on a fast moving heat source model, and calculating the interface flash temperature by adopting an iteration method; and the obtained interface flash temperature is reversely brought back to the non-Newtonian fluid model, and repeated iteration is carried out until convergence.

Description

Technical Field

[0001] The invention relates to the field of calculation of friction-flash temperature characteristics of helical gear pairs, and in particular to a calculation method of friction-flash temperature characteristics of helical gear pairs under real machined surface conditions. Background Art

[0002] With the continuous improvement of the power density of marine diesel engines, the complex and harsh working conditions make the working state of the internal timing helical gears extremely harsh. During the working process of the gear, the relative sliding between the two tooth surfaces and the mutual contact of the micro-convex bodies on the surface will cause friction between the two tooth surfaces, and then generate a large amount of heat at the moment of meshing, causing the temperature of the local position in the meshing domain to rise. This phenomenon is called tooth surface flash temperature. Excessive tooth surface flash temperature can cause the gear to have a bonding failure. The instantaneous temperature rise in the gear contact area is a factor that cannot be ignored. Therefore, it is of great significance to study the distribution of flash temperature during the working process of the gear. Based on the one-dimensional simplified equation for calculating the heat flow, the steady-state instantaneous temperature solution formula and the gear instantaneous temperature rise formula are derived, which provides a theoretical basis for subsequent gear flash temperature research.

[0003] Some existing scholars have established a friction flash temperature analysis model for helical gears based on the theory of mixed elastohydrodynamic lubrication of infinite line contact, analyzed the mixed lubrication state and friction flash temperature distribution of the central meshing area under the real machined surface, and explored the distribution law of tooth surface flash temperature along the meshing line and the influence of friction factor on tooth surface flash temperature. Until another group of scholars carried out contact interface friction-flash temperature calculation based on the finite element model of gears. There are also methods that combine finite element analysis method with theoretical analysis and calculation, and use the test measurement results of gear temperature as a reference to establish a calculation method and analysis model for the instantaneous contact temperature of the tooth surface of high-speed transmission gears, and systematically analyze the distribution of the gear body temperature and tooth surface flash temperature of standard involute and tooth top modified involute gears; others combine thermal analysis finite element method with mixed elastohydrodynamic lubrication model, propose a direct calculation method for the tooth surface flash temperature of spiral bevel gears under mixed lubrication state, calculate the friction flash temperature of bevel gear tooth surface under real three-dimensional machined surface, and simultaneously analyze the distribution law of tooth surface flash temperature under different lubrication states.

[0004] However, since ship helical gears have been serving in harsh transient working conditions such as heavy load, variable load, and speed change for a long time, the tooth surface is engaged in a non-steady-state mixed lubrication environment, and the local rough peak contact between the two interfaces and the relative sliding between the interfaces further cause local dry friction and flash temperature on the tooth surface, and even cause high-temperature bonding failure on the helical gear tooth surface. Summary of the invention

[0005] The purpose of the present invention is to provide a method for calculating the friction-flash temperature characteristics of a helical gear pair under real machined surface conditions, and to solve the following technical problems:

[0006] Ship helical gears have been serving in harsh transient working conditions such as heavy load, variable load, and speed change for a long time, causing the tooth surface to mesh in a non-steady-state mixed lubrication environment. The local rough peak contact between the two interfaces and the relative sliding between the interfaces further lead to local dry friction and flash temperature on the tooth surface, and even cause technical problems such as high-temperature bonding failure of the helical gear tooth surface.

[0007] The purpose of the present invention can be achieved through the following technical solutions:

[0008] The calculation method of friction-flash temperature characteristics of helical gear pairs under real machined surface conditions includes:

[0009] S1. Preliminarily establish a mixed lubrication state analysis model for a finite-length line contact of a marine helical gear pair based on a real machined surface, and use key mixed lubrication characteristic parameters and transient operating condition parameters as input to solve the friction coefficient and contact flash temperature during the meshing process of the helical gears; the key mixed lubrication characteristic parameters include oil film pressure and oil film thickness; the transient operating condition parameters include entrainment velocity and sliding-rolling ratio;

