Design method of planetary reducer heat dissipation system

Abstract

The present application relates to a kind of design methods of planetary reducer heat dissipation system, comprising: determining the input parameter of planetary reducer and heat dissipation system design;Estimate the total thermal power of planetary reducer;Determine the design scheme of heat dissipation system;The design scheme of heat dissipation system includes heat dissipation system heat dissipation capacity evaluation and design, heat dissipation system heat dissipation capacity evaluation and design include the evaluation of total heat dissipation area and the arrangement of heat dissipation pipe.The calculation of heat transfer coefficient in the calculation model of the total heat dissipation area of the method of the application uses the equivalent heat transfer coefficient of heat dissipation pipe, when calculating equivalent heat transfer coefficient, the surface heat transfer coefficient of heat dissipation pipe air side, sliding oil side is fully considered in combination with the length of heat dissipation pipe in oil pool and air cavity, then according to the total heat dissipation area, the arrangement of heat dissipation pipe is carried out, compared with the heat transfer coefficient of heat exchange pipe selected by experience in the traditional way, the calculation of total heat dissipation area is more practical and more accurate, can realize quickly and efficiently in the planetary reducer with limited space to carry out heat dissipation system optimization design.

Description

Technical Field

[0001] This invention relates to the field of planetary gearbox lubrication and cooling technology, and specifically to a design method for a planetary reducer heat dissipation system. Background Technology

[0002] Scraper conveyors are core transportation equipment in fully mechanized coal mining faces. With the development of ultra-high mining heights and ultra-long working faces, the operating power of scraper conveyors is gradually increasing, which significantly increases the power requirements for planetary reducers. The development of intelligent and permanent magnet frequency conversion technology has made permanent magnet integrated drive technology a new development direction. In the limited and compact installation environment of underground scraper conveyors, new challenges are also posed to the structural space of planetary reducers.

[0003] Based on this, with the increase in power, in order to ensure sufficient allowable thermal power, the lubrication and cooling of mining planetary gearboxes has evolved from air cooling to dedicated high-efficiency coolers, and water-cooled finned tube heat dissipation systems suitable for underground use are widely adopted. The general design process for the heat dissipation system involves a preliminary assessment of the planetary gearbox's heat generation, selection of the heat transfer coefficient based on experience, designing the system's dimensions and specifications, and finally verifying the system's model, specifications, and operating conditions through bench efficiency tests on the planetary gearbox. Conventional planetary gearboxes have ample space for oil storage and heat dissipation system placement, allowing for a high design margin. However, compact planetary gearboxes with integrated drives have extremely limited space. Traditional design methods make it very difficult to arrange the heat dissipation system within the limited space of the planetary gearbox, and the uneven heat dissipation effect of lubricating oil in different areas of the planetary gearbox makes it difficult to meet the requirements of sufficient heat dissipation efficiency within the limited space of the planetary gearbox.

[0004] Therefore, it is necessary to design a more detailed partitioned cooling system for the compact planetary gearbox to ensure its continuous, efficient, stable, and safe operation. Summary of the Invention

[0005] In view of this, the present invention provides a design method for a planetary gear reducer heat dissipation system, thereby solving or at least alleviating one or more of the above-mentioned problems and other problems existing in the prior art.

[0006] To achieve the aforementioned objectives, the technical solution adopted by the present invention is as follows:

[0007] A design method for a heat dissipation system of a planetary gear reducer, the design method comprising the following steps:

[0008] Step 1: Determine the input parameters for the planetary gearbox and cooling system design;

[0009] Step 2: Estimate the total thermal power of the planetary reducer based on the input parameters;

[0010] Step 3: Determine the design scheme of the heat dissipation system based on the total heat power;

[0011] The design scheme of the heat dissipation system includes the evaluation and design of the heat dissipation capacity of the heat dissipation system, which includes the evaluation of the total heat dissipation area and the arrangement of heat dissipation pipes.

[0012] The total heat dissipation area is calculated based on the total heat power, and the arrangement of the heat dissipation pipes is obtained by conversion based on the total heat dissipation area.

[0013] The calculation model for the total heat dissipation area is as follows:

[0014]

[0015] Among them, S g The total heat dissipation area of ​​the heat pipes, in meters (m²). 2 ;P Q Heat dissipation, measured in W; T omax Maximum permissible oil temperature, unit: K; T wi T wo These are the inlet and outlet temperatures of the cooling medium, respectively, in Kelvin (K).

[0016] The heat dissipation P Q The value is the total thermal power of the planetary reducer obtained in step two;

[0017]

[0018] Among them, h eff The equivalent surface heat transfer coefficient of the heat pipe is expressed in W / (m²). 2 .K); L gt The total length of a single heat pipe, in meters (m); L ga L is the length of a single heatsink tube within the air cavity, in meters (m). go The length of a single heat dissipation pipe submerged in the oil bath, in meters (m).

[0019]

[0020] Among them, h a1 The heat transfer coefficient of the air-side surface of the heat pipe, in W / (m²). 2 .K); λ a1 d is the thermal conductivity of air, in W / (mK); fi The inner diameter of the heat pipe is in meters (m); Re is the Reynolds number; Pr is the Prandtl number; c h1 n h1 m h1The coefficient is selected based on the correlation formula of the average heat transfer coefficient across the surface of the circular tube, according to the range of the Re value.

[0021]

[0022] Among them, h o1 The heat transfer coefficient of the lubricating oil side surface of the heat sink is expressed in W / (m²). 2 .K); λ o1 c is the thermal conductivity of lubricating oil, in W / (mK); h2 n h2 m h2 The coefficient is selected based on the correlation formula of the average heat transfer coefficient across the surface of the circular tube, according to the range of the Re value.

[0023] The required number of heat dissipation tubes is determined based on the total heat dissipation area and the selected specifications of the heat dissipation tubes, and then arranged in the corresponding planetary cavity.

[0024] Optionally, in step one, the input parameters include establishing a 3D model of the planetary reducer assembly structure;

[0025] The total heat power is the sum of the total power loss of a single-stage or multi-stage planetary gear train and the total power loss of the sealing element. When the total heat power is the sum of the total power loss of a multi-stage planetary gear train and the total power loss of the sealing element, the arrangement of the heat dissipation pipes is proportionally allocated according to the total power loss of the multi-stage planetary gear train.

[0026] If the internal cavity of the planetary reducer cannot accommodate the required number of heat dissipation pipes, return to step one to readjust the 3D model of the planetary reducer structure to ensure sufficient space for heat dissipation pipes and that the total heat dissipation area meets the requirements.

