Method for improving comprehensive performance of joint reduction device of industrial robot

Abstract

The application discloses a method for improving comprehensive performance of an industrial robot joint reduction device, and belongs to the technical field of industrial robot reducers. The device comprises an input shaft with an input gear, three groups of crankshafts with eccentric sections, a planetary gear transmission mechanism and a cycloidal pin wheel transmission mechanism. The method provided by the application optimizes the design of the material, surface strengthening parameters, matching size, meshing side gap range of the input gear and the planetary gear, and the needle profile in the needle and holder assembly of each core component, reduces the wear and seizure failures in the transmission process, improves the hardness, matching precision, positioning precision, transmission stability and heat dissipation performance of the industrial robot joint reduction device, solves the problems of insufficient service life of the industrial robot joint reduction device, unstable backlash and low transmission precision, meets the demand of long-term stable operation of high-end industrial robots, and has a good industrial application prospect.

Description

Technical Field

[0001] This invention belongs to the field of industrial robot reducer technology, and in particular relates to a method for improving the overall performance of industrial robot joint deceleration devices. Background Technology

[0002] The joint reduction gear of an industrial robot is a core component of the robot's motion actuator. Its main function is to convert the high-speed rotational motion of the motor into low-speed rotational motion and high-torque motion of the robot joint, while ensuring the positioning accuracy and motion stability of the joint. Its performance directly determines the operating accuracy, load-bearing capacity, and service life of the industrial robot. Currently, industrial robot joint reduction gears have advantages such as large transmission ratio, high transmission accuracy, high rigidity, and good impact resistance, making them an indispensable core component of high-end industrial robots.

[0003] However, current industrial robot joint reduction devices still have many technical defects in practical applications, making it difficult to meet the requirements of long-term stable operation of high-end heavy-duty industrial robots. First, the raceways of tapered roller bearings and the surfaces of the eccentric sections of the crankshaft are prone to fatigue spalling under long-term heavy loads, alternating load conditions, and the action of iron powder, leading to decreased accuracy and shortened lifespan of the reduction device. Second, gear backlash control is unstable; the backlash design between the input shaft and planetary gears lacks quantitative standards, easily resulting in problems such as excessive backlash causing transmission shock and insufficient backlash causing jamming, affecting robot positioning accuracy. Third, during the operation of the needle roller and cage assembly, the cage is prone to wear, affecting the efficiency and lifespan of the reduction device. Fourth, the spline pair fit between the planetary gears and the crankshaft lacks consistency, easily resulting in large differences in tightness after assembly, which can reduce the transmission efficiency of the reduction device and generate noise. Fifth, the fit between the roller grooves, needle pins, and cycloidal gears is unreasonable; the fit between these three components is difficult to control precisely, easily resulting in large backlash / return, exacerbating transmission losses, and further reducing the transmission accuracy and lifespan of the reduction device. Summary of the Invention

[0004] To solve the above problems, the present invention adopts the following technical solution.

[0005] This invention first provides a method for improving the overall performance of a joint deceleration device for industrial robots. The deceleration device includes a pin gear housing, an input planetary carrier, an output planetary carrier, planetary gears, a crankshaft with an eccentric section, and an input shaft with an input gear. One side of the pin gear housing is connected to the input planetary carrier through an outer ring assembly of an angular contact ball bearing. The pin gear housing contains two cycloidal gears with a phase difference of 180°. The cycloidal gears are hinged to the eccentric section of the crankshaft through a needle roller and cage assembly.

[0006] The method includes:

[0007] 1) Determine the material and surface hardening parameters of each component of the speed reduction device according to the requirements of heavy-duty working conditions;

[0008] 2) Determine the inner diameter of the outer ring assembly of the angular contact ball bearing based on the root circle diameter of the cycloidal gear and the eccentricity of the cycloidal gear forming parameters;

[0009] 3) Based on the uniformity requirements of the transmission clearance, adjust the center distance tolerance between the input gear and the planetary gear, and the span distance between the input gear and the planetary gear. Use the extreme value method based on the base circle tooth thickness to determine the meshing backlash range between the input gear and the planetary gear.

[0010] 4) Determine the matching method of each component of the speed reduction device;

[0011] 5) Adjust the module, number of teeth, pressure angle, addendum coefficient and clearance coefficient of the planetary gear and the input gear so that the gear pair parameters of the planetary gear and the input gear meet the overlap ratio ≥1.2, and adjust the theoretical displacement coefficient of the planetary gear and the input gear so that the difference between the maximum slip ratio of the planetary gear and the input gear does not exceed the preset threshold.