[0010] S2. Use the non-Newtonian fluid model and the micro-asperity contact model to calculate the oil film shear stress, shear modulus and the micro-asperity contact shear stress of the tooth surface during the meshing process of the helical gears; sum the oil film shear stress and the micro-asperity shear stress, and integrate them to solve the friction force in the contact domain of the helical gear meshing process and the friction coefficient of the helical gear single tooth meshing cycle;

[0011] S3, coupling the oil film shear stress, the asperity contact shear stress and the relative slip between the two interfaces, and solving the heat generated in the tooth surface contact area during the meshing process through the heat distribution model;

[0012] S4, based on the fast moving heat source model, the heat is introduced into the second kind of Volterra integral equation, and the interface flash temperature is calculated by the iterative method;

[0013] S5. Bring the obtained interface flash temperature back to the non-Newtonian fluid model and iterate repeatedly until convergence.

[0014] Preferably, in step S2, the linear equation for the oil film shear stress in the contact domain is solved by the non-Newtonian fluid model as follows:

[0015]

[0016] Among them, τ is the oil film shear stress, τ L is the ultimate shear stress, G ∞ is the limiting shear modulus, h is the oil film thickness in the contact domain of the helical gear pair during the meshing process; the shear stress distribution in the entire solution domain can be obtained by solving the nonlinear equation at each node.

[0017] Preferably, step S2 predicts the ultimate shear stress and ultimate shear modulus of the oil film in the contact domain of the helical gear by using a micro-convex contact model:

[0018]

[0019] τ L (p,T)=0.25G ∞ ,h / σ≤1.0

[0020] τ L (p,T)=0.25G ∞ [1-0.2(h / σ-1)],h / σ>1.0

[0021] Among them, T is the interface flash temperature, σ is the comprehensive roughness of the two surfaces, and p is the oil film pressure in the contact area of ​​the helical gear pair during the meshing process.

[0022] Preferably, in step S2, after obtaining the oil film shear stress in the contact area, the oil film transient friction force of the helical gear contact pair is obtained by integrating the oil film shear stress, and then the oil film friction coefficient of the helical gear single tooth meshing period is obtained by ratioing the oil film shear stress with each meshing transient contact load:

[0023]

[0024] Dry friction coefficient F contact ∈[0.07, 0.15], the shear stress under this dry friction coefficient is solved as follows:

[0025] τ contact =f contact p

[0026] Among them, P is the oil film pressure in the contact area of ​​the helical gear pair at each meshing moment.

[0027] Preferably, in step S3:

[0028] To judge the relative slip degree of the micro-convex bodies on the two interfaces, the sliding-rolling ratio of the two contact interfaces can be known:

[0029]

[0030] Preferably, in step S3, based on the heat distribution model, the generated heat is divided into two parts according to the heat distribution coefficient, and the heat distribution coefficient between the two contact surfaces of the gear pair is determined according to the local lubrication velocity distribution and the temperature change curve:

[0031]

[0032] Among them, K f is the thermal conductivity of lubricating oil.

[0033] Preferably, the second kind of Volterra integral equation is solved using an iterative method:

[0034]

[0035]

[0036] Among them, T1 and T2 are the contact surface temperatures of the two helical gears, Tb1 and Tb2 are the initial surface temperatures, ρ1 and ρ2 are solid densities, c1 and c2 are solid specific heat capacities, k1 and k2 are solid thermal conductivity coefficients, and q(λ) is the heat generated by friction.

[0037] Preferably, the method further includes evaluating the degree of friction flash temperature failure during the meshing process of the marine helical gear pair, and calculating the limit integral temperature when the tooth surface and tooth direction glue failure occurs:

[0038] θ sint =(0.85+1.4X w )·(ρν oil ) -0.05 ·S FZG 2

[0039] Among them, X w is the material welding coefficient, v oil is the kinematic viscosity of the lubricating oil, S FZG It is the bonding load grade of oil FZG.

[0040] Beneficial effects of the present invention:

[0041] (1) The maximum flash temperature during the meshing period of a single tooth of a helical gear on a smooth surface was compared with the BLOK flash temperature formula. The results show that the flash temperature analysis method proposed in this paper is consistent with the data obtained by the empirical formula, with a maximum error of 15.83%, which proves the accuracy and feasibility of the numerical model.