[0027] In this invention, "multi-level" in multi-level planetary gear train refers to a planetary gear train with two or more levels, such as a two-level planetary gear train, a three-level planetary gear train, or more levels.

[0028] Optionally, the heat dissipation tube is selected as a finned tube, and the design scheme of the heat dissipation system also includes the design of heat dissipation system structural parameters, including the material, diameter and fin specifications of the finned tube.

[0029] In the aforementioned design method for the planetary gear reducer heat dissipation system, the design method may optionally include a step of determining the planetary gear reducer lubricating oil cooling parameters based on the total thermal power;

[0030] The lubricating oil cooling parameters include the lubricating oil demand, the oil level in the sump, and the effective volume of the oil sump; wherein...

[0031] The required amount of lubricating oil was estimated using the following method:

[0032]

[0033] Among them, P Q Heat dissipation, in W; q vo ρ represents the flow rate of lubricating oil, in L / min. o This refers to the density of the lubricating oil, expressed in kg / m³. 3 c po ΔT is the isobaric heat capacity of lubricating oil, expressed in J / (kg·K); o Temperature rise of lubricating oil, in K;

[0034] The required amount of lubricating oil is calculated by converting the flow rate and circulation cooling time.

[0035] The principle for determining the oil level height is: the oil level height should not be lower than 1 / 3 of the inner diameter height of the mechanical end face seal, and should not exceed the lowest point of the inner diameter of the lip seal.

[0036] The effective volume of the oil tank is calculated based on the 3D model established in step one, according to different oil tank liquid levels.

[0037] Optionally, the selection of the circulating cooling time is as follows: for external circulation of lubricating oil, the circulating cooling time is selected as 4~5 minutes; for oil sump storage without external circulation, the circulating cooling time can be reduced.

[0038] When the effective lubricating oil volume is lower than the required lubricating oil volume, return to step one, readjust the 3D model of the planetary reducer structure, and increase the cavity space to ensure that the lubricating oil volume meets the requirements.

[0039] Optionally, the arrangement of the heat dissipation pipes is proportionally allocated based on the total power loss of the first-stage planetary gear train and the total power loss of the second-stage planetary gear train. If the required number of finned tubes cannot be arranged in the internal cavity of the planetary reducer, the process returns to step one to readjust the 3D model of the planetary reducer structure to ensure sufficient space for the heat dissipation pipes and that the effective heat dissipation area meets the requirements.

[0040] Optionally, the heat dissipation capacity assessment and design of the heat dissipation system also includes an assessment of the cooling medium requirement, which is estimated through the following methods:

[0041]

[0042] Where, q vw ρ represents the flow rate of the cooling medium, in L / min. w The density of the cooling medium is expressed in kg / m³. 3 c pw ΔT is the isobaric heat capacity of the cooling medium, expressed in J / (kg·K);w The temperature rise of the cooling medium is expressed in Kelvin (K).

[0043] In the aforementioned design method for the planetary gearbox heat dissipation system, optionally, in step one, the input parameters include planetary gearbox performance parameters, planetary gearbox structural parameters, and planetary gearbox cooling parameters. The planetary gearbox structural parameters include structural parameters of each stage of the planetary gear train, bearing structural parameters, and sealing element structural parameters.

[0044] The performance parameters of the planetary reducer include input power and speed;

[0045] The structural parameters of each level of the planetary gear train include gear train type, number of planetary gears, gear tooth profile parameters and working center distance, wherein the gear tooth profile parameters include number of teeth, module, pressure angle, helix angle, tooth width, addendum coefficient, clearance coefficient and displacement coefficient;

[0046] The bearing structural parameters include bearing type, quantity, inner diameter, and outer diameter;

[0047] The structural parameters of the sealing element include sealing type, material, sealing diameter, and rotational speed;

[0048] The cooling parameters of the planetary reducer include the cooling medium of the heat dissipation system, inlet temperature, pressure, and flow rate requirements.

[0049] In the aforementioned design method for the planetary gear reducer heat dissipation system, optionally, in step two, the total thermal power is the sum of the total power loss of each stage of the planetary gear train and the total power loss of the sealing element, wherein the total power loss of each stage of the planetary gear train is the sum of the no-load power loss of the planetary gear reducer and the load power loss of the planetary gear reducer.

[0050] The total power loss of the sealing element is the frictional power loss of the contact oil seal, which is the sum of the power loss of the lip seal and the power loss of the mechanical end face seal.

[0051] Optionally, the no-load power loss of the planetary reducer in each stage of the planetary gear train includes the sum of the oil churning power loss of the bearings and the oil churning power loss of the gears.

[0052] Optionally, the calculation model for the oil churning power loss of the bearing is as follows:

[0053]

[0054] Among them, P BO Power loss of bearing oil churning, unit: kW; M BO The bearing's no-load torque is expressed in Nm; n B The rotational speed of the bearing about its own axis, in r / min;

[0055] The no-load torque of the bearing is calculated as follows:

[0056] When ν O n B <2000,

[0057]

[0058] When ν O n B When ≥2000,

[0059]

[0060] Where, ν O The unit is the kinematic viscosity of the lubricating oil, in mm. 2 / s;f BO The bearing oil immersion factor is selected according to different bearing types, referring to standard ISO 14179; d Bm The bearing's average diameter is in mm.

[0061] Optionally, the calculation model for the oil churning power loss of the gear is as follows:

[0062]

[0063] Among them, P MO Power loss of the gear churning system, unit: kW; P CS P CP P CC These represent the churning power losses of the sun gear, planet gears, and planet carrier, respectively, in kW, and their calculation models are as follows:

[0064]

[0065] Among them, A g f is the arrangement constant; S n is the sun gear oil immersion coefficient, with values ​​taken from standard ISO 14179; S The rotational speed of the sun gear is expressed in r / min; d OS b is the outer diameter of the sun gear, in mm; WS R is the total tooth width of the sun gear, in mm. f R is the roughness coefficient. f =7.93-4.648 / m t , where m t β is the end face module, in mm; β is the helix angle, in (︒).

[0066]

[0067] Among them, f Pn is the oil immersion coefficient of the planetary gears. C P d represents the rotational speed of the planetary gears relative to the planet carrier, in r / min. OP b is the outer diameter of the planetary gear, in mm. WP N represents the total tooth width of the planetary gear, in mm. CP The number of planetary gears;

[0068]

[0069] Among them, f C n is the oil immersion coefficient of the planetary carrier; C D is the rotational speed of the planetary carrier, in r / min; C W is the outer diameter of the planetary carrier, in mm. C The width of the planetary carrier is in mm.