[0012] 6) Establish the force equation of the needle roller and cage assembly in the reduction gear, and use a logarithmic curve to determine the needle roller profile in the needle roller and cage assembly based on the force.

[0013] 7) Assemble the speed reduction device.

[0014] Preferably, the extreme value method based on the base circle tooth thickness in step 3) specifically refers to:

[0015] When the center distance between the input gear and the planetary gear is the minimum center distance The base circle tooth thickness of the input gear is the base circle tooth thickness corresponding to the maximum span distance under the specified gauge bar. The base circle tooth thickness of a planetary gear is the base circle tooth thickness corresponding to the maximum span distance under a specified gauge bar. At this time, the meshing backlash between the input gear and the planetary gear is minimized, and the minimum backlash... satisfy:

[0016]

[0017] When the center distance between the input gear and the planetary gear is the maximum center distance The base circle tooth thickness of the input gear is the base circle tooth thickness corresponding to the minimum span distance under the specified gauge bar. The base circle tooth thickness of a planetary gear is the base circle tooth thickness corresponding to the minimum span distance under a specified gauge bar. At this time, the meshing backlash between the input gear and the planetary gear is at its maximum, and the maximum backlash is... for:

[0018]

[0019] In the formula: , This is the center distance between the input gear and the planetary gear. When finding the minimum side clearance, it is... When calculating the maximum side clearance, it is , The base circle radius of the input gear. Let be the base circle radius of the planetary gear. The base circle tooth pitch.

[0020] Preferably, the needle tooth housing has circumferentially distributed grooves, and needle tooth pins are provided on the grooves. The planetary gear and the crankshaft cooperate to form an involute spline pair.

[0021] Preferably, the component fits determined in step 4) are as follows: the involute spline pair of the planetary gear and the crankshaft adopts a transition fit; the pin pin, cycloidal gear and pin housing groove adopt an interference fit; the crankshaft and tapered roller bearing adopt an interference fit; the outer ring assembly of the angular contact ball bearing and the pin housing adopt a transition fit; the input planetary carrier, output planetary carrier and tapered roller bearing adopt a transition fit.

[0022] Preferably, the spline pitch tolerance of the planetary gear is set to three levels, and the spline pitch tolerance of the crankshaft is also set to the corresponding three levels, so that the mating tolerance zones of the planetary gears and crankshaft in the same group are consistent.

[0023] Preferably, the mating method determined in step 4) includes the pin tooth center not coinciding with the groove center, and the pin tooth radius being smaller than the groove radius.

[0024] Preferably, in step 6), the force on the needle roller and cage assembly in the reduction gear... Satisfies the vector relation:

[0025]

[0026] In the formula: To input the number of gear teeth, This refers to the number of teeth on a planetary gear. Eccentricity is the forming parameter of the cycloidal gear; For crankshaft rotation angle, To output the planetary carrier rotation angle, For the weight of a single cycloidal gear, This refers to the number of teeth on a cycloidal gear. For the pressure angle of the cycloidal gear, For input shaft torque, This is the center distance between the input gear and the planetary gear.

[0027] Compared with existing technologies, the present invention has a reasonable and compact structure, and has the following beneficial effects:

[0028] (1) The inner and outer rings of tapered roller bearings were prepared using 20CrMo alloy carburized steel. The residual austenite content was controlled to be ≤5%, and the residual compressive stress and effective stress layer depth on the raceway surface were optimized. Combined with carburizing quenching + low-temperature tempering + deep cryogenic process, the bearing contact fatigue life was improved by more than 1.2 times compared with conventional products, which can meet the service life requirements of industrial robot joints. By optimizing the cage material of the needle roller and cage assembly and the cage material of the tapered roller bearing, 20CrMo was used. After carburizing quenching, tempering and phosphating, the surface hardness reached 600HV0.2. 800HV0.2 reduces cage wear and improves the lifespan and transmission efficiency of the needle roller and cage assembly;

[0029] (2) The inner diameter of the outer ring assembly of the angular contact ball bearing is specified, which restricts the two cycloidal gears between the outer ring assembly of the bearing, thereby reducing the activity space and installation error of the cycloidal gears;

[0030] (3) By quantitatively designing the backlash between the input shaft and the planetary gear, and using the process of three-level spline tolerance grouping and one-time clamping grinding of the planetary gear involute gear, the unevenness of the transmission clearance is eliminated.

