Maintenance-free rail transit axle box bearing

Abstract

The invention provides a maintenance-free rail transit axle box bearing. The bearing comprises an inner ring, an outer ring, a roller, a retainer, an outer flange and an oil seal. The invention relates to an axle box bearing of a high-speed train, a metro train and a common passenger train. The bearing has the advantages of being small in number of parts, good in assembly performance, compact in structure, large in load, low in noise, low in torque, ultralow in temperature rise, ultralow in friction torque, ultralow in cost, integrated, super-energy-saving, always free of pre-tightening (adjustment) zero clearance, ultra-long in mileage, free of maintenance, ultra-long in service life, high in strength, rigidity, reliability and stability and the like. 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 energy conservation and environmental protection are realized.

Description

Technical Field

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

[0002] High-speed rail axle box bearings are key components of high-speed trains, and their safety and reliability are crucial to the safety of passengers. Figure 7 , existing imported bearings have a series of systematic design defects. These systematic design defects not only affect the performance and life of the bearings, but also increase the maintenance cost and operation risk of high-speed rail:

[0003] 1. Outer raceway spalling: The highest failure rate is the outer raceway spalling, accounting for up to 44%. The spalling area is concentrated in the area with the largest load on the outer raceway, which is caused by the stress concentration problem. The outer ring is stationary during operation, and the inner ring drives the roller to rotate along the outer raceway. Due to the existence of axial clearance, the outer raceway actually bears a local load rather than a uniform load.

[0004] 2. Inner raceway spalling: Inner raceway spalling ranks second in the failure type, accounting for 12%. Although the inner raceway should theoretically bear the cyclic load evenly, due to the influence of axial clearance, it actually bears stress concentration cyclic load rather than uniform cyclic load.

[0005] 3. Temperature rise and wear problems: Temperature rise and wear account for the third and fifth proportion of failures, accounting for 10% and 8% respectively. Due to the original structural design defects of imported axle box bearings, temperature rise and "pumping effect" will inevitably lead to temperature rise. The temperature rise causes the viscosity of the grease to decrease. The "pumping effect" and high speed cause the grease to move to both ends of the bearing, resulting in a lack of grease on the roller head, and then loss. This causes dry friction and rapid temperature rise in the contact area between the inner and outer raceways and the roller head. When there is almost no grease in the working area, wear and ablation are the stages that the bearing continues to operate.

[0006] 4. Retainer breakage problem: The third most common problem is retainer breakage, also accounting for 10%. Although the plastic retainer currently used has many advantages over the iron retainer, it also has insurmountable disadvantages in design and process. For example, the design requires that the single-sided gap between the retainer and the roller is only 0.125 mm. If the retainer is made of low-carbon steel plate, the above gap requirement can be achieved by controlling and adjusting the shrinkage of the iron retainer during assembly. Since the plastic retainer cannot shrink in design and process, it is naturally impossible to control the gap between the roller and the retainer. In addition, the radial manufacturing error of the retainer, the manufacturing error of the outer diameter of the small flange and the radial clearance tolerance of the bearing, these errors and tolerances accumulate together, so that the centrifugal force generated by the roller during rotation continuously impacts the retainer, resulting in a gradual increase in the gap between the roller and the retainer (see the specific implementation method below). The roller is no longer simply rolling in the window of the retainer, but rolling, sliding and swinging exist at the same time. As the gap between the retainer and the roller becomes larger and larger, the motion trajectory of the retainer also changes. The retainer is no longer rotating along the center line of the bearing. The center line of the retainer and the center line of the bearing no longer coincide. The working conditions between the retainer and the roller are getting worse and worse until it breaks.

[0007] 5. Problems related to abnormal operating conditions: Although faults such as abnormal noise, inner ring breakage, and roller breakage account for a small proportion, they are usually related to abnormal operating conditions of the bearing.

[0008] 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 lubricating oil to volatilize and oxidize, which will also affect the life of the bearing. Since the temperature rise is caused by friction energy consumption, and the high friction energy consumption is caused by the design of the bearing, the high energy consumption also increases the operating cost of the high-speed rail.