[0070] Optionally, the load power loss of the planetary reducer in each stage of the planetary gear train includes the sum of the frictional power loss of the bearings and the frictional power loss of the gear meshing.

[0071] Optionally, the calculation model for the frictional power loss of the bearing is as follows:

[0072]

[0073] Among them, P BL The frictional power loss of the bearing is expressed in kW (m). L1 The bearing load friction torque is expressed in Nm (m). L2 The frictional torque of the bearing axial load, in Nm;

[0074] Among them, M L1 M L2 The calculation models are as follows:

[0075]

[0076] Among them, f L1 A1 and A2 are the bearing friction coefficients; P is the exponent. d1 This refers to the dynamic load on the bearing, in N.

[0077]

[0078] Among them, f L2 F is the coefficient of axial friction of the bearing. a This represents the axial component of the dynamic load on the bearing, in N.

[0079] bearing friction coefficient f L1 f L2The indices A1 and A2, as well as the bearing dynamic load, are selected according to different bearing types, with reference to standard ISO14179.

[0080] Optionally, the calculation model for the frictional power loss of the gear meshing is as follows:

[0081]

[0082] Among them, P ML The frictional power loss during gear meshing, expressed in kW (P). MLE P MLI These are the frictional power losses for the external meshing of the sun gear and planet gears, and the internal meshing of the planet gears and internal gear rings, respectively, in kW;

[0083] P MLE P MLI The calculation models are as follows:

[0084]

[0085] Among them, f Me T is the coefficient of external meshing friction. Me n represents the sun gear torque per branch, in Nm. CS β is the rotational speed of the sun gear relative to the planet carrier, in r / min; we The pitch circle helix angle of the sun gear and planet gear meshing, in (°); M Me For the mechanical benefits of external meshing;

[0086]

[0087] Among them, f Mi T is the coefficient of internal meshing friction. Mi The planetary gear torque for each branch is expressed in Nm; β wi The pitch circle helix angle of the meshing planetary gear and the internal ring gear, in (°); M Mi For the mechanical benefits of internal meshing;

[0088] friction coefficient f Me f Mi and mechanical benefits M Me M Mi The calculation method is based on standard ISO 14179.

[0089] Optionally, the calculation model for the power loss of the lip seal is as follows:

[0090]

[0091] Among them, P s1 Power loss of lip seal, in kW; T s1The frictional torque of the lip seal is expressed in Nm; n s1 The rotational speed of the shaft in contact with the lip seal, in r / min.

[0092] Optionally, the calculation model for the power loss of the mechanical end face seal is as follows:

[0093]

[0094] Among them, P s2 Power loss of mechanical end face seal, in watts (W); A s2 The contact area of ​​the sealing end face, in meters. 2 ;P cs2 This refers to the end-face specific pressure, in Pa; v s2 f is the average linear velocity of the sealing end face, in m / s. s2 The coefficient of friction of the sealing end face.

[0095] In the aforementioned design method for the planetary gear reducer cooling system, optionally, the design method further includes the step of determining the lubrication method and lubricating oil selection based on the input parameters. The method for determining the lubrication method and lubricating oil selection is as follows:

[0096] The lubrication method and lubricating oil viscosity grade are determined based on the gear pitch radius and ambient temperature.

[0097] The load rating of the lubricating oil is determined based on the maximum contact stress on the tooth surface and the oil film thickness.

[0098] Due to the application of the above technical solution, the present invention has the following advantages compared with the prior art:

[0099] The design method of this invention calculates the total heat dissipation area based on the total thermal power of the planetary reducer. In the calculation model of the total heat dissipation area, the heat transfer coefficient is calculated using the equivalent heat transfer coefficient of the heat exchange tube. When calculating the equivalent heat transfer coefficient, the surface heat transfer coefficients of the air side and the lubricating oil side of the heat exchange tube are fully considered, as well as the length of the heat exchange tube in the oil sump and the air cavity. Then, the heat exchange tube is arranged according to the total heat dissipation area. Compared with the traditional method of selecting the heat transfer coefficient of the heat exchange tube based on experience, the calculation of the total heat dissipation area is more realistic and more accurate, and can realize the rapid and efficient optimization design of the heat dissipation system in the planetary reducer with limited space.

[0100] Furthermore, the design method of the present invention comprehensively considers the total heat power loss of each planetary gear train and sealing element of the planetary reducer, and obtains the total heat dissipation area of ​​the heat dissipation system based on the total heat power. Then, based on the total heat dissipation area and the heat production ratio of each planetary gear train, the heat dissipation system is optimized and designed, which is more in line with reality and more accurate.

[0101] The design method of this invention can be used not only for the design of water-cooled heat dissipation systems for planetary reducers in high-power compact scraper conveyors for mining, but also for the design of water-cooled heat dissipation systems for other high-power industrial planetary reducers with limited space that require optimized design. Attached Figure Description

[0102] The disclosure of this invention will become more apparent from the accompanying drawings. It should be understood that these drawings are for illustrative purposes only and are not intended to limit the scope of protection of this invention.

[0103] Figure 1 A flowchart illustrating the optimized design method of the planetary gear reducer heat dissipation system provided in an embodiment of the present invention. Detailed Implementation

[0104] For any single technical feature described or implied in the embodiments submitted herein, or any single technical feature shown or implied in the various drawings, the present invention still operates in any combination or deletion among these technical features or their equivalents without any technical obstacle, and thus these further embodiments according to the present invention should also be considered within the scope of the description herein.

[0105] The present invention provides a design method for a planetary gear reducer heat dissipation system, which can be used for the structural design of a heat dissipation system for a high-power compact planetary gear reducer. The design method includes the following steps:

[0106] Step S1: Determine the input parameters for the planetary reducer and cooling system design;

[0107] Step S2: Determine the lubrication method and lubricant type based on the input parameters;

[0108] Step S3: Estimate the total thermal power of the planetary reducer. The total thermal power is the sum of the total power loss of the single or multi-stage planetary gear train and the total power loss of the sealing elements.

[0109] Step S4: Determine the lubricating oil cooling parameters of the planetary reducer based on the total thermal power;

[0110] Step S5: Determine the design scheme of the heat dissipation system based on the total heat power;

[0111] The design scheme of the heat dissipation system includes the assessment and design of the heat dissipation capacity of the heat dissipation system, which includes the assessment of the total heat dissipation area and the arrangement of heat dissipation pipes.