[0042] (2) The effects of three different real machined surfaces on the friction coefficient and flash temperature distribution during the meshing cycle of a single tooth of a helical gear were studied. The results show that the maximum surface flash temperature of the shaved surface at the moment of partial meshing far exceeds the bonding limit temperature, and tooth surface bonding failure is very likely to occur; the overall friction coefficient and surface flash temperature of the polished surface are the smallest. In order to reduce the occurrence of pitting and bonding failure during the meshing process of the helical gear, the polished surface should be selected as much as possible.

[0043] (3) The influence of different RMS roughness of the same machined surface and different machined surface textures under the same roughness on the lubrication state and friction-flash temperature during the meshing process of helical gears was discussed. With the increase of the RMS roughness of the surface, the fluctuation amplitude of the central oil film pressure and the oil film thickness increased significantly, the friction coefficient fluctuated more violently, and the maximum flash temperature between the contact interfaces increased. Due to the influence of the roughness peak distribution characteristics of the machined surface texture, the friction coefficient and maximum flash temperature of the honing texture are still large under the same roughness, while the friction coefficient and maximum flash temperature of the polishing texture are still small. It can be seen that the surface processing texture determines the distribution of the flash temperature and friction coefficient within the contact surface, and the size of the RMS roughness determines the fluctuation amplitude under this distribution.

[0044] (4) The influence of speed conditions on the friction coefficient and surface temperature rise of the two tooth surfaces of the helical gear was analyzed. The results show that as the speed of the helical gear increases, the average oil film thickness increases, the average oil film pressure decreases, the lubrication state tends to be good, and the helical gear completes the transition from mixed lubrication to full film lubrication. As the speed increases, the friction coefficient between the teeth and the flash temperature distribution of the two tooth surfaces decrease and then increase.

[0045] Of course, any product implementing the present invention does not necessarily need to achieve all of the advantages described above at the same time. BRIEF DESCRIPTION OF THE DRAWINGS

[0046] In order to more clearly illustrate the technical solutions of the embodiments of the present invention, the accompanying drawings required for describing the embodiments will be briefly introduced below. Obviously, the accompanying drawings described below are only some embodiments of the present invention. For ordinary technicians in this field, other accompanying drawings can be obtained based on these accompanying drawings without paying creative work.

[0047] Figure 1 It is a schematic diagram of the friction-flash temperature calculation flow of the helical gear pair of the present invention;

[0048] Figure 2 This is a step diagram of the helical gear pair friction-flash temperature calculation method of the present invention;

[0049] Figure 3 It is a schematic diagram of the change of the friction coefficient of the contact area of ​​the helical gear pair under different real processing surfaces of the present invention;

[0050] Figure 4 It is a schematic diagram of the maximum flash temperature variation of the two contact surfaces of the helical gear pair under different real processing surfaces of the present invention;

[0051] Figure 5 Schematic diagram of flash temperature distribution of the helical gear driving wheel at time A under different real processing surfaces of the present invention. DETAILED DESCRIPTION

[0052] The following will be combined with the drawings in the embodiments of the present invention to clearly and completely describe the technical solutions in the embodiments of the present invention. Obviously, the described embodiments are only part of the embodiments of the present invention, not all of the embodiments. Based on the embodiments of the present invention, all other embodiments obtained by ordinary technicians in this field without creative work are within the scope of protection of the present invention.

[0053] See also Figure 1 As shown, the present invention is a method for calculating the friction-flash temperature characteristics of a helical gear pair under real machined surface conditions, and the method includes:

[0054] S1. Preliminarily establish a mixed lubrication state analysis model for a finite-length line contact of a marine helical gear pair based on a real machined surface, and use key mixed lubrication characteristic parameters and transient operating condition parameters as input to solve the friction coefficient and contact flash temperature during the meshing process of the helical gears; the key mixed lubrication characteristic parameters include oil film pressure and oil film thickness; the transient operating condition parameters include entrainment velocity and sliding-rolling ratio;

[0055] S2. Use the non-Newtonian fluid model and the micro-asperity contact model to calculate the oil film shear stress, shear modulus and the micro-asperity contact shear stress of the tooth surface during the meshing process of the helical gears; sum the oil film shear stress and the micro-asperity shear stress, and integrate them to solve the friction force in the contact domain of the helical gear meshing process and the friction coefficient of the helical gear single tooth meshing cycle;

[0056] S3, coupling the oil film shear stress, the asperity contact shear stress and the relative slip between the two interfaces, and solving the heat generated in the tooth surface contact area during the meshing process through the heat distribution model;

[0057] S4, based on the fast moving heat source model, the heat is introduced into the second kind of Volterra integral equation, and the interface flash temperature is calculated by the iterative method;

[0058] S5. Bring the obtained interface flash temperature back to the non-Newtonian fluid model and iterate repeatedly until convergence.