[0031] (4) The upper deviation of the root circle diameter of the pin tooth housing groove adopts a negative deviation. By selecting pin tooth pins and cycloidal gears, they are made to have an interference fit with the pin tooth housing groove in the radial direction. This causes the groove to undergo elastic deformation under the interference, changing from the original line contact to a circular arc surface contact. The center of the pin tooth pin does not coincide with the actual center of the groove, and the actual radius of the pin tooth pin is smaller than the radius of the groove, which provides a grease storage space and is beneficial for heat dissipation of the cycloidal gear tooth surface. At the same time, the matching method of the pin tooth housing groove, pin tooth pin and cycloidal gear is optimized. The precise matching of the three is achieved by selecting them in groups, which improves the matching accuracy and ensures the positioning accuracy and transmission stability of the reducer.

[0032] (5) Based on the force on the needle roller and cage assembly in the deceleration device, optimize the needle roller profile to a logarithmic curve to improve the needle roller's load-bearing capacity and anti-eccentric load performance.

[0033] (6) Reasonably design the fit method and surface strengthening parameters of each part to greatly improve the overall rigidity and anti-overturning torque of the device and adapt to the requirements of heavy-duty working conditions; by optimizing the material, process and fit parameters of key parts, reduce wear, jamming and other faults in the transmission process and improve the operational reliability of the industrial robot joint deceleration device. Attached Figure Description

[0034] Figure 1 This is a front view of a schematic diagram of an embodiment of the present invention;

[0035] Figure 2 yes Figure 1 BB section view;

[0036] Figure 3 yes Figure 2 The view to the view;

[0037] Figure 4 This is a schematic diagram of the crankshaft eccentric section;

[0038] Figure 5 This is a schematic diagram of a planetary gear;

[0039] Figure 6 This is a schematic diagram of the root circle diameter and groove radius of the needle tooth shell.

[0040] Figure 7 This is a schematic diagram of the assembly of the pin tooth, the groove, and the cycloidal gear.

[0041] Figure 8 It is the force exerted by the needle roller and cage assembly on the bearing hole of the cycloidal gear during the operation of the reduction gear.

[0042] In the diagram, 1-output planetary carrier; 2-first sealing cover; 3-tapered roller bearing; 4-bore retaining ring; 5-second sealing cover; 6-washer; 7-needle roller and cage assembly; 8-seal ring; 9-outer ring assembly of angular contact ball bearing; 10-pin housing; 11-pin pin; 12-first cycloidal gear; 13-second cycloidal gear; 14-input planetary carrier; 15-internal threaded tapered pin; 16-input shaft; 17-shaft retaining ring; 18-planetary gear; 19-crankshaft; 20-Socket headstock screw. Detailed Implementation

[0043] The present invention will be further described and illustrated below with reference to specific embodiments. The technical features of each embodiment of the present invention can be combined accordingly, provided that there is no mutual conflict.

[0044] This invention provides a method for improving the overall performance of a joint reduction device for industrial robots. First, the industrial robot joint reduction device is described with reference to the accompanying drawings. This industrial robot joint reducer includes a cycloidal pinwheel transmission mechanism, a planetary gear transmission mechanism, an input shaft 16, and three crankshafts 19, as shown below. Figures 1 to 3 As shown.

[0045] The planetary gear transmission mechanism includes an input planetary carrier 14 with raceways, an output planetary carrier 1 with raceways, an input shaft 16, and three sets of planetary gears 18. The input planetary carrier 14 and the output planetary carrier 1 are fixedly connected as a single unit. In this embodiment, the input planetary carrier 14 and the output planetary carrier 1 are positioned by internally threaded tapered pins 15 and connected as a single unit by hexagonal head screws 20. The planetary gears 18 are involute gears, located on the outside of the input planetary carrier 14. The three sets of planetary gears 18 are evenly distributed circumferentially along the input shaft 16, and the external teeth of all three sets of planetary gears 18 mesh with the input gear located at the input end of the input shaft 16. Both the input planetary carrier 14 and the output planetary carrier 1 have three crankshaft mounting holes evenly distributed circumferentially around their central axes. The crankshaft 19 is mounted in the crankshaft mounting holes via tapered roller bearings 3, with an interference fit between the crankshaft 19 and the tapered roller bearings 3. The end face of the tapered roller bearing 3 is provided with a retaining ring 4 for axial positioning. The input planetary carrier 14 and the output planetary carrier 1 are each driven by three pairs of tapered roller bearings 3, which transmit the rotation vectors on the first cycloidal wheel 12 and the second cycloidal wheel 13 to the output planetary carrier 1 in a 1:1 ratio.