[0112] The total heat dissipation area is calculated based on the total heat power, and the arrangement of the heat dissipation pipes is obtained by conversion based on the total heat dissipation area.

[0113] The design method of the planetary gear reducer heat dissipation system of this application will be described in detail below through specific embodiments. The following embodiments take the design of the heat dissipation system of a two-stage planetary gear reducer as an example.

[0114] Combination Figure 1 As shown, the specific implementation process of the design method of the planetary reducer heat dissipation system of this invention is as follows:

[0115] Step S1: Determine the input parameters for the planetary reducer and cooling system design.

[0116] In this step, the input parameters mainly include the planetary reducer performance parameters, planetary reducer structural parameters, and planetary reducer cooling parameters.

[0117] The performance parameters of a planetary gearbox include input power and speed.

[0118] The structural parameters of the planetary gear reducer include the structural parameters of the first-stage planetary gear train, the structural parameters of the second-stage planetary gear train, the bearing structural parameters, the structural parameters of the sealing elements, and a 3D model of the planetary gear reducer assembly structure.

[0119] The structural parameters of the first-stage planetary gear train and the second-stage planetary gear train include gear train type, number of planetary gears, gear tooth profile parameters and working center distance, respectively. The gear tooth profile parameters include number of teeth, module, pressure angle, helix angle, tooth width, addendum coefficient, clearance coefficient and displacement coefficient.

[0120] Bearing structural parameters include bearing type, quantity, inner diameter, and outer diameter;

[0121] The structural parameters of a sealing element include the sealing type, material, sealing diameter, and rotational speed.

[0122] The cooling parameters of a planetary gearbox include the cooling medium of the heat dissipation system, inlet temperature, pressure, and flow rate requirements.

[0123] Step S2: Determine the lubrication method and lubricant type based on the input parameters.

[0124] In this step, the range of gear pitch circle linear velocity is calculated based on the input parameters of the planetary reducer, and the lubrication method and lubricating oil viscosity grade are determined based on the gear pitch circle linear velocity and the actual operating environment temperature.

[0125] Based on the calculation results of the maximum contact stress on the tooth surface of the planetary reducer and the oil film thickness, the load rating of the lubricating oil is determined.

[0126] In some specific embodiments, such as when the gear pitch circle velocity is below 15 m / s, oil bath lubrication is recommended; the oil film thickness must ensure that the tooth surface is in an elastohydrodynamic lubrication state. Relevant parameters and criteria can be selected according to the requirements specified in industry standards, such as ISO 6336 or gear transmission design manuals, etc., and will not be detailed here.

[0127] In this step, the gear pitch circle velocity, maximum contact stress on the tooth surface, and oil film thickness can be obtained using existing commonly used calculation methods, such as ISO 6336, which will not be described in detail here.

[0128] Step S3: Estimate the total thermal power of the planetary reducer. The total thermal power is the sum of the total power loss of the first-stage planetary gear train, the total power loss of the second-stage planetary gear train, and the total power loss of the sealing elements.

[0129] The total power loss of the first-stage planetary gear train is the sum of the no-load power loss and the load power loss of the planetary reducer; the total power loss of the second-stage planetary gear train is the sum of the no-load power loss and the load power loss of the planetary reducer; the total power loss of the sealing elements is the friction power loss of the contact oil seal, which is the sum of the power loss of the lip seal and the power loss of the mechanical end face seal.

[0130] Among them, the no-load power loss of the planetary reducers in the first-stage planetary gear train and the second-stage planetary gear train is the sum of the oil churning power loss of the rolling bearings and the oil churning power loss of the gears, respectively; the load power loss of the planetary reducers in the first-stage planetary gear train and the second-stage planetary gear train is the sum of the frictional power loss of the rolling bearings and the frictional power loss of the gear meshing, respectively.

[0131] After calculating the total power loss of the first-stage planetary gear train, the total power loss of the second-stage planetary gear train, and the total power loss of the sealing elements, the total power loss of the three is added together, and the sum is the total thermal power of the planetary reducer.

[0132] Specifically, the calculation models for each power loss are as follows:

[0133] (1) Calculation model of power loss of lip seal

[0134]

[0135] Among them, P s1 Power loss of lip seal, in kW; T s1 The frictional torque of the lip seal is expressed in Nm; n s1 The rotational speed of the shaft in contact with the lip seal, in r / min.

[0136] For T s1 The value of varies depending on the sealing material. For example, the value corresponding to T s1The value for fluororubber is 3.737 x 10. -3 d s1 Nitrile rubber is 2.429x10 -3 d s1 ; where d s1 The sealing diameter is in mm.

[0137] (2) Calculation model for power loss of mechanical end face seal

[0138]

[0139] Among them, P s2 Power loss of mechanical end face seal, in watts (W); A s2 The contact area of ​​the sealing end face, in meters. 2 ;P cs2 This refers to the end-face specific pressure, in Pa; v s2 f is the average linear velocity of the sealing end face, in m / s. s2 The coefficient of friction of the sealing end face.

[0140] Mechanical end face sealing end face specific pressure P cs2 The value range is 0.4 MPa ~ 0.6 MPa; the coefficient of friction f of the sealing end face s2 The value range is 0.005 to 0.15.

[0141] (3) Calculation model of oil churning power loss of rolling bearing

[0142]

[0143] Among them, P BO Power loss of rolling bearing oil churning, unit: kW; M BO The bearing's no-load torque is expressed in Nm; n B The rotational speed of the bearing around its own axis, in r / min.

[0144] The bearing's no-load torque is calculated as follows:

[0145] When ν O n B <2000,

[0146]

[0147] When ν O n B When ≥2000,

[0148]

[0149] Where, ν O The unit is the kinematic viscosity of the lubricating oil, in mm.2 / s;f BO The bearing oil immersion factor is selected according to different bearing types, referring to standard ISO 14179; d Bm The bearing's average diameter is in mm.

