Optimization design method for wet brake of hydraulic motor with high braking torque

Abstract

The invention discloses an optimization design method for a wet brake of a hydraulic motor with high braking torque, which comprises the following steps of: firstly, determining basic performance parameters of the brake according to the braking performance and the structural size requirement of the hydraulic motor, and providing a design process from the angle of model selection design of key parts such as a friction pair and a loading assembly; and then a brake is designed according to the braking performance requirement of the low-speed large-torque inner curve hydraulic motor. And finally, a brake performance test is carried out on a test platform, a research result shows that performance parameters such as brake torque of the brake meet design requirements, and the effectiveness of the optimization design method is verified. The design method has certain guiding significance on optimization design research of the low-speed large-torque hydraulic motor brake.

Description

Technical Field

[0001] The invention belongs to the technical field of hydraulic motors, and in particular relates to an optimization design method for a high braking torque hydraulic motor wet brake. Background Art

[0002] As a radial piston hydraulic motor, the inner curve hydraulic motor is an important actuator in hydraulic transmission. Due to its advantages of large torque, low speed and good stability, it can be directly used as the actuator terminal of the mechanical system, saving transmission links and improving energy conversion efficiency. The inner curve hydraulic motor is widely used in military and civilian fields, such as the rotary drive system of large equipment such as destroyers, dredgers, tunneling machines and injection molding machines. Affected by uncertain factors such as oil leakage, the deceleration braking of such hydraulically driven ultra-high power mechanical equipment during operation has the safety hazard of brake failure, so it has always been an important research topic.

[0003] In order to ensure the effectiveness and stability of the deceleration braking of the inner curve hydraulic motor, it is necessary to equip it with a brake with high braking torque and high stability. There are three main types of brakes currently used: caliper disc brakes, drum brakes and wet brakes. Since the former two have problems such as small contact area of ​​friction pairs, poor automatic compensation gap ability of friction pairs, poor air heat dissipation effect, and easy thermal failure, wet brakes have the advantages of good oil heat dissipation performance and strong anti-pollution ability. Therefore, wet brakes are more suitable for low-speed and high-torque hydraulic motors. At present, the wet brakes studied by domestic and foreign scholars are mainly used on the wheels of engineering machinery such as mining vehicles and loaders, while there are fewer studies on wet brakes used on inner curve hydraulic motors.

[0004] Domestic and foreign scholars have mainly focused their research on wet brakes for vehicles, but have done little research on wet brakes for hydraulic motors. At the same time, the braking torque required for braking conditions on inner curve hydraulic motors is much higher than that for vehicle brakes, sometimes even dozens or even hundreds of times higher. Therefore, the design requirements for such high braking torque hydraulic motor wet brakes are higher and more difficult. Summary of the invention

[0005] The purpose of this invention is to solve the deceleration and braking problems of inner curve hydraulic motors and the lack of related research on wet brakes for hydraulic motors, and propose a design method for a high braking torque hydraulic motor wet brake that can be assembled in a modular manner. Based on the basic structural composition and design requirements of the brake, key components including friction pairs and loading components are designed, the overall structural design of the brake is verified, and a complete design scheme is proposed from the perspectives of material design, structural design and size design.

[0006] The objective of the present invention is achieved through the following technical solution, a method for optimizing the design of a high braking torque hydraulic motor wet brake, the method comprising the following steps:

[0007] S1: Determine the structural parameters of the brake according to the actual working requirements and specific working parameters of the hydraulic motor wet brake.

[0008] S2: Perform standard tests on the model under different working conditions to obtain the actual friction coefficient of the friction plate.

[0009] S3: Design and verify the specific parameters of friction pairs and loading components

[0010] Furthermore, in step S1, various structural parameters include: maximum parking braking torque, dynamic output torque, brake dimensions, friction plate material, friction plate inner diameter, friction plate outer diameter, and friction plate thickness. The above structural parameters are calculated and selected according to actual working parameters.

[0011] Furthermore, in step S2, the ball and disk are rubbed, which is different from the actual surface friction. Therefore, it is necessary to use different loading forces and different rotation speeds to conduct tests, and roughly determine the contact pressure through scratches. Then, a set of tests equal to the actual use pressure is selected as the reference value of the friction coefficient.

[0012] Furthermore, in step S2, the specific operations for obtaining the actual friction coefficient of the friction plate are as follows: first, install the rotation drive module and the circular sample block; the first step is to control the loading module to load so that the load reaches the preset value and stabilizes for 10 seconds; the second step is to control the drive module to rotate at a constant speed and record the data during the friction process at a sampling frequency of 100 Hz. This step lasts for 5 minutes; then replace the sample block, perform three repeated tests, take the average of the three groups of friction coefficients and record the data; finally, repeat the above steps for each working condition.