[0059] In the above technical scheme, based on the mixed lubrication state analysis model of the real machined surface marine helical gear pair, the key mixed lubrication characteristic parameters (oil film thickness, oil film pressure, etc.) and transient operating parameters (slide-roll ratio, entrainment speed, etc.) are used as input to solve the friction coefficient and contact flash temperature during the meshing process of the helical gear. First, considering the non-Newtonian fluid effect of the lubricating oil film during the meshing process of the helical gear, the Bai r-Wi ner rheological model (non-Newtonian fluid model) and the modified Dowson model (micro-convex contact model) are used to calculate the oil film shear stress and shear modulus during the meshing process. In addition, considering the influence of the real machined surface roughness, the contact shear stress of the micro-convex body on the tooth surface during the meshing process is solved at the same time, the oil film shear stress and the micro-convex body shear stress are summed, and the friction force in the contact domain of the helical gear meshing process is integrated. Then, the friction coefficient of the helical gear within the meshing cycle can be obtained by ratioing it with the contact domain load. At the same time, the oil film shear stress, the micro-convex contact shear stress and the relative slip of the two interfaces are coupled to solve the heat generated in the tooth surface contact domain during the meshing process. Then, based on the theory of rapidly moving heat sources, the obtained heat is brought into the second-kind Volterra integral equation, and the interface flash temperature is calculated using an iterative method. Finally, the obtained interface flash temperature is brought back to the non-Newtonian fluid model, and it is iterated repeatedly until convergence. The specific process is as follows: Figure 1-2 shown.

[0060] As an embodiment of the present invention, the friction of the helical gear tooth surface is mainly composed of two parts: the oil film friction force generated by the oil film shear effect and the dry friction caused by the rough peak contact between the two contact surfaces. When the oil film is in high-speed and high-pressure conditions, the shear stress no longer changes linearly with the change of the shear rate. At this time, it is necessary to consider the non-Newtonian fluid effect of the lubricating oil and establish a non-Newtonian fluid model to calculate the shear stress during the meshing of the helical gears. The Bair-Winer rheological model (non-Newtonian fluid model) is used to solve the oil film shear stress in the contact domain:

[0061]

[0062] Among them, τ is the oil film shear stress, τ L is the ultimate shear stress, G ∞ is the limiting shear modulus, h is the oil film thickness in the contact domain of the helical gear pair during the meshing process; the shear stress distribution in the entire solution domain can be obtained by solving the nonlinear equation at each node.

[0063] As an implementation of the present invention, considering that the lubricating oil used in the calculation is a typical mineral oil, the existing modified Dowson model is used to predict the ultimate shear stress and ultimate shear modulus of the oil film in the contact domain of the helical gear. Specifically:

[0064]

[0065] τ L (p,T)=0.25G ∞ ,h / σ≤1.0 (3)

[0067] τ L (p,T)=0.25G ∞ [1-0.2(h / σ-1)],h / σ>1.0

[0068] Among them, T is the interface flash temperature, σ is the comprehensive roughness of the two tables, and p is the oil film pressure in the contact area of ​​the helical gear pair during the meshing process. From the calculation formula, it can be seen that considering the non-Newtonian fluid effect of the lubricating oil during the meshing process, the lubricating oil performance parameters are functions of pressure and temperature.

[0069] As an implementation mode of the present invention, after obtaining the oil film shear stress in the contact area, the oil film transient friction force of the helical gear contact pair is obtained by integrating it, and then the oil film friction coefficient of the helical gear single tooth meshing period is obtained by ratioing it with each meshing transient contact load:

[0070]

[0071] The boundary lubrication friction coefficient generated by the contact of the tooth surface roughness peak is relatively easy to measure. In engineering practice, the dry friction coefficient generally fluctuates within a smaller range. Existing technology shows that the dry friction coefficient F contact It fluctuates within the range of 0.07-0.15. The material contact of the specific gear can be simulated on the friction and wear tester to measure its dry friction coefficient. The literature is comprehensively considered to make it closer to the actual working conditions of the helical gear contact. This paper selects the dry friction coefficient as 0.14, and the shear stress under this dry friction coefficient is solved as follows:

[0072] τ contact =f contact p (5)

[0073] Among them, P is the oil film pressure in the contact area of ​​the helical gear pair at each meshing moment.