[0150] (4) Calculation model for oil stirring power loss of gears

[0151]

[0152] Among them, P MO P represents the power loss of the gear's oil churning mechanism, expressed in kW. CS P CP P CC These represent the churning power losses of the sun gear, planet gears, and planet carrier, respectively, in kW, and their calculation models are as follows:

[0153]

[0154] Among them, A g As a placement constant, we take 0.2; f S n is the sun gear oil immersion coefficient, with values ​​taken from standard ISO 14179; S The rotational speed of the sun gear is expressed in r / min; d OS b is the outer diameter of the sun gear, in mm; WS R is the total tooth width of the sun gear, in mm. f R is the roughness coefficient. f =7.93-4.648 / m t , where m t β is the end face module, in mm; β is the helix angle, in (︒).

[0155]

[0156] Among them, f P n is the oil immersion coefficient of the planetary gears. C P d represents the rotational speed of the planetary gears relative to the planet carrier, in r / min. OP b is the outer diameter of the planetary gear, in mm. WP N represents the total tooth width of the planetary gear, in mm. CP This represents the number of planetary gears.

[0157]

[0158] Among them, f C n is the oil immersion coefficient of the planetary carrier; C D is the rotational speed of the planetary carrier, in r / min; C W is the outer diameter of the planetary carrier, in mm. CThe width of the planetary carrier is in mm.

[0159] (5) Calculation model for frictional power loss of rolling bearings

[0160]

[0161] Among them, P BL The frictional power loss of rolling bearings is expressed in kW (m). L1 The bearing load friction torque is expressed in Nm (m). L2 This refers to the frictional torque of the bearing under axial load (cylindrical roller bearing), measured in Nm. L1 M L2 The calculation models are as follows:

[0162]

[0163] Among them, f L1 A1 and A2 are the bearing friction coefficients; P is the exponent. d1 This represents the dynamic load on the bearing, in N.

[0164]

[0165] Among them, f L2 F is the coefficient of axial friction of the bearing. a This represents the axial component of the dynamic load on the bearing, expressed in nanometers (N).

[0166] bearing friction coefficient f L1 f L2 The indices A1 and A2, as well as the bearing dynamic load, are selected according to different bearing types, with reference to standard ISO14179.

[0167] (6) Calculation model for total frictional power loss in gear meshing

[0168]

[0169] Among them, P ML This represents the total frictional power loss of the gear meshing in this stage of the planetary gear train, expressed in kW (P). MLE P MLI These represent the frictional power losses in each branch, specifically the external meshing of the sun gear and planet gears, and the internal meshing of the planet gears and the internal gear ring, expressed in kW. MLE P MLI The calculation models are as follows:

[0170]

[0171] Among them, f Me T is the coefficient of external meshing friction. Me n represents the sun gear torque per branch, in Nm.CS β is the rotational speed of the sun gear relative to the planet carrier, in r / min; we The pitch circle helix angle of the sun gear and planet gear meshing, in (°); M Me For the mechanical benefits of external meshing.

[0172]

[0173] Among them, f Mi T is the coefficient of internal meshing friction. Mi The planetary gear torque for each branch is expressed in Nm; β wi The pitch circle helix angle of the meshing planetary gear and the internal ring gear, in (°); M Mi For the mechanical benefits of internal meshing.

[0174] friction coefficient f Me f Mi and mechanical benefits M Me M Mi The calculation method is based on standard ISO 14179.

[0175] Step S4: Determine the lubricating oil cooling parameters of the planetary reducer based on the total heat power.

[0176] The parameters for lubricating oil cooling include the amount of lubricating oil required, the oil level in the oil sump, and the effective volume of the oil sump.

[0177] In this step, the lubricating oil demand is estimated based on the thermal balance of the planetary reducer when it operates at a steady-state oil bath temperature, that is, the total heat generation of the planetary reducer is equal to the total heat dissipation of the lubricating oil. The total thermal power of the planetary reducer is the total heat generation of the planetary reducer.

[0178] Specifically, the demand for lubricating oil is estimated in the following ways:

[0179] (1) First estimate the flow rate of the lubricating oil.

[0180]

[0181] Among them, P Q Heat dissipation, in W; q vo ρ represents the flow rate of lubricating oil, in L / min. o This refers to the density of the lubricating oil, expressed in kg / m³. 3 c po ΔT is the isobaric heat capacity of lubricating oil, expressed in J / (kg·K); o The temperature rise of the lubricating oil is expressed in K. The initial temperature of the lubricating oil is usually the ambient temperature of the planetary reducer. Generally, the temperature rise of the lubricating oil can be taken as 25℃~60℃.

[0182] Heat dissipation P QThe value is the total thermal power of the planetary reducer obtained in step S3.

[0183] (2) The lubricating oil demand is calculated by converting the lubricating oil flow rate and the circulation cooling time.

[0184] Selection of circulating cooling time: For external circulation of lubricating oil, the circulating cooling time is generally taken as 4~5 minutes; for oil sump storage without external circulation, the circulating cooling time can be reduced, and can be taken as less than the circulating cooling time taken for external circulation of lubricating oil, thereby calculating the lubricating oil demand. Optionally, the design method of this application embodiment is to design a heat dissipation system in the planetary reducer cavity, and the circulating cooling time can be selected when there is no external circulation in the oil sump storage, and the circulating cooling time is selected as less than 4~5 minutes.

[0185] In this step, the calculation of the oil level and effective volume of the oil sump is determined by the inner lip seal and mechanical end face seal of the planetary reducer. The general design principle is that the oil level in the oil sump should not be lower than 1 / 3 of the inner diameter height of the mechanical end face seal, and should not exceed the lowest point of the inner diameter of the lip seal.

[0186] The effective volume of the oil sump is calculated using the 3D model of the planetary reducer assembly structure (established in step S1) and the corresponding effective lubricating oil volume based on different oil sump levels. When the effective lubricating oil volume is lower than the designed lubricating oil requirement, the process returns to step S1, and the 3D model structure of the planetary reducer is readjusted. By increasing the cavity space, the lubricating oil volume is ensured to meet the requirements.

[0187] Step S5: Determine the design scheme of the heat dissipation system based on the total heat power;

[0188] The design scheme of the heat dissipation system includes the design of the structural parameters of the heat dissipation system and the evaluation and design of the heat dissipation capacity of the heat dissipation system. The evaluation and design of the heat dissipation capacity of the heat dissipation system includes the evaluation of the total heat dissipation area and the arrangement of heat dissipation pipes. Finned tubes are selected for heat dissipation pipes.

[0189] In this step, the structural parameters of the heat dissipation system are designed, taking into account factors such as design performance, manufacturing process, cycle time and cost, and selecting appropriate finned tube materials, tube diameters and fin specifications.

[0190] This step also includes assessing the cooling medium requirement, with cooling water being used as the cooling medium, in order to evaluate and design the heat dissipation capacity of the cooling system.