[0046] The cycloidal pinwheel transmission mechanism includes a pin housing 10, a pin pin 11, a first cycloidal gear 12, and a second cycloidal gear 13. The pin housing 10 has grooves evenly distributed circumferentially. The pin pin 11 is positioned on the grooves. One side of the pin housing 10 is connected to the input planetary carrier 14 via an angular contact ball bearing outer ring assembly 9, providing positioning for the input planetary carrier 14. The other side of the pin housing 10 forms a rotating pair with the output planetary carrier 1. Both the input planetary carrier 14 and the output planetary carrier 1 contain bearing raceways, and the angular contact ball bearing outer ring assembly 9 is separated from the bearing raceways. The first cycloidal gear 12 and the second cycloidal gear 13 are 180° out of phase. The first cycloidal gear 12 and the second cycloidal gear 13 are hinged to three crankshafts 19 via a needle roller and cage assembly 7. Washers 6 are provided at both ends of the needle roller and cage assembly 7. The first cycloidal gear 12 and the second cycloidal gear 13 mesh with the pin pin 11 placed in the grooves. A sealing ring 8 is installed between the output planetary carrier 1 and the pin tooth housing 10. A center hole is provided at the center axis of both the input planetary carrier 14 and the output planetary carrier 1. A second sealing cover 5 is installed on the outer end face of the center hole of the output planetary carrier 1, and a first sealing cover 2 is installed on the outer end face of the three crankshaft mounting holes.

[0047] like Figure 4 and Figure 5 As shown, the crankshaft 19 has an involute external spline machined on the end surface near the input planetary carrier 14, and the hub bore of the planetary gear 18 has an involute internal spline machined. The two form an involute spline mating pair with a transition fit. A shaft elastic retaining ring 17 is installed on the outer end face of the planetary gear 18 for axial positioning of the planetary gear 18.

[0048] The crankshaft 19 is an eccentric crankshaft. The first cycloidal wheel 12 and the second cycloidal wheel 13 are hinged to the eccentric sections at both ends of the crankshaft 19. The two eccentric sections have equal offsets and are 180° out of phase, thereby driving the first cycloidal wheel 12 and the second cycloidal wheel 13 to rotate eccentrically.

[0049] The working principle of the joint reduction device of this industrial robot is explained below: When the pin gear housing 10 is fixed, if power is input clockwise from the splined input shaft 16, the input gear on the input shaft 16 simultaneously meshes with the three circumferential planetary gears 18. The planetary gears 18 are fixed to the crankshaft 19, driving the crankshaft 19 to move eccentrically. The first cycloidal gear 12 and the second cycloidal gear 13, which are 180° out of phase, are hinged to the three crankshafts 19 through needle rollers and the cage assembly 7, and mesh with the pin pin 11 placed on the output planetary carrier. While the first cycloidal gear 12 and the second cycloidal gear 13 revolve under the action of the eccentric section of the crankshaft 19, they generate a torque on the pin pin 11 in the groove of the pin gear housing 10 in the same direction as the revolution of the cycloidal gear, causing the output planetary carrier 1 to rotate clockwise.

[0050] The method for improving the overall performance of industrial robot joint deceleration devices provided by this invention solves problems such as insufficient lifespan, unstable backlash / return, and low transmission accuracy in industrial robot joint deceleration devices by optimizing the materials, surface strengthening parameters, mating dimensions, and transmission design of core components. Specifically, it includes:

[0051] I. In order to improve the overall rigidity and anti-overturning moment capability of the device, adapt to the requirements of heavy-duty working conditions, reduce wear, jamming and other failures in the transmission process, and improve the operational reliability of the industrial robot joint deceleration device, it is necessary to optimize the material and surface strengthening parameters of each core component.

[0052] 1) The effective hardened layer on the tooth surfaces of the first cycloidal gear 12 and the second cycloidal gear 13 and the bearing bore is 0.6 mm. 0.9mm thick, surface hardness 59 63 HRC, heart hardness 30 40HRC; the surface roughness of the tooth surfaces of the first cycloidal gear 12 and the second cycloidal gear 13 is Ra0.2, and the surface roughness of the bearing hole is Ra0.1.

[0053] 2) The effective hardened layer on the surface of the eccentric section of crankshaft 19 is 0.8 mm. 1.1mm thick, surface hardness 61 64 HRC, cardiac hardness 40 50 HRC, material support ratio 75%, surface roughness Ra 0.12, waviness Wz1 maximum height ≤ 0.1 .

[0054] 3) The cage material of the needle roller and cage assembly 7 and the tapered roller bearing 3 is 20CrMo, which, after carburizing, quenching, tempering, and phosphating, achieves a surface hardness of 600HV0.2. 800HV0.2.

[0055] 4) The outer and inner rings of the tapered roller bearing 3 are made of 20CrMo material, with retained austenite controlled to within 5%, and the residual compressive stress on the raceway surface of the inner and outer rings is -700. -900MPa, with an effective compressive stress layer of 0.2. The 0.4mm difference in thickness results in a 3-contact fatigue life of the tapered roller bearing that is more than 1.2 times longer than that of conventional products, thus meeting the service life requirements of industrial robot joints.