[0013] Furthermore, in step S3, the specific parameters of the friction pair, the number of friction surfaces and the axial clamping force, are selected according to the following formula:

[0014] M z =nkμF a R d

[0015] Where n is the number of friction surfaces, n = 2N (N is the number of friction plates); k is the reduction coefficient, which is related to the number of friction surfaces. The more friction surfaces there are, the smaller k is. The k value is determined according to the manual. μ is the friction coefficient, F a is the axial pressing force, R d is the equivalent radius of the friction surface

[0016] Furthermore, in step S3, the basic parameters of the loading component are determined according to the following method:

[0017] Step 1: Preliminary selection of disc springs according to the mechanical design manual. The selection parameters of disc springs include D—outer diameter of disc springs d—inner diameter of disc spring; t—thickness of disc spring; h0—calculated value of the deformation of disc spring when flattened; H0—height of disc spring; P—load of disc spring; f—deformation of disc spring; σ—calculated stress at a certain point on the surface of disc spring.

[0018] Step 2: Calculate the load P when flattening the disc spring c :

[0019]

[0020] Where, E is elastic modulus; μ is Poisson's ratio; K1 and K4 are calculation coefficients. K4 is 1 for disc springs without support surface, and K1 is calculated according to the following formula:

[0021]

[0022] Wherein, C = D / d.

[0023] Step 3: To adjust the deformation requirements of the disc spring during further compression, you can use a disc spring composite or matched disc spring method. When the disc spring matching method is selected, for j disc spring composite combinations, calculate the load ratio P' / Pc=P / (jPc). Consult the characteristic curve of the single disc spring, and obtain f / h0 based on the parameters h0 / t and P' / Pc, and then calculate the disc spring deformation f. The maximum allowable deformation of the disc spring is f'=0.75h0, and the piston return amount Lh=f'-f is calculated. The average clearance hp of each friction pair is:

[0024]

[0025] If h p If the conditions are met, the disc spring selection can be completed. p Too small, the friction torque loss is too large when releasing the brake, the disc spring should be re-selected, the disc spring deformation f should be increased and re-checked until h p Once the conditions are met, the disc spring selection is completed and the loading component design is completed.

[0026] When the disc spring is used, the deformation f will be doubled after the disc spring is combined. Calculate the piston return stroke L h , the average clearance of each friction pair h p , meet the design requirements and complete the loading component design.

[0027] The beneficial results of the present invention are:

[0028] 1. A complete design solution is proposed to solve the deceleration and braking problem of inner curve hydraulic motor and wet brake for hydraulic motor.

[0029] 2. It has certain guiding significance for the subsequent theoretical research and design research on wet brakes for hydraulic motors. BRIEF DESCRIPTION OF THE DRAWINGS

[0030] Figure 1 Flowchart of the optimization design method for a high braking torque hydraulic motor wet brake

[0031] Figure 2 A schematic diagram of the specific structure and size limitations of a certain internal curve hydraulic motor

[0032] Figure 3 Schematic diagram of the variation curve of friction coefficient with sliding speed and loading force

[0033] Figure 4 Schematic diagram of the effect of oil on the friction coefficient when the loading force is 40N

[0034] Figure 5 Design schematics for loading scenarios

[0035] In the accompanying drawings, the main reference numerals are as follows:

[0037] 1. Brake cylinder, 2. O-ring, 3. Brake piston, 4. Friction plate, 5. Dual steel plate, 6. Adjustment gasket, 7. Brake shaft, 8. Disc spring, 9. Guide sleeve, 10. Hexagon socket screw, 11. Rear cover. DETAILED DESCRIPTION

[0038] The preferred embodiments of the present invention are described in detail below in conjunction with the accompanying drawings so that the advantages and features of the present invention can be more easily understood by those skilled in the art, thereby making a clearer and more definite definition of the protection scope of the present invention.

[0039] The present invention is further described below with reference to the accompanying drawings. The present invention provides an optimization design method for a high braking torque hydraulic motor wet brake. Figure 1 As shown, the following steps are included:

[0040] S1: According to the actual working requirements and specific working parameters of the hydraulic motor wet brake, determine the various structural parameters of the brake: maximum parking braking torque, dynamic output torque, brake dimensions, friction plate material, friction plate inner diameter, friction plate outer diameter, and friction plate thickness.