[0191] The assessment of cooling water demand is based on the thermal balance of the planetary reducer when it operates at a steady oil bath temperature under a stable water supply in the cooling system. That is, the total heat generation of the lubricating oil in the planetary reducer is equal to the total heat dissipation of the finned tube cooling system, and the total heat generation of the lubricating oil in the planetary reducer is equal to the total heat generation of the planetary reducer. The required cooling water demand of the finned tube cooling system is estimated.

[0192] Cooling water demand is estimated using the following methods:

[0193]

[0194] Where, q vw ρ is the flow rate of cooling water, in L / min; w The density of cooling water, in kg / m³ 3 c pw The isobaric heat capacity of cooling water, in J / (kg·K); ΔT w This refers to the temperature rise of the cooling water, measured in K. Generally, the temperature rise of cooling water can be taken as 5°C to 10°C.

[0195] The method for evaluating the total heat dissipation area in this step is as follows:

[0196] (1) Calculate the equivalent surface heat transfer coefficient of the finned tube (i.e., heat dissipation tube).

[0197] The calculation of the equivalent surface heat transfer coefficient of finned tubes is based on the application conditions of finned tubes in the oil bath of planetary gearboxes, including the total length of the finned tubes and the oil immersion depth. The air-side surface heat transfer coefficient and the oil-side surface heat transfer coefficient of the finned tubes are calculated separately, and then converted to obtain the equivalent surface heat transfer coefficient. The calculation models are as follows:

[0198]

[0199] Among them, h a1 The heat transfer coefficient of the air-side surface of the finned tube is expressed in W / (m²). 2 .K); λ a1 d is the thermal conductivity of air, in W / (mK); fi fin inner diameter, in meters; Re is the Reynolds number; Pr is the Prandtl number; c h1 n h1 m h1 The coefficient is selected based on the correlation formula of the average heat transfer coefficient across the surface of the circular tube, according to the range of Re values.

[0200]

[0201] Among them, h o1 The heat transfer coefficient of the oil-side surface of the finned tube is expressed in W / (m²). 2 .K); λ o1 c is the thermal conductivity of lubricating oil, in W / (mK); h2 n h2 m h2 The coefficient is selected based on the correlation formula of the average heat transfer coefficient across the surface of the circular tube, according to the range of Re values.

[0202]

[0203] Among them, h eff The equivalent surface heat transfer coefficient of the finned tube is expressed in W / (m²). 2 .K); L gt The total length of a single annular finned tube, in meters (m); L ga L is the length of a single annular finned tube in the air cavity, in meters (m). go The length of a single annular finned tube submerged in the oil bath, in meters (m) or liters (L). gt =L ga +L go .

[0204] (2) Calculate the total heat dissipation area of ​​the required finned tube heat dissipation system.

[0205]

[0206] Among them, S g The total heat dissipation area of ​​the finned tube is m. 2 ;T omax Maximum permissible oil temperature, unit: K; T wi T wo These are the inlet and outlet temperatures of the cooling water, respectively, in Kelvin (K). Where T... omax The maximum allowable oil temperature is usually determined based on the grade of the lubricating oil.

[0207] In this step, the optimized arrangement of finned tubes involves proportionally allocating the required total heat dissipation area based on the total power loss of the first-stage planetary gear train and the second-stage planetary gear train. The required number of finned tubes is determined by the heat dissipation area of ​​a single tube based on the selected specifications of the finned tubes, and then arranged as required in the corresponding planetary cavities.

[0208] If the internal cavity of the planetary reducer cannot accommodate the required number of finned tubes, return to step S1, readjust the 3D model of the planetary reducer structure to ensure sufficient space for the heat dissipation system, so that the total heat dissipation area meets the requirements, and obtain the final 3D model of the planetary reducer structure and the number of selected heat dissipation tubes in each planetary cavity.

[0209] Traditional design methods present significant challenges in arranging the cooling system within the limited space of a compact planetary gearbox driven by an integrated unit. These challenges include the difficulty of fitting the cooling system within the limited space of the planetary gearbox and the uneven cooling effect of the lubricating oil in different areas within the gearbox. This presents a contradiction between the limited design space of a compact planetary gearbox and the requirement for high-efficiency cooling of the cooling system. The optimized design method proposed in this invention effectively resolves this contradiction. Furthermore, compared to traditional design methods, this invention discloses an optimized design calculation process for the cooling system of a high-power compact planetary gearbox. Based on the calculation of the total heat dissipation area of ​​the cooling system, it allocates the heat dissipation area of ​​the cooling system according to the actual heat generation ratio within the limited cavity space of different planetary levels within the gearbox, thus calculating the required number of finned tubes. This method can fully and efficiently utilize the limited space within the planetary gearbox for the cooling system, balance the lubricating oil cooling effect in different cavity spaces, and rapidly and effectively improve the efficiency of design and experimental iteration.

[0210] This design method can be used for the design of water-cooled heat dissipation systems for high-power compact scraper conveyors in mining, and can also be applied to the design of water-cooled heat dissipation systems for other high-power industrial planetary gearboxes with limited space that require optimized design.

[0211] Traditional design methods typically involve a simple centralized arrangement of the heat dissipation system within a single cavity of the planetary gear reducer. However, this can lead to uneven lubricant cooling in different planetary cavity regions within a multi-stage planetary gear reducer. The design method described in this application distributes the heat dissipation system according to the actual heat generation ratio of the first-stage and second-stage planetary gear trains, balancing the lubricant cooling effect in different cavity spaces and ensuring uniform lubricant cooling.

[0212] The design method of this application comprehensively considers the total heat power loss (i.e., total heat generation) of each planetary gear train and sealing element of the planetary reducer. At the same time, it calculates the equivalent heat transfer coefficient of the heat exchanger tubes by combining the actual 3D model structure. The equivalent heat transfer coefficient fully considers the surface heat transfer coefficients of the air side and the lubricating oil side of the heat exchanger tubes, as well as the length of the heat exchanger tubes in the oil sump and air cavity, and finally obtains the total heat dissipation area of ​​the heat dissipation system. Then, the heat exchanger tubes are arranged according to the total heat dissipation area. Compared with the traditional method of selecting the heat transfer coefficient of the heat exchanger tubes based on experience, the calculation of the total heat dissipation area is more realistic and more accurate, and can realize the rapid and efficient optimization design of the heat dissipation system in the space-constrained planetary reducer.