[0056] II. Cooperation Relationship

[0057] 1) In order to confine the two cycloidal gears between the outer ring assembly 9 of the angular contact ball bearing and reduce the play and installation error of the cycloidal gears, the inner diameter of the outer ring assembly 9 of the angular contact ball bearing is... The following relation must be satisfied:

[0058]

[0059] In the formula: The root circle diameter of the cycloidal gear is... This refers to the eccentricity of the forming parameters of the cycloidal gear.

[0060] 2) Based on the uniformity requirements of the transmission clearance, the center distance tolerance between the input gear and planetary gear 18, and the span distance between the input gear and planetary gear 18 are reasonably adjusted. The meshing backlash range between the input gear and planetary gear 18 is determined using the extreme value method based on the base circle tooth thickness; specifically:

[0061] When the center distance between the input gear and the planetary gear 18 is the minimum center distance The base circle tooth thickness of the input gear is the base circle tooth thickness corresponding to the maximum span distance under the specified gauge bar. The base circle tooth thickness of planetary gear 18 is the base circle tooth thickness corresponding to the maximum span distance under a specified gauge bar. At this time, the meshing backlash between the input gear and the planetary gear 18 is minimized, and the minimum backlash... satisfy:

[0062]

[0063] When the center distance between the input gear and planetary gear 18 is the maximum center distance The base circle tooth thickness of the input gear is the base circle tooth thickness corresponding to the minimum span distance under the specified gauge bar. The base circle tooth thickness of planetary gear 18 is the base circle tooth thickness corresponding to the minimum span distance under a specified gauge bar. At this time, the meshing backlash between the input gear and planetary gear 18 is at its maximum, and the maximum backlash... for:

[0064]

[0065] In the formula: , This is the center distance between the input gear and the planetary gear 18. When finding the minimum side clearance, it is... When calculating the maximum side clearance, it is , The base circle radius of the input gear. The base circle radius of planetary gear 18 The base circle tooth pitch.

[0066] In this embodiment, the meshing backlash between the input gear and the planetary gear 18 is between 0.04 mm and 0.105 mm, thereby eliminating the unevenness of the transmission backlash.

[0067] 3) The involute spline pair between planetary gear 18 and crankshaft 19 adopts a transition fit. The tolerance of the spline span of the planetary gear 18 is set to three levels, and the tolerance of the spline span of the crankshaft 19 is also set to the corresponding three levels. The fit tolerance zones of planetary gear 18 and crankshaft 19 in the same group are consistent. When grinding the external gear, planetary gear 18 with the same spline span is ground simultaneously in a single clamping. This optimization of fit dimensions is also to eliminate the unevenness of transmission clearance.

[0068] 4) The diagram shows the groove radius and groove root circle diameter of the needle tooth housing 10. Figure 6 The upper deviation of the root circle diameter of the tooth groove of the needle tooth housing 10 adopts a negative deviation, and the tolerance zone of the root circle diameter is 0.021mm; the lower deviation of the groove radius is 0, and the upper deviation is a positive deviation.

[0069] The root circle diameter of the grooved tooth is divided into three grades. On the one hand, by selecting the pin 11 and cycloidal gear, an interference fit is made radially with the groove of the pin housing 10, so that the groove undergoes elastic deformation under the interference, changing from the original line contact to a circular arc surface contact, such as... Figure 7 As shown; on the other hand, by optimizing the matching method of the pin tooth housing 10 groove, the pin tooth pin 11 and the cycloidal gear, the precise matching of the three can be achieved through group selection, thereby improving the matching accuracy and ensuring the positioning accuracy and transmission stability of the speed reduction device.

[0070] The center of the pin 11 does not coincide with the center of the groove, and the radius of the pin 11 is smaller than the radius of the groove, so as to leave space for grease storage, which is beneficial to heat dissipation of the cycloidal gear tooth surface.

[0071] 5) The fit between crankshaft 19 and tapered roller bearing 3 is an interference fit. In this embodiment, the minimum interference is 0.05% of the shaft diameter and the maximum interference is 0.16% of the shaft diameter.

[0072] 6) Both the input planetary carrier 14 and the output planetary carrier 1 contain bearing raceways. The outer ring assembly 9 of the angular contact ball bearing is separated from the planetary carrier raceway. The outer ring assembly 9 of the angular contact ball bearing and the needle tooth housing 10 adopt an transition fit. In this embodiment, the maximum interference is 0.015% of the fitting bore diameter and the maximum clearance is 0.02% of the fitting bore diameter.