[0009] 7. Bearing clearance problem: The bearing clearance has a significant impact on fatigue life, temperature rise, noise, vibration, etc. The existing structure makes it impossible to adjust the bearing clearance to the optimal state, resulting in performance degradation, shortened life, and regular overhaul and maintenance.

[0010] 8. The above problems are not only problems of axle box bearings of high-speed railways at home and abroad, but also common problems of wheelset bearings of ordinary freight trains and passenger trains at home and abroad. These problems have not been solved since the birth of the train in 1804.

[0011] Therefore, developing rail transit axle box bearings with originality and completely independent intellectual property rights and solving these design defects are of great strategic significance for improving the autonomy and controllability of high-speed rail technology. Summary of the invention

[0012] The purpose of the present invention 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 present invention is as follows: a maintenance-free rail transit axle box bearing includes 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 arranged on the circumferential outer circle of the middle section of the inner ring. The part of the inner ring located at the left end of the inner raceway is marked as a small head, and the part at the right end is marked as a large head. The inner raceway is in the shape of a conical surface. Both the small head and the large head are cylindrical.

[0015] The small rib is an annular structure as a whole. The small rib is sleeved on the periphery of the small head. A working surface is arranged on one 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 in a conical shape. The small diameter end of the outer raceway is connected to the cylindrical section.

[0017] The outer rib comprises an outer rib body and a step portion. The outer rib is arranged at 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 portion enclose a friction working portion. The retainer is arranged in the friction working portion. A tapered roller is arranged in the window of the retainer. The tapered roller rolls in the friction working portion. The retainer drives the tapered roller to rotate along the outer raceway, inner raceway and step portion in the circumferential direction. The tapered surface of the tapered roller contacts the inner raceway and the outer raceway, and the ball base surface contacts the step portion.

[0019] The outer oil seal is arranged in the gap between the big head and the outer rib. The outer oil seal is arranged on the right end. The outer circle of the outer oil seal and the inner hole of the outer rib body are interference fit, and the lip of the outer oil seal and the outer circle surface of the big head are interference seal fit.

[0020] The inner oil seal is arranged in the gap between the small rib and the cylindrical section of the inner hole of the outer ring. 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 are interference fit, and the lip and the outer circle of the small rib are interference seal fit.

[0021] Furthermore, the small rib is fixed on 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 effect of the present invention is unquestionable: it has the advantages of small number of parts, good assembly performance, compact structure, large load, low noise, low torque, low cost, ultra-low friction torque, lightweight, integration, energy saving, always no pre-tightening (adjustment) 0 clearance, ultra-long mileage maintenance-free, ultra-long life, high strength, high rigidity, high reliability, high stability, etc., and significantly improves 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, greatly 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, and has remarkable energy-saving and environmental protection effects, and is simple and easy to disassemble and install during maintenance. BRIEF DESCRIPTION OF THE DRAWINGS

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

[0027] Figure 2 The working principle diagram of maintenance-free rail transit axle box bearing without adjustment (control) and zero clearance;

[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 It is a schematic diagram of the ratio of bearing radial clearance and bearing life;

[0031] Figure 6 Axial clearance variation diagram of maintenance-free rail transit axle box bearing under thermal expansion and contraction;

[0032] Figure 7 This is a diagram of the fault types of axle box bearings of 50 imported EMUs that were randomly sampled and scrapped prematurely;

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

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

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

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

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

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

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

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

[0041] Fig.16 Product diagram of the outer rib of a maintenance-free rail transit axle box bearing;

[0042] Fig.17 This is a schematic diagram of the loss of grease in the existing high-speed rail axle box bearing due to the "pumping effect";

[0043] Fig.18 Schematic diagram of lubrication conditions for maintenance-free rail transit axle box bearings.

[0044] Fig.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, ball base surface 702, inner oil seal 8. DETAILED DESCRIPTION

[0046] The present invention is further described below in conjunction with the embodiments, but it should not be understood that the above subject matter of the present invention is limited to the following embodiments. Without departing from the above technical ideas of the present invention, various substitutions and changes are made according to the common technical knowledge and customary means in the art, which should all be included in the protection scope of the present invention.