[0041] In a specific example, the structure of a certain type of hydraulic motor optimized by the present invention is as follows Figure 2 As shown, the parameters are shown in Table 1:

[0042] Table 1 Internal curve hydraulic motor parameters

[0043]

[0044] The structural parameters are calculated and rounded off: the maximum parking braking torque M1 is 42 kN·m, the dynamic braking torque M2 is 27.5 kN·m, the size of the brake part is limited to less than Φ450 mm × 198 mm, and the size and material of the friction plate and the dual steel plate are shown in Table 2:

[0045] Table 2 Size and material of friction plate and dual steel plate

[0046]

[0047] S2: Perform standard tests on the model under different working conditions to obtain the actual friction coefficient of the friction plate. In a specific example, the working conditions are set to 20N, 40N, 60N, and 80N, and the friction coefficient curve obtained by the test is shown in Figure 3 In the wet brake studied in this study, the actual surface contact pressure P of the copper-based friction plate is a The initial setting is 16MPa, so the test under a loading force of 40N is more in line with the actual working conditions. The loading force is set to 40N, and the copper-based friction plate is tested with and without oil. The results are shown in Figure 4 , and the friction coefficient is stabilized at 0.08.

[0048] S3: Design and verify the specific parameters of the friction pair and loading components. After determining the materials and dimensions of the friction plate and the dual plate, it is necessary to design and determine the number of friction components that make up the friction pair. The friction plate has a certain pressure limit, and its maximum pressure must not exceed 20MPa. In order to ensure safety and extend life, the initial pressure on a single friction surface is Pa=16MPa. The dynamic braking torque of the brake is also the total braking torque M2=Mz=27.5kN·m. The dynamic friction coefficient is initially selected as μ=0.08, R1=59.5mm, and R2=80.5mm based on the test results in the previous section. Substitute the above basic parameters into the following formula:

[0049] M z =nkμF a R d

[0050] It can be obtained that n1=40.96, n2=161.04. Considering a certain error, add 1 to n1 and round it to the integer, and determine the number of friction surfaces n=42, then the number of friction plates is 21, and the number of dual steel plates is 22. Further, the dynamic braking torque is 28.1kN·m, which meets the braking requirements. Substitute Pa=16MPa into the following formula:

[0051]

[0052] The axial clamping force Fa = 147.8 kN.

[0053] The maximum braking torque is the parking braking torque. The friction coefficient at this time is the static friction coefficient. According to relevant information, the ratio of static and dynamic friction coefficients is 1.5 to 2. The static friction coefficient is taken as μ=0.12. After calculation, the maximum braking torque is 42.15kN·m, which is greater than the design requirement of 42kN·m.

[0054] In summary, the friction pair of the wet brake has 21 friction plates and 22 dual steel plates, and the required axial clamping force is 147.8 kN.

[0055] Furthermore, the selection of loading components is mainly based on the disc spring structure size and basic parameters. Set the opening pressure P0 = 1.7MPa; Considering the safety margin, the brake axial clamping force Fa is set to 150kN, and bring in The outer diameter D of the brake piston is 0.334m. The calculated result is rounded off and the outer diameter D of the brake piston is taken as 330mm. Considering that the axial clamping force is too large, four sets of disc springs are selected to be evenly spaced and distributed on the circumference of the brake piston. The specific form is as follows Figure 5 shown.

[0056] The basic parameters of the disc spring are further determined below. According to the axial compression force and size, the initial disc spring parameters are shown in Table 3.

[0057] Table 3 Disc spring parameters

[0058]

[0059] Considering that the disc spring needs to be further compressed to release the brake, a composite combination of two disc springs is used. Then each disc spring bears a load of P' = F a / (i×j)=15000 / (2×4)=18750N. Determine the mechanical parameters of the disc spring: elastic modulus E=2.06×105MPa; Poisson's ratio μ=0.3, Calculate the load Pc when flattening the disc spring = 62725N, then the load ratio P / Pc≈0.3. Check the single disc spring characteristic curve, when h0 / t≈0.4, P / P c When ≈0.3, f / h0=0.28. The disc spring deformation f=0.28h0=0.28×2.2=0.616mm.

[0060] Calculate the piston return stroke L h =f'-f=1.65-0.616=1.034mm. The average clearance h of each friction pair p=1.034 / 42=0.025mm. Since the average clearance is too small, the friction torque loss when releasing the brake is too large. The deformation of the disc spring should be increased. This can be achieved by directly engaging the disc spring under the condition that the disc spring parameters remain unchanged. Then the deformation of the disc spring during braking and the deformation of the disc spring during releasing the brake are directly doubled, f=1.232mm, f′=3.3mm. Determine the piston return amount L h =2.068mm, the average clearance hp of each friction pair is 0.05mm, which meets the design requirements and completes the loading component design.