[0074] As an embodiment of the present invention, since the limiting shear stress of the lubricating oil is a function of temperature, the interface friction is obviously affected by the surface temperature, and the friction generates heat, which is then reflected in the temperature. The temperature will affect the oil film shear modulus and react to the friction. In the helical gear pair meshing system, tooth surface friction and heat are interdependent. In the helical gear meshing process considering the real machined surface, heat is generated by the viscous shear of the lubricating oil film and the sliding contact of the micro-protrusions. The generated heat is carried away by the conduction of the lubricating oil or the two contact surfaces. In order to judge the relative slip degree of the micro-protrusions on the two interfaces, it can be known from the slide-roll ratio of the two contact interfaces:

[0075]

[0076] Furthermore, based on the existing heat distribution model of Francis and Plint, 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:

[0077]

[0078] Among them, K f is the thermal conductivity of lubricating oil.

[0079] As an embodiment of the present invention, the calculation of flash temperature is based on the theory of moving heat source on a semi-infinite body. Based on this theory, a transient temperature calculation model of the contact surface of a marine helical gear pair is established, and the second kind Volterra integral equation is solved using an iterative method:

[0080]

[0081] Among them, T1 and T2 are the contact surface temperatures of the two helical gears, Tb1 and Tb2 are the initial surface temperatures, ρ1 and ρ2 are solid densities, c1 and c2 are solid specific heat capacities, k1 and k2 are solid thermal conductivity coefficients, and g(λ) is the heat generated by friction.

[0082] As an embodiment of the present invention, it also includes evaluating the degree of friction flash temperature failure during the meshing process of the marine helical gear pair, and calculating the limit integral temperature when the tooth surface and tooth direction glue failure occurs:

[0083] θ sint =(0.85+1.4X w )·(ρv oil ) -0.05 ·S FZG 2 (10)

[0084] Among them, X w is the material welding coefficient, v oil is the kinematic viscosity of the lubricating oil, S FZG It is the bonding load grade of oil FZG.

[0085] Based on the mixed lubrication friction flash temperature model of three-dimensional line contact helical gear pairs, the iterative method is used to solve the time-varying friction coefficient and interface flash temperature of the contact area. Due to the consideration of the real machined surface roughness in the solution process, the lubrication state between the contact pairs will show strong time-varying characteristics. In order to improve the calculation accuracy, this paper takes the derivative of the rough morphology matrix at the highest level grid, and then converts the processed results to the Reynolds solution domain grid by interpolation method for solution. The calculation domain range used in this paper is -2.5a≤x≤1.5a, -1.1b≤y≤1.1b, 0≤z≤1.5a. The solution domain grid is 256×128×32, and the convergence criterion is that the relative error of the temperature before and after the iteration is less than 10-4. Numerical iteration is performed at each instant, and the maximum number of iterations is 12000.

[0086] The experimental results of the above-mentioned real machined surface helical gear pair friction flash temperature calculation process are analyzed as follows: the influence of friction coefficient and tooth surface temperature rise under different real machined surfaces, specifically:

[0087] The friction coefficient and flash temperature distribution of three helical gear pairs with different machined surfaces within a single tooth meshing cycle are analyzed in detail under the same working conditions. Figure 3 The variation of friction coefficient in the contact area of ​​the tooth surface at each meshing instant during the operation of the helical gear under three different real machined surfaces is given. Figure 3 It can be seen that the distribution trend of the friction coefficient under different real machined surfaces is similar. The friction coefficient fluctuates significantly during the single tooth meshing cycle, and the friction coefficient suddenly changes to zero at the node position. This is because the sliding-rolling ratio at the node position is extremely small, and the two contact surfaces of the helical gear pair hardly slide relative to each other, which is a pure rolling process, resulting in a sharp decrease in the oil film shear stress and oil film friction at this position, and a corresponding decrease in the friction coefficient. In addition, the values ​​of the friction coefficient of different real machined surfaces during the single tooth meshing cycle are different. The overall friction coefficient of the shaved surface is the largest, with a maximum value of 0.067; the honing surface is second, with a maximum friction coefficient of 0.053 during the meshing cycle; the overall friction coefficient of the polished surface is the smallest, with a maximum friction coefficient of 0.033 during the meshing cycle.