[0213] In other alternative embodiments, this design method for the planetary gear reducer heat dissipation system can also be used to design the heat dissipation system for a single-stage planetary gear reducer, a three-stage planetary gear reducer, or a planetary gear reducer with more stages. When designing the heat dissipation system for a single-stage planetary gear reducer, the total thermal power loss of the planetary gear reducer is calculated by summing the total power loss of the first-stage planetary gear train and the total power loss of the sealing elements. When designing the heat dissipation system for a three-stage or more planetary gear reducer, the total thermal power loss of the planetary gear reducer is calculated by summing the total power loss of each stage of the planetary gear train and the total power loss of the sealing elements.

[0214] The above embodiments are only used to illustrate the embodiments of the present invention, and are not intended to limit the embodiments of the present invention. Those skilled in the art can make various changes and modifications without departing from the spirit and scope of the embodiments of the present invention. Therefore, all equivalent technical solutions also fall within the scope of the embodiments of the present invention, and the patent protection scope of the embodiments of the present invention should be defined by the claims.

Claims

1. A design method for a heat dissipation system of a planetary gear reducer, characterized in that, The design method includes the following steps: Step 1: Determine the input parameters for the planetary gearbox and cooling system design; Step 2: Estimate the total thermal power of the planetary reducer based on the input parameters; Step 3: Determine the design scheme of the heat dissipation system based on the total heat power; The design scheme of the heat dissipation system includes the evaluation and design of the heat dissipation capacity of the heat dissipation system, which includes the evaluation of the total heat dissipation area and the arrangement of heat dissipation pipes. The total heat dissipation area is calculated based on the total heat power, and the arrangement of the heat dissipation pipes is obtained by conversion based on the total heat dissipation area. The calculation model for the total heat dissipation area is as follows: ； Among them, S g The total heat dissipation area of ​​the heat pipes, in meters. 2 ;P Q Heat dissipation, measured in W; T omax Maximum permissible oil temperature, unit: K; T wi T wo These are the inlet and outlet temperatures of the cooling medium, respectively, in Kelvin (K). The heat dissipation P Q The value is the total thermal power of the planetary reducer obtained in step two; ； Among them, h eff The equivalent surface heat transfer coefficient of the heat pipe is expressed in W / (m²). 2 .K); L gt The total length of a single heat pipe, in meters (m); L ga L is the length of a single heatsink tube within the air cavity, in meters (m). go The length of a single heat dissipation pipe submerged in the oil bath, in meters (m). ； Among them, h a1 The heat transfer coefficient of the air-side surface of the heat pipe, in W / (m²). 2 .K); λ a1 d is the thermal conductivity of air, in W / (mK); fi The inner diameter of the heat pipe is in meters (m); Re is the Reynolds number; Pr is the Prandtl number; c h1 n h1 m h1 The coefficient is selected based on the correlation formula of the average heat transfer coefficient across the surface of the circular tube, according to the range of the Re value. ； Among them, h o1 The heat transfer coefficient of the lubricating oil side surface of the heat sink is expressed in W / (m²). 2 .K); λ o1 c is the thermal conductivity of lubricating oil, in W / (mK); h2 n h2 m h2 The coefficient is selected based on the correlation formula of the average heat transfer coefficient across the surface of the circular tube, according to the range of the Re value. The required number of heat dissipation tubes is determined based on the total heat dissipation area and the selected specifications of the heat dissipation tubes, and then arranged in the corresponding planetary cavity.

2. The design method for the heat dissipation system of the planetary gear reducer according to claim 1, characterized in that, In step one, the input parameters include establishing a 3D model of the planetary reducer assembly structure; The total heat power is the sum of the total power loss of a single-stage or multi-stage planetary gear train and the total power loss of the sealing element. When the total heat power is the sum of the total power loss of a multi-stage planetary gear train and the total power loss of the sealing element, the arrangement of the heat dissipation pipes is proportionally allocated according to the total power loss of the multi-stage planetary gear train. If the internal cavity of the planetary reducer cannot accommodate the required number of heat dissipation pipes, return to step one to readjust the 3D model of the planetary reducer structure to ensure sufficient space for heat dissipation pipes and that the total heat dissipation area meets the requirements.

3. The design method for the heat dissipation system of the planetary gear reducer according to claim 2, characterized in that, The design method also includes the step of determining the lubricating oil cooling parameters of the planetary reducer based on the total thermal power; The lubricating oil cooling parameters include the lubricating oil demand, the oil level in the sump, and the effective volume of the oil sump; wherein... The required amount of lubricating oil was estimated using the following method: ； Among them, P Q Heat dissipation, in W; q vo ρ represents the flow rate of lubricating oil, in L / min. o This refers to the density of the lubricating oil, expressed in kg / m³. 3 c po ΔT is the isobaric heat capacity of lubricating oil, expressed in J / (kg·K); o Temperature rise of lubricating oil, in K; The required amount of lubricating oil is calculated by converting the flow rate and the circulation cooling time. The selection of the circulation cooling time is as follows: for external circulation of lubricating oil, the circulation cooling time is selected as 4 to 5 minutes; for oil storage in the oil sump without external circulation, the circulation cooling time can be reduced. The principle for determining the oil level height is: the oil level height should not be lower than 1 / 3 of the inner diameter height of the mechanical end face seal, and should not exceed the lowest point of the inner diameter of the lip seal. The effective volume of the oil tank is calculated based on the 3D model established in step one, according to different oil tank liquid levels. When the effective lubricating oil volume is lower than the required lubricating oil volume, return to step one, readjust the 3D model of the planetary reducer structure, and increase the cavity space to ensure that the lubricating oil volume meets the requirements.

4. The design method of the planetary gear reducer heat dissipation system according to claim 1, characterized in that, The assessment and design of the heat dissipation capacity of the heat dissipation system also includes an assessment of the cooling medium requirement, which is estimated using the following methods: ； Where, q vw ρ represents the flow rate of the cooling medium, in L / min. w The density of the cooling medium is expressed in kg / m³. 3 c pw ΔT is the isobaric heat capacity of the cooling medium, expressed in J / (kg·K); w The temperature rise of the cooling medium is expressed in Kelvin (K).

