A combined bearing with high efficiency, low energy consumption and long service life

Abstract

This invention discloses a high-efficiency, low-energy-consumption, and long-life combined bearing, comprising a first row of cylindrical roller bearings and a second row of thrust tapered roller bearings, which can be used in main reducer assemblies and differential assemblies. When used in a main reducer assembly, the inner ring of the combined bearing mates with the journal of the gear shaft, and the outer ring mates with the bearing housing, resulting in only rolling friction and almost zero sliding friction in the main reducer assembly. This significantly improves the transmission efficiency of the main reducer assembly, reduces energy consumption and temperature rise to an extreme level, greatly reduces NVH (noise, vibration, and harshness), effectively extends the life of gear oil and the main reducer assembly, and significantly reduces manufacturing costs and maintenance expenses. This invention significantly improves the rigidity, precision, efficiency, and lifespan of the main reducer and differential assemblies, enabling the installation of higher-horsepower engines and higher-torque retarders.

Description

Technical Field

[0001] This invention relates to the field of gear-bearing transmission mechanisms such as motor vehicle drive axles, specifically a combined bearing that is highly efficient, energy-saving, and has a long service life. Background Technology

[0002] Existing main reducer and differential assembly ( Figure 5 , 6 7, 19) are core components of the drive axle. Whether it's a passenger car or a commercial vehicle's main reducer assembly, and regardless of whether it's a cantilever support structure or a straddle support structure, existing motor vehicle main reducer assemblies have two major design flaws: First, the driving and driven bevel gear pairs cannot mesh properly due to the temperature rise of the tapered roller bearings on the gear shaft. Existing main reducer assemblies... Figure 6 The cantilever structure shown is the most common. When the main reducer assembly is working, the main gear shaft 1505 applies huge axial and radial forces to the upper tapered roller bearing 1502 and the lower tapered roller bearing 1504. In particular, the huge axial force is mainly transmitted to the large end face of the roller by the large flange of the inner ring of the lower tapered roller bearing 1504. Since there is sliding friction between the large flange of the inner ring and the large end face of the roller, the coefficient of sliding friction is 10-30 times that of rolling friction. This results in the frictional resistance, heat generation and temperature rise between the large flange of the inner ring and the large end face of the roller being much greater than the frictional resistance, heat generation and temperature rise generated by rolling friction between the roller and the inner and outer ring raceways. Therefore, the expansion of the inner ring is much greater than that of the outer ring. If the clearance between the two bearings is adjusted too small during assembly, it is easy to seize up shortly after operation. If the clearance is adjusted too large, the rigidity of the gear shaft will be too small, and the main and auxiliary gears will not mesh properly. The entire main reducer assembly and differential assembly will not work. Therefore, during assembly, it is necessary to ensure that the axial clearance of the two bearings is large enough, while also ensuring the rigidity of the gear shaft and the normal meshing of the gear pair. This is a contradictory and impossible problem to solve simultaneously. Therefore, the current solution is to add tapered roller bearing I1701 and tapered roller bearing II1702 (see...) to the gear shaft. Figure 9 and 10 This contradiction can be resolved by adjusting the distance between the tapered roller bearings I1701 and II1702. If the tapered roller bearings I1701 and II1702 are too close together, the center of force on the main bevel gear will shift when subjected to radial force. Such a large offset cannot guarantee the rigidity of the gear shaft, and the gear pair cannot mesh properly; only by increasing the distance between tapered roller bearing I1701 and tapered roller bearing II1702 can the center point of the main bevel gear shift when subjected to the same radial force. The distance, and compared to, The size has been reduced significantly. Even if this barely meets the rigidity requirements of the main bevel gear shaft and the normal meshing requirements of the gear pair, it's only a temporary solution. Existing main reducers are designed based on this principle. For example, passenger cars, primarily the Santana, have increased the distance between the cylindrical roller bearing 1402 and a set of double-row tapered roller bearings 1401 on the main reducer gear shaft (see...). Figure 5 Commercial vehicles increase the distance between the upper tapered roller bearing 1502 and the lower tapered roller bearing 1504 on the gear shaft (see...). Figure 6 Because the bearing clearance has not been reduced, vibration, noise, and temperature rise during the operation of the bearing and gear pair still exist, seriously affecting the working condition and lifespan of the main reducer and differential assembly.

[0003] For heavy-duty trucks, the existing main reducer assembly adopts a straddle-mounted support structure (see...). Figure 7 This structure adds a guide bearing 1603 to the small end of the main bevel gear, ensuring bearing support at both ends. This significantly increases support rigidity and reduces the load on the first and second row tapered bearings 1605 and 1608 on the main bevel gear shaft, improving gear meshing conditions and thus increasing the gear's load-bearing capacity compared to the cantilever type. However, the straddle-type support requires a support bearing housing 1604 on the main reducer housing to support the guide bearing 1603, making the main reducer housing 1602 structurally complex and costly to manufacture. Furthermore, although the distance between the first and second row tapered bearings 1605 and 1608 on the large end of the main bevel gear is relatively small, the axial load borne by the bearings is not reduced. The temperature rise from sliding friction between the rollers and the inner ring flange remains significant, as does the expansion of the inner ring. The clearance between the first and second row tapered bearings 1605 and 1608 must be large, resulting in significant noise and vibration from the gear pair and bearings. This design is merely a stopgap measure that does not address the root cause of the problem.

[0004] Existing differential assemblies Figure 19The two side bearings of the differential assembly also use single-row tapered roller bearings. Because the driven gear 9 applies huge radial and axial forces to the side bearings, similarly, huge sliding friction resistance is generated between the existing outer flange 1105 of the existing tapered inner ring 1106 and the large end face of the existing roller 1108. This results in the heat generation and temperature rise of the existing tapered inner ring 1106 being much higher than that of the existing tapered outer ring 1107. Since the inner hole of the existing tapered inner ring 1106 is connected to the differential housing 12, which is located at the center of the axle housing, and the entire differential assembly is connected to the main reduction housing 10 via two side bearings, the heat on the existing tapered inner ring 1106 and the differential housing 12 can only be carried to the oil sump by splashed gear oil and then transferred out through the axle housing. Conversely, because the raceways of the existing roller 1108 and the existing tapered outer ring 1107 experience rolling friction, the heat generated is much less than that of the existing roller 1108. The heat generated between the roller 1108 and the existing conical inner ring 1106, plus the fact that the outer surface area of ​​the existing conical outer ring 1107 is about twice the inner diameter of the existing conical inner ring 1106, and the outer surface of the existing conical outer ring 1107 is tightly mounted on the main reduction housing 10, with the outer surface of the main reduction housing 10 directly exposed to the air and moving with the vehicle, makes it very easy for the heat generated between the existing conical outer ring 1107 and the existing roller 1108 to be transferred away through the main reduction housing 10. This results in the temperature of the existing conical outer ring 1107 being much lower than that of the existing conical inner ring 1106, and the expansion of the existing conical inner ring 1106 being much greater than that of the existing conical outer ring 1107. In order to prevent the two side bearings and the main and driven gear pairs from jamming, a certain amount of clearance must be provided between the main and driven gear pairs when assembling the differential assembly. This leads to excessive deterioration of the NVH, stiffness, and load-bearing capacity of the main and driven gear pairs and the two side bearings.

[0005] In summary, existing motor vehicle main reducers and differential assemblies all use tapered roller bearings to bear the combined load. Due to the sliding friction between the large end face of the roller and the large flange of the inner ring, huge sliding friction resistance is generated, resulting in enormous energy consumption and heat generation. This causes the temperature rise and expansion of the inner ring of the bearing to be much greater than that of the outer ring. In order to prevent the bearing from seizing during operation, a sufficiently large clearance must be reserved for the bearing and gear pair during assembly. However, a large clearance has a significant impact on the vibration, noise, temperature rise, load and fatigue life of the main reducer and differential assembly, and also significantly shortens the life of the gear oil. Summary of the Invention

[0006] The purpose of this invention is to provide a combined bearing that is highly efficient, energy-saving, and has a long service life, comprising a first row of cylindrical roller bearings and a second row of thrust tapered roller bearings.

[0007] The first row of cylindrical roller bearings includes an inner ring, flange I, cylindrical rollers, an outer ring, and flange II. The cylindrical rollers are located between the inner and outer rings, and radially, the inner cylindrical surface of the roller contacts the inner raceway surface of the inner ring, and the outer cylindrical surface contacts the outer raceway surface of the outer ring. Axially, the right end face of the cylindrical roller faces flange II and has a clearance between it and flange II, and the left end face of the cylindrical roller faces flange I and has a clearance between it and flange I. The second row of thrust tapered roller bearings includes a shaft ring, rolling elements, and a housing ring. The rolling elements are located between the shaft ring and the housing ring. The shaft ring is positioned near flange I of the first row of cylindrical roller bearings. This combined bearing can be used in main reducer assemblies and differential assemblies.

[0008] Furthermore, the rolling element is a tapered roller or a steel ball.

[0009] When the rolling elements are tapered rollers, the second-row thrust tapered roller bearing also includes flange III. Axially, the right conical surface of the tapered roller contacts the right raceway surface of the shaft ring, and the left conical surface contacts the left raceway surface of the housing ring. Radially, the large end face of the tapered roller contacts flange III. When the rolling elements are steel balls, annular grooves are provided on the right raceway surface of the shaft ring and the left raceway surface of the housing ring, and several steel balls are embedded in the annular grooves.

[0010] Furthermore, when the combined bearing is used in the main reducer assembly, the main reducer assembly includes a main bevel gear, a flange cover, a bearing housing, a locking washer, and a gear shaft. The main bevel gear meshes with the driven bevel gear of the differential assembly and is integrally formed with the gear shaft.

[0011] The bearing housing is mounted on the outside of the gear shaft, and a flange cover and a locking gasket are respectively mounted on the side closer to the main bevel gear and the side farther from the main bevel gear. A cavity is formed between the flange cover, bearing housing, locking gasket, and gear shaft, and the combined bearing is assembled in this cavity and axially positioned by the locking gaskets and flange cover at both ends.

[0012] The first row of cylindrical roller bearings is located on the side closest to the main bevel gear.