[0088] In addition, Table 1 shows the average friction coefficients during the meshing cycle under different real machining surfaces. It can be seen from the table that the average friction coefficient of the honing surface is 1.54 times that of the polishing surface; the average friction coefficient of the shaved surface is 1.95 times that of the polishing surface. The overall friction coefficient of the polished surface is small during the meshing process, and the wear failure of the helical gear tooth surface is lower than that of other machined surfaces.

[0089] Table 1 Average friction coefficient of helical gears during meshing cycle under different real machining surfaces

[0090]

[0091] Figure 4 The maximum temperature rise variation of the contact area of ​​the two contact surfaces of the helical gears during the meshing cycle under different real machined surface roughness is given. Figure 4 It can be seen that the flash temperature distribution trend of the helical gear during operation is similar to that of the smooth surface, and the overall distribution is "V" shaped. The overall temperature rise of the shaved surface is the largest, and the maximum temperature rise can reach 206.86℃; the honing surface is second; the overall flash temperature of the polished surface is the smallest. In addition, it can be calculated by formula (4-10) that the gluing limit temperature of the helical gear under this working condition is 127.68℃, and the result is compared with the maximum temperature rise of each machined surface. For the shaved surface, due to the large roughness of the two surfaces, the film thickness in the contact area is relatively small, the contact area ratio and contact load are relatively large, and dry contact is very likely to occur, which will cause the oil film to generate excessive friction heat, resulting in an increase in the flash temperature of the tooth surface. At some meshing moments, the flash temperature of the tooth surface far exceeds the gear gluing limit temperature, bringing about the risk of gluing failure.

[0092] In addition, Table 2 shows the average values ​​of the maximum flash temperature during the meshing cycle of a single tooth of the helical gear under different machined surfaces. The average value of the maximum flash temperature during the meshing cycle of the polished surface is the smallest, the average value of the maximum flash temperature of the honing surface is 1.76 times that of the polished surface, and the average value of the maximum flash temperature of the shaved surface is 3.53 times that of the polished surface. The overall flash temperature distribution during the meshing process of the shaved surface is large, and the tooth surface bonding failure is very likely to occur.

[0093] Table 2 Average maximum temperature rise during the meshing cycle of helical gears under different real machining surfaces

[0094]

[0095] In order to more intuitively compare the temperature rise changes in the contact area of ​​the helical gears under the three real machined surfaces, Figure 5 The flash temperature distribution in the instantaneous contact domain at the characteristic meshing moment A is further given. Due to the typical characteristics of the roughness size and distribution of each machined surface and the texture of the machined surface, the flash temperature distribution under different real machined surfaces is different. For the shaved surface, due to the strong randomness of the roughness peak distribution and the roughness size under its surface processing texture, its maximum flash temperature appears at a certain position in the contact domain, and its value is 142.5℃, and the overall flash temperature is about 71.25℃. The overall flash temperature distribution of the polished surface is the smallest, and the overall flash temperature distribution at this meshing moment is about 18.5℃. Therefore, in order to reduce the occurrence of wear and bonding failures during the meshing process of helical gears, polishing is a better choice.

[0096] The above contents are merely examples and explanations of the concept of the present invention. The technicians in this technical field may make various modifications or additions to the specific embodiments described or replace them in a similar manner. As long as they do not deviate from the concept of the invention or exceed the scope defined by the claims, they should all fall within the protection scope of the present invention.