[0061] Taking a hydraulic motor as an example, the design results are as follows:

[0062] The size of the wet brake is Φ375mm×162mm, which is within the allowable range of Φ450mm×198mm. The designed wet brake has a maximum braking torque of 42kN·m, a minimum opening pressure of 1.75MPa, and a maximum opening pressure of 4.2MPa.

Claims

1. A high braking torque hydraulic motor wet brake optimization design method, characterized in that: The method comprises the following steps: S1: Determine the structural parameters of the brake according to the actual working requirements and specific working parameters of the hydraulic motor wet brake; S2: Conduct standard tests on the model under different working conditions to obtain the actual friction coefficient of the friction plate; S3: Design and verify the specific parameters of friction pair and loading components.

2. The optimization design method of a high braking torque hydraulic motor wet brake according to claim 1, characterized in that: In step S1, various structural parameters include: maximum parking brake torque, dynamic output torque, brake dimensions, friction plate material, friction plate inner diameter, friction plate outer diameter, and friction plate thickness. The above structural parameters are calculated and selected according to actual working parameters.

3. The method for optimizing the design of a high-efficiency, low-pulsation cycloid motor flow distribution system according to claim 1, characterized in that: In step S2, friction is performed between a ball and a disk. Different from actual surface-to-surface friction, tests are performed with different loading forces and rotation speeds. The contact pressure is roughly determined by scratches, and then a set of tests equal to the actual operating pressure is selected as a reference value for the friction coefficient.

4. The method for optimizing the design of a high-efficiency, low-pulsation cycloid motor flow distribution system according to claim 1, characterized in that: In step S2, the specific operations for obtaining the actual friction coefficient of the friction plate are as follows: first, install the rotating drive module and the circular sample block; in the first step, control the loading of the loading module so that the load reaches the preset value and stabilizes; in the second step, control the driving module to rotate at a constant speed and record the data during the friction process; then replace the sample block, perform three repeated tests, take the average of the three groups of friction coefficients and record the data; finally, repeat the above steps for each working condition.

5. The method for optimizing the design of a high-efficiency, low-pulsation cycloid motor flow distribution system according to claim 1, characterized in that: In step S3, the specific parameters of the friction pair, the number of friction surfaces and the axial clamping force, are selected according to the following formula: M z =nkμF a R d Where n is the number of friction surfaces, n = 2N (N is the number of friction plates); k is the reduction coefficient, which is related to the number of friction surfaces. The more friction surfaces there are, the smaller k is. The k value is determined according to the manual. μ is the friction coefficient, F a is the axial pressing force, R d is the equivalent radius of the friction surface 6. The method for optimizing the design of a high-efficiency, low-pulsation cycloid motor flow distribution system according to claim 1, characterized in that: In step S3, the basic parameters of the loading component are determined according to the following method: Step 1: Preliminary selection of disc springs according to the mechanical design manual. The selection parameters of disc springs include D—outer diameter of disc springs d—inner diameter of disc spring; t—thickness of disc spring; h0—calculated value of disc spring deformation when flattened; H0—height of disc spring; P—load of disc spring; f—deformation of disc spring; σ—calculated stress at a certain point on the surface of disc spring; Step 2: Calculate the load Pc when flattening the disc spring: Where, E is elastic modulus; μ is Poisson's ratio; K1 and K4 are calculation coefficients. K4 is 1 for disc springs without support surface, and K1 is calculated according to the following formula: Where, C = D / d; Step 3: Adjust the deformation requirements of the disc spring during further compression, and use a disc spring composite or matched disc spring method; if the disc spring matching method is selected, for j disc spring composite combinations, calculate the load ratio P' / P c =P / (jP c ); Consult the characteristic curve of the single disc spring, based on the parameters h0 / t and P' / P c Get f / h0, and then find the disc spring deformation f; the maximum allowable disc spring deformation f'=0.75h0, calculate the piston return stroke L h =f'-f, average clearance of each friction pair h p : If h p If the conditions are met, the disc spring selection can be completed; if h p If the disc spring is too small, the friction torque loss is too large when the brake is released. The disc spring should be re-selected, the disc spring deformation f should be increased and re-checked until h p If the conditions are met, the disc spring selection is completed and the loading component design is completed; When the disc spring is used, the deformation f will be doubled after the disc spring is combined. Calculate the piston return stroke L h , the average clearance of each friction pair h p , meet the design requirements and complete the loading component design.

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