[0073] 7) The fit between the input planetary carrier 14, the output planetary carrier 1 and the tapered roller bearing 3 is a transition fit. In this embodiment, the maximum interference is 0.04% of the fit hole diameter and the maximum clearance is 0.05% of the fit hole diameter.

[0074] III. Transmission Design

[0075] 1) Maximum slip ratio design of planetary gear 18 and input gear

[0076] Adjust the module, number of teeth, pressure angle, addendum coefficient, and clearance coefficient of the planetary gear 18 and the input gear on the input shaft 16 to ensure that the gear pair parameters of the planetary gear 18 and the input gear on the input shaft 16 meet the requirement of a contact ratio ≥ 1.2. Also adjust the theoretical displacement coefficient of the planetary gear 18 and the input gear on the input shaft 16 to make the maximum slip ratio of the tooth surface of the planetary gear 18 as close as possible to the maximum slip ratio of the tooth surface of the input gear.

[0077] 2) Needle roller profile design in needle roller and cage assembly 7

[0078] Establish the force equations for the needle roller and cage assembly 7 in the reduction gear, optimize the needle roller profile to a logarithmic curve, and determine the forces acting on the needle roller and cage assembly 7 in the reduction gear. Satisfies the vector relation:

[0079]

[0080] In the formula: To input the number of gear teeth, The planetary gear has 18 teeth. This refers to the eccentricity of the forming parameters of the cycloidal gear. For the rotation angle of crankshaft 19, To output the rotation angle of planetary carrier 1, For the weight of a single cycloidal gear, This represents the number of teeth on the cycloidal gear. The pressure angle of the cycloidal gear. For input shaft 16 torque, This is the center distance between the input gear and the planetary gear 18.

[0081] The logarithmic curve conforms to the following equation:

[0082]

[0083] In the formula: For roller profile shaping; This represents the number of rollers; This is the effective length of the needle roller; This is the eccentric load factor; The elastic modulus of the needle roller; 1 / J is Poisson's ratio for the needle roller; 1 / J is taken as 4.08. The coordinates of the contact surfaces between the needle rollers and the cage are given, with values ​​ranging from [value range missing]. .

[0084] The industrial robot joint deceleration device is then assembled, thus completing the overall performance optimization.

[0085] In this embodiment, the parameters of the input gear and planetary gear 18 on the input shaft 16, determined by the method provided by the present invention, are shown in Table 1:

[0086] Table 1. Parameters of First-Stage Gear Transmission

[0087]

[0088] In this embodiment, the root circle diameter of the cycloidal gear is set to 128.4 mm, the eccentricity is 1.3 mm, and the inner diameter of the outer ring assembly 9 of the angular contact ball bearing is taken as the reference. It is 128mm.

[0089] The diameter tolerance of crankshaft 19, which mates with tapered roller bearing 3, is set to 0.008. The value is 0.013, with a minimum interference of 0.008 mm and a maximum interference of 0.025 mm.

[0090] The bore diameter tolerance for the mating of the needle tooth housing 10 and the outer ring assembly 9 of the angular contact ball bearing is set to 0.004. -0.021, maximum interference is 0.021mm, maximum clearance is 0.029mm.

[0091] The bore diameter tolerance for the input planetary carrier 14, output planetary carrier 1, and tapered roller bearing 3 is set to -0.005. -0.015, maximum interference is 0.015 mm, maximum clearance is 0.011 mm.

[0092] The tolerance for the root circle diameter of the 10-grooved needle tooth housing is set to -0.01. -0.031 is divided into three levels: A1: -0.01 -0.017, A2: -0.017 -0.024, A3: -0.024 -0.031, the grooving radius tolerance is set to 0. -0.015, in the optional configuration, select a pin 11 with a size smaller than the groove radius, and then select one of the three gears of the pin housing groove according to the cycloidal gear and the pin 11.

[0093] The tolerance for the spline spacing of crankshaft 19 is set to ±0.03, which is divided into three grades: d: -0.01. -0.03, e: -0.01 0.01, f: 0.01 The tolerance for the internal spline span of planetary gear 18 is set to 0.03. 0.07 is divided into three levels: D: 0.025 0, E: 0.025 0.05, F:0.05 0.07, then pair using Dd, Ff, and Ee.

[0094] The effective hardened layer on the tooth surfaces and bearing holes of the first cycloidal wheel 12 and the second cycloidal wheel 13 is 0.6 mm. 0.9mm thick, surface hardness 59 63 HRC, heart hardness 30 40HRC, the surface roughness of the teeth of the first cycloidal wheel 12 and the second cycloidal wheel 13 is Ra0.2, and the surface roughness of the bearing hole is Ra0.1.