5. The design method for the heat dissipation system of the planetary gear reducer according to any one of claims 1 to 4, characterized in that, In step one, the input parameters include planetary reducer performance parameters, planetary reducer structural parameters, and planetary reducer cooling parameters. The planetary reducer structural parameters include the structural parameters of each stage of the planetary gear train, the structural parameters of the bearings, and the structural parameters of the sealing elements. The performance parameters of the planetary reducer include input power and speed; The structural parameters of each level of the planetary gear train include gear train type, number of planetary gears, gear tooth profile parameters and working center distance, wherein the gear tooth profile parameters include number of teeth, module, pressure angle, helix angle, tooth width, addendum coefficient, clearance coefficient and displacement coefficient; The bearing structural parameters include bearing type, quantity, inner diameter, and outer diameter; The structural parameters of the sealing element include sealing type, material, sealing diameter, and rotational speed; The cooling parameters of the planetary reducer include the cooling medium of the heat dissipation system, inlet temperature, pressure, and flow rate requirements.

6. The design method of the planetary gear reducer heat dissipation system according to any one of claims 1 to 4, characterized in that, In step two, the total thermal power is the sum of the total power loss of each stage of the planetary gear train and the total power loss of the sealing elements, wherein... The total power loss of each planetary gear train is the sum of the no-load power loss of the planetary reducer and the load power loss of the planetary reducer; The total power loss of the sealing element is the frictional power loss of the contact oil seal, which is the sum of the power loss of the lip seal and the power loss of the mechanical end face seal.

7. The design method of the planetary gear reducer heat dissipation system according to claim 6, characterized in that, The no-load power loss of the planetary reducer in each stage of the planetary gear train includes the sum of the oil churning power loss of the bearings and the oil churning power loss of the gears; the calculation model for the oil churning power loss of the bearings is as follows: ； Among them, P BO Power loss of bearing oil churning, unit: kW; M BO The bearing's no-load torque is expressed in Nm; n B The rotational speed of the bearing about its own axis, in r / min; The no-load torque of the bearing is calculated as follows: When ν O n B <2000, ； When ν O n B When ≥2000, ； Where, ν O The viscosity is the kinematic viscosity of the lubricating oil, in mm. 2 / s;f BO The bearing oil immersion factor is selected according to different bearing types, referring to standard ISO 14179; d Bm The bearing's average diameter is in mm. The calculation model for the oil churning power loss of the gear is as follows: ； Among them, P MO Power loss of the gear churning system, unit: kW; P CS P CP P CC These represent the churning power losses of the sun gear, planet gears, and planet carrier, respectively, in kW, and their calculation models are as follows: ； Among them, A g f is the arrangement constant; S n is the sun gear oil immersion coefficient, with values ​​taken from standard ISO 14179; S The rotational speed of the sun gear is expressed in r / min; d OS b is the outer diameter of the sun gear, in mm; WS R is the total tooth width of the sun gear, in mm. f R is the roughness coefficient. f =7.93-4.648 / m t , where m t β is the end face module, in mm; β is the helix angle, in ︒. ； Among them, f P n is the oil immersion coefficient of the planetary gears. C P d represents the rotational speed of the planetary gears relative to the planet carrier, in r / min. OP b is the outer diameter of the planetary gear, in mm. WP N represents the total tooth width of the planetary gear, in mm. CP The number of planetary gears; ； Among them, f C n is the oil immersion coefficient of the planetary carrier; C D is the rotational speed of the planetary carrier, in r / min; C W is the outer diameter of the planetary carrier, in mm. C The width of the planetary carrier is in mm.

8. The design method of the planetary gear reducer heat dissipation system according to claim 6, characterized in that, The load power loss of the planetary reducer in each stage of the planetary gear train includes the sum of the frictional power loss of the bearings and the frictional power loss of the gear meshing. The calculation model for the frictional power loss of the bearing is as follows: ； Among them, P BL The frictional power loss of the bearing is expressed in kW (m). L1 The bearing load friction torque is expressed in Nm (m). L2 The frictional torque of the bearing axial load, in Nm; Among them, M L1 M L2 The calculation models are as follows: ； Among them, f L1 A1 and A2 are the bearing friction coefficients; P is the exponent. d1 This refers to the dynamic load on the bearing, in N. ； Among them, f L2 F is the coefficient of axial friction of the bearing. a This represents the axial component of the dynamic load on the bearing, in N. bearing friction coefficient f L1 f L2 The indices A1 and A2, as well as the bearing dynamic load, are selected according to different bearing types, with reference to standard ISO14179. The calculation model for the frictional power loss during gear meshing is as follows: ； Among them, P ML The frictional power loss during gear meshing, expressed in kW (P). MLE P MLI These are the frictional power losses for the external meshing of the sun gear and planet gears, and the internal meshing of the planet gears and internal gear rings, respectively, in kW; P MLE P MLI The calculation models are as follows: ； Among them, f Me T is the coefficient of external meshing friction. Me n represents the sun gear torque per branch, in Nm. CS β is the rotational speed of the sun gear relative to the planet carrier, in r / min; we The pitch circle helix angle of the sun gear and planet gear meshing, measured in ︒; M Me For the mechanical benefits of external meshing; ； Among them, f Mi T is the coefficient of internal meshing friction. Mi The planetary gear torque for each branch is expressed in Nm; β wi M is the pitch circle helix angle of the meshing planetary gear and the internal ring gear, expressed in α; Mi For the mechanical benefits of internal meshing; friction coefficient f Me f Mi and mechanical benefits M Me M Mi The calculation method is based on standard ISO 14179.

9. The design method of the planetary gear reducer heat dissipation system according to claim 6, characterized in that, The calculation model for the power loss of the lip seal is as follows: ； Among them, P s1 Power loss of lip seal, in kW; T s1 The frictional torque of the lip seal is expressed in Nm; n s1 The rotational speed of the shaft in contact with the lip seal, in r / min; The calculation model for the power loss of the mechanical end face seal is as follows: ； Among them, P s2 Power loss of mechanical end face seal, in watts (W); A s2 The contact area of ​​the sealing end face, in meters. 2 ;P cs2 This refers to the end-face specific pressure, in Pa; v s2 f is the average linear velocity of the sealing end face, in m / s. s2 The coefficient of friction of the sealing end face.

10. The design method of the planetary gear reducer heat dissipation system according to claim 1, characterized in that, The design method further includes the step of determining the lubrication method and lubricating oil selection based on the input parameters. The method for determining the lubrication method and lubricating oil selection is as follows: The lubrication method and lubricating oil viscosity grade are determined based on the gear pitch radius and ambient temperature. The load rating of the lubricating oil is determined based on the maximum contact stress on the tooth surface and the oil film thickness.

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