Claims

1. A method for calculating the friction-flash temperature characteristics of helical gear pairs under real machined surface conditions, characterized in that: The method comprises: S1. Preliminarily establish a mixed lubrication state analysis model for a finite-length line contact of a marine helical gear pair based on a real machined surface, and use key mixed lubrication characteristic parameters and transient operating condition parameters as input to solve the friction coefficient and contact flash temperature during the meshing process of the helical gears; the key mixed lubrication characteristic parameters include oil film pressure and oil film thickness; the transient operating condition parameters include entrainment velocity and sliding-rolling ratio; S2. Use the non-Newtonian fluid model and the micro-asperity contact model to calculate the oil film shear stress, shear modulus and the micro-asperity contact shear stress of the tooth surface during the meshing process of the helical gears; sum the oil film shear stress and the micro-asperity shear stress, and integrate them to solve the friction force in the contact domain of the helical gear meshing process and the friction coefficient of the helical gear single tooth meshing cycle; S3, coupling the oil film shear stress, the asperity contact shear stress and the relative slip between the two interfaces, and solving the heat generated in the tooth surface contact area during the meshing process through the heat distribution model; S4, based on the fast moving heat source model, the heat is introduced into the second kind of Volterra integral equation, and the interface flash temperature is calculated by the iterative method; S5. Bring the obtained interface flash temperature back to the non-Newtonian fluid model and iterate repeatedly until convergence.

2. The method for calculating the friction-flash temperature characteristics of helical gear pairs under real machined surface conditions according to claim 1 is characterized in that: In step S2, the linear equation for solving the oil film shear stress in the contact domain by using the non-Newtonian fluid model is: Among them, τ is the oil film shear stress, τ L is the ultimate shear stress, G ∞ is the limiting shear modulus, h is the oil film thickness in the contact domain of the helical gear pair during the meshing process; the shear stress distribution in the entire solution domain can be obtained by solving the nonlinear equation at each node.

3. The method for calculating the friction-flash temperature characteristics of helical gear pairs under real machined surface conditions according to claim 2 is characterized in that: The step S2 predicts the ultimate shear stress and ultimate shear modulus of the oil film in the contact domain of the helical gear by using the asperity contact model: t L (p,T)=0.25G ∞ ,h / σ≤1.0 t L (p,T)=0.25G ∞ [1-0.2(h / s-1)],h / s>1.0 Among them, T is the interface flash temperature, σ is the comprehensive roughness of the two surfaces, and p is the oil film pressure in the contact area of ​​the helical gear pair during the meshing process.

4. The method for calculating the friction-flash temperature characteristics of helical gear pairs under real machined surface conditions according to claim 1, characterized in that: In step S2, after obtaining the oil film shear stress in the contact area, the oil film transient friction force of the helical gear contact pair is obtained by integrating the oil film shear stress, and then the oil film friction coefficient of the helical gear single tooth meshing period is obtained by ratioing the oil film shear stress with each meshing transient contact load: Dry friction coefficient F contact ∈[0.07, 0.15], the shear stress under this dry friction coefficient is solved as follows: t contact =f contact p Among them, P is the oil film pressure in the contact area of ​​the helical gear pair at each meshing moment.

5. The method for calculating the friction-flash temperature characteristics of helical gear pairs under real machined surface conditions according to claim 1, characterized in that: In step S3: To judge the relative slip degree of the micro-convex bodies on the two interfaces, the sliding-rolling ratio of the two contact interfaces can be known:

6. The method for calculating the friction-flash temperature characteristics of helical gear pairs under real machined surface conditions according to claim 2, characterized in that: In step S3, based on the heat distribution model, the generated heat is divided into two parts according to the heat distribution coefficient, and the heat distribution coefficient between the two contact surfaces of the gear pair is determined according to the local lubrication velocity distribution and the temperature change curve: Among them, K f is the thermal conductivity of lubricating oil.

7. The method for calculating the friction-flash temperature characteristics of helical gear pairs under real machined surface conditions according to claim 6, characterized in that: The second kind of Volterra integral equation is solved by iterative method: Among them, T1 and T2 are the contact surface temperatures of the two helical gears, Tb1 and Tb2 are the initial surface temperatures, ρ1 and ρ2 are solid densities, c1 and c2 are solid specific heat capacities, k1 and k2 are solid thermal conductivity coefficients, and q(λ) is the heat generated by friction.

8. The method for calculating the friction-flash temperature characteristics of helical gear pairs under real machined surface conditions according to claim 1, characterized in that: It also includes the evaluation of the degree of friction flash failure during the meshing process of marine helical gear pairs, and the calculation of the limiting integral temperature when the tooth surface and tooth direction are glued: i sint =(0.85+1.4X w )·(ρν oil ) -0.05 ·S FZG 2 Among them, X w is the material welding coefficient, v oil is the kinematic viscosity of the lubricating oil, S FZG It is the bonding load grade of oil FZG.

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