Maintenance-free rail transit axle box bearing

Abstract

The utility model provides a maintenance-free rail transit axle box bearing. The bearing comprises an inner ring, an outer oil seal, an outer flange, a retainer, an outer ring, a small flange, a tapered roller and an inner oil seal. And the outer ring coaxially sleeves the outer side of the inner ring. And the outer raceway, the inner raceway, the small flange and the step part are encircled to form a friction working part. The retainer is arranged in the friction working part. And a tapered roller is arranged in the window of the retainer. And the tapered roller rolls in the friction working part. And the retainer drives the tapered roller to rotate along the outer raceway, the inner raceway and the step part in the circumferential direction. The conical surface of the tapered roller is in contact with the inner raceway and the outer raceway, and the spherical base surface is in contact with the step part. The bearing has the advantages that the effective load of the axle box bearing is obviously improved, the ineffective load is obviously reduced by more than 60%, the efficiency of a transmission system is obviously improved, the operation cost of a high-speed rail is obviously reduced, the energy consumption is obviously reduced, and the bearing is energy-saving and environment-friendly.

Description

Technical Field

[0001] The utility model relates to the field of high-speed train sets, in particular to a maintenance-free rail transit axle box bearing. Background Art

[0002] High-speed rail axlebox bearings are key components of high-speed trains, and their safety and reliability are crucial to the safety of passengers.

[0003] See also Figure 7 Existing bearings have a series of systematic design flaws. These systematic design flaws not only affect the performance and life of the bearings, but also increase the maintenance costs and operational risks of high-speed rail, seriously restricting the development of the high-speed rail industry:

[0004] 1. Outer raceway spalling: Spalling of the outer raceway has the highest failure rate, accounting for 44%. Spalling is concentrated in the outer raceway's heaviest load areas, a phenomenon caused by stress concentration. During operation, the outer ring remains stationary, while the inner ring drives the rollers along the outer raceway. Due to axial play, the outer raceway bears a localized load, rather than a uniform load.

[0005] 2. Inner raceway spalling: This is the second most common type of failure, accounting for 12%. Although the inner raceway should theoretically bear cyclic loads evenly, due to the influence of axial clearance, it actually bears stress-concentrated cyclic loads rather than uniform cyclic loads.

[0006] 3. Temperature rise and wear: Temperature rise and wear rank third and fifth in the failure rate, accounting for 10% and 8% respectively. Due to the original structural design defects of imported axle box bearings, temperature rise and the "pumping effect" are inevitable. The temperature rise causes the grease viscosity to decrease. The "pumping effect" and high speed cause the grease to move to the ends of the bearing, resulting in grease starvation and subsequent loss of grease in the roller heads. This causes dry friction and rapid temperature rise in the contact area between the inner and outer raceways and the roller heads. When the working area is almost free of grease, wear and ablation are the stages that the bearing enters when it continues to operate.

[0007] 4. Retainer breakage: The third most common cause of failure is retainer breakage, also accounting for 10%. While currently used plastic retainers offer many advantages over iron retainers, they also have insurmountable design and manufacturing shortcomings. For example, the design requires a single-sided clearance of only 0.125 mm between the retainer and the roller. If the retainer is made of mild steel, the required clearance can be achieved during assembly by controlling and adjusting the shrinkage of the iron retainer. However, since plastic retainers cannot shrink in both design and manufacturing, the clearance between the roller and retainer cannot be controlled. Furthermore, due to radial manufacturing errors in the retainer, manufacturing errors in the outer diameter of the small flange, and the radial clearance tolerance of the bearing, these errors and tolerances accumulate, causing the centrifugal force generated by the rotating roller to continuously impact the retainer, leading to a gradual increase in the clearance between the roller and retainer (see the detailed embodiment below). The roller no longer simply rolls within the retainer's window, but instead experiences a combination of rolling, sliding, and oscillating motion. As the clearance between the retainer and roller increases, the retainer's trajectory changes. The retainer no longer rotates along the bearing's centerline, and the retainer's centerline no longer coincides with the bearing's centerline. Consequently, the working conditions between the retainer and roller become increasingly severe, ultimately leading to fracture.

[0008] 5. Problems related to abnormal operating conditions: Although the proportion of faults is small, they are usually related to abnormal operating conditions of bearings.

[0009] 6. Temperature rise and high energy consumption: Imported bearings have the problem of temperature rise, which not only causes the temperature sensor to alarm and the high-speed rail driver to slow down, affecting the passengers' journey, but also causes the high temperature to volatilize and oxidize the lubricating oil, which will also affect the life of the bearings. Since the temperature rise is caused by friction energy consumption, and the high friction energy consumption is caused by the design of the bearings, the high energy consumption also increases the operating cost of the high-speed rail.

[0010] 7. Bearing clearance: Bearing clearance significantly affects fatigue life, temperature rise, noise, vibration, and other factors. The existing structure makes it impossible to adjust bearing clearance to the optimal state, resulting in reduced performance and shortened life, and requiring regular overhauls and maintenance.

[0011] Therefore, it is of great significance to develop a maintenance-free rail transit axle box bearing. Utility Model Content

[0012] The purpose of the utility model is to provide a maintenance-free rail transit axle box bearing to solve the problems existing in the prior art.

[0013] The technical solution adopted to achieve the purpose of the utility model is as follows: a maintenance-free rail transit axle box bearing, including an inner ring, an outer oil seal, an outer rib, a retainer, an outer ring, a small rib, a tapered roller and an inner oil seal.

[0014] An inner raceway is provided on the outer circumference of the middle section of the inner ring. The portion of the inner ring at the left end of the inner raceway is labeled "small end," and the portion at the right end is labeled "large end." The inner raceway is conical in shape. Both the small end and the large end are cylindrical.

[0015] The small rib is annular in structure as a whole. The small rib is sleeved on the periphery of the small head. A working surface is provided on the end surface of the small rib facing the inner raceway.

[0016] The inner hole of the outer ring is marked as an outer raceway and a cylindrical section from the right end to the left end. The outer raceway is conical in shape. The small diameter end of the outer raceway is connected to the cylindrical section.

[0017] The outer rib includes an outer rib body and a step portion. The outer rib is arranged on the right end side of the outer ring. The step portion extends into the outer raceway.

[0018] The outer ring is coaxially mounted on the outer side of the inner ring. The outer raceway, inner raceway, small rib, and step define a friction working section. The retainer is positioned within the friction working section. Tapered rollers are housed within the retainer's window. These tapered rollers roll within the friction working section. The retainer circumferentially drives the tapered rollers along the outer raceway, inner raceway, and step. The tapered surfaces of the tapered rollers contact the inner and outer raceways, while the spherical base surfaces contact the step.

[0019] The outer oil seal is arranged in the gap between the large head and the outer rib. The outer oil seal is arranged on the right end side. The outer circle of the outer oil seal and the inner hole of the outer rib body have an interference fit, and the lip of the outer oil seal and the outer circle of the large head have an interference sealing fit.

[0020] The inner oil seal is arranged in the gap between the small rib and the cylindrical section of the outer ring inner hole. The inner oil seal is arranged on the left end side. The outer circle of the inner oil seal and the inner hole of the cylindrical section have an interference fit, and the lip and the outer circle of the small rib have an interference sealing fit.

[0021] Furthermore, the small rib is fixed to the small head in a mechanical manner.

[0022] Furthermore, the retainer is made of plastic.

[0023] Furthermore, the semi-cone angle of the outer raceway is a=1° to 9°.

[0024] Furthermore, maintenance-free axlebox bearings are used in high-speed railways, subways, ordinary passenger trains or ordinary freight trains.

[0025] The technical effects of the present invention are unquestionable: it has the advantages of a small number of parts, good assembly performance, compact structure, large load, low noise, low torque, low cost, ultra-low friction torque, lightweight, integrated, energy-saving, no need for pre-tightening (adjustment), zero clearance, ultra-long mileage maintenance-free, ultra-long life, high strength, high rigidity, high reliability, and high stability. In addition, it significantly increases the effective load of the axle box bearing, significantly reduces the invalid load of the axle box bearing by more than 60%, significantly improves the transmission efficiency of the axle box bearing, greatly improves the energy utilization rate, significantly reduces the temperature rise of the axle box bearing by more than 60%, and makes the temperature rise of the entire bearing uniform, significantly reduces the operating cost of the high-speed rail, has significant energy-saving and environmental protection effects, and is simple and easy to disassemble and install during maintenance. At the same time, it also benefits the people of all countries in the world who own high-speed rail, subways, ordinary passenger and freight trains, and all motor vehicles, and contributes to the carbon neutrality, carbon peak, energy conservation and emission reduction, and environmental protection of the global village. BRIEF DESCRIPTION OF THE DRAWINGS

[0026] Figure 1 Schematic diagram of maintenance-free rail transit axle box bearing products;

[0027] Figure 2 For maintenance-free rail transit axle box bearings, no adjustment (control) and zero clearance working principle diagram;

[0028] Figure 3 Transmission route diagram of the axial force to the left for maintenance-free rail transit axle box bearings;

[0029] Figure 4 Transmission route diagram of the rightward axial force on the axle box bearing of maintenance-free rail transit

[0030] Figure 5 This is a schematic diagram of the ratio of bearing radial clearance to bearing life;

[0031] Figure 6 A diagram showing the change in axial clearance of maintenance-free rail transit axle box bearings under thermal expansion and contraction conditions;

[0032] Figure 7 This is a diagram showing the fault types of axle box bearings from 50 imported EMU trains that were randomly inspected and scrapped prematurely.

[0033] Figure 8 This is a schematic diagram of existing imported high-speed rail axle box bearing products;

[0034] Figure 9 This is the force analysis diagram of existing imported high-speed rail axle box bearing products;

[0035] Figure 10 Force analysis diagram of maintenance-free rail transit axle box bearing products;

[0036] Figure 11This is a product diagram of the outer ring of a maintenance-free rail transit axle box bearing;

[0037] Figure 12 Schematic diagram of roller products for maintenance-free rail transit axle box bearings;

[0038] Figure 13 Schematic diagram of small rib product for maintenance-free rail transit axle box bearings;

[0039] Figure 14 Schematic diagram of rail transit axle box bearing products with different internal structures;

[0040] Figure 15 This is a product diagram of the inner ring of a maintenance-free rail transit axle box bearing;

[0041] Figure 16 This is a schematic diagram of the outer rib of a maintenance-free rail transit axle box bearing;

[0042] Figure 17 Schematic diagram of grease loss in existing high-speed rail axle box bearings due to the "pumping effect";

[0043] Figure 18 Schematic diagram of the lubrication working conditions of maintenance-free rail transit axle box bearings.

[0044] Figure 19 Schematic diagram of the retainer product for maintenance-free rail transit axle box bearings.

[0045] In the figure: inner ring 1, inner raceway 101, small head 102, large head 103, outer oil seal 2, outer rib 3, outer rib body 301, step portion 302, retainer 4, outer ring 5, outer raceway 501, cylindrical section 502, small rib 6, tapered roller 7, spherical base surface 702, inner oil seal 8. DETAILED DESCRIPTION

[0046] The present invention will be further described below with reference to the following embodiments. However, it should not be understood that the scope of the present invention is limited to the following embodiments. Without departing from the above technical concept of the present invention, various substitutions and modifications based on common technical knowledge and customary means in the field should be included in the scope of protection of the present invention.

[0047] Example 1:

[0048] See also Figures 1 to 19 This embodiment provides a maintenance-free rail transit axle box bearing, including an inner ring 1, an outer oil seal 2, an outer rib 3, a retainer 4, an outer ring 5, a small rib 6, a tapered roller 7 and an inner oil seal 8.

[0049] An inner raceway 101 is provided on the outer circumference of the middle section of the inner ring 1. The portion of the inner ring 1 at the left end of the inner raceway 101 is designated as a small end 102, while the portion at the right end is designated as a large end 103. The inner raceway 101 is conical in shape. Both the small end 102 and the large end 103 are cylindrical.

[0050] The small rib 6 is annular in structure and is sleeved around the small head 102 of the inner ring 1. A working surface 601 is provided on the end surface of the small rib 6 facing the inner raceway 101.

[0051] The inner hole of the outer ring 5 is marked as an outer raceway 501 and a cylindrical section 502 from the right end to the left end. The outer raceway 501 is in a conical shape. The small diameter end of the outer raceway 501 is connected to the cylindrical section 502.

[0052] The outer rib 3 includes an outer rib body 301 and a step portion 302. The outer rib 3 is arranged on the right end side of the outer ring 5. The step portion 302 extends into the outer raceway 501.

[0053] The outer ring 5 is coaxially mounted on the outer side of the inner ring 1. The outer raceway 501, inner raceway 101, small rib 6, and step 302 together form a friction working area. The retainer 4 is positioned within the friction working area. The retainer 4 has windows housing tapered rollers 7, which roll within the friction working area. The retainer 4 circumferentially rotates the tapered rollers 7 along the inner and outer raceways. The tapered surfaces of the tapered rollers 7 contact the inner raceway 101 and the outer raceway 501, while the spherical base surfaces 702 contact the step 302.

[0054] The outer oil seal 2 is positioned in the gap between the large end 103 of the inner ring 1 and the outer rib 3. The outer oil seal 2 is positioned on the right end. The outer diameter of the outer oil seal 2 forms an interference fit with the inner hole of the outer rib body 301, while the lip of the outer oil seal 2 forms an interference, sealing fit with the outer diameter of the large end 103 of the inner ring 1.

[0055] The inner oil seal 8 is positioned in the gap between the small rib 6 and the inner cylindrical section 502 of the outer ring 5. The inner oil seal 8 is positioned on the left end. The outer diameter of the inner oil seal 8 forms an interference fit with the inner bore of the cylindrical section 502, while the lip of the inner oil seal 8 forms an interference, sealing fit with the outer diameter of the small rib 6.

[0056] The maintenance-free rail transit axle box bearing of this embodiment has the advantages of a small number of parts, good assembly performance, compact structure, large load, low noise, low torque, ultra-low friction torque, lightweight, integrated, super energy-saving, always no preload (adjustment) and zero clearance, ultra-long mileage maintenance-free, ultra-long life, ultra-low cost, high strength, high rigidity, high reliability, high stability, etc., and significantly increases the effective load of the axle box bearing, significantly reduces the invalid load by more than 60%, significantly reduces the temperature rise by more than 60%, significantly improves the transmission efficiency, significantly reduces energy consumption, and significantly reduces the operating cost of high-speed rail. Moreover, it is simple and easy to install and disassemble during maintenance, indicating that the axle box bearing has excellent novelty and creativity.

[0057] The axlebox bearings of this embodiment are suitable not only for high-speed rail, but also for conventional passenger trains, subways, freight trains, and all other modes of transportation. Due to flaws in the original design, after mass installation, the wheelset bearings themselves caused "hot axle" and "hot axle" problems, resulting in numerous serious accidents such as derailments and overturning. However, due to improvements in bearing manufacturing precision and quality, as well as improved hot axle monitoring and prevention measures, the number of major "hot axle" accidents has dropped to single digits annually since the beginning of the new century, and the number of hot axle accidents has also decreased significantly. However, the source of the "hot shaft" has not been eliminated. Only by adopting the axle box bearing of this innovative design can the root cause of the "hot shaft" be eliminated from the source. The root cause of the "hot shaft" is that the contact angle of the tapered roller bearing is too large. The large contact angle causes the bearing that originally bears pure radial load to bear a large derived axial force. The friction torque generated by the derived axial force, especially the sliding friction torque between the roller ball base surface 702 and the large rib, is huge. According to the torque calculation formula T=F*r, where F represents the axial force (see Table 6 and Figure 10), r represents the distance from the point of force application to the center line of the bearing, and the axial friction torque between the roller and the inner and outer raceways and the large rib can be calculated; according to the relationship between power and torque: P = T * ω (P represents power, T represents torque, ω represents angular velocity), the power generated by the above three axial forces can be calculated, and then according to the calculation formula of work and power: W = P * t (t represents time), the above three formulas are sorted out to obtain the following formula: W = F * r * ω * t. According to the first law of thermodynamics, that is, the law of conservation of energy, 100% of the work done by these derived axial forces is converted into heat. From this formula, the following conclusion can be drawn: the temperature rise of the high-speed rail axle box bearing is proportional to the load, radial size, speed and running time borne by the bearing. One of the innovative designs of the axle box bearing of this patent is to significantly reduce the contact angle of the bearing, reducing the contact angle of the axle box bearing from the existing 10° to below 3.2°. As a result, the temperature rise is significantly reduced by more than 60%, completely eliminating the root cause of the "hot shaft". Moreover, the source of the "hot shaft" of existing bearings is that the heat generated during the operation of the bearing is concentrated on the inner ring, because there are two friction pairs between the inner raceway of the inner ring and the conical surface of the roller, and between the large rib of the inner ring and the spherical base surface of the roller, which generate more heat. Because there is clearance between the outer raceway of the outer ring and the conical surface of the roller, the outer raceway and the conical surfaces of all rollers are not in full contact, resulting in relatively little heat generated between the roller and the outer raceway, and the heat generated between the roller and the large rib and the inner raceway cannot be transferred to the outer ring. The bearing of this patent has only one friction pair between the inner ring and the roller, that is, the rolling friction pair between the inner raceway of the inner ring and the conical surface of the roller. Compared with the existing structure, there is one less sliding friction pair, so the heat generated by the inner ring is less than that of the existing structure. In terms of structure, the outer rib and the small end face of the outer ring are in close contact. Even though there are two friction pairs between the outer raceway 501 and the conical surface of the roller, and between the spherical base surface 702 of the roller and the end face of the step 302 of the outer rib 3, the heat generated is much greater than that between the inner ring and the roller. However, since there is no clearance between the outer raceway of the outer ring and the conical surface of the roller, the outer raceway and all the conical surfaces of the rollers are in uniform and comprehensive contact, resulting in a uniform temperature rise of the outer ring 5 and the outer rib 3. In addition, the volume and surface area of ​​the outer ring 5 are much larger than those of the inner ring. Even though the heat generated between the outer ring 5 and the outer rib 3 between the outer ring 5 and the roller 7 and between the roller 7 and the outer rib 3 is much greater than that of the inner ring 1, the temperature rise of the outer ring and the outer rib is still relatively low, and the heat dissipation is relatively fast. Furthermore, the temperature rise is greatly reduced, the evaporation and oxidation rates of grease are slowed down or even eliminated, the life of the bearings is greatly improved, and the energy consumption of driving the train is reduced, saving the country huge energy costs every year and making a huge contribution to environmental protection. It really kills two birds with one stone.

[0058] Since the temperature rise of the bearings of this patent is greatly reduced, the impact of ambient temperature on the temperature rise of the bearings is no longer so obvious. It is possible to consider adding direct EMUs between two areas with relatively large temperature differences, such as adding EMUs from South China to Northeast China.

[0059] It is worth noting that for various motor vehicles such as cars, tricycles, motorcycles, agricultural machinery, tractors, and construction machinery, the stress properties of their hub bearings are exactly the same as those of high-speed railways, subways, ordinary passenger and freight trains, except that the magnitude of the stress and the speed of the two are different. The hub bearings of the above-mentioned motor vehicles are only smaller in size and weight than the axle box bearings of trains. Therefore, the axle box bearings of the patent in this embodiment are also suitable for all the above-mentioned motor vehicles, and the hub bearings of the above-mentioned motor vehicles are also within the scope of protection of this patent.

[0060] Example 2:

[0061] The main contents of this embodiment are the same as those of embodiment 1, wherein the small rib 6 is mechanically fixed to the small head 102. In this embodiment, the parts diagram of the small rib 6 of the axle box bearing is shown in FIG. Figure 13 Its inner hole is positioned and fixed on the small head 102 of the inner ring 1 by means of thread or interference fit. According to the function, the small rib 6 can be divided into two types: fixed and removable. The fixed type is further divided into two types. One is to fix the small rib 6 by the spinning riveting process to the small end face of the inner ring 1 (see Figure 14 ), one is to fix the small rib 6 by means of interference fit between the inner hole of the small rib 6 and the small head 102 of the inner ring 1. In order to prevent the axial movement of the small rib 6, a few screw holes can be drilled in the radial direction of the small rib 6, and the small rib 6 can be fastened to the inner ring 1 with set screws. The fixed small rib 6 has an obvious feature, that is, the retainer 4 of the existing axle box bearing has a detachable column or beam, so that it can be removed and inspected during "planned repairs" to inspect the inner raceway 101 of the inner ring 1. After the inspection, all parts must be cleaned and re-lubricated with grease. Since the inner ring 1, roller 7 and retainer 4 form an internal assembly together, it is difficult to clean them, and it is also inconvenient to inspect the working surfaces of the roller 7 and the inner raceway 101, which lays a hidden danger for the subsequent loading and operation of the bearing. Moreover, the function of a detachable column or beam is definitely different from that of the other integrated columns or beams, which also affects the normal operation of the bearing. If all the columns or beams of the retainer 4 are made into one piece (see Figure 19), there are no longer any detachable columns or beams, and the retainer 4 is removed as a whole during the "planned repair", and the retainer 4, roller 7 and inner ring 1 are cleaned and inspected separately. This way, the cleaning is relatively clean and the inspection is convenient. After the inspection, they are assembled. This "planned repair" inspection plan is the best choice, and the utility model can achieve this. The small rib 6 of this maintenance-free axle box bearing is fixed to the outer circle 102 of the small head of the inner ring 1 through an internal thread, and several screw holes are drilled radially on the small rib 6. There is a cylindrical pin inside the screw hole. There is a small interference or transition fit between the cylindrical pin and the screw hole. Before turning the internal thread of the small rib 6, use a set screw to push the cylindrical pin to move in the screw hole, ensure that the cylindrical pin slightly protrudes from the inner hole, and then use a thread cutter to turn the internal thread hole of the small rib 6. When "planned repair" is needed, first loosen the set screw, and the cylindrical pin and the external thread 102 of the small head of the inner ring 1 are no longer locked, unscrew the small rib 6, and the roller 7 and retainer 4 slide out of the inner raceway 101 of the inner ring 1. Clean the roller 7, retainer 4 and inner ring 1 separately, and assemble the roller 7 and retainer 4 to the inner raceway 101 of the inner ring 1 after checking that they are correct. Screw the small rib 6 onto the external thread of the small head 102 of the inner ring 1, tighten the set screw, and the small rib 6 is firmly fixed on the small head 102 of the inner ring 1.

[0062] In fact, since this patented bearing has the advantages of ultra-long mileage maintenance-free, ultra-long life, as well as high rigidity, high reliability, and high stability, the maintenance mode of high-speed trains will gradually change from "planned maintenance" to "preventive maintenance" mode, and finally to "maintenance-free" mode. Therefore, no matter how the small rib 6 and the inner ring 1 are connected, the effect is the same, that is, from the time the axle box bearing is installed on the train to the end of its life, it can be directly replaced, and no major maintenance is required in the middle, truly achieving all-round, full-process, and full-life cycle crushing of imported products.

[0063] Example 3:

[0064] This embodiment is similar in principle to Embodiments 1 or 2, except that retainer 4 is made of plastic. In actual production, retainer 4 is typically made of a combination of one or more of polycarbonate, polyamide 6, polyamide 66, polyacetal, modified polyphenylene ether, butylene terephthalate, GF-reinforced polyethylene terephthalate, ultra-high molecular weight polyethylene, or polyphenylene sulfide. Retainer 4 is injection molded in a single step using an injection mold. The primary technical requirement is to ensure that the clearance between retainer 6 and roller 7 meets national standards.

[0065] Example 4:

[0066] The main content of this embodiment is the same as any one of Embodiments 1 to 3, wherein the semi-cone angle of the outer raceway 501 is a=1° to 9°.

[0067] Example 5:

[0068] The main contents of this embodiment are the same as any one of the embodiments 1 to 4, wherein the greatest feature of the outer ring 5 of the axle box bearing is that the contact angle is less than 1 / 2 of the current imported bearings. For example, the utility model inventor has designed a bearing with the same inner and outer diameters of 130 mm and 240 mm as the existing bearings, and the semi-cone angle a of the outer raceway 501 is only 3.2° (see Figure 11 ), if the contact angle of the outer raceway 501 is smaller, the additional ineffective derived axial force of the entire bearing will be smaller (see Table 6), the friction torque will be smaller, and the energy consumption and heat generation will be less. Not only energy is saved, the temperature rise of the bearing is also reduced, the radial dynamic load rating is also increased, and the service life is also increased accordingly.

[0069] The biggest feature of this axle box bearing roller 7 is that it is particularly slender (see Figure 12 ), the conical surface of the roller cannot be processed by through-grinding, and can only be processed by plunge grinding. In this way, as long as the grinding wheel is trimmed into the shape of a logarithmic curve, the roller with logarithmic generatrix convexity can be ground, avoiding the defects of no convexity in the current through-grinding and the inability to process the logarithmic curve convexity through through-finishing. After the roller 7 is super-finished, it passes 100% full inspection and the roller diameter size is grouped according to 0.001mm as a group. In this way, the roller 7 of each set of bearings is evenly stressed during operation without stress concentration, which greatly extends the life of the bearing.

[0070] Example 6:

[0071] This embodiment discusses the technical effects of any one of the embodiments 1 to 4. When the high-speed train is running straight on the track, the axle box bearing only bears radial force. Only when the high-speed train turns, the axle box bearing bears both radial force and axial force. In order to allow the axle box bearing to withstand axial force in two directions, the existing high-speed train is either equipped with a set of imported double-row tapered roller bearings (see Figure 8 ), or install a set of imported double-row cylindrical roller bearings, while the actual domestic installation is mostly double-row tapered roller bearings. The following content only selects double-row tapered roller bearings as an example for comparative discussion.

[0072] Statistics show that as of 2019, the country imported a total of 220,000 sets of axlebox bearings for high-speed trains, of which 550 sets failed prematurely and were scrapped before reaching their design life, with a failure rate of 0.25%. The National Bearing Quality Supervision and Inspection Center tested 50 sets of axlebox bearings between 2013 and 2020, and the types of failures were as follows: Figure 7 As shown, these 50 sets of bearings are randomly selected from 550 sets of scrapped bearings and are representative. The utility model inventor of this patent now analyzes the causes and solutions of the failures one by one according to the failure types of imported axle box bearings.

[0073] 1. The top two failure types are spalling of the outer raceway 501 and the inner raceway 101. The outer raceway 501 has the highest spalling rate, reaching 44%. Moreover, the spalling is located at the same azimuth angle as the outer ring when it is working. That is, the spalling is concentrated in the area with the largest load on the circumference of the outer raceway 501. This is the most typical spalling caused by stress concentration. Because the outer ring 5 of the axle box bearing is stationary during operation, the inner ring 1 drives the roller 7 to rotate along the outer raceway 501. According to the process document requirements of the document number "sfet-77-48300" and the document name "EMU Maintenance Axlebox Bearing Pressing Operation Instructions", NTN and NSK companies specify that the axial clearance values ​​of the bogie axlebox bearings of the Harmony 2C Phase II / 380A (L) EMU are in the range of 0.15~0.62 and 0.15~0.561 respectively (see the table below). Table 1 shows the axial clearance values ​​of the axlebox bearings supplied by NTN and NSK companies of Japan after installation.

[0074] Table 1

[0075]

[0076] Now calculate the maximum value of radial clearance based on the maximum value of axial clearance. The outer raceway semi-cone angle of the axle box bearings currently imported from NTN and NSK is 10°. From the table above, we can see that the maximum axial clearance values ​​of NTN and NSK are 0.62mm and 0.561mm respectively. Take NSK's 0.561mm for calculation. According to the relationship between the radial clearance and axial clearance of tapered roller bearings, it can be calculated that the maximum radial clearance value of NSK's axle box bearings is: 0.561*2*tan10°=0.2mm. Since the imported axle box bearings are double-row tapered roller bearings, the maximum value of the radial clearance of each row is 0.1mm. The maximum value of the axial clearance of NTN is 0.62mm, which is 0.5 higher than the maximum value of the axial clearance of NSK. 61mm is 0.06mm more, which means that the maximum radial clearance value of each row of NTN's double-row tapered roller bearing is larger than 0.1mm. Therefore, we can conclude that the outer raceway 501 of the stationary outer ring 5 does not bear the load evenly on the circumference. The outer raceway 501 of the outer ring 5 actually bears a local load. The load that should have been evenly borne by the entire outer raceway 501 is actually borne by a local position of the outer raceway 501 due to the existence of a large radial clearance. This local position has been under excessive load for a long time and is eventually overwhelmed. It is natural and inevitable that the outer raceway 501 will peel off at this position. This is the most typical design problem. The detailed principle is as follows: The size of the clearance during bearing operation has a huge impact on fatigue life, temperature rise, noise, vibration, etc. Figure 5The following graph shows the bearing life ratios for different radial clearances. When the radial clearance of the bearing is less than -0.002mm, negative clearance or an interference fit between the rollers and the inner and outer raceways creates a significant frictional torque, which in turn causes wear and significantly shortens the fatigue life of the wheel hub bearing. As the radial clearance varies from -0.002mm to 0, the bearing life significantly increases. When the bearing clearance approaches 0, the fatigue life of the bearing is longest. As the bearing clearance gradually increases from 0, the rolling elements experience uneven forces, and the load on the inner and outer raceways is also unevenly distributed, further reducing the bearing's fatigue life. Therefore, it can be inferred that only by adjusting the wheel hub bearing's clearance to 0 can the fatigue life be maximized and the effects of temperature rise, noise, and vibration be minimized. Currently, imported axlebox bearings, or double-row tapered roller bearings, cannot achieve this clearance adjustment, either in principle or due to process or assembly requirements. The axle box bearing of the utility model is only a set of single-row tapered roller bearings. Its internal structure determines that it does not need to be adjusted or preloaded, and its clearance is always at 0 (see Figure 2 ), for the convenience of explaining the problem, the inventor only retained several parts that affect the clearance of the axle box bearing of this patent, and deleted all other parts that are not related to the clearance. Since the weight of the high-speed rail itself, that is, the radial force, is applied to the outer raceway 501 through the outer ring 5, the outer raceway 501 of the outer ring 5 transfers the load to the roller 7. This radial force generates a derived axial force on the outer raceway 501. The radial force and the derived axial force force the roller 7 to move toward the large end (or the right side) of the bell mouth formed by the outer raceway 501 and the inner raceway 101. When the large end face of the roller 7 moves to the end face of the step portion 302 of the outer rib 3, the axial clearance between the large end face of the roller 7 and the end face of the step portion 302 of the outer rib 3 is 0. The radial clearance between the conical surface of the roller 7 and the inner raceway 101 of the inner ring 1 and the outer raceway 501 of the outer ring 5 is also 0. Regardless of whether the high-speed rail is stopped or in motion, as long as there is weight pressing on the axle box bearing, the axle box bearing is always in the optimal working state of zero clearance.

[0077] The second most common fault type is the spalling of the inner raceway 101 of the inner ring 1, accounting for 12%. The inner raceway 101 of the inner ring 1 of the imported axle box rotates along the outer raceway 501 with the roller 7. Although in principle the inner raceway 101 of the inner ring 1 is subjected to a cyclic load, indicating that the load is evenly distributed on the circumference of the inner raceway 101, in fact, due to the existence of the axial clearance of the imported axle box mentioned above, every part of the inner raceway 101 in the circumferential direction is always subjected to a super cyclic load, rather than the conical surface of the inner raceway 101 evenly bearing the same cyclic load. Even so, the cyclic stress condition of the inner raceway 101 is much better than the local stress condition of the outer raceway 501. This is why the spalling of the inner raceway 101 of the inner ring 1 is reduced by more than two-thirds than the spalling of the outer raceway 501 of the outer ring 5.

[0078] The axle box bearing of the utility model is only a set of single-row tapered roller bearings. Its internal structure determines that it does not need to be adjusted or preloaded, and its clearance is always at 0 (see Figure 2 ), that is, the inner raceway 101 and the roller 7 are always in a working state of zero clearance, and the inner raceway 101 bears the cyclic load evenly on the circumference, and its service life is doubled.

[0079] 2. Temperature rise and wear ranked third and fifth in the failure rate, accounting for 10% and 8% respectively. Using "temperature rise" to define the failure type is very inaccurate, because the results of a temperature rise of 1° and a temperature rise of 1000° are very different. The accurate description of the failure caused by temperature rise should be called ablation or high-temperature burning. The most obvious manifestation of ablation is the blueing of the rollers and inner and outer rings due to high temperature. This is because high temperature will reduce the viscosity of the grease, accelerate the evaporation and oxidation rate of the grease, and the "pumping effect" (see Figure 17 ), an oil-poor area is created at the small end of the roller, causing dry friction between the inner and outer rings and the rollers, which in turn causes the inner and outer rings and rollers to heat up rapidly and turn blue; as for wear, it is because the axlebox bearings are not replaced in time after turning blue, and the grease is burned. At this time, the bearings continue to operate, which must have entered the wear stage. The grease of imported axlebox bearings is screened through a rigorous process and can meet the requirements of high-speed rail axlebox bearings. In addition to the above-mentioned "pumping effect", could it be that the high-temperature ablation and wear are caused by the design of the imported bearings themselves?

[0080] As we all know, my country has a vast territory and a large span of high-speed rail lines. EMUs run over long routes, for a long time, at high speed and continuously. At the same time, they will face harsh operating environments such as high and low temperatures, environmental corrosion, and wind and sand damage. In particular, the temperature difference between the north and the south is large, and the viscosity of the grease changes dramatically with the change of temperature. This makes the service environment temperature of the EMU bearing system complex and changeable, thereby increasing the probability of failure of the EMU axle box bearings, which brings challenges to the operation and manufacturing of the EMU bearings. The country has realized this problem, so in order to ensure the safety and reliability of high-speed rail, high-speed rail has not been opened in countries between regions with huge north-south distances and temperature differences. For example, there is no high-speed rail between South China and the three northeastern provinces. It seems that the impact of excessive environmental temperature difference on the operation of the axle box bearings has also been ruled out.

[0081] Since the effects of product design and ambient temperature on axlebox bearing erosion and wear are excluded, the only cause of axlebox bearing erosion and wear can only be found in the technical requirements. According to the process document requirements with the document number "sfet-77-48300" and the file name "EMU Maintenance Axlebox Bearing Pressing Operation Instructions", NSK and NTN both specify a minimum value of 0.15mm for the axial clearance of the axlebox bearings of the Harmony 2C Phase II / 380A(L) EMU bogie (see Table 1). Now, let's quantitatively calculate the reduction in the axial clearance of the axlebox bearings caused by temperature rise. Will it be smaller than the minimum value of 0.15mm specified by NSK and NTN? For the sake of simplicity and conservative calculation, it is assumed that the axle box bearing runs continuously for more than 6 hours, and the temperature of the inner ring and roller is only 20℃ higher than that of the outer ring. Take a certain type of imported double-row tapered roller bearing as an example (partial parameter values ​​are shown in Table 2 below, partial parameter values ​​of axle box bearings supplied by Japan's NTN and NSK companies). The length and big end diameter of its roller are 45.8 and 23.903 respectively, the width of the inner ring and the average diameter of the inner raceway are 70 and 170 respectively. Figure 8 It can be seen that the length of the contact area between the roller and the inner ring in the longitudinal direction is almost the same, so the length expansion caused by the temperature rise is the same, resulting in the roller and the inner raceway having the same length after the temperature rises by 20°C. According to the relevant design standards of tapered roller bearings, the distance between the large and small ribs of the inner ring is more than 0.4mm longer than the roller length. Therefore, after the roller becomes longer, it is still between the large and small ribs of the inner ring, and there is no clearance for the clearance of the whole bearing. There are three factors that affect the axial clearance of the whole bearing, namely the width of the inner ring, the diameter of the inner raceway and the roller. It is only necessary to calculate the expansion of these three factors after the temperature rise of 20°C.

[0082] Table 2

[0083]

[0084] The thermal expansion coefficient of steel is 11.7×10 -6 The change in the width of the inner ring directly affects the clearance of the entire bearing at a ratio of 1:1. An inner ring becomes wider when the temperature is 20°C higher than that of the outer ring. The specific value of the expansion of the inner ring width is:

[0085] ΔL 10 =70*11.7×10 -6 *20=0.016(mm),

[0086] The roller diameter increases because the roller temperature is 20°C higher than the outer ring temperature. The specific value of the expansion is:

[0087] ΔL1=23.903*11.7×10 -6 *20=0.006(mm)

[0088] The inner raceway diameter increases because the inner ring temperature is 20°C higher than the outer ring. The specific value of the expansion is: ΔL2=170*11.7×10 -6 *20=0.04(mm)

[0089] The increase in roller and inner raceway diameters directly affects the decrease in axial clearance δ. Their calculation process is as follows:

[0090] The decrease in axial clearance δ caused by the increase in roller diameter has a certain relationship with the angle between the inner raceway and the roller. From Table 2 above, we can see that the cone angle b of the inner raceway is 8°, the cone angle ф of the roller is 1°, and the decrease in axial clearance δ caused by the increase in roller diameter is:

[0091] ΔL 11 =(ΔL1*cosф) / sinb=(0.006*cos1°) / sin8°=0.043mm

[0092] The decrease in axial clearance δ caused by the increase in the diameter of the inner raceway has a certain relationship with the angle of the inner raceway. The decrease in axial clearance δ caused by the increase in the diameter of the inner raceway is:

[0093] ΔL 22 =ΔL2 / (2*tanb)=0.04 / (2*tan8°)=0.142mm

[0094] Since the imported axle box bearing is a double-row tapered roller bearing, the axial clearance between the small end face and the small rib of the roller is reduced by:

[0095] δ=2*(ΔL 10 +ΔL11+ΔL22)=2*(0.016+0.043+0.142)

[0096] =0.4(mm)

[0097] From this, we can conclude that once the temperature of the inner ring and the roller is 20°C higher than that of the outer ring, the axial clearance value of the imported axlebox bearing will become 0.4mm smaller, which is 2.7 times the minimum value of 0.15mm specified by NSK and NTN for the axial clearance of the axlebox bearing of the Harmony 2C Phase II / 380A (L) EMU bogie. This indicates that the clearance value between the roller and the inner and outer raceways is negative, and the bearing is in a "stuck" state. In order to quantitatively calculate the severity of the "stuck" state of the bearing, the minimum value of 0.15mm specified by NSK and NTN for the axial clearance of the axlebox bearing of the Harmony 2C Phase II / 380A (L) EMU bogie is converted into radial clearance. According to the design method of tapered roller bearings, this radial clearance is 0.15mm. The minimum clearance is: 2*0.15*tan10°=0.053(mm), and then evenly distributed to the two rows of bearings. The minimum radial clearance of each row of bearings is 0.026mm. The expansion of each row of inner rings and rollers due to being 20℃ higher than the outer ring has been calculated before, which is ΔL1+ΔL2=0.006+0.04=0.046(mm). This radial expansion is subtracted from the minimum clearance value of each row specified by NSK and NTN, 0.046-0.026=0.02(mm). This 0.02mm is the interference fit of each row of bearings, or the negative clearance value. This negative clearance value is much smaller than the minimum negative clearance value required for normal operation of each row of bearings, i.e. -0.002mm (see Figure 5 ), indicating that the friction torque between the rollers and the large ribs, as well as the inner and outer raceways, is enormous, naturally leading to bluing and wear failures. At this point, the inference at the beginning of this chapter that the double-row tapered roller bearing structure is reasonable is self-defeating. This is because even if the EMU travels on mostly straight tracks, this double-row tapered roller design will still generate additional ineffective derived axial forces (see Table 6 for specific values). The friction torque generated by this additional ineffective derived axial force is 100% converted into heat, causing the roller base surface and large rib, as well as the roller's conical surface and inner raceway, to rise the fastest. This further heat is then transferred to the entire roller and inner ring, causing the roller and inner ring temperatures to be higher than the outer ring. Detailed theoretical and practical data support this conclusion in Chapter 4. According to relevant international standards for high-speed rail, the temperature rise of the axle box bearing is ≤ 80℃. Therefore, if the axle box bearing runs continuously for more than 6 hours under high-speed and heavy-load conditions, the temperature rise will inevitably exceed 20℃, and the temperature difference between the main operating parts and the stationary parts of the bearing must also exceed 20℃. This is the reason why the bearing "gets stuck" and then "gets stuck", which eventually leads to ablation and wear.

[0098] In addition to the above-mentioned erosion (temperature rise) and wear caused by too small axial clearance, the working conditions of grease can also cause erosion (temperature rise) and wear of imported axle box bearings. Although the grease of imported axle box bearings is screened through a strict procedure and can bear the functions required by the axle box bearings of the starting train, and the grease filled in each set of bearings is the same weight, even so, if the grease is applied improperly, it will cause ablation (temperature rise) and wear. Due to the high temperature, the viscosity of the grease will decrease, the evaporation rate will accelerate, and the oxidation rate will accelerate. The high speed of the existing axle box bearings and the running speed of the EMU are all above 250 kilometers per hour. The corresponding axle box bearings are also in a high-speed operation state. In particular, the rollers must revolve around the center line in the inner and outer raceways, and rotate around their own center lines. The centrifugal force is huge, and the grease adhering to the rollers, inner raceways and retainers will be thrown onto the outer raceways. Since the generatrix of the tapered rollers is inclined, the oblique component of force generated forms a "pumping effect" - moving the grease from the small end of the bell mouth formed by the inner and outer raceways to the large end of the bell mouth, that is, the grease is concentrated on the seals on both sides of the bearing (see Figure 17 ), forming a pressure zone at the seal, which is prone to oil leakage and causing a lack of grease at the small heads of the two rows of rollers, affecting the lubrication performance. After running for a long time, the small heads of the rollers will first enter the dry friction state, and the dry friction area will slowly spread to the big heads of the rollers, eventually leading to ablation (temperature rise) and wear.

[0099] Since the contact angle of the axle box of the utility model is below 4° (see Figure 18 ), its structure is similar to that of a cylindrical roller bearing, and it has no obvious bell-mouth structure in the axial direction, and almost no "pumping effect". It has good grease retention ability, and the grease no longer moves to the big end of the roller 7. The roller 7 and the inner raceway 101 and the outer raceway 501 are always in a good lubrication state, ensuring that the life of the bearing is maximized.

[0100] The bearing of this patent is always in the best working state of zero clearance, which means that the clearance between the ball base surface 702 of the roller 7 and the end surface of the step portion 302 of the outer rib 3, and between the conical surface of the roller 7 and the inner raceway 101 of the inner ring 1 and the outer raceway 501 of the outer ring 5 is always zero (see Figure 2 ), the axial clearance of the axle box bearing described below in this patent has nothing to do with the load and life of the bearing. It refers to the value of the axial clearance δ between the small end face 701 of the roller 7 and the working face 602 of the small rib 6 (see Figure 6 ).

[0101] Since the spherical base surface 702 of the axle box bearing roller 7 of the present invention contacts the step end surface 302 of the outer rib 3, and the contact angle a of a certain type of axle box bearing of the present invention is only 3.2° (see Figure 11), resulting in the additional ineffective derived axial force generated between the spherical base surface 702 of the roller 7 and the step end surface 302 of the outer rib 3 being only 0.4 times less than that of the imported bearing (see Table 6). Therefore, the heat and temperature rise generated in this area are only within 40% of the imported bearing, and the friction torque between the roller 7 and the inner raceway 101 of the inner ring 1 and the outer raceway 501 of the outer ring 5 is also reduced by more than 60%, resulting in a temperature rise of the entire bearing being reduced by more than 60%. Since the size of the axial clearance δ of the axle box of this patent has no effect on the load and life of the bearing (see Figure 6 ), so it is only necessary to calculate the axial expansion of the entire bearing according to the maximum temperature allowed by the axlebox bearing, and use this expansion as a reference value for the minimum clearance δ of the axlebox bearing. According to the relevant international standards for high-speed rail, the temperature rise of the axlebox bearing is less than 80°C. Taking an axlebox bearing designed by the utility model with an inner and outer diameter and width of 130, 240 and 160 respectively as an example, for safety reasons, it is assumed that in extreme cases, the inner ring 1 and the roller 7 have a temperature difference of 80°C with the outer ring 5 ( Under normal circumstances, if the actual temperature difference between the outer ring 5, roller 7 and inner ring 1 is less than 80°C, the axial clearance value is greater than the expansion amount when the temperature difference between the outer ring 5, roller 7 and inner ring 1 is 80°C, and the bearing will not get stuck. The maximum outer diameter of the inner raceway 101 of this patented bearing is 167.856, the taper angle of the inner raceway 101 is 2.4°, the taper angle a of the outer raceway 501 is 3.2°, the big end diameter of the roller 7 is 27.936, and the taper angle ф of the roller 7 is 0.4°; Figure 6 It can be seen that the length expansion of the roller 7 and the inner ring 1 in the contact area in the length direction due to the increase in temperature is the same, resulting in the simultaneous elongation of the contact part between the roller 7 and the inner raceway 101. The gap between the small end face 701 of the roller 7 and the working surface 602 of the small rib 6 remains unchanged. Therefore, it is only necessary to calculate the change in axial clearance corresponding to the change in the diameter of the roller 7 and the inner ring 1.

[0102] The diameter of roller 7 increases because the temperature is 80°C higher than that of outer ring 5. The specific values ​​are:

[0103] ΔL1=27.936*11.7×10 -6 *80=0.026(mm)

[0104] The diameter of the inner raceway 101 increases because the temperature is 80°C higher than that of the outer ring 5. The specific value is: ΔL2=167.856*11.7×10 -6 *80=0.157(mm)

[0105] The increase in the diameter of the roller 7 and the inner raceway 101 directly affects the decrease in the axial clearance δ. Their calculation formula is as follows:

[0106] The decrease in axial clearance δ caused by the increase in the diameter of roller 7 is related to the angle between the outer raceway 501 and roller 7. The decrease in axial clearance δ caused by the increase in the diameter of roller 7 is:

[0107] ΔL 11 =(ΔL1*cosф) / sina=(0.026*cos0.4°) / sin3.2°

[0108] =0.47(mm)

[0109] The decrease in axial clearance δ caused by the increase in the diameter of the inner raceway 101 has a certain relationship with the angle of the inner raceway 101. The decrease in axial clearance δ caused by the increase in the diameter of the inner raceway 101 is:

[0110] ΔL 22 =ΔL2 / (2*tanb)=0.157 / (2*tan2.4°)=1.87(mm)

[0111] Since the axle box bearing of this patent is a single-row tapered roller bearing, the axial clearance reduction caused by the roller 7 and the inner ring 1 being 80°C higher than the outer ring is the sum of the two, that is,

[0112] δ=ΔL11+ΔL22=0.47+1.87=2.34(mm), take one decimal place and you get δ=2.3mm

[0113] That is to say, when the minimum axial clearance δ of this patented bearing product at room temperature is 2.3mm (see Figure 6 ), under extreme working conditions, if the temperature of the inner ring 1 and the roller 7 is 80°C higher than that of the outer ring, the sum of the expansion of the inner ring 1 and the roller 7 increases by 2.3mm in the axial direction. At this time, the axial clearance δ becomes 0, and the bearing can still work normally; according to the current technological status of each part that affects the axial clearance of the patented bearing, the manufacturing tolerances of each part that affects the axial clearance are reasonably determined, and the change in the axial clearance caused by the sum of the manufacturing tolerances of each part is determined to be 0.7mm, that is, the variation range of the axial clearance δ value of this patented bearing is 2.3-3mm.

[0114] As can be seen from Chapters 1 and 2 above, in terms of design, the upper and lower limits of the axial clearance of imported axle box bearings are seriously out of tolerance, resulting in some axial clearances of the same batch of bearings being too large and some being too small. Excessive clearance not only affects the uneven force on the inner and outer rings of the bearing, causing stress concentration, and ultimately leading to the peeling of the inner raceway 101 and the outer raceway 501, but also affects the running axis of the inner ring 1 and the roller 7, increasing noise, increasing wear and causing high temperature; when the axial clearance δ is too small, once the inner ring 1 and the roller 7 have a temperature rise of 20°C during operation, the friction between the roller and the inner and outer ring raceways will increase, causing heat and leading to high temperature negative clearance seizure. The national standard for radial clearance of double-row tapered roller bearings, JB / T, revised in 1996, 8236-96 (see Table 3), it was found that even if measured according to this national standard, the clearance value of imported double-row tapered roller axle box bearings was seriously out of tolerance. According to the data in Table 1, the axial clearance ranges of NTN and NSK are 0.15~0.62mm and 0.15~0.561mm respectively. Since the contact angles of the axle box bearings of these two companies are both 10°, according to the conversion relationship between the radial clearance and axial clearance of tapered roller bearings, that is, radial clearance = 2*axial clearance*tan10°, the axial clearance values ​​of the above two companies are substituted into the calculation, and the radial clearance ranges of NTN and NSK are 0.053~0.219mm and 0.053~0.198mm respectively, that is, the radial clearance ranges of NTN and NSK are 53~219μm and 53~198μm respectively. The inner hole size of the axle box bearings of these two companies is 130mm. Compared with the corresponding nominal diameter of 12 in the above table, From the column 0 to 140, it can be seen that the radial clearance range of the axle box bearings of these two companies spans the clearance values ​​of Group 2, Group 0, Group 3, Group 4, and Group 5. Such a large range of clearance values ​​has led to the simultaneous occurrence of batch inner and outer raceway spalling and ablation failures in the bearings supplied by these two companies. Comparing the radial clearance values ​​of each group in the size segment with a nominal diameter of 120 to 140 above, it can be seen that the range of variation between the maximum and minimum values ​​of the radial clearance of Group 1, Group 2, Group 0 and Group 3 is 30μm, and the range of variation between the maximum and minimum values ​​of the radial clearance of Group 4 and Group 5 is 50μm. The range of variation between the maximum and minimum values ​​of the radial clearance of NTN and NSK bearings is 219-53=166μm and 198-53=145μm respectively. In this way, the clearance values ​​of these two Japanese companies are between 2.9 and 5.5 times the Chinese national standard values, which are seriously exceeded, resulting in systematic and batch bearing failures. Domestic railway freight car wheelset bearing manufacturers have been introducing and producing freight train wheelset bearings since 1978. Some models have all the same parameter values ​​as imported high-speed rail axle box bearings. Even if the clearance values ​​are strictly controlled in accordance with national standards, "hot shaft" and other faults cannot be avoided. The root cause is that temperature rise and clearance seriously affect the normal operation of the bearings.The greatest advantage of the present invention is that no matter whether the axial clearance δ is large or small, as long as δ ≥ 0 is maintained under extreme temperature rise conditions, the size of the δ value has nothing to do with faults such as peeling, temperature rise, friction and wear of bearing parts. The large head ball base surface 702 of the roller 7 and the end face of the step portion 302 of the outer rib 3, as well as the conical surface of the roller 7 and the inner raceway 101 and the outer raceway 501 are always in the optimal working state of zero clearance.

[0115] Table 3

[0116] Double row and four row tapered roller bearing radial clearance μm

[0117]

[0118] Approved by the Ministry of Machinery Industry on September 3, 1996, and implemented on July 1, 1997

[0119] 3. The third most common failure rate is retainer breakage, which is also 10%. Since the author of this patent cannot obtain the drawings of the plastic retainers of imported axle box bearings, he can only analyze the domestic railway freight train wheelset bearing 352226X2, which is of the same size as the imported high-speed rail axle box bearings. The inner and outer diameters and structure of 352226X2 are exactly the same as those of an existing imported high-speed rail axle box bearing, and the retainer material and manufacturing process are also the same. Although the plastic retainer used in 352226X2 has many advantages over the iron retainer, it also has insurmountable shortcomings in design and process. Table 5 shows the parameter tolerance values ​​of some national standards for tapered roller bearing retainers (technical standard ZYB9-93). For example, the design requires a bilateral clearance ε between the retainer and the roller. △The clearance is only 0.25 mm (see Table 5). During assembly, the required clearance can be achieved by controlling and adjusting the shrinkage of the iron retainer. However, since plastic retainers cannot expand and contract due to design and manufacturing processes, it is naturally impossible to control the gap between the roller and the retainer by controlling the shrinkage of the retainer. Table 4 shows the main dimensions, radial play, water content, and other quality requirements for the plastic retainer used in freight train wheelset bearings 352226X2, which have the same structure and inner and outer diameter dimensions as EMU axlebox bearings (Technical Standard TJCL079-2013). At present, the radial play between the plastic retainer and the roller is 0.15mm~0.55mm (see Table 4), far exceeding the national standard value of 0.25mm. In addition, the error of the inner diameter of the small end of the retainer is ±0.4mm (see Table 4), the manufacturing error of the outer diameter of the small rib is 0.4mm (see d3 in the "Tapered Roller Bearing Design Method" compiled by Luoyang Bearing Research Institute), and the radial clearance error of the bearing is 0.053mm~0.219mm for NTN and 0.053mm~0.198mm for NSK. These errors and tolerances accumulate together, making the gap between the roller and the retainer seriously exceed the standard. The retainer is no longer along the bearing. The center line rotates, and the center of gravity of the retainer deviates from the center line of the bearing, resulting in radial runout and collision between the rollers. Since the radial clearance of the bearing itself also exceeds the standard, there is clearance between the tapered surface of the roller and the raceway of the outer ring. The roller continues to impact the retainer under the action of centrifugal force, causing the gap between the roller and the retainer to gradually increase. The roller is no longer simply rolling in the window of the retainer, but rolling, sliding and swinging at the same time. As the gap between the retainer and the roller becomes larger and larger, the center line of the retainer deviates seriously from the center line of the bearing, the radial swing of the retainer also becomes larger and larger, and the working conditions between the retainer and the roller become worse and worse until the retainer breaks. In addition, the existing axle box bearing retainer has an obvious defect, that is, this retainer has a detachable column or beam, so that it can be removed to inspect the inner raceway of the inner ring during "planned maintenance". After the inspection, all parts must be cleaned and re-greased. Since the inner ring, roller and retainer form an internal assembly together, it is difficult to clean them, and it is also inconvenient to inspect the working surface of the roller, which poses a hidden danger to the subsequent bearing loading and operation. Moreover, this detachable column or beam is definitely different in function from the other integrated columns or beams, and it also affects the normal operation of the bearing.

[0120] Table 4

[0121]

[0122] Table 5

[0123]

[0124] This maintenance-free axle box bearing retainer 4 Figure 19 As shown, in terms of design, the entire retainer 4 is an integral structure, and no detachable column or beam is designed. Since the axle box bearing of this patent has a "0" clearance in the structural design (see Figure 2 ), when all rollers 7 rotate along the inner raceway 101 and the outer raceway 501, the gap between their conical surface and the inner and outer raceways is always 0, and the gap between their spherical base surface 702 and the step end face 302 of the outer rib 3 is always 0. In this way, when the rollers 7 rotate at high speed, their centrifugal force is borne by the outer raceway 501 of the outer ring 5, and their axial force is borne by the step end face 302 of the outer rib 3, thus avoiding the impact of the rollers on the cage due to the centrifugal force caused by the radial clearance of the existing bearing being greater than the gap between the rollers and the cage; in addition, by controlling the accuracy of the cage injection mold, the cage The tolerances of the inner diameter size and window width size of the small end of the retainer are controlled within an appropriate range, so that the play between the retainer and the roller meets the industry standard value; another factor that affects the gap between the retainer and the roller of the existing bearing is the small rib. Since the plastic retainer cannot expand and contract, the small rib must retain the roller inside the inner raceway and prevent it from falling out. The outer diameter of the small rib must be larger than the small end diameter of the inner raceway 501. Since the outer diameter of the small rib itself has tolerance requirements, coupled with the aforementioned inner diameter size tolerance of the retainer small end and the window width size tolerance, as well as the radial clearance of the bearing, these tolerances cause the gap between the retainer and the roller to be seriously out of tolerance. The small rib 6 of this patent eliminates the influence of its outer diameter size on the gap between the retainer and the roller from the design (see Figure 1 ), the parts diagram of the small rib 6 is shown in Figure 13 , whose inner hole is threaded (see Figure 1 ) or screw riveting (see Figure 14 ) is positioned and fixed on the outer thread 102 of the small head of the inner ring 1. When assembling, first assemble the roller and retainer assembly to the inner raceway. At this time, the gap between the retainer and the roller meets the design standard, and then the small rib 6 is threaded (see Figure 1 ) or screw riveting (see Figure 14 ) is positioned and fixed on the outer cylindrical surface 102 of the small end of the inner ring 1; if according to the current maintenance plan of the EMU company, when the axle box bearing runs to a certain mileage and needs "planned repair", it is only necessary to remove the small rib 6 (see Figure 13) Loosen several radial set screws, unscrew the small rib 6 counterclockwise, remove the retainer 4 and roller 7, and clean and inspect the retainer 4, roller 7 and inner ring 1 separately. This way, the cleaning is cleaner and the inspection is convenient. After the inspection, if there is no problem, first put the roller 7 into the window of the retainer 4, put the inner ring 1 into the assembly of roller 7 and retainer 4, and screw the small rib 6 clockwise onto the small head external thread 102 of the inner ring 1. Tighten several radial set screws of the small rib 6, and the entire inner assembly is assembled. Since the axle box bearing of this patent is a maintenance-free bearing, when the bearing is found to be intact every time it is "planned repaired", it will gradually no longer be overhauled and maintained. The bearing will run from installation to the designed life and then be directly scrapped.

[0125] The remaining faults with relatively low failure rates, such as abnormal noise, inner ring breakage, roller fragmentation, etc., are most likely related to abnormal operating conditions of the bearings. Since they account for a small proportion and are not representative, they cannot be analyzed.

[0126] 4. In addition to the above-mentioned design defects exposed by practice, existing imported axle box bearings also have problems with temperature rise and high energy consumption. The temperature rise reduces the viscosity of the grease, accelerates evaporation and oxidation, and causes a "pumping effect" that causes grease loss, thereby affecting the life of the bearing. High energy consumption pushes up the operating costs of high-speed rail. The following is a detailed calculation of the temperature rise and energy consumption of the two bearings. The structural parameter values ​​of the imported double-row tapered roller bearings are shown in Table 2 above. From the parameters in this table, it can be seen that the contact angle between the roller and the rib is 8° (the actual contact angle is 8.5°). Since there is sliding friction between the spherical base surface of the roller and the large rib, its friction coefficient is much greater than the rolling friction coefficient. Even if the high-speed rail only travels in a straight line for most of the distance and does not apply any axial force to the axle box bearing, due to the structural reasons of the double-row tapered roller bearing, a large additional invalid derived axial force is still generated. The specific calculation process is as follows:

[0127] (1) The existing axle box bearing force analysis (see Figure 9). Currently, the imported axle box bearings are double-row tapered roller bearings. When the EMU body applies a pure radial force Qa to the outer ring, the outer ring transfers Qa evenly to the two inner rings through the two rows of rollers. Since all the parameters of the two inner and outer raceways and the two rows of rollers are the same, it is only necessary to calculate the force conditions of one inner and outer raceway and roller, and then multiply it by 2 to get the force conditions of the entire bearing. In this case, it can be further simplified and the two inner and outer raceways and the two rows of rollers can be calculated as one inner and outer raceway and one row of rollers. Calculation is equivalent to a single-row bearing under pure radial force Qa. According to the parameters listed in the table above, the semi-cone angle a of the outer raceway of the existing bearing is 10°, the semi-cone angle b of the inner raceway is 8°, and the angle c of the large rib is 8.5°. When the outer ring is subjected to pure radial force Qa, the outer ring decomposes the pure radial force Qa into vertical and horizontal forces through the rollers and applies them to the raceway and rib of the inner ring respectively. The roller is in a state of equilibrium under the action of these three forces of the inner and outer raceways and the large rib, so:

[0128] Qn-Qi-Qy=0 (1)

[0129] Qa-Qe+Qx=0 (2)

[0130] Among them, Qn=Qm*sina, Qa=Qm*cosa, Qi=Qb*sinb, Qe=Qb*cosb, Qx=Qz*sinc, Qy=Qz*cosc,

[0131] Substituting the angle values ​​of a, b, and c and these six equations into equations (1) and (2) and simplifying them, we obtain:

[0132] Qb=Qm*(sina*sinc+cosa*cosc) / (sinb*sinc+cosc*cosb)≈Qm

[0133] Qz=Qm*(sina*cosb-cosa*sinb) / (sinb*sinc+cosc*cosb)

[0134] =0.035Qm

[0135] The radial force borne by the inner ring is:

[0136] Qe+Qx=Qb*cosb+Qz*sinc=Qm*0.99+0.035Qm*0.148=Qm=Qa / cosa=1Qa

[0137] The axial force borne by the inner ring raceway is:

[0138] Qi=Qb*sinb=Qm*sinb=0.139Qm=0.139Qa / cosa=0.141Qa

[0139] The axial force on the outer ring raceway is Qn = Qa * taga = 0.176Qa

[0140] The axial force on the large flange is Qy = Qz*cosc = 0.035Qm*cosc = 0.035Qa

[0141] The axial force on the roller is Qn+Qi+Qy=Qa*taga+Qb*sinb+Qz*cosc

[0142] =Qa(taga+sinb / cosa+0.035cosc / cosa)=0.352Qa

[0143] (2) The relevant parameters of the axle box bearing designed in this embodiment are as follows: Figure 6 , see the force analysis Figure 10 The outer raceway semi-cone angle a of this axle box bearing is 3.2°, the inner raceway semi-cone angle b is 2.4°, and the outer rib angle c is 3.15°. The outer ring decomposes the pure radial force Qa into vertical and horizontal forces through the rollers and applies them to the inner ring and outer rib respectively. The roller is in a state of equilibrium under the action of these three forces of the inner and outer raceways and the outer rib, so:

[0144] Qn-Qi-Qy=0 (1)

[0145] Qa-Qe+Qx=0 (2)

[0146] Among them, Qn=Qm*sina, Qa=Qm*cosa, Qi=Qb*sinb, Qe=Qb*cosb, Qx=Qz*sinc, Qy=Qz*cosc,

[0147] Substituting the angle values ​​of a, b, and c and these six equations into equations (1) and (2) and simplifying them, we obtain:

[0148] Q b =Qm*(sina*sinc+cosa*cosc) / (sinb*sinc+cosc*cosb)=Qm

[0149] Q z =Qm(sina*cosb-cosa*sinb) / (sinb*sinc+cosc*cosb)

[0150] =0.014Qm

[0151] The radial force borne by the inner ring is:

[0152] Q e=Qa+Qx=Qa+Qz*sinc=Qa+0.014Qm*sinc=Qa+0.00077Qm

[0153] =Qa(1+0.00077 / cosa)=Qa

[0154] The axial force borne by the inner ring is:

[0155] Qn-Qy=Qm*sina-Qz*cosc=Qm*(sina-0.014cosc)

[0156] =0.042Qm=0.042Qa / cosa=0.042Qa

[0157] The axial force on the outer ring is Qn = Qa * taga = 0.056Qa

[0158] The axial force on the outer flange is Qy = Qz*cosc = 0.014Qm*cosc = 0.014Qa

[0159] The axial force on the roller is Qn+Qi+Qy=Qa*taga+Qb*sinb+Qz*cosc

[0160] =Qa(taga+sinb / cosa+0.014cosc / cosa)=0.11Qa

[0161] (3) The following conclusions can be drawn from the calculations in (1) and (2):

[0162] A. When the outer rings of the two types of axle box bearings are subjected to pure radial force, the outer rings transmit the same magnitude of radial force to the inner rings through the rollers, that is, Qe = Qa.

[0163] B. From the additional invalid derived axial forces borne by each component in Table 6, it can be seen that through innovative design, the various components of the axle box of this patent can reduce 59-78% of the additional invalid derived axial forces, as well as the same percentage of friction torque and heat, greatly reducing temperature rise and energy consumption.

[0164] C. Under the same radial load, the derived axial force generated by the outer rib of the axle box bearing of this patent is 0.41 times that of the large rib of the existing bearing. Overall, the patented axle box bearing not only reduces the derived axial force by 59%, but more importantly, because there is sliding friction between the step portion 302 of the rib 3 and the roller ball base surface 702, its friction coefficient is much greater than the rolling friction coefficient. This weak link is affected by the derived axial force and friction torque reduced by 59%, and the heat and temperature generated by the sliding friction torque are also reduced by about 59%, which greatly reduces the temperature rise of the entire bearing, ensures that the grease does not deteriorate due to temperature rise, extends the life of the bearing, and eliminates the misunderstanding that bearing overheating is attributed to unqualified grease; the significant reduction in derived axial force improves energy utilization and reduces the operating cost of high-speed rail. The above theoretical analysis and calculation results have been verified in practice. According to relevant literature, taking the experiments of Geng Zilin, Song Dongli, Zhang Weihua and others as an example, they took a certain type of high-speed EMU as the test object, established a vehicle-track coupling dynamics model and a double-row tapered roller dynamics model, and calculated the friction power consumption of each contact part in the bearing based on the friction heat generation theory. The results show that the part with the highest temperature inside the axle box is the contact area between the rolling element and the inner ring rib, that is, the contact area between the roller ball base surface and the inner ring large rib.

[0165] Table 6

[0166]

[0167] As is known to all, the two inner rings of the currently imported axle box bearings, i.e. double row tapered roller bearings, bear radial forces and transmit axial forces in two directions respectively. Although the internal structure of the axle box bearing of the present invention is a set of single row tapered roller bearings, it can still transmit radial forces and axial forces in two directions at the same time. When the inner ring 1 is subjected to an axial force in the left direction, the axial force is transmitted through the AA path, i.e. inner ring 1-inner raceway 101-roller 7-outer raceway 501-outer ring 5 (see Figure 3 ); When the inner ring 1 is subjected to an axial force in the right direction, the axial force is transmitted through the BB path, that is, the inner ring 1-small rib 6-working surface 601 of the small rib 6-small end surface 701 of the roller 7-spherical base surface 702 of the roller 7-step end surface 302 of the outer rib 3-outer rib 3 (see Figure 4 ).

[0168] This embodiment innovates and optimizes the currently imported double-row tapered roller bearings into single-row tapered roller bearings, and optimizes the two-row rollers into single-row rollers. The number of chamfers and the length of the roller 7 are reduced by 50%, and the corresponding effective working area of ​​the roller 7 is increased by 50%. Moreover, since the length of the roller 7 reaches more than 100mm, the radial basic rated dynamic load and fatigue life are increased by increasing the assembly height, thereby compensating for the reduction in the radial basic rated dynamic load and fatigue life of the bearing caused by insufficient raw materials and processing accuracy. The utility model is a certain Take a high-speed rail axle box bearing as an example. Its inner and outer diameters are 130 and 240 respectively, which are the same as the current imported axle box bearings. The other main parameters are as follows: Lw = 120, a = 3.2°, Z = 20, roller big end diameter Dw = 27.936, roller semi-cone angle ф = 0.4°. According to the data and calculation, bm = 1.1, fc = 87.7, Dwe = 27.1, Lwe = 119, single-row bearing i = 1. Substituting the above parameters into the calculation formula of the basic radial dynamic load rating Cr of the tapered roller bearing:

[0169] C r =b m *f c *(i*L we *cosa) 0.778 *Z 0.75 *D we 1.074

[0170] =1.1*87.7*(119*cos3.2°) 0.778 *20 0.75 *27.1 1.074

[0171] =1298(KN)

[0172] Under the same conditions as the inner and outer diameters of imported axle box bearings, the number of parts of the axle box bearing of this patent is reduced from 10 to 8, while the radial basic rated dynamic load is significantly higher than that of the imported bearings. Furthermore, if the imported axle box bearing also adopts the design of this patent, the existing double-row tapered roller bearing is converted into a single-row tapered roller bearing, and the above model is used as an example, the dimensions of the various parts of the imported bearing are also the same as those in the example above of this patent, with only the contact angle being different (see Table 2), i.e., a=10° for the imported bearing, the semi-cone angle ф of the roller is 1°, and the other main dimensional parameters remain unchanged as follows: L w =120, Z=20, roller big end diameter D w =27.936, according to the data and calculation, b m =1.1,f c =85.4, D we =25.86, L we=119, single-row bearing i=1, substitute the above parameters into the calculation formula of the basic radial dynamic load rating Cr of the tapered roller bearing:

[0173] C r =b m *f c *(i*L we *cosa) 0.778 *Z 0.75 *D we 1.074

[0174] =1.1*85.4*(119*cos10°) 0.778 *20 0.75 *25.86 1.074

[0175] =1190(KN),

[0176] And the radial basic rated dynamic load C of a certain model of this patent r =1298KN, the imported bearing only has the same contact angle as the existing a=10°, and its radial basic rated dynamic load becomes C r =1190KN, a=3.2° of the patented bearing, and its radial basic rated dynamic load is increased by (1298-1190) / 1190=9% compared with the imported bearing, indicating that the radial load-bearing capacity of the patented bearing has increased. Moreover, the increase in its radial load-bearing capacity just corresponds to the decrease in its additional invalid derived axial force. When the vehicle does not turn sharply, the bearing does not bear the axial force. The imported bearing generates more additional invalid derived axial force simply because the contact angle is larger than that of the patented bearing, all of which is converted into heat. This is the root cause of the excessive temperature rise of the imported bearing and the high energy consumption of the imported bearing.

[0177] In terms of maintenance, since this patented bearing has the advantages of ultra-long mileage maintenance-free, ultra-long life, high rigidity, high reliability, ultra-low temperature rise, and high stability, the maintenance mode of high-speed trains will gradually move from "planned maintenance" to "preventive maintenance" mode, and finally to "maintenance-free" mode. From loading to scrapping and replacement, no form of maintenance is required for the bearings.

Claims

1. A maintenance-free rail transit axle box bearing, characterized by: It includes an inner ring (1), an outer oil seal (2), an outer rib (3), a retainer (4), an outer ring (5), a small rib (6), a tapered roller (7) and an inner oil seal (8); An inner raceway (101) is provided on the circumferential outer circle of the middle section of the inner ring (1); the portion of the inner ring (1) located at the left end of the inner raceway (101) is marked as a small head (102), and the portion at the right end is marked as a large head (103); the inner raceway (101) is in the shape of a conical surface; the small head (102) and the large head (103) are both in the shape of a cylinder; The small rib (6) is annular in structure as a whole; the small rib (6) is sleeved on the periphery of the small head (102); a working surface (601) is provided on the end surface of the small rib (6) facing the inner raceway (101); The inner hole of the outer ring (5) is marked as an outer raceway (501) and a cylindrical section (502) from the right end to the left end; the outer raceway (501) is in a conical shape; the small diameter end of the outer raceway (501) is connected to the cylindrical section (502); The outer rib (3) comprises an outer rib body (301) and a step portion (302); the outer rib (3) is arranged on the right end side of the outer ring (5); the step portion (302) extends into the outer raceway (501); The outer ring (5) is coaxially sleeved on the outer side of the inner ring (1); the outer raceway (501), the inner raceway (101), the small rib (6) and the step portion (302) enclose a friction working portion; the retainer (4) is arranged in the friction working portion; a tapered roller (7) is arranged in the window of the retainer (4); the tapered roller (7) rolls in the friction working portion; the retainer (4) drives the tapered roller (7) to rotate along the outer raceway (501), the inner raceway (101) and the step portion (302) in the circumferential direction; the tapered surface of the tapered roller (7) contacts the inner raceway (101) and the outer raceway (501), and the spherical base surface contacts the step portion (302); The outer oil seal (2) is arranged in the gap between the large end (103) of the inner ring (1) and the outer rib (3); the outer oil seal (2) is arranged on the right end side; the outer circle of the outer oil seal (2) and the inner hole of the outer rib body (301) are interference fit, and the lip of the outer oil seal (2) and the outer circle surface of the large end (103) of the inner ring (1) are interference sealed fit; The inner oil seal (8) is arranged in the gap between the small rib (6) and the cylindrical section (502) of the outer ring (5); the inner oil seal (8) is arranged on the left end side; the outer circle of the inner oil seal (8) and the inner hole of the cylindrical section (502) are interference fit, and the lip and the outer circle of the small rib (6) are interference sealing fit.

2. The maintenance-free rail transit axle box bearing according to claim 1, characterized in that: The small rib (6) is fixed on the small head (102) in a mechanical manner.

3. The maintenance-free rail transit axle box bearing according to claim 1, characterized in that: The retainer (4) is made of plastic.

4. The maintenance-free rail transit axle box bearing according to claim 1, characterized in that: The semi-cone angle of the outer raceway (501) is a=1° to 9°.

5. A maintenance-free rail transit axle box bearing according to any one of claims 1 to 4, characterized in that: Maintenance-free axlebox bearings are used in the wheel hubs of high-speed railways, subways, ordinary passenger trains, ordinary freight trains or motor vehicles.

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