[0095] The effective hardened layer on the surface of the eccentric section of crankshaft 19 is 0.8 mm. 1.1mm thick, surface hardness 61 64 HRC, cardiac hardness 40 50 HRC, material support ratio 75%, surface roughness Ra 0.12, waviness Wz1 maximum height ≤ 0.1 .

[0096] Take the torque of input shaft 16 The maximum force exerted by the needle roller and cage assembly 7 on the bearing bore of the cycloidal gear is 4527.6 N, which is 9.8 Nm. Figure 8 As shown, the cage material of the needle roller and cage assembly 7 and the tapered roller bearing 3 is 20CrMo. After carburizing, quenching, tempering, and phosphating, the surface hardness of the cage of the needle roller and cage assembly 7 and the tapered roller bearing 3 reaches 600HV0.2. 800HV0.2; The outer and inner rings of the tapered roller bearing 3 are made of 20CrMo, with retained austenite controlled to within 5%, and the residual compressive stress on the raceway surface of the inner and outer rings is -700. -900MPa, with an effective compressive stress layer of 0.2. 0.4mm.

[0097] By optimizing the design of materials, surface strengthening parameters, mating dimensions, meshing backlash range of input gear and planetary gear, and needle roller profile in needle roller and cage assembly, the hardness, mating accuracy, positioning accuracy, transmission stability and heat dissipation performance of industrial robot joint deceleration device are improved.

[0098] The effects of this invention are explained below with reference to experimental data:

[0099] Three samples each of speed reducer A (a conventional product) and speed reducer B (a product with optimized overall performance using the method provided by this invention) were taken and installed on a speed reducer performance test bench, with an input speed of 2135 r / min for each sample. The relevant data are as follows:

[0100] Table 2 Comprehensive performance parameters of the speed reduction device

[0101]

[0102] The data in the table above shows that, in terms of temperature rise, speed reducer B is 3.4℃ lower on average than speed reducer A, transmission efficiency is increased by an average of 6%, and noise is reduced by 2.4dB. The starting torque of the speed reducer was measured by the performance testing system and was reduced by about 0.2NM compared to speed reducer A. Through comparative experiments, accelerated overload tests were conducted on speed reducers A(1), A(2), A(3), B(1), B(2), and B(3). After the tests, the cycloidal gear tooth surface, crankshaft eccentric section, and tapered roller bearing raceway in speed reducers A(1), A(2), and A(3) showed varying degrees of fatigue spalling and wear. The cage of the needle roller and cage assembly and the cage of the tapered roller bearing outer ring assembly also showed wear. However, no fatigue problems were found in the corresponding positions of speed reducers B(1), B(2), and B(3), indicating that the lifespan of speed reducer B was improved to a certain extent. The above data demonstrates that the industrial robot joint deceleration device, after its overall performance is improved by the method provided by this invention, has achieved significant improvements in temperature rise, transmission efficiency, noise, starting torque, and lifespan.

[0103] The embodiments described above are merely illustrative of implementation methods of the present invention, and while the descriptions are specific and detailed, they should not be construed as limiting the scope of the present invention. It should be noted that those skilled in the art can make various modifications and improvements without departing from the concept of the present invention, and these modifications and improvements all fall within the scope of protection of the present invention. Therefore, the scope of protection of the present invention should be determined by the appended claims.

Claims

1. A method for improving the overall performance of a joint deceleration device in an industrial robot, characterized in that, The reduction gear includes a pin gear housing, an input planetary carrier, an output planetary carrier, planetary gears, a crankshaft with an eccentric section, and an input shaft with an input gear; one side of the pin gear housing is connected to the input planetary carrier through an outer ring assembly of an angular contact ball bearing; two cycloidal gears with a phase difference of 180° are provided inside the pin gear housing, and the cycloidal gears are hinged to the eccentric section of the crankshaft through a needle roller and cage assembly; The method includes: 1) Determine the material and surface hardening parameters of each component of the speed reduction device according to the requirements of heavy-duty working conditions; 2) Determine the inner diameter of the outer ring assembly of the angular contact ball bearing based on the root circle diameter of the cycloidal gear and the eccentricity of the cycloidal gear forming parameters; 3) Based on the uniformity requirements of the transmission clearance, adjust the center distance tolerance between the input gear and the planetary gear, and the span distance between the input gear and the planetary gear. Use the extreme value method based on the base circle tooth thickness to determine the meshing backlash range between the input gear and the planetary gear. 4) Determine the matching method of each component of the speed reduction device; 5) Adjust the module, number of teeth, pressure angle, addendum coefficient and clearance coefficient of the planetary gear and the input gear so that the gear pair parameters of the planetary gear and the input gear meet the overlap ratio ≥1.2, and adjust the theoretical displacement coefficient of the planetary gear and the input gear so that the difference between the maximum slip ratio of the planetary gear and the input gear does not exceed the preset threshold. 6) Establish the force equation of the needle roller and cage assembly in the reduction gear, and use a logarithmic curve to determine the needle roller profile in the needle roller and cage assembly based on the force. 7) Assemble the speed reduction device.