[0013] Furthermore, the inner bore of the inner ring is interference-fitted with the cylindrical surface of the gear shaft. The gear shaft is a stepped shaft, and the stepped surface on the side of the stepped shaft closest to the gear shaft is designated as step surface II. Step surface II contacts the end face I of the inner ring. Step surface II serves as the axial positioning surface for both the inner ring and the shaft ring. The step surface I of the flange cover contacts the end face II of the flange II, and step surface I serves as the axial positioning surface for the outer ring. The outer circle of the bearing seat ring has an external thread surface. The inner bore of the bearing housing has an internal thread section at the end furthest from the flange cover. The external thread surface of the bearing seat ring mates with the internal thread section of the inner bore of the bearing housing. Tightening the external thread surface clamps the rolling element between the left raceway surface of the bearing seat ring and the right raceway surface of the shaft ring.

[0014] Furthermore, the main reducer assembly also includes a main reducer housing. The differential assembly includes a driven bevel gear, a bearing unit assembly, a differential housing, and adjusting shims. The main reducer housing is connected to the bearing housing, and the main bevel gear is located in the main reducer housing and meshes with the driven bevel gear. The driven bevel gear is connected to the differential housing, and bearing unit assemblies are fitted at both ends of the differential housing. The bearing unit assembly includes a conical inner ring, a conical outer ring, rollers, and an outer flange.

[0015] The inner conical ring has an interference fit with one end of the differential housing, and the outer conical ring fits with the main reduction gear housing. The outer conical ring has an outer flange near the driven bevel gear. The outer flange and the outer conical ring are integrally formed. Rollers are positioned between the inner and outer conical rings. The large end face of the outer conical ring contacts the adjusting shim.

[0016] The adjusting shims are used for axial positioning of the bearing unit assembly, which directly determines the meshing clearance between the main bevel gear and the driven bevel gear. Tightening the adjusting shims allows the conical outer ring to move forward within the housing bore of the main reducer housing, thereby adjusting the position of the driven bevel gear and changing the meshing clearance between the driven and main bevel gears. When the meshing clearance between the main and driven bevel gears meets the design requirements, the adjusting shims are stopped, and the meshing clearance adjustment between the main and driven bevel gears is complete.

[0017] Furthermore, when the combined bearing is used in the differential assembly, the differential assembly includes the combined bearing, the driven bevel gear, the differential housing, and adjusting shims. The main reducer assembly includes the main bevel gear and the main reducer housing. The main bevel gear is disposed within the main reducer housing and meshes with the driven bevel gear. The driven bevel gear is connected to the differential housing, and combined bearings are fitted at both ends of the differential housing. The first row of cylindrical roller bearings of the combined bearing is located near the driven bevel gear. The inner ring bore is interference-fitted with one end of the differential housing, and the outer ring fits with the main reducer housing. The second row of thrust tapered roller bearings, located away from the first row of cylindrical roller bearings, contacts the adjusting shims.

[0018] Furthermore, the outer ring and flange II of the first row of cylindrical roller bearings are an integral structure. The inner ring and flange I of the first row of cylindrical roller bearings, as well as the shaft ring of the second row of thrust tapered roller bearings, are an integral structure. Alternatively, the outer ring and flange II of the first row of cylindrical roller bearings are two separate components. The inner ring and flange I of the first row of cylindrical roller bearings are two separate parts.

[0019] Furthermore, when the rolling elements of the second-row thrust tapered roller bearing are tapered rollers, there are three possible configurations for the flange III: Configuration 1: Flange III is an integral structure with the shaft ring. Configuration 2: Flange III is an integral structure with the housing ring. Configuration 3: Both the shaft ring and the housing ring are provided with flange III, and the two flanges III are integral structures with the shaft ring and the housing ring respectively.

[0020] Another objective of this invention is to provide a combined bearing similar to the aforementioned combined bearing, but with a different position, a combined bearing with high efficiency, low energy consumption, and long service life, comprising a first row of cylindrical roller bearings and a second row of thrust tapered roller bearings.

[0021] The first row of cylindrical roller bearings includes an inner ring, flange I, cylindrical rollers, an outer ring, and flange II. The cylindrical rollers are located between the inner and outer rings, and radially, the inner cylindrical surface of the cylindrical roller contacts the inner raceway surface of the inner ring, and the outer cylindrical surface contacts the outer raceway surface of the outer ring. Axially, the right end face of the cylindrical roller faces flange II and has a gap with flange II, and the left end face of the cylindrical roller faces flange I and has a gap with flange I. The second row of thrust tapered roller bearings includes a shaft ring, rolling elements, and a housing ring. The housing ring is located near flange I of the first row of cylindrical roller bearings. The rolling elements are located between the shaft ring and the housing ring. The rolling elements are tapered rollers or steel balls. When the rolling elements are tapered rollers, the second row of thrust tapered roller bearings also includes flange III.

[0022] The combined bearing can be used in the main reducer assembly and the differential assembly.

[0023] When combined bearings are used in the main reducer assembly, the first row of cylindrical roller bearings is located away from the main bevel gear. The second row of thrust tapered roller bearings is located closer to the main bevel gear.

[0024] The technical effects of this invention are undeniable, and its beneficial effects are as follows:

[0025] A. When this invention is used in the main reducer of a motor vehicle, it has advantages such as good assembly performance, compact structure, light weight, large load capacity, almost zero clearance without preload, no need for spacers, no need for adjusting shims, and ultra-long fatigue life.

[0026] B. The structure of the device in this invention results in only rolling friction in the main reducer assembly, with virtually zero sliding friction. This significantly improves the transmission efficiency of the main reducer assembly, reduces energy consumption and temperature rise to an extreme level, NVH is greatly reduced, manufacturing costs and maintenance expenses are significantly reduced, and the lifespan of gear oil and the main reducer assembly is significantly extended. The two side bearings of the differential assembly use two bearing unit assemblies, allowing the heat generated by the bearings themselves to be quickly transferred away through the main reducer housing, greatly improving heat dissipation efficiency. The thermal expansion and contraction of the various parts of the two side bearings are basically consistent, thus ensuring that the clearance values ​​of the bearings and the driving and driven bevel gears are kept at a minimum, keeping the driving and driven bevel gears and the two side bearings in optimal working condition at all times. The maintenance-free mileage and lifespan of the vehicle's main reducer and differential assembly are significantly extended.

[0027] C. This invention significantly reduces the installation height of the main reducer assembly, which is beneficial for the overall layout of the rear-engine configuration. Furthermore, it greatly optimizes and simplifies the installation and adjustment procedures for the main bevel gear support bearing, significantly improving the rigidity, precision, efficiency, and lifespan of the main reducer and differential assembly. It enables the installation of higher-horsepower engines and higher-torque retarders, providing numerous conveniences and support for rear-engine vehicles.

[0028] D. When this invention is used in motor vehicle differentials, the differential bearings experience only rolling friction, with almost zero sliding friction, which significantly improves the transmission efficiency of the differential, reduces energy consumption and temperature rise to the extreme, and greatly reduces NVH. Attached Figure Description

[0029] Figure 1 This is a schematic diagram of a combined bearing (one-piece molding) when the outer rolling element is a tapered roller.

[0030] Figure 2 This is a schematic diagram of the main reducer and differential assembly of the motor vehicle according to the present invention;

[0031] Figure 3 This is a schematic diagram of the main reducer assembly for motor vehicles according to the present invention;

[0032] Figure 4 This is a schematic diagram of the load transfer of the combined bearing unit assembly of the present invention;

[0033] Figure 5 This is a schematic diagram of an existing passenger vehicle cantilevered final drive and differential assembly.

[0034] Figure 6 A schematic diagram of an existing cantilevered main reducer assembly for commercial vehicles;

[0035] Figure 7 A schematic diagram of an existing straddle-mounted main reducer assembly for commercial vehicles;

[0036] Figure 8 The diagram shows the force analysis of two single-row tapered roller bearings on the main gear shaft of an existing commercial vehicle; where (a) is the force analysis diagram of the second row bearing under radial load, and (b) and (c) are the force analysis diagrams of the first row bearing under radial load and axial load, respectively.

[0037] Figure 9 This is a schematic diagram showing the effect of the small distance between the two bearings in the existing cantilever main reducer assembly on the stiffness of the main bevel gear and the meshing of the main and driven gears.

[0038] Figure 10 This is a schematic diagram showing the effect of increasing the distance between the two tapered roller bearings in the existing cantilever main reducer assembly on the stiffness of the main bevel gear and the meshing of the main and driven gears.

[0039] Figure 11 A schematic diagram showing the dimensional chain that affects the thickness B0 of the adjusting shim in the existing main reducer assembly;

[0040] Figure 12 A schematic diagram of the main reducer assembly drawn in accordance with the national standard QC / T1099-2018;

[0041] Figure 13 A schematic diagram of the force analysis of an existing main reducer assembly;

[0042] Figure 14 This is a schematic diagram illustrating the transmission of radial load and axial load in one direction in the main reducer assembly of the present invention.

[0043] Figure 15 This is a schematic diagram illustrating the transmission of axial load in another direction.

[0044] Figure 16 This is a schematic diagram of the axial clearance adjustment of the first row of bearings in the main reducer assembly of the present invention;

[0045] Figure 17 For a kind of Figure 1 The diagram shows alternative solutions for the main reducer assembly.

[0046] Figure 18 This is a schematic diagram of the differential assembly side bearing section of this patent.

[0047] Figure 19 This is a schematic diagram of the existing differential assembly structure;

[0048] Figure 20 This is a schematic diagram showing the relationship between the contact angle and total frictional resistance of the differential bearing in this patent.

[0049] Figure 21 This is a schematic diagram showing the relationship between the contact angle and load of the differential side bearing (including existing bearing unit assemblies and those used in this patent).

[0050] Figure 22 As a kind Figure 18 The diagram shows alternative solutions for the differential assembly.

[0051] Figure 23 This is a schematic diagram of a combined bearing (one-piece molding) when the outer rolling element is a steel ball.

[0052] In the diagram: 1. Main bevel gear; 2. Flange cover; 3. Bearing housing; 4. Combined bearing; 5. Locking shim; 6. Gear shaft; 7. Flange I; 8. Pressure plate; 9. Driven bevel gear; 10. Main reducer housing; 11. Bearing unit assembly; 12. Differential housing; 13. Adjusting shim.

[0053] The first row of cylindrical roller bearings consists of: 41 inner ring, 411 inner ring, 412 flange I, 413 cylindrical roller, 414 outer ring, 415 flange II, 4111 end face I, 4112 inner raceway, 4131 right end face, 4132 inner cylindrical surface, 4133 left end face, 4134 outer cylindrical surface, 4141 outer raceway, 4142 outer cylindrical surface, and 4151 end face II; the second row of thrust tapered roller bearings consists of: 42 shaft ring, 421 tapered roller, 422 seat ring, 423 flange III, 424 steel ball, 425 right raceway, 4211 right tapered surface, 4221 left tapered surface, 4222 large end face, 4223 left raceway, and 4232 external thread surface.

[0054] Stepped surface I 201, internal thread section 301, inner hole 302, stepped surface II 601, cylindrical surface 602, conical inner ring 1101, conical outer ring 1102, roller 1103, outer flange 1104; existing outer flange 1105; existing conical inner ring 1106, existing conical outer ring 1107, existing roller 1108;

[0055] Double row tapered roller bearing 1401, cylindrical roller bearing 1402; adjusting shim I 1501, upper tapered roller bearing 1502, spacer I 1503, lower tapered roller bearing 1504, main gear shaft 1505;

[0056] 1601 grooved nut, 1602 main reduction housing, 1603 guide bearing, 1604 support bearing seat, 1605 first row tapered bearing, 1606 spacer II, 1607 adjusting shim II, 1608 second row tapered bearing, 1609 flange II; 1701 tapered roller bearing I, 1702 tapered roller bearing II; 1801 second row large tapered angle bearing, 1802 first row large tapered angle bearing, 1803 spacer III, 1804 adjusting shim III, 1805 compressible elastic spacer. Detailed Implementation

[0057] The present invention will be further described below with reference to embodiments, but it should not be construed that the scope of the present invention is limited to the following embodiments. Various substitutions and modifications made based on ordinary technical knowledge and common practices in the art without departing from the above-described technical concept of the present invention should be included within the scope of protection of the present invention.

[0058] Example 1: A combined bearing with high efficiency, low energy consumption and long service life, comprising a first row of cylindrical roller bearings 41 and a second row of thrust tapered roller bearings 42.

[0059] The first row of cylindrical roller bearings 41 includes an inner ring 411, flange I 412, cylindrical rollers 413, an outer ring 414, and flange II 415. The cylindrical rollers 413 are located between the inner ring 411 and the outer ring 414. In the radial direction, the inner cylindrical surface 4132 of the cylindrical roller 413 contacts the inner raceway 4112 of the inner ring 411, and the outer cylindrical surface 4134 contacts the outer raceway 4141 of the outer ring 414. In the axial direction, the right end face 4131 of the cylindrical roller 413 faces flange II 415 and has a gap between it and flange II 415, while the left end face 4133 of the cylindrical roller 413 faces flange I 412 and has a gap between it and flange I 412.

[0060] The second row of thrust tapered roller bearings 42 includes a shaft ring 421, rolling elements, and a housing ring 423. The rolling elements are located between the shaft ring 421 and the housing ring 423. The shaft ring 421 is disposed on the side of the flange I 412 near the first row of cylindrical roller bearings 41.

[0061] The combined bearing 4 can be used in the main reducer assembly and the differential assembly.

[0062] Example 2: The main structure of this example is the same as that of Example 1. Further, the rolling element is a tapered roller 422 or a steel ball 425. When the rolling element is a tapered roller 422, the second row of thrust tapered roller bearings 42 also includes a flange III 424. In the axial direction, the right conical surface 4221 of the tapered roller 422 contacts the right raceway 4211 surface of the bearing ring 421, and the left conical surface 4222 contacts the left raceway 4231 surface of the seat ring 423. In the radial direction, the large end face 42523 of the tapered roller 422 contacts the flange III 424. When the rolling element is a steel ball 425, annular grooves are provided on the right raceway 4211 surface of the bearing ring 421 and the left raceway 4231 surface of the seat ring 423, and several steel balls 425 are embedded in the annular grooves.

[0063] Example 3: The main structure of this example is the same as any one of Examples 1 and 2. Further, when the combined bearing 4 is used in the main reducer assembly, the main reducer assembly includes a main bevel gear 1, a flange cover 2, a bearing seat 3, a locking washer 5, and a gear shaft 6. The main bevel gear 1 meshes with the driven bevel gear 9 of the differential assembly and is an integral part of the gear shaft 6.

[0064] The bearing housing 3 is mounted on the outside of the gear shaft 6, and a flange cover 2 and a locking gasket 5 are respectively mounted on the side closer to the main bevel gear 1 and the side farther away from the main bevel gear 1. A cavity is formed between the flange cover 2, the bearing housing 3, the locking gasket 5, and the gear shaft 6. The combined bearing 4 is mounted in this cavity and is axially positioned by the locking gaskets 5 and the flange cover 2 at both ends. The first row of cylindrical roller bearings 41 is located on the side closer to the main bevel gear 1.

[0065] Example 4: The main structure of this example is the same as that of Example 3. Further, the inner hole of the inner ring 411 is interference-fitted with the cylindrical surface 602 of the gear shaft 6. The gear shaft 6 is a stepped shaft, and the stepped surface near the gear shaft 6 is designated as step surface II 601. Step surface II 601 contacts the end face I 4111 of the inner ring 411. Step surface II 601 serves as the axial positioning surface for the inner ring 411 and the shaft ring 421.

[0066] The stepped surface I201 of the flange cover 2 contacts the end face II4151 of the flange II415, and the stepped surface I201 is the axial positioning surface of the outer ring 414.

[0067] The outer circle of the bearing seat 423 is an external threaded surface 4232. The inner hole 302 of the bearing housing 3 is provided with an internal threaded section 301 at the end away from the flange cover 2. The external threaded surface 4232 of the bearing seat 423 mates with the internal threaded section 301 of the inner hole 302 of the bearing housing 3.

[0068] Tighten the external thread surface 4232, and the left raceway 4231 of the seat ring 423 and the right raceway 4211 of the shaft ring 421 clamp the rolling element. When the rolling element is a tapered roller 422, tighten the external thread surface 4232, and the left raceway 4231 of the seat ring 423 contacts the left tapered surface 4222 of the tapered roller 422, the right tapered surface 4221 of the tapered roller 422 contacts the right raceway 4211 of the shaft ring 421, and the large end face 4223 of the tapered roller 422 contacts the flange III 424. The flange III 424 is the radial positioning surface of the large end face 4223 of the tapered roller 422.

[0069] Example 5: The main structure of this example is the same as any one of Examples 3-4. Further, the main reducer assembly also includes a main reducer housing 10. The differential assembly includes a driven bevel gear 9, a bearing unit assembly 11, a differential housing 12, and adjusting shims 13. The main reducer housing 10 is connected to the bearing seat 3. The main bevel gear 1 is located in the main reducer housing 10 and meshes with the driven bevel gear 9. The driven bevel gear 9 is connected to the differential housing 12, and the bearing unit assemblies 11 are fitted at both ends of the differential housing 12. The bearing unit assembly 11 includes a conical inner ring 1101, a conical outer ring 1102, a roller 1103, and an outer flange 1104.

[0070] The inner conical ring 1101 has an interference fit with one end of the differential housing 12, and the outer conical ring 1102 fits with the main reduction gear housing 10. The outer conical ring 1102 has an outer flange 1104 near the driven bevel gear 9. The outer flange 1104 and the outer conical ring 1102 are integrally formed. The roller 1103 is disposed between the inner conical ring 1101 and the outer conical ring 1102. The large end face of the outer conical ring 1102 contacts the adjusting shim 13.

[0071] The adjusting shim 13 is used to axially position the bearing unit assembly 11, and the axial position of the bearing unit assembly 11 directly determines the meshing clearance between the main bevel gear 1 and the driven bevel gear 9.

[0072] Tightening the adjusting shim 13 allows it to push the conical outer ring 1102 forward within the housing hole of the main reducer housing 10, thereby adjusting the position of the driven bevel gear 9 and changing the meshing clearance between the driven bevel gear 9 and the main bevel gear 1. When the meshing clearance between the main bevel gear 1 and the driven bevel gear 9 meets the design requirements, stop adjusting the adjusting shim 13; the meshing clearance between the main bevel gear 1 and the driven bevel gear 9 is now adjusted.

[0073] Example 6: The main structure of this example is the same as any one of Examples 1 to 4. Further, when the combined bearing 4 is used in the differential assembly, the differential assembly includes the combined bearing 4, the driven bevel gear 9, the differential housing 12, and the adjusting shim 13.

[0074] The main reducer assembly includes a main bevel gear 1 and a main reducer housing 10. The main bevel gear 1 is disposed within the main reducer housing 10 and meshes with a driven bevel gear 9. The driven bevel gear 9 is connected to a differential housing 12, and both ends of the differential housing 12 are fitted with combined bearings 4. The first row of cylindrical roller bearings 41 of the combined bearings 4 is located near the driven bevel gear 9. The inner ring 411 of the first row of cylindrical roller bearings 41 has an interference fit with one end of the differential housing 12, and the outer ring 414 fits with the main reducer housing 10.

[0075] The side of the second row of thrust tapered roller bearings 42 away from the first row of cylindrical roller bearings 41 contacts the adjusting shim 13.

[0076] Example 7: The main structure of this example is the same as any one of Examples 1 to 6. Further, the outer ring 414 and flange II 415 of the first row of cylindrical roller bearings 41 are an integral structure. The inner ring 411 and flange I 412 of the first row of cylindrical roller bearings 41, and the shaft ring 421 of the second row of thrust tapered roller bearings 42 are also an integral structure.

[0077] Example 8: The main structure of this example is the same as any one of Examples 1 to 6. Further, the outer ring 414 and flange II 415 of the first row of cylindrical roller bearings 41 are two separate components. The inner ring 411 and flange I 412 of the first row of cylindrical roller bearings 41 are two separate parts.

[0078] Example 9: The main structure of this example is the same as any one of Examples 2 to 8. Furthermore, when the rolling element of the second row of thrust tapered roller bearings 42 is a tapered roller 422, there are three possible arrangements for the flange III 424:

[0079] Setting method 1: The flange Ⅲ424 and the shaft ring 421 are integrated into one structure.

[0080] Option 2: The flange Ⅲ424 and the seat ring 423 are integrated into one structure.

[0081] Method 3: Both the shaft ring 421 and the seat ring 423 are provided with a retaining edge Ⅲ 424, and the two retaining edges Ⅲ 424 are integral with the shaft ring 421 and the seat ring 423 respectively.

[0082] Example 10: A combined bearing with high efficiency, low energy consumption and long service life, comprising a first row of cylindrical roller bearings 41 and a second row of thrust tapered roller bearings 42.

[0083] The first row of cylindrical roller bearings 41 includes an inner ring 411, flange I 412, cylindrical rollers 413, an outer ring 414, and flange II 415. The cylindrical rollers 413 are located between the inner ring 411 and the outer ring 414. In the radial direction, the inner cylindrical surface 4132 of the cylindrical roller 413 contacts the inner raceway 4112 of the inner ring 411, and the outer cylindrical surface 4134 contacts the outer raceway 4141 of the outer ring 414. In the axial direction, the right end face 4131 of the cylindrical roller 413 faces flange II 415 and has a gap between it and flange II 415, while the left end face 4133 of the cylindrical roller 413 faces flange I 412 and has a gap between it and flange I 412.

[0084] The second row of thrust tapered roller bearings 42 includes a shaft ring 421, rolling elements, and a housing ring 423. The housing ring 423 is disposed on the side near the flange I 412 of the first row of cylindrical roller bearings 41. The rolling elements are located between the shaft ring 421 and the housing ring 423. The rolling elements are tapered rollers 422 or steel balls 425. When the rolling elements are tapered rollers 422, the second row of thrust tapered roller bearings 42 also includes a flange III 424.

[0085] The combined bearing 4 can be used in both the main reducer assembly and the differential assembly. When the combined bearing 4 is used in the main reducer assembly, the first row of cylindrical roller bearings 41 is located away from the main bevel gear 1. The second row of thrust tapered roller bearings 42 is located closer to the main bevel gear 1.

[0086] Example 11: The main structure of this example is the same as any one of Examples 1 to 10, and further includes a main bevel gear 1, a combined bearing 4, a bearing housing 3, a flange cover 2, and a gear shaft 6. The gear shaft 6 and the main bevel gear 1 are an integral structure. The combined bearing 4 is mounted on the gear shaft 6. The combined bearing 4 includes a first row of cylindrical roller bearings 41 and a second row of thrust tapered roller bearings 42.

[0087] The first row of cylindrical roller bearings 41 is located on the side near the main bevel gear 1.

[0088] The first row of cylindrical roller bearings 41 includes an inner ring 411, cylindrical rollers 413, and an outer ring 414. The outer ring 414 includes a flange II 415, an outer raceway 4141, an end face II 4151, and an outer cylindrical surface 4142. The inner ring 411 includes an end face I 4111, an inner raceway 4112, a flange I 412, and an inner bore. The cylindrical roller 413 is located between the outer ring 414 and the inner ring 411. In the radial direction, the outer cylindrical surface 4134 of the cylindrical roller 413 contacts the outer raceway 4141 of the outer ring 414, and the inner cylindrical surface 4132 of the cylindrical roller 413 contacts the inner raceway 4112 of the inner ring 411. The right end face 4131 of the cylindrical roller 413 is adjacent to the flange II 415 of the outer ring 414, and the left end face 4133 of the cylindrical roller 413 is adjacent to the flange I 412 of the inner ring 411. The cylindrical roller 413 is positioned radially and axially between the outer ring 414 and the inner ring 411 and cannot move radially or axially.

[0089] The second row of thrust tapered roller bearings 42 includes a shaft ring 421, tapered rollers 422, and a housing ring 423. The housing ring 423 includes a left raceway 4231, a flange Ⅲ 424, and an external thread surface 4232, which mates with the internal thread section 301 of the bearing housing 3's inner bore. The shaft ring 421 includes a right raceway 4211, and the shaft ring 421 and the inner ring 411 of the first row of cylindrical roller bearings 41 are integrally formed. The tapered roller 422 is located between the seat ring 423 and the shaft ring 421. Axially, the right tapered surface 4221 of the tapered roller 422 contacts the right raceway 4211 of the shaft ring 421, and the left tapered surface 4222 of the tapered roller 422 contacts the left raceway 4231 of the seat ring 423. Radially, the large end face 4223 of the tapered roller 422 contacts the flange III 424 of the seat ring 423. The tapered roller 422 is positioned radially and axially between the seat ring 423 and the shaft ring 421 and cannot move radially or axially.

[0090] The end face I 4111 of the inner ring 411 of the first row of cylindrical roller bearings 41 contacts the stepped surface II 601 of the gear shaft 6. The stepped surface II 601 is both the axial positioning surface of the inner ring 411 of the first row of cylindrical roller bearings 41 and the axial positioning surface of the shaft ring 421 of the second row of thrust tapered roller bearings 42. The inner hole of the inner ring 411 is interference-fitted onto the cylindrical surface 602 of the gear shaft 6. The outer circular surface 4142 of the outer ring 414 of the first row of cylindrical roller bearings 41 is installed in the inner hole 302 of the bearing housing 3. The end face II 4151 contacts the stepped surface I 201 of the flange cover 2. The stepped surface I 201 is the axial positioning surface of the outer ring 414.

[0091] The external thread surface 4232 of the seat ring 423 of the second-row thrust tapered roller bearing 42 mates with the internal thread section 301 of the bearing housing 3. When the external thread surface 4232 is tightened, the left raceway 4231 contacts the left conical surface 4222, and the right conical surface 4221 contacts the right raceway 4211 of the bearing ring 421. The flange III 424 of the seat ring 423 of the second-row thrust tapered roller bearing 42 contacts the large end face 4223 of the tapered roller 422. The flange III 424 serves as the radial locating surface of the large end face 4223 of the tapered roller 422.

[0092] The shaft ring 421 and inner ring 411 can also be two separate parts. The outer ring 414 and flange II 415 can be an integral structure or two separate parts. The flange III 424 of the second row thrust tapered roller bearing 42 can be on the housing ring 423, on the shaft ring 421, or both the housing ring 423 and the shaft ring 421 can have flanges.

[0093] The positions of the first row of cylindrical roller bearings 41 and the second row of thrust tapered roller bearings 42 can also be interchanged.

[0094] The differential assembly mainly includes a driven bevel gear 9, a main reduction gear housing 10, a bearing unit assembly 11, a differential housing 12, and adjusting shims 13. The meshing clearance between the main bevel gear 1 and the driven bevel gear 9 is adjusted by the adjusting shims 13. Tightening the adjusting shims 13 pushes the outer ring 1102 forward in the housing hole of the main reduction gear housing 10. When the meshing clearance between the main bevel gear 1 and the driven bevel gear 9 reaches the design requirements, the adjustment of the adjusting shims 13 is stopped, and the meshing clearance between the main bevel gear 1 and the driven bevel gear 9 is adjusted.

[0095] Example 12: The main structure of this example is the same as any one of Examples 1 to 11. Furthermore, this example assembles two single-row tapered roller bearings on the existing gear shaft (see Example 1). Figure 13 As is well known, when the main bevel gear of the existing main reducer assembly carries the radial load... When the load is transferred to the two single-row tapered roller bearings on the gear shaft, the first and second rows of bearings respectively bear the load. and The radial load of the size, but due to the structural characteristics of tapered roller bearings, and Additional derived axial loads were generated at the contact surfaces of the inner and outer rings and the rollers, respectively. , , and , , When the main bevel gear applies an axial force to the inner ring of the first row of bearings At this time, the axial force is transmitted from the inner raceway and the large flange to the rollers, and then from the rollers to the outer ring. This generates significant sliding friction resistance between the large end face of the rollers and the large flange, and also generates significant, additional, harmful, and derived radial rolling friction resistance between the rollers and the inner and outer raceways. Similar to the radial load analysis of the bearing above, the outer ring, in addition to bearing the axial force transmitted from the inner ring and rollers, also experiences significant rolling friction resistance. In addition, friction was also generated between the three friction pairs. , and As can be seen from the above analysis, the radial load is derived from the two rows of bearings on the gear shaft of the existing main reducer assembly. Under the action of the radial and axial loads of the main bevel gear, the thermal expansion of the inner ring is much greater than that of the outer ring. In order to prevent the bearings from seizing, there must be a certain amount of clearance between the two rows of bearings.

[0096] One advantage of the main reducer assembly of the present invention is the elimination of derived loads (see...). Figure 4 That is, there is only rolling friction inside the combined bearing 4, and the main bevel gear transmits the radial load. The material is transferred to the first row of cylindrical roller bearings 41, and then from the first row of cylindrical roller bearings 41 to the bearing housing. The specific transfer path is shown in [reference needed]. Figure 14 As shown in ABC, that is, radial load. The load is transmitted from the main bevel gear 1 to the gear shaft 6, then from the gear shaft 6 to the inner ring 411, then from the inner ring 411 to the cylindrical roller 413, then from the cylindrical roller 413 to the outer ring 414, and finally from the outer ring 414 to the bearing housing 3. Since the bearing housing 3 is fixed to the axle housing, the radial load is ultimately distributed. The purpose is to transfer the load to the stationary bridge housing. Since the first row of cylindrical roller bearings 41 only experiences rolling friction when bearing radial loads, there is no derived load. The main bevel gear transfers the axial load. The thrust is transferred to the second row of tapered roller bearings 42 (see...) Figure 4 Then, the force is transmitted from the second row of thrust tapered roller bearings 42 to the bearing housing. The specific transmission path is shown in [reference needed]. Figure 14 As shown in DEFGH, this represents the axial load. The axial load is transmitted from the main bevel gear 1 to the gear shaft 6, then from the gear shaft 6 to the inner ring 411, then from the inner ring 411 to the shaft ring 421, then from the shaft ring 421 to the tapered roller 422, then from the tapered roller 422 to the housing ring 423, and finally from the housing ring 423 to the bearing housing 3. Since the bearing housing 3 is fixed to the axle housing, the axial load is ultimately distributed. The purpose is to transfer the load to the stationary bridge housing. Since the second-row thrust tapered roller bearing 42 experiences almost only rolling friction when bearing axial load, there is no derived load.

[0097] When the vehicle reverses, the axial load on the main bevel gear is... like Figure 15 As shown, the specific transmission path is shown in the diagram as ABCDDEFGHKLMNP, which represents the axial load. The load is transmitted from the main bevel gear 1 to the gear shaft 6, then from the gear shaft 6 to the pressure plate 8, then from the pressure plate 8 to the flange I 7, then from the flange I 7 to the inner ring 411, then from the inner ring 411 to the flange I 412, then from the flange I 412 to the cylindrical roller 413, then from the cylindrical roller 413 to the flange II 415, then from the flange II 415 to the outer ring 414, then from the outer ring 414 to the flange cover 2, and finally from the flange cover 2 to the bearing housing 3. Since the bearing housing 3 is fixed to the axle housing, the radial load is ultimately distributed. The purpose is to transmit the signal to the stationary bridge housing.

[0098] Since the bearing clearance significantly affects bearing load, lifespan, temperature rise, noise, and vibration, this embodiment's innovative invention eliminates the sliding friction pair. The first row of cylindrical roller bearings 41 and the second row of thrust tapered roller bearings 42 achieve pure rolling friction operation. The temperature rise of each component in the two rows of bearings is almost identical. As a result, the assembly clearance of the two rows of bearings can be significantly compressed to 0 or a very small positive clearance range, greatly improving the rigidity and load-bearing capacity of the main reducer assembly, significantly improving the transmission efficiency of the main reducer assembly, significantly reducing energy consumption and temperature rise, and significantly reducing NVH. Due to the compact structure and small size of the combined bearing 4, manufacturing costs can be significantly reduced. In addition, maintenance costs can be significantly reduced, and the lifespan of gear oil and the main reducer assembly can be significantly extended.

[0099] Example 13: The main structure of this example is the same as any one of Examples 1 to 12. Furthermore, this example quantitatively compares and measures the transmission efficiency, energy consumption, and temperature rise of the existing main reducer assembly and the inventive main reducer assembly. The most common light commercial vehicle main reducer assembly is used as an example to illustrate the problem. The existing commercial vehicle main reducer has two tapered roller bearings installed on the gear shaft (see...). Figure 6 The bearings closest to the large end of the main bevel gear (first row) are generally 31300 series tapered roller bearings (see...). Figure 6 Part No. 1511), located away from the large end (second row) of the main bevel gear, generally uses 32300 series tapered roller bearings (see Part No. 1511). Figure 6 (Item No. 1509). The contact angle of the 31300 series tapered roller bearing is 29°, and the contact angle of the 32300 series tapered roller bearing is 13°. For specific angle parameters, please refer to Table 1.

[0100] Table 1. List of contact angles for roller bearings

[0101]

[0102] 1. Load analysis of each component in the two single-row tapered roller bearings of the existing main reducer. See [reference needed]. Figure 8 The existing main bevel gear applies pressure to the inner ring of the first row of bearings. radial force and The axial force only applies to the inner ring of the second row of bearings. The radial force, in a two-row bearing, is decomposed into vertical and horizontal forces by the rollers on the inner ring, which are then applied to the raceways and flanges of the outer ring. The rollers are in equilibrium under the action of these three forces: the inner and outer raceways, and the flanges. Therefore:

[0103] (1) See Figure 8 (b) The first row of bearings bears the load. The radial force, the roller is in equilibrium under the action of the inner and outer raceways and the large flange force, so we have:

[0104] (1)

[0105] (2)

[0106] in, = , = , = , = , = , = Substituting these 6 equations into formulas (1) and (2) and simplifying, we get:

[0107] = (sin × + × ) / ( + (3)

[0108] = ( × - × ) / ( × + × (4)

[0109] The front , and Substituting the angle values ​​into (3) and (4), we obtain:

[0110] = (sin29°×sin21.5°+cos29°×cos21.5°) / (sin21°×sin21.5°+cos21°×cos21.5°)= = / cos = / cos29°=1.14

[0111] = (sin29°×cos21°-cos29°×sin21°) / (sin21°×sin21.5°+cos21°cos21.5°)=0.14 =0.14 / cos =0.14 / cos29°=0.16

[0112] (2) See Figure 8 (c) The first row of bearings bears the load. The axial force, the roller is in equilibrium under the action of the inner and outer raceways and the large flange force, so we have:

[0113] - - =0 (5)

[0114] - + =0 (6)

[0115] in, = × , = × , = × , = × , = × 1, = × 1,

[0116] Substituting these 6 equations into equations (5) and (6) and simplifying, we get:

[0117] = (sin × + × ) / ( 1+ (7)

[0118] = ( × - × ) / ( × + × (8)

[0119] The front , and Substituting the angle values ​​into (7) and (8), we obtain:

[0120] = (sin29°×sin21.5°+cos29°×cos21.5°) / (sin21°×sin21.5°+cos21°×cos21.5°)= = / = / sin29°=2.06

[0121] = (sin29°×cos21°-cos29°×sin21°) / (sin21°×sin21.5°+

[0122] cos21°cos21.5°) = 0.14 =0.14 / =0.14 / sin29°=0.29

[0123] (3) See Figure 8 (a) The second row of bearings bears the load. The axial force, the roller is in equilibrium under the action of the inner and outer raceways and the large flange force, so we have:

[0124] (9)

[0125] (10)

[0126] in, = , = , = , = , = , = Substituting these 6 equations into formulas (9) and (10) and simplifying them, we get:

[0127] = ( × + × ) / ( × + × (11)

[0128] = ( × - × ) / ( × + × (12)

[0129] The front , and Substituting the angle values ​​into (11) and (12), we get:

[0130] = (sin13°×sin9.5°+cos13°×cos9.5°) / (sin9°×sin9.5°+cos9°×cos9.5°)= = / = / cos13°=1.03

[0131] = (sin13°×cos9°-cos13°×sin9°) / (sin9°×sin9.5°+cos9°×cos9.5°)=0.07 =0.07 / =0.07 / cos13°=0.07

[0132] 2. Load analysis of individual components of the combined bearing in the main reducer assembly. (See also...) Figure 4 The main bevel gear 1 receives the radial load. The load is transmitted to the gear shaft 6, which in turn transmits it to the inner ring 411. The inner ring 411 then transmits it to the cylindrical roller 413, which in turn transmits it to the outer ring 414. Pure rolling friction loads exist between the cylindrical roller 413 and the inner and outer raceways. Similarly, the main bevel gear 1 will bear the axial load. The load is transmitted to the gear shaft 6, which in turn transmits it to the inner ring 411. The inner ring 411 then transmits it to the shaft ring 421, which in turn transmits it to the tapered roller 422. The tapered roller 422 then transmits it to the seat ring 423. There is pure rolling friction load between the tapered roller 422, the shaft ring 421, and the seat ring 423. In addition, sliding frictional resistance is generated between the tapered roller 422 and the retaining flange Ⅲ 424 of the seat ring 423 (see...). Figure 2 and 3 Since the cone angle of the tapered roller 422 is relatively small, the radial sliding friction load between the tapered roller 422 and the flange III 424 of the seat ring 423 is calculated to be 0.03. .

[0133] 3. Based on the calculations in sections 1 and 2 above, the sliding friction load and rolling friction load borne by each component of the existing and the present invention bearings are listed below.

[0134] Table 2. Sliding friction load and rolling friction load borne by bearings

[0135]

[0136] The existing main reducer's main bevel gear carries the radial load. Assigned to the first column of bearings Assigned to the second column of bearings Due to the radial load of the second row of bearings relative to the main bevel gear Since the points are at a certain distance, the total load and bending moment borne by the existing two rows of bearings are greater than (assuming) the load and bending moment borne by only the first row of bearings. Furthermore, because the first row of bearings is closer to the radial load of the main bevel gear... At the stress points, the radial load borne by the first row of bearings is much greater than that of the second row, and for the same radial load... Whether the load is borne by one bearing or by two paired bearings, the difference in frictional resistance, heat generation, temperature rise, and thermal expansion within the bearing is negligible. Therefore, for ease of analysis, we can simplify the problem so that the radial load of the main bevel gear is borne solely by the first row of bearings. = Therefore, the sum of the rolling friction loads generated between the rollers and the inner and outer raceways of the existing first-row bearing is 1.14. +1.14 =2.28 The main reducer assembly is relatively simple, with the sum of the rolling friction loads generated between the cylindrical rollers and the inner and outer raceways being 2. In comparison, the innovative invention of the main reducer assembly combined bearing in bearing radial loads demonstrates superior performance. At that time, the rolling friction load borne by each rolling friction pair of the first row of cylindrical roller bearings 41 was only 88% of that of existing bearings (2 / 2.28 The corresponding frictional resistance, heat generation, energy consumption, and temperature rise are somewhat advantageous compared to existing bearings; however, the existing first-row bearings still generate 0.16 between the large end face of the roller and the large flange of the inner ring. For the same bearing, the sliding friction load is calculated as follows: since the internal sliding friction coefficient is 10-20 times that of the rolling friction coefficient, taking the median value of 15 times, the existing first bearing under radial load... When converting the sliding friction load between the large end face of the roller and the large flange into a rolling friction load, it becomes 15 × 0.16. =2.4 , which is the rolling friction load between the roller and the inner and outer raceways, 1.14. More than twice that of the existing first-row bearings indicates that the sliding friction resistance generated between the large end face of the roller and the large flange of the inner ring is greater than the sum of the rolling friction resistance generated between the roller and the inner and outer raceways. Correspondingly, the friction resistance, heat generation, and temperature rise between the large end face of the roller and the large flange of the inner ring are also greater than the sum of those between the roller and the inner and outer raceways. Because the large flange is on the inner ring, in... Under the action of , the ratio of the rolling friction load on the inner ring to the rolling friction load on the outer ring is: (2.4) +1.14 ) / 1.14 =3.1, indicating that the rolling friction load of the inner ring is more than three times that of the outer ring. Therefore, the frictional resistance, heat generation, and temperature rise of the inner ring of the existing first-row bearing are more than three times that of the outer ring. In addition, the surface area of ​​the outer circle of the outer ring is about twice that of the inner hole. The outer ring is connected to the outside world through the bearing housing, and the heat on the outer ring can be easily dissipated through the bearing housing. With the movement of the vehicle and the convection of air in the bearing housing, the heat dissipation effect is even more significant. Conversely, since the inner ring of the first-row bearing is mounted on the gear shaft, which is located in the center of the main reducer, the heat on the inner ring can only be thrown into the oil sump by splashed gear oil and then transferred out through the axle housing. The heat dissipation effect is extremely poor, so the temperature rise of the inner ring is much greater than that of the outer ring. The radial load borne by the first-row cylindrical roller bearing 41 of this combined bearing unit assembly. Rolling friction loads are generated at the inner and outer raceways, respectively. The total rolling friction load is 2 The existing first-row bearings under radial load Under the influence of this, the total rolling friction load generated inside is 1.14. +1.14 +2.4 =4.68 In comparison, the rolling friction load of the combined bearing unit assembly is only 43% of that of the existing bearings (2 / 4.68 The main advantage of the combined bearing unit assembly is that the first row of cylindrical roller bearings in the combined bearing unit assembly does not generate derived sliding friction loads, and the rolling friction loads of the inner and outer rings are the same. The rolling friction resistance, heat generation, temperature rise and thermal expansion are also the same. In comparison, the superiority of the combined bearing unit assembly is evident.

[0137] As shown in the table above, the main bevel gear carries the axial load. The sum of the rolling friction loads applied to the first row of bearings, between the rollers and the inner and outer raceways, is 2.06. +2.06 =4.12 A 0.29mm diameter was generated between the large end face of the roller and the large flange of the inner ring. The sliding friction load; for the combined bearing on the gear shaft of the main reducer assembly of the inventive invention, the main bevel gear will transfer the axial load. The sum of the rolling friction loads generated between the tapered rollers 422, the shaft ring 421, and the housing ring 423, when applied to the second row of thrust tapered roller bearings 42, is 2. A 0.03mm diameter was generated between the large end face 4223 of the tapered roller 422 and the retaining flange Ⅲ 424 of the seat ring 423. The sliding friction load; in comparison, the total axial rolling friction load inside the bearing of this patent is only 49% (2) of that of existing bearings. / 4.12 The axial sliding friction load is only 8% (0.03 / 0.29) of that of existing bearings. Then, converting the sliding friction load of the two types of bearings into rolling friction load, the existing bearing load is 15 × 0.29. =4.35 The patented bearing is 15×0.03. =0.45 Then, the total rolling friction load inside the two types of bearings is calculated, and the load on the existing bearing is 2.06. +2.06 +4.35 =8.47 The patented bearing is + +0.45 =2.45 The total load of the patented bearing is only 29% (2.45) of that of existing bearings. / 8.47 From this, we can draw two conclusions: First, under the same axial load... At the same time, the frictional resistance, heat generation, energy consumption, and temperature rise of this patented bearing are significantly reduced compared to existing bearings; secondly, it is related to the radial load borne by the previous bearing. In comparison, existing bearings can withstand axial loads. At that time, the load generated inside the bearing, whether it is rolling friction load, sliding friction load, or axial load, The ratios have increased significantly, for example, radial load. 2.28 was produced only inside the first row of bearings. Rolling friction load and 0.16 Sliding friction load, while axial load However, a 4.12mm diameter was generated inside the first row of bearings. Rolling friction load and 0.29 Sliding friction load, this is because the first row of bearings is subjected to axial load. When this happens, a huge sliding friction load and a huge, additional, harmful, and derived radial load are generated inside the bearing. On the one hand, when existing bearings are subjected to axial loads, the overall heat generation, energy consumption, and temperature rise of the bearing are greater than when subjected to radial loads; on the other hand, the heat generation, energy consumption, and temperature rise of the inner ring are also greater than those of the outer ring.

[0138] In summary, this embodiment shows that because the combined bearings only have rolling friction, the rolling friction resistance, heat generation, energy consumption, and temperature rise of the main reducer assembly of this inventive invention are greatly reduced. As a result, the clearance of the combined bearings can be controlled to be very small. Depending on the load of the motor vehicle, as long as the clearance of the two rows of bearings after assembly is between 0 and a very small positive clearance, the great superiority of this inventive invention is theoretically proven. In practice, many well-known foreign bearing companies have long conducted experimental verification of the friction coefficients of various bearings. Since the sliding friction coefficient and rolling friction coefficient inside tapered roller bearings cannot be calculated separately through experiments and practice, only the overall friction coefficient of various bearings can be given (see Table 3). As can be seen from the table, when the friction coefficients of the same type of bearings from four well-known foreign bearing companies are added together and averaged, the friction coefficient of cylindrical roller bearings is only about half that of tapered roller bearings. This indicates that the friction coefficient of tapered roller bearings is about doubled simply because they have one more large flange than cylindrical roller bearings. This verifies that although the sliding friction load between the large end face of the roller and the large flange of the inner ring of tapered roller bearings is small, the sliding friction resistance is huge, which is the root cause of the much larger heat generation, temperature rise, and thermal expansion of the inner ring compared to the outer ring. Since the contact angle of the second row of thrust tapered roller bearings in this patent is about 4°, which is close to that of thrust cylindrical roller bearings with pure rolling friction, together with the first row of cylindrical roller bearings with pure rolling friction, this directly verifies the great superiority of this invention.

[0139] Table 3. Comparison of friction coefficients for different types of bearings

[0140]

[0141] Example 14: The main structure of this example is the same as any one of Examples 1 to 13. Furthermore, since the outer ring 414 of the first row of cylindrical roller bearings 41 has only one flange II 415 and the inner ring 411 has only one flange I 412, if the distance between these two flanges is much larger than or the same as the length of the cylindrical roller 413, it will affect the normal operating condition of the first row of cylindrical roller bearings 41. Therefore, the axial clearance of the first row of cylindrical roller bearings 41 must be controlled (see...). Figure 1 and 16 The process of controlling the axial clearance of the first row of cylindrical roller bearings 41 is as follows: First, install the combined bearing 4 into the bearing housing 3. The external thread surface 4232 of the housing ring 423 is slightly screwed into the internal thread section 301 of the bearing housing 3. Use a press to assemble the bearing housing 3 and the combined bearing 4 onto the gear shaft 6. When pressed into place, the stepped surface II 601 on the gear shaft 6 and the end face I 4111 of the inner ring 411 are tightly fitted together. The stepped surface II 601 is the axial clearance of the combined bearing 4 on the gear shaft. Positioning surface; continue tightening the external thread surface 4232 of the seat ring 423, causing the seat ring 423 and the bearing housing 3 to move towards each other. On one hand, the left raceway 4231 of the seat ring 423 drives the tapered roller 422 to move to the right along the EF direction. When the right tapered surface 4221 moves to contact the right raceway 4211 of the shaft ring 421, since the shaft ring 421 and the inner ring 411 are an integral structure, the clearance of the second row of thrust tapered roller bearings 42 is 0. 42 is in optimal working condition; on the other hand, the bearing housing 3 drives the stepped surface I201 of the flange cover 2 to move to the left, the stepped surface I201 pushes the end face II4151 of the outer ring 414 to move to the left, and the flange II415 of the outer ring 414 pushes the right end face 4131 and the left end face 4133 of the cylindrical roller 413 to move to the left along the axial direction of BCD. When the left end face 4133 of the cylindrical roller 413 moves to contact the flange I412 of the inner ring 411, the first row of cylindrical rollers... When the axial clearance of the sub-bearing 41 is 0, the external thread surface 4232 of the seat ring 423 is slightly loosened in the opposite direction. This gives the first row of cylindrical roller bearings 41 axial clearance, ensuring that when the first row of cylindrical roller bearings 41 is working, the left end face 4133 of the cylindrical roller 413 does not contact the flange I 412 of the inner ring 411, and the right end face 4131 of the cylindrical roller 413 does not contact the flange II 415 of the outer ring 414. In this way, the first row of cylindrical roller bearings 41 is in the best working condition.

[0142] When both the radial and axial clearances of the combined bearing 4 are within their optimal ranges, tighten the locking shim 5. If it cannot be tightened further, loosen it a few turns in the opposite direction. Then, use screws to secure the locking shim 5 and the bearing race 423 together (see...). Figure 2 At this point, the positioning and clamping of the combined bearing 4 is completed, and the combined bearing 4 is in its optimal working condition.

[0143] Example 15: The main structure of this example is the same as any one of Examples 1 to 14. Furthermore, a second major design flaw in existing motor vehicle main reducer assemblies is that spacers and adjusting shims must be installed between the inner rings of the two single-row tapered roller bearings on the main reducer gear shaft. To precisely control the stiffness of the main reducer gear and the normal meshing of the gear pair, spacers and adjusting shims (cantilever structure) must be installed between the inner rings of the two tapered roller bearings on the existing gear shaft. Figure 6 Spacer I1503, adjusting shim I1501; straddle-type structure Figure 7 The spacer II1606 and adjusting shim II1607 control the axial clearance by adjusting the width of the adjusting shims, thereby controlling the stiffness of the main reducer gears and the normal meshing of the gear pair. However, controlling the width of the adjusting shims between the inner rings of the two tapered roller bearings is labor-intensive, inefficient, and costly. Even with automatic machine detection of various parameters, defective products cannot be completely eliminated. For defective products, repeated disassembly and installation, as well as the selection of adjusting shims of different thicknesses, are required. Details are as follows: Figure 11As shown, to accurately measure the clearance B0 between the inner rings of the two bearings, i.e., the width of the adjusting shim III1804 added during assembly, it is necessary to measure the dimensions of A1, A2, A3, B1, B2, and B3. Then, the clearance B0 is calculated using the dimensional chain formula B0 = (A1 + A2 + A3) - (B1 + B2 + B3). The size of B0 is the thickness of the adjusting shim. The dimensions of A1, A2, A3, B1, B2, and B3 can all be accurately measured before assembly, so the size of B0 can be accurately calculated before assembly. However, since the inner rings of the two bearings and the gear shaft, and the outer rings and the bearing bores of the bearing housing are both interference fits, the clearance B0 is calculated accordingly. Manufacturing errors in the inner and outer diameters of the two bearings, as well as the dimensions of the bearing holes in the mating gear shaft and bearing housing, all contribute to the increased size of A1 and A3 after assembly. Other factors affecting the size of B0 include the coaxiality error of the journals of the two bearing positions on the gear shaft, the perpendicularity error of the step surface P on the gear shaft (which axially positions the large end face of the inner ring of the first row of bearings) to the center line of the gear shaft, the coaxiality error of the two bearing holes inside the bearing housing, and the perpendicularity error of the step surfaces M and N (which axially position the large end faces of the two outer rings) to the center line of the gear shaft. These various form and position errors have the greatest impact on B0, resulting in a significant difference between the calculated or automatically measured B0 and the actual B0 after assembly. In summary, the size of B0 can only be accurately measured after assembly. However, if B0 is found to be out of tolerance after assembly, the internal components of the second-row large tapered bearing 1801 must be disassembled. The size of B0 can only be adjusted by adding or subtracting adjusting shims III 1804 of different thicknesses. Current processes easily damage the bearings when disassembling the internal components of the second-row large tapered bearing 1801. Therefore, many factories, to save time and avoid causing secondary damage to the bearings, rely on adjusting the grooved nut 1601 (see...). Figure 7 Using the slotted nut 1601 to adjust the axial clearance between the two bearings results in a large error, affecting the normal operation of the gears and bearings. Furthermore, the slotted nut 1601 is originally used to clamp the flange, spacer, adjusting shim, and the inner rings of the first row tapered bearing 1605 and the second row tapered bearing 1608 onto the gear shaft. Once the clearance between the two bearings is adjusted by the slotted nut, it indicates that the spacer and adjusting shim are no longer effective. This not only results in waste but also seriously affects the rigidity of the main reducer assembly and the normal meshing of the driving and driven gear pairs.

[0144] To address the above issues, the national automotive industry standard "QC / T 1099-2018 Technical Conditions for Compressible Elastic Spacers in Automotive Main Reducer Assemblies" was formulated in 2018. The purpose of this standard was to replace the original spacers and adjusting shims with a compressible elastic spacer component, achieving accurate control of bearing clearance without repeatedly adjusting the shim thickness. However, quite the opposite has occurred; this compressible elastic spacer not only fails to accurately control bearing clearance but also makes the bearing more prone to seizing (see...). Figure 12The reasons are as follows: During assembly, the outer rings of the second-row large tapered angular bearing 1801 and the first-row large tapered angular bearing 1802 are first press-fitted into the bearing housing. Then, the inner assembly of the first-row large tapered angular bearing 1802 is press-fitted onto the main gear shaft. Finally, the inner assembly of the second-row large tapered angular bearing 1801 is press-fitted onto the gear shaft. Because the press cannot accurately control the axial position of the inner assembly of the first-row large tapered angular bearing 1802, coupled with the compressibility of the elastic spacer in the axial direction, the clearance of the two bearings is in a state of 0 to negative clearance after assembly. When the two bearings are working, the heat generated by sliding friction between the large end face of the roller and the large flange is greater than the heat generated by rolling friction between the roller and the inner and outer raceways. The heat causes the rollers and inner rings of the two bearings to expand much more radially and axially than the outer rings. The radial expansion of the rollers and inner rings of the two-row bearings leads to a decrease in axial clearance. Due to the axial compressibility of the elastic spacer, the thermal expansion of the inner rings of the two-row bearings causes them to compress the compressible elastic spacer in two axial directions. The inner raceways of the two-row bearings move towards each other, similarly reducing the clearance between the two bearings. Therefore, if a compressible elastic spacer is used, regardless of whether the inner rings of the two bearings expand radially or axially, the result is a decrease in the clearance between the two bearings. Since the two bearings are already in a zero or negative clearance state after assembly, this will cause them to seize quickly after operation. If an incompressible spacer III1803 (see...) that has been used internationally for decades is used... Figure 11 Because the spacer III1803 has a relatively thick wall, the thermal expansion of the inner rings of the two-row bearings cannot compress the spacer and adjusting shims axially. Therefore, the inner rings of the two-row bearings cannot move towards each other axially due to thermal expansion. Instead, significant internal stress is generated within the two inner rings and the parts in axial contact with them. When the vehicle is not in operation, this internal stress gradually dissipates, meaning the existing spacer III1803 is not compressed. Consequently, the clearance between the two bearings does not decrease due to axial thermal expansion. The decrease in clearance between the two bearings is due to the radial thermal expansion of the inner rings being greater than that of the outer rings. Although this national standard was promulgated as early as 2018, in practice, no factory has yet implemented it.

[0145] This combined bearing perfectly solves all the problems existing in current main reducer bearings. In addition to eliminating the need for spacers, adjusting shims, and preload with zero clearance, the length of the bearing housing and gear shaft is reduced to a fraction of the existing length. The two bearing holes inside the existing bearing housing are compressed into one, and the two stepped journals of the existing gear shaft are compressed into one. The process and manufacturing difficulty of the entire main reducer assembly are greatly optimized and simplified, significantly improving the precision of the main reducer assembly and significantly reducing the manufacturing cost of the main reducer.

[0146] Example 16: The main structure of this example is the same as any one of Examples 1 to 15. Furthermore, as can be seen from the previous analysis, under the combined action of radial and axial loads, the heat generation, energy consumption, and temperature rise of the existing inner tapered roller bearing 1106 are much greater than those of the existing outer tapered roller bearing 1107. Figure 19 To prevent the main bevel gear, driven gear, and two side bearings from seizing, a reasonable clearance must be maintained between the main and driven gear pairs and the two side bearings. This is why existing differential assemblies must maintain a certain clearance between the main and driven gear pairs during assembly. This clearance is the root cause of NVH (noise, vibration, and harshness) deterioration in the main and driven gear pairs and the two side bearings. Therefore, significantly reducing the clearance between the main and driven gear pairs and the two side bearings is the only key factor in improving the quality of the differential assembly.

[0147] The key inventive invention of this differential assembly lies in replacing the existing two differential side bearings with a bearing unit assembly invented by our company (see...). Figure 18 As the outer flange moves from the inner conical ring 1101 to the outer conical ring 1102, when the driven bevel gear 9 applies a combined radial and axial load to the inner conical ring 1101, the frictional resistance, heat generation, and temperature rise between the inner conical ring 1101 and the roller 1103 are relatively small due to rolling friction. However, the frictional resistance, heat generation, and temperature rise between the outer flange 1104 and the large end face of the roller 1103 are relatively large due to sliding friction. Furthermore, the frictional resistance, heat generation, and temperature rise generated by the rolling friction between the outer conical ring 1102 and the roller 1103 result in a much greater heat generation and temperature rise for the outer conical ring 1102 and the outer flange 1104 than for the inner conical ring 1101. Since the outer conical ring 1102 and the outer flange 1104 are an integral structure, the thermal expansion of the outer conical ring 1102 is much greater than that of the inner conical ring 1101. However, since the surface area of ​​the outer circle of the conical outer ring 1102 is about twice that of the inner hole of the conical inner ring 1101, and the conical outer ring 1102 is fixed in the inner hole of the main reduction housing 10, the outer surface of the main reduction housing 10 is in contact with the air. When the vehicle is in motion, the outer surface of the main reduction housing 10 convects with the air, and the heat on the conical outer ring 1102 is easily dissipated through the main reduction housing 10. As a result, the temperature rise of the conical outer ring 1102 is almost the same as that of the conical inner ring 1101. During assembly, the clearance between the main and driven gear pairs and the two side bearings can be significantly compressed, thereby significantly reducing the NVH of the main and driven gear pairs and the two side bearings, improving the meshing accuracy of the main and driven gear pairs, reducing fatigue wear and spalling of the main and driven gear pairs, extending the life of the main reducer assembly and the differential assembly, extending the life of the gear oil, and saving maintenance costs.

[0148] Example 17: The main structure of this example is the same as any one of Examples 1 to 16. Furthermore, existing differential side bearings generally use two identical standard single-row tapered roller bearings. The dimensions and angles of each bearing component are standardized and have specific values. However, when the two tapered roller bearings are subjected to combined radial and axial loads, the bearing contact angle directly determines the magnitude of the total internal frictional resistance. To achieve the minimum total internal frictional resistance, the optimal contact angle value must be calculated based on the magnitude of the radial and axial loads borne by the bearing (see...). Figure 20 , 21 When the contact angle is 0°, it is... Figure 21 The cylindrical roller bearing on the left has a main bevel gear that applies a radial load to the inner ring of the bearing. The inner ring of the bearing The friction is transmitted to the cylindrical rollers, which then transmit it to the outer ring. There is only rolling friction between the rollers and the raceways of the inner and outer rings, resulting in minimal total frictional resistance. As the contact angle increases (see...), the frictional resistance decreases. Figure 21 (Right side), the inner ring holds the same radial load. The load is transferred to the tapered rollers, which then transfer it to the outer ring. The outer ring, in addition to bearing the radial load Q1, also bears... The derived axial load, the outer flange also has to bear The inner ring also bears the derived axial load. The derived axial load, as the total load inside the bearing increases, the total frictional resistance inside the bearing gradually increases, and the radial load... The relationship between the bearing contact angle and the total frictional resistance under the influence of [various factors] is shown in [the figure]. Figure 20 Thin solid lines.

[0149] When the contact angle is 0°, the main bevel gear applies an axial load to the inner ring of the bearing. (See Figure 21 (Left) The flange of the inner ring of the bearing bears the axial load. The frictional resistance is transmitted to the right end face of the cylindrical roller, and then to the outer ring flange from the left end face. There is only sliding friction between the roller and the flanges of the inner and outer rings, resulting in the maximum total frictional resistance. As the contact angle increases (see...),... Figure 21 (Right side), the inner ring holds the same axial load. The load is transferred to the tapered rollers, which then transfer it to the outer ring. The outer ring bears the axial load. In addition, they also have to bear The derived radial load is borne by the outer flange. Reduced load The inner circle also has to bear The derived radial load, although the total load inside the bearing increases with the increase of the contact angle, the total frictional resistance inside the bearing gradually decreases. This is because, with the increase of the contact angle, the sliding friction load between the large end face of the roller and the large flange of the inner ring gradually decreases, while the rolling friction load gradually increases. Since the coefficient of sliding friction is much greater than the coefficient of rolling friction, as the contact angle increases, the total load inside the bearing gradually increases, while the total frictional resistance gradually decreases. This results in a decrease in axial load. The relationship between the bearing contact angle and the total frictional resistance under the influence of [various factors] is shown in [the figure]. Figure 20 The thick solid line.

[0150] In summary, when designing the two side bearings of the differential, the total load or frictional resistance corresponding to different contact angles should be calculated according to the radial and axial loads borne by the UNIT-ASSY bearing unit assembly 11, based on the aforementioned formulas, and then plotted. Figure 20 The optimal contact angle is the contact angle when the total radial load or frictional resistance plus the total axial load or frictional resistance is minimized.

Claims

1. A combined bearing characterized by high efficiency, low energy consumption, and long service life, wherein: It includes a first row of cylindrical roller bearings (41) and a second row of thrust tapered roller bearings (42); The first row of cylindrical roller bearings (41) includes an inner ring (411), flange I (412), cylindrical rollers (413), an outer ring (414), and flange II (415); the cylindrical rollers (413) are located between the inner ring (411) and the outer ring (414), and in the radial direction, the inner cylindrical surface (4132) of the cylindrical rollers (413) contacts the inner raceway (4112) of the inner ring (411), and the outer cylindrical surface (4134) contacts the outer raceway (4141) of the outer ring (414); in the axial direction, the right end face (4131) of the cylindrical rollers (413) faces the flange II (415) and has a gap with the flange II (415), and the left end face (4133) of the cylindrical rollers (413) faces the flange I (412) and has a gap with the flange I (412); The second row of thrust tapered roller bearings (42) includes a shaft ring (421), rolling elements and a seat ring (423); the rolling elements are located between the shaft ring (421) and the seat ring (423); the shaft ring (421) is disposed on the side of the flange I (412) near the first row of cylindrical roller bearings (41); The combined bearing (4) can be used in the main reducer assembly and the differential assembly.

2. The combined bearing with high efficiency, low energy consumption, and long service life according to claim 1, characterized in that: The rolling element is a tapered roller (422) or a steel ball (425); When the rolling element is a tapered roller (422), the second row of thrust tapered roller bearings (42) also includes a flange III (424); in the axial direction, the right tapered surface (4221) of the tapered roller (422) contacts the right raceway (4211) of the shaft ring (421), and the left tapered surface (4222) contacts the left raceway (4231) of the seat ring (423); in the radial direction, the large end face (42523) of the tapered roller (422) contacts the flange III (424); When the rolling element is a steel ball (425), the right raceway (4211) surface of the shaft ring (421) and the left raceway (4231) surface of the seat ring (423) are provided with annular grooves, and a number of steel balls (425) are embedded in the annular grooves.

3. The combined bearing with high efficiency, low energy consumption and long service life according to claim 1 or 2, characterized in that: When the combined bearing (4) is used in the main reducer assembly, the main reducer assembly includes a main bevel gear (1), a flange cover (2), a bearing housing (3), a locking gasket (5), and a gear shaft (6); the main bevel gear (1) meshes with the driven bevel gear (9) of the differential assembly, and is an integral structure with the gear shaft (6); The bearing housing (3) is mounted on the outside of the gear shaft (6), and a flange cover (2) and a locking gasket (5) are respectively mounted on the side closer to the main bevel gear (1) and the side farther away from the main bevel gear (1); a cavity is formed between the flange cover (2), the bearing housing (3), the locking gasket (5) and the gear shaft (6), and the combined bearing (4) is mounted in this cavity and is axially positioned by the locking gaskets (5) and the flange cover (2) set at both ends; The first row of cylindrical roller bearings (41) is located on the side near the main bevel gear (1).

4. The combined bearing of claim 3, wherein: The inner hole of the inner ring (411) is interference-fitted with the cylindrical surface (602) of the gear shaft (6); the gear shaft (6) is a stepped shaft, and the stepped surface of the stepped shaft near the gear shaft (6) is denoted as step surface II (601). The step surface II (601) is in contact with the end face I (4111) of the inner ring (411); the step surface II (601) is the axial positioning surface of the inner ring (411) and the shaft ring (421); The stepped surface I (201) of the flange cover (2) contacts the end face II (4151) of the flange II (415), and the stepped surface I (201) is the axial positioning surface of the outer ring (414); The outer circle of the bearing seat (423) is an external threaded surface (4232); the inner hole (302) of the bearing seat (3) is provided with an internal threaded section (301) at the end away from the flange cover (2). The external threaded surface (4232) of the seat ring (423) mates with the internal threaded section (301) of the inner hole (302) of the bearing seat (3); when the external threaded surface (4232) is tightened, the left raceway (4231) of the seat ring (423) and the right raceway (4211) of the shaft ring (421) clamp the rolling element.

5. The combined bearing of claim 3, wherein: The main reducer assembly also includes a main reducer housing (10); the differential assembly includes a driven bevel gear (9), a bearing unit assembly (11), a differential housing (12), and an adjusting shim (13). The main reduction housing (10) is connected to the bearing seat (3), and the main bevel gear (1) is located in the main reduction housing (10) and meshes with the driven bevel gear (9); The driven bevel gear (9) is connected to the differential housing (12), and the differential housing (12) is fitted with bearing unit assemblies (11) at both ends. The bearing unit assembly (11) includes a tapered inner ring (1101), a tapered outer ring (1102), a roller (1103), and an outer flange (1104). The inner hole of the conical inner ring (1101) is interference-fitted with one end of the differential housing (12), and the outer conical ring (1102) is fitted with the main reduction housing (10). The outer conical ring (1102) has an outer flange (1104) near the driven bevel gear (9). The outer flange (1104) and the outer conical ring (1102) are integral structures. The roller (1103) is disposed between the inner conical ring (1101) and the outer conical ring (1102). The large end face of the outer conical ring (1102) is in contact with the adjusting shim (13). The adjusting shim (13) is used to axially position the bearing unit assembly (11), and the axial position of the bearing unit assembly (11) directly determines the meshing clearance between the main bevel gear (1) and the driven bevel gear (9). Tightening the adjusting shim (13) will cause the adjusting shim (13) to push the outer conical ring (1102) forward in the housing hole of the main reducer housing (10), thereby adjusting the position of the driven bevel gear (9) and changing the meshing clearance between the driven bevel gear (9) and the main bevel gear (1); When the meshing clearance between the main bevel gear (1) and the driven bevel gear (9) meets the design requirements, stop adjusting the adjusting shim (13), and the meshing clearance between the main bevel gear (1) and the driven bevel gear (9) is adjusted.

6. The combined bearing of high efficiency, low energy consumption and long service life according to claim 1 or 2, characterized in that: When the combined bearing (4) is used in the differential assembly, the differential assembly includes the combined bearing (4), the driven bevel gear (9), the differential housing (12), and the adjusting shim (13); the main reducer assembly includes the main bevel gear (1) and the main reducer housing (10); the main bevel gear (1) is disposed in the main reducer housing (10) and meshes with the driven bevel gear (9); The driven bevel gear (9) is connected to the differential housing (12), and the two ends of the differential housing (12) are fitted with combined bearings (4); the first row of cylindrical roller bearings (41) of the combined bearings (4) is close to the driven bevel gear (9), and the inner ring (411) of the first row of cylindrical roller bearings (41) is interference-fitted with one end of the differential housing (12), and the outer ring (414) is fitted with the main reduction housing (10); The side of the second row of thrust tapered roller bearings (42) away from the first row of cylindrical roller bearings (41) contacts the adjusting shim (13).

7. The combined bearing of high efficiency, low energy consumption and long service life according to claim 1, characterized in that: The outer ring (414) and flange II (415) of the first column cylindrical roller bearing (41) are an integral structure; The inner ring (411) and flange I (412) of the first row of cylindrical roller bearings (41) and the shaft ring (421) of the second row of thrust tapered roller bearings (42) are an integral structure.

8. The combined bearing of high efficiency, low energy consumption and long service life according to claim 1, characterized in that: The outer ring (414) and flange II (415) of the first column of cylindrical roller bearings (41) are two separate components; The inner ring (411) and flange I (412) of the first column of cylindrical roller bearings (41) are two separate parts.

9. The combined bearing of claim 2, wherein: When the rolling elements of the second-row thrust tapered roller bearing (42) are tapered rollers (422), there are three possible arrangements for the flange III (424): Setting method 1: The flange Ⅲ (424) and the shaft ring (421) are an integral structure; Setting method 2: The flange Ⅲ (424) and the seat ring (423) are an integral structure; Setting method 3: Both the shaft ring (421) and the seat ring (423) are provided with a retaining edge Ⅲ (424), and the two retaining edges Ⅲ (424) are integrated with the shaft ring (421) and the seat ring (423) respectively.

10. A combined bearing characterized by high efficiency, low energy consumption, and long service life, wherein: It includes a first row of cylindrical roller bearings (41) and a second row of thrust tapered roller bearings (42); The first row of cylindrical roller bearings (41) includes an inner ring (411), flange I (412), cylindrical rollers (413), an outer ring (414), and flange II (415); the cylindrical rollers (413) are located between the inner ring (411) and the outer ring (414), and in the radial direction, the inner cylindrical surface (4132) of the cylindrical rollers (413) contacts the inner raceway (4112) of the inner ring (411), and the outer cylindrical surface (4134) contacts the outer raceway (4141) of the outer ring (414); in the axial direction, the right end face (4131) of the cylindrical rollers (413) faces the flange II (415) and has a gap with the flange II (415), and the left end face (4133) of the cylindrical rollers (413) faces the flange I (412) and has a gap with the flange I (412); The second row of thrust tapered roller bearings (42) includes a shaft ring (421), rolling elements, and a seat ring (423); the seat ring (423) is disposed on the side of the flange I (412) near the first row of cylindrical roller bearings (41); the rolling elements are located between the shaft ring (421) and the seat ring (423); the rolling elements are tapered rollers (422) or steel balls (425); when the rolling elements are tapered rollers (422), the second row of thrust tapered roller bearings (42) also includes flange III (424); The combined bearing (4) can be used in the main reducer assembly and the differential assembly; When the combined bearing (4) is used in the main reducer assembly, the first row of cylindrical roller bearings (41) is on the side away from the main bevel gear (1); the second row of thrust tapered roller bearings (42) is on the side close to the main bevel gear (1).

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