[0047] Embodiment 1:

[0048] See also Figures 1 to 19 The present 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 circumferential outer circle 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 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 a conical shape. Both the small head 102 and the large head 103 are cylindrical.

[0050] The small rib 6 is an annular structure as a whole. The small rib 6 is sleeved on the outer periphery of the small head 102 of the inner ring 1. A working surface 601 is arranged 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 comprises an outer rib body 301 and a step portion 302. The outer rib 3 is arranged at 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 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. There is a tapered roller 7 in the window of the retainer 4, and the tapered roller 7 rolls in the friction working portion. The retainer 4 drives the tapered roller 7 to rotate along the inner and outer raceways in the circumferential direction. The tapered surface of the tapered roller 7 contacts the inner raceway 101 and the outer raceway 501, and the ball base surface 702 contacts the step portion 302.

[0054] The outer oil seal 2 is arranged in the gap between the large head 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 head 103 of the inner ring 1 are interference fit.

[0055] The inner oil seal 8 is arranged in the gap between the small rib 6 and the inner hole 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 seal fit.

[0056] The maintenance-free rail transit axle box bearing of this embodiment has the advantages of few parts, good assembly performance, compact structure, large load, low noise, low torque, ultra-low friction torque, lightweight, integrated, super energy-saving, always free of preload (adjustment) 0 clearance, ultra-long mileage maintenance-free, ultra-long life, ultra-low cost, high strength, high rigidity, high reliability, high stability, etc., and significantly improves 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 axle box bearing of this embodiment is not only suitable for high-speed rail, but also for ordinary passenger trains, subways, freight trains and all motor vehicles. At present, the wheel bearings of ordinary passenger and freight trains are the same as the currently imported axle box bearings in terms of type and structure. At the beginning, they were 100% imitations of imported freight train wheel bearings. Due to defects in the original design, after large-scale loading, the "hot axle" and "hot axle" phenomena occurred due to the design problems of the wheel bearings themselves, resulting in many major accidents of vehicle derailment and subversion. Later, due to the improvement of bearing manufacturing precision and quality, and the improvement of "hot axle" monitoring and prevention measures, after entering the new century, major accidents such as "hot axle" have been reduced to single digits each year, and "hot axle" accidents have also dropped significantly. However, the source of the "hot shaft" has not been eliminated. Only by adopting the axle box bearing of the 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 formed 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 Fig.10), r represents the distance from the point of force application to the center line of the bearing, 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, rotation 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". Not only that, 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 tapered surface of the roller and between the large rib of the inner ring and the ball base surface of the roller, which generate more heat. Because of the clearance between the outer raceway of the outer ring and the tapered surface of the roller, the outer raceway and the tapered surfaces of all rollers are not in full contact, resulting in less 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. There is only one friction pair between the inner ring and the roller of the bearing of this patent, that is, the rolling friction pair between the inner raceway of the inner ring and the conical surface of the roller, which has one less sliding friction pair than the existing structure. Therefore, 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 if there are two friction pairs between the outer raceway 501 and the conical surface of the roller, and between the ball base surface 702 of the roller and the end face of the step portion 302 of the outer rib 3, the heat generated is much more 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 uniform temperature rise of the outer ring 5 and the outer rib 3, and the volume and surface area of ​​the outer ring 5 are much larger than those of the inner ring. Even if the heat generated between the outer ring 5 and the roller 7 and between the roller 7 and the outer rib 3 is much more 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 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 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, construction machinery, etc., the force properties of their wheel hub bearings are exactly the same as those of high-speed railways, subways, ordinary passenger and freight trains, except that the magnitude of the force and the speed of the two are different. The wheel 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 wheel hub bearings of the above-mentioned motor vehicles are also within the protection scope of this patent.

[0060] Embodiment 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. Fig.13 The inner hole is positioned and fixed on the small head 102 of the inner ring 1 by means of threads or interference fit. According to the function, the small rib 6 can be divided into two types: fixed type and removable type; the fixed type is further divided into two types, one of which is to fix the small rib 6 by turning the small end face of the inner ring 1 by the spinning riveting process (see Fig.14 ), one is to fix the small rib 6 by the 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, several 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 repair" to check 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 component together, it is difficult to clean them, and it is not convenient to check the working surfaces of the roller 7 and the inner raceway 101, which buries hidden dangers 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 Fig.19), there is no longer a detachable column or beam, 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. In 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 present invention can achieve this. The small rib 6 of the 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 in the radial direction of 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 to 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, 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, 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, it will run until the end of its life date and can be directly replaced. No major overhaul or maintenance is required in the middle, truly crushing imported products in all aspects, all processes, and the entire life cycle.

[0063] Embodiment 3:

[0064] The main contents of this embodiment are the same as those of embodiment 1 or 2, wherein the retainer 4 is made of plastic. In actual production, the material of the retainer 4 is selected from 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. The retainer 4 is once-injected by an injection mold, and the main technical standard is to ensure that the gap between the retainer 6 and the roller 7 meets the standard.

[0065] Embodiment 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] Embodiment 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 that of the current imported bearings. For example, the inventor 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 Fig.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, but the temperature rise of the bearing is also reduced, the radial dynamic load rating is increased, and the service life is correspondingly increased.

[0069] The biggest feature of this axle box bearing roller 7 is that it is particularly slender (see Fig.12 ), the conical surface of the roller cannot be processed by through-grinding, and can only be processed by cut-in grinding. In this way, as long as the grinding wheel is dressed into the shape of a logarithmic curve, a roller with a 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 is 100% inspected 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] Embodiment 6:

[0071] This embodiment discusses the technical effect of any one of the embodiments 1 to 4, wherein when the high-speed rail is running straight on the track, the axle box bearing only bears radial force, and only when the high-speed rail turns, the axle box bearing bears both radial force and axial force. In order to allow the axle box bearing to bear axial force in two directions, the existing high-speed rail is either equipped with a set of imported double-row tapered roller bearings (see Figure 8 ), the following content only selects double-row tapered roller bearings as an example for comparison and discussion.

[0072] Statistics show that as of 2019, a total of 220,000 sets of axle box bearings for high-speed EMUs were imported, and 550 sets of bearings failed prematurely and were scrapped before their design life, with a failure rate of 0.25%. The failure types of 50 sets of axle box bearings tested by the Bearing Quality Supervision and Inspection Center between 2013 and 2020 are as follows: Figure 7 As shown, these 50 sets of bearings are randomly selected from 550 sets of scrapped bearings and are representative. The 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 fault types are the spalling of the outer raceway 501 and the inner raceway 101 of the bearing. In particular, the spalling of the outer raceway 501 accounts for the largest proportion, reaching 44%. Moreover, the spalling locations are all located at the same azimuth 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 with the file number "sfet-77-48300" and the file name "EMU Maintenance Axle Box Bearing Pressing Operation Instructions", NTN and NSK companies have specified the axial clearance values ​​of the bogie axle box bearings of the Harmony 2C Phase II / 380A (L) EMU as 0.15~0.62 and 0.15~0.561 respectively (see the table below). Table 1 shows the axial clearance values ​​of the axle box bearings supplied by NTN and NSK after installation.

[0074] Table 1

[0075]

[0076] Now, we can calculate the maximum radial clearance based on the maximum axial clearance. The outer raceway semi-cone angle of the axle box bearings 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. Taking NSK's 0.561mm for calculation, according to the relationship between the radial clearance and axial clearance of tapered roller bearings, we can calculate 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 radial clearance of each row is 0.1mm, and the maximum axial clearance of NTN is 0.62mm, which is 0.5 higher than the maximum 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 subjected to super load for a long time and finally cannot bear it. It is natural and inevitable for the outer raceway 501 to peel off at this position. This is the most typical design problem. The detailed principle is as follows: The size of the clearance during the operation of the bearing has a huge impact on fatigue life, temperature rise, noise, vibration, etc. Figure 5This is a curve of bearing life ratio corresponding to different radial clearances. When the radial clearance of the bearing is less than -0.002mm, due to the negative clearance between the roller and the inner and outer raceways, or the interference fit state, a huge friction torque is generated between the roller and the inner and outer raceways and the large flange, which then causes wear, resulting in a sharp shortening of the fatigue life of the hub bearing. When the radial clearance changes from -0.002mm to 0, the life of the bearing increases significantly. When the bearing clearance is close to 0, the fatigue life of the bearing is the longest; as the bearing clearance gradually increases from 0, the rolling element is unevenly stressed, and the load on the inner and outer raceways is also unevenly distributed, and the fatigue life of the bearing becomes smaller again. Therefore, it can be inferred that only when the clearance of the hub bearing is adjusted to 0 can the fatigue life be guaranteed to be the longest, and the influence of temperature rise, noise, vibration, etc. can be guaranteed to be the lowest. However, the currently imported axle box bearings, that is, double-row tapered roller bearings, cannot adjust the clearance to 0 in principle, process, or assembly. The axle box bearing of the present invention 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 retains several parts that affect the clearance of the axle box bearing of this patent, and deletes the 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, and 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 to the big 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 clearance between the large end face of the roller 7 in the axial direction and the end face of the step portion 302 of the outer rib 3 is 0, and 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 best working state of 0 clearance.

[0077] The second 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 borne on the circumference of the inner raceway 101, but 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 present invention 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. The third and fifth fault proportions are temperature rise and wear, accounting for 10% and 8% respectively. It is very inaccurate to define the fault type by "temperature rise", because the results of a temperature rise of 1° and a temperature rise of 1000° are very different. The accurate description of the fault caused by temperature rise should be called ablation or high-temperature scorching. The most obvious manifestation of ablation is that the rollers and inner and outer rings turn blue due to high temperature. This is because high temperature will reduce the viscosity of the grease, accelerate the evaporation and oxidation of the grease, and the "pumping effect" (see Fig.17 ), an oil-poor area is generated 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 the rollers to heat up rapidly and turn blue; as for wear, it is because the axle box bearings are not replaced in time after turning blue, and the grease is burnt. At this time, the bearings are still running, and they must have entered the wear stage. The grease of imported axle box bearings is screened through a strict procedure and can meet the requirements of high-speed rail axle box 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? Because the imported double-row tapered roller bearings, this type of bearing structure is very mature, stable and reliable, and this type of double-row tapered roller bearing is very common and has no technical content at all, so we can only temporarily look for reasons other than design.

[0080] As we all know, high-speed rail lines have large spans, and EMUs run continuously over long routes, for long periods of time, at high speeds, and at the same time face harsh operating environments such as high and low temperatures, environmental corrosion, and wind and sand damage. In particular, there is a large temperature difference between the north and the south, and the viscosity of the grease changes dramatically with temperature changes. This makes the service environment temperature of the EMU bearing system complex and changeable, which increases the probability of failure of the EMU axle box bearing, bringing challenges to the operation and manufacturing of the EMU bearings.

[0081] Since the effects of product design and ambient temperature on the ablation and wear of axle box bearings are excluded, the only cause of ablation and wear of axle box bearings can only be found in the technical requirements. According to the process document requirements with the file number "sfet-77-48300" and the file name "EMU Maintenance Axle Box Bearing Pressing Operation Instructions", NSK and NTN companies stipulate that the minimum value of axial clearance of the axle box bearings of the Harmony 2C Phase II / 380A (L) EMU bogie is 0.15mm (see Table 1). Now quantitatively calculate the reduction in axial clearance of the axle box bearing 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 NTN and NSK). The length and big end diameter of the 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 length direction is almost the same, so the length expansion caused by the temperature increase 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 length of the roller. 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 rises by 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 width of the inner ring is:

[0085] ΔL10 =70*11.7×10 -6 *20=0.016(mm),

[0086] The diameter of the roller 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 temperature. The specific value of the expansion is:

[0089] ΔL2=170*11.7×10 -6 *20=0.04(mm)

[0090] The increase in roller and inner raceway diameters directly affects the decrease in axial clearance δ. Their calculation process is as follows:

[0091] 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, it can be seen 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:

[0092] ΔL 11 =(ΔL1*cosф) / sinb=(0.006*cos1°) / sin8°=0.043mm

[0093] 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:

[0094] ΔL 22 =ΔL2 / (2*tanb)=0.04 / (2*tan8°)=0.142mm

[0095] Since the imported axle box bearing is a double-row tapered roller bearing, the temperature of the second row rollers and the inner ring is 20°C higher than that of the outer ring, causing the axial clearance of the small end face and the small rib of the roller to decrease by:

[0096] δ=2*(ΔL 10 +ΔL11+ΔL22)=2*(0.016+0.043+0.142)=0.4(mm)

[0097] It can be concluded 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 axle box bearing will decrease by 0.4mm, which is 2.7 times the minimum value of 0.15mm specified by NSK and NTN for the axial clearance of the axle box bearing of the Harmony 2C Phase II / 380A (L) EMU bogie. This shows 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 axle box 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, and the expansion of each row of inner rings and rollers due to being 20℃ higher than the outer ring has been calculated before, that is, ΔL1+ΔL2=0.006+0.04=0.046(mm), and 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 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 each row of bearings to work normally, that is -0.002mm (see Figure 5 ), indicating that the friction torque between the roller and the large rib and the inner and outer raceways is huge, and naturally the failures of ablation, bluing and wear occur. So far, the inference of the reasonable structure of the double-row tapered roller bearing described in the beginning of this chapter has been refuted, because this double-row tapered roller design will still produce additional invalid derived axial forces (see Table 6 for specific values) even if most of the sections on which the EMU runs are straight rails. The friction torque generated by this additional invalid derived axial force is 100% converted into heat, causing the temperature of the roller ball base surface and the large rib, as well as the roller's conical surface and the inner raceway to rise the fastest. Further heat is then transferred to the entire roller and inner ring, causing the roller and inner ring temperature to be higher than the outer ring. The following Chapter 4 contains detailed theoretical and practical data to support this conclusion. 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 moving parts and the stationary parts of the bearing will 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 ablation (temperature rise) and wear caused by the small axial clearance mentioned above, the working conditions of grease can also cause ablation (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 not used properly, it will cause ablation (temperature rise) and wear. Due to the high temperature, the viscosity of the grease will decrease, the evaporation rate will increase, and the oxidation rate will increase. 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, the inner raceways and the retainer will be thrown onto the outer raceway. Since the generatrix of the tapered roller is inclined, the oblique component 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 Fig.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. Long-term operation will cause the small heads of the rollers to enter a dry friction state first, and the dry friction area will slowly spread to the large heads of the rollers, eventually leading to ablation (temperature rise) and wear.

[0099] Since the contact angle of the axle box of the present invention is below 4° (see Fig.18 ), whose structure is similar to that of a cylindrical roller bearing, and has no obvious bell-mouth structure in the axial direction, and almost no "pumping effect", and has good grease retention ability. The grease no longer moves to the big end of the roller 7, and 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, the 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 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, and 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 ball base surface 702 of the roller 7 of the axle box bearing 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 Fig.11), resulting in the additional ineffective derived axial force generated between the ball 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 that 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 axle box bearing, and use this expansion as the reference value of the minimum clearance δ of the axle box bearing. According to the relevant international standards for high-speed rail, the temperature rise of the axle box bearing is ≤80℃. Taking an axle box bearing designed by the inventor 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℃ to the outer ring 5 (positive Under normal circumstances, the actual temperature difference between the outer ring 5 and the roller 7 and the inner ring 1 is less than 80°C, and the axial clearance value is greater than the expansion amount when the temperature difference between the outer ring 5 and the roller 7 and the 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 head 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 caused by the increase in temperature in the contact area between the roller 7 and the inner ring 1 in the length direction is the same, resulting in the simultaneous elongation of the contact portion 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, so it is only necessary to calculate the change in the axial clearance corresponding to the change in the diameter of the roller 7 and the inner ring 1.

[0102] The diameter of roller 7 becomes larger because the temperature is 80°C higher than that of outer ring 5. The specific value is:

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

[0104] The diameter of the inner raceway 101 becomes larger 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 formulas are as follows:

[0106] The decrease in the axial clearance δ caused by the increase in the diameter of the roller 7 has a certain relationship with the angle between the outer raceway 501 and the roller 7. The decrease in the axial clearance δ caused by the increase in the diameter of the roller 7 is:

[0107] ΔL 11 =(ΔL1*cosф) / sina=(0.026*cos0.4°) / sin3.2°=0.47(mm)

[0108] The decrease in the 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 the axial clearance δ caused by the increase in the diameter of the inner raceway 101 is:

[0109] ΔL 22 =ΔL2 / (2*tanb)=0.157 / (2*tan2.4°)=1.87(mm)

[0110] Since the axle box bearing of the present 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,

[0111] δ=ΔL11+ΔL22=0.47+1.87=2.34(mm), take one decimal place and you get δ=2.3mm

[0112] That is to say, when the minimum axial clearance of this patented bearing product at room temperature is 2.3 mm (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, and the axial clearance δ becomes 0 at this time, and the bearing can still work normally; according to the technological status of each part affecting the axial clearance of the patented bearing, the manufacturing tolerances of each part affecting 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 change range of the axial clearance δ value of this patented bearing is 2.3-3mm.

[0113] It can be seen from the above Chapters 1 and 2 that, 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 the axial clearance of the same batch of bearings being too large or too small. In addition to affecting the uneven force on the inner and outer rings of the bearing, causing stress concentration, and finally leading to the peeling of the inner raceway 101 and the outer raceway 501, the excessive clearance will also affect the running axis of the inner ring 1 and the roller 7, increase noise, increase wear and cause 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 jamming. The standard JB / T8236-96 for radial clearance of double-row tapered roller bearings revised in 1996 (see Table 3) found that even if measured according to this standard, the clearance value of imported double-row tapered roller axle box bearings is 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 angle of the axle box bearings of these two companies is 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 brought 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 holes of the axle box bearings of these two companies are The size is 130mm. Compared with the corresponding nominal diameter column of 120-140 in the table above, 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 clearance value range has caused the bearings supplied by these two companies to have batches of inner and outer raceway peeling and ablation failures. By comparing the radial clearance values ​​of each group in this size segment of the nominal diameter column of 120-140 above, it can be seen that the range 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 between the maximum and minimum values ​​of the radial clearance of Group 4 and Group 5 is 50μm. The range between the maximum and minimum values ​​of the radial clearance of NTN and NSK bearings is 219-53=166μm and 198-53=145μm respectively. The clearance value is seriously exceeded, resulting in systematic and batch bearing failures. Domestic railway freight car wheelset bearing manufacturers have introduced and produced 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 according to the 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 surface 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.

[0114] Table 3 Radial clearance of double-row and four-row tapered roller bearings μm

[0115]

[0116] Approved by the Ministry of Machinery Industry on September 3, 1996 and implemented on July 1, 1997

[0117] 3. The third most common fault is retainer breakage, which is also 10%. Since the inventor of this patent could not obtain the drawings of the plastic retainers of imported axle box bearings, he could 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 standard parameter tolerance values ​​of some tapered roller bearing retainers (technical standard ZYB9-93). For example, the design requires a bilateral clearance ε between the retainer and the roller. △Only 0.25mm (see Table 5). During assembly, the above clearance requirements can be achieved by controlling and adjusting the shrinkage of the iron retainer. Since the plastic retainer cannot expand and shrink in design and process, 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 of the plastic retainer used in the freight train wheel set bearing 352226X2 with the same structure and inner and outer diameter dimensions as the EMU axle box bearing (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 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 center of the bearing. The center line of the bearing rotates, and the center of gravity of the retainer deviates from the center line of the bearing, resulting in radial runout and collision with the roller. Since the radial clearance of the bearing itself 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 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 check 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 component together, it is difficult to clean them, and it is not convenient to check the working surface of the roller, which poses a hidden danger to the subsequent loading and operation of the bearing. Moreover, this detachable column or beam is definitely different from the other integrated columns or beams in function, which also affects the normal operation of the bearing.

[0118] Table 4

[0119]

[0120] Table 5

[0121]

[0122] This maintenance-free axle box bearing retainer 4 Fig.19 As shown, in design, the entire retainer 4 is an integral structure, and no detachable column or beam is designed. Since the axle box bearing of the patent achieves "0" clearance in structural design (see Figure 2 ), when all rollers 7 rotate along the inner raceway 101 and the outer raceway 501, the gap between the conical surface and the inner and outer raceways is always 0, and the gap between the spherical base surface 702 and the step end surface 302 of the outer rib 3 is always 0. In this way, when the rollers 7 rotate at high speed, the centrifugal force is borne by the outer raceway 501 of the outer ring 5, and the axial force is borne by the step end surface 302 of the outer rib 3, thereby avoiding the impact of the rollers on the cage due to the centrifugal force of the existing bearings because the radial clearance is 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 the appropriate range, so that the play between the retainer and the roller meets the industry standard value; another factor that affects the clearance 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 block 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 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 clearance 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 clearance between the retainer and the roller from the design (see Figure 1 ), the parts diagram of the small rib 6 is shown in Fig.13 , whose inner hole is threaded (see Figure 1 ) or screw riveting (see Fig.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 Fig.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 maintenance", only the small rib 6 (see Fig.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 relatively clean and the inspection is also 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 the roller 7 and retainer 4, and screw the small rib 6 clockwise onto the small head external thread 102 of the inner ring 1, and lock several radial set screws of the small rib 6. 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 loading to the designed life, and then directly scrapped.

[0123] The remaining faults with a relatively small failure rate, such as abnormal noise, inner ring breakage, roller breakage, etc., are most likely related to abnormal operating conditions of the bearing. Since they account for a small proportion and are not representative, they cannot be analyzed.

[0124] 4. In addition to the above design defects exposed by practice, the existing imported axle box bearings also have problems of 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, which in turn affects the life of the bearing. High energy consumption pushes up the operating cost of the 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 ball 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 journey 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:

[0125] (1) The existing axle box bearing force analysis (see Fig. 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 to the two inner rings evenly 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 of one inner and outer raceway and roller, and then multiply it by 2 to get the force of the whole set of bearings. 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. It 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 balanced state under the action of the three forces of the inner and outer raceways and the large rib, so:

[0126] Qn-Qi-Qy=0 (1)

[0127] Qa-Qe+Qx=0 (2)

[0128] Among them, Qn=Qm*sina, Qa=Qm*cosa, Qi=Qb*sinb, Qe=Qb*cosb, Qx=Qz*sinc, Qy=Qz*cosc,

[0129] Substituting the angle values ​​of a, b and c and these six equations into equations (1) and (2) and simplifying them, we obtain:

[0130] Qb=Qm*(sina*sinc+cosa*cosc) / (sinb*sinc+cosc*cosb)≈Qm

[0131] Qz=Qm*(sina*cosb-cosa*sinb) / (sinb*sinc+cosc*cosb)=0.035Qm

[0132] The radial force on the inner ring is:

[0133] Qe+Qx=Qb*cosb+Qz*sinc=Qm*0.99+0.035Qm*0.148=Qm=Qa / cosa=1Qa

[0134] The axial force borne by the inner ring raceway is:

[0135] Qi=Qb*sinb=Qm*sinb=0.139Qm=0.139Qa / cosa=0.141Qa

[0136] The axial force on the outer ring raceway is Qn = Qa*taga = 0.176Qa

[0137] The axial force on the large flange is Qy = Qz*cosc = 0.035Qm*cosc = 0.035Qa

[0138] The axial force on the roller is Qn+Qi+Qy=Qa*taga+Qb*sinb+Qz*cosc

[0139] =Qa(taga+sinb / cosa+0.035cosc / cosa)=0.352Qa

[0140] (2) The relevant parameters of an axle box bearing designed in this embodiment are as follows Figure 6 , see Fig.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 the outer rib respectively. The roller is in a state of equilibrium under the action of the three forces of the inner and outer raceways and the outer rib, so:

[0141] Qn-Qi-Qy=0 (1)

[0142] Qa-Qe+Qx=0 (2)

[0143] Among them, Qn=Qm*sina, Qa=Qm*cosa, Qi=Qb*sinb, Qe=Qb*cosb, Qx=Qz*sinc, Qy=Qz*cosc,

[0144] Substituting the angle values ​​of a, b and c and these six equations into equations (1) and (2) and simplifying them, we