2. The method according to claim 1, characterized in that, The material and surface strengthening parameters of each component determined in step 1) include: The effective hardened layer on the tooth surface and bearing bore of the cycloidal gear is 0.6~0.9mm, with a surface hardness of 59~63HRC and a core hardness of 30. 40 HRC; The effective hardened layer on the surface of the eccentric section of the crankshaft is 0.8 mm. 1.1mm thick, surface hardness 61 64 HRC, cardiac hardness 40 50HRC, waviness Wz1 maximum height ≤ 0.1 ; The crankshaft is circumferentially distributed with tapered roller bearings around the input shaft. The residual austenite content in the outer and inner rings of the tapered roller bearings is controlled to within 5%, and the residual compressive stress on the raceway surfaces of the inner and outer rings is -700. -900MPa, with an effective compressive stress layer of 0.

2. 0.4mm; The surface hardness of the needle roller and cage assembly and the cage material of tapered roller bearings is 600HV0.2~800HV0.

2.

3. The method according to claim 1, characterized in that, The extreme value method based on the base circle tooth thickness in step 3) is specifically as follows: When the center distance between the input gear and the planetary gear is the minimum center distance The base circle tooth thickness of the input gear is the base circle tooth thickness corresponding to the maximum span distance under the specified gauge bar. The base circle tooth thickness of a planetary gear is the base circle tooth thickness corresponding to the maximum span distance under a specified gauge bar. At this time, the meshing backlash between the input gear and the planetary gear is minimized, and the minimum backlash... satisfy: ； When the center distance between the input gear and the planetary gear is the maximum center distance The base circle tooth thickness of the input gear is the base circle tooth thickness corresponding to the minimum span distance under the specified gauge bar. The base circle tooth thickness of a planetary gear is the base circle tooth thickness corresponding to the minimum span distance under a specified gauge bar. At this time, the meshing backlash between the input gear and the planetary gear is at its maximum, and the maximum backlash is... for: ； In the formula: , This is the center distance between the input gear and the planetary gear. When finding the minimum side clearance, it is... When calculating the maximum side clearance, it is , The base circle radius of the input gear. Let be the base circle radius of the planetary gear. The base circle tooth pitch.

4. The method according to claim 3, characterized in that, The meshing backlash between the input gear and the planetary gear is between 0.04 mm and 0.105 mm.

5. The method according to claim 1, characterized in that, The needle tooth housing has circumferentially distributed grooves, and needle tooth pins are provided on the grooves. The planetary gear and the crankshaft cooperate to form an involute spline pair.

6. The method according to claim 5, characterized in that, The fit of each component determined in step 4) is as follows: the involute spline pair of the planetary gear and the crankshaft adopts a transition fit; the pin pin, cycloidal gear and pin housing groove adopt an interference fit; the crankshaft and tapered roller bearing adopt an interference fit; the outer ring assembly of the angular contact ball bearing and the pin housing adopt a transition fit; the input planetary carrier, output planetary carrier and tapered roller bearing adopt a transition fit.

7. The method according to claim 6, characterized in that, The tolerance of the spline span of the planetary gear is set to three levels, and the tolerance of the spline span of the crankshaft is also set to the corresponding three levels. The planetary gears and crankshafts in the same group have the same tolerance zone.

8. The method according to claim 6, characterized in that, The mating method determined in step 4) includes the pin tooth center not coinciding with the groove center, and the pin tooth radius being smaller than the groove radius.

9. The method according to claim 1, characterized in that, In step 6), the force on the needle roller and cage assembly in the reduction gear... Satisfies the vector relation: ； In the formula: To input the number of gear teeth, This refers to the number of teeth on a planetary gear. Eccentricity is the forming parameter of the cycloidal gear; For crankshaft rotation angle, To output the planetary carrier rotation angle, For the weight of a single cycloidal gear, This refers to the number of teeth on a cycloidal gear. For the pressure angle of the cycloidal gear, For input shaft torque, This is the center distance between the input gear and the planetary gear.

10. The method according to claim 1, characterized in that, The input planetary carrier and the output planetary carrier are positioned by internally threaded tapered pins and connected and fixed as one unit by internal hexagonal head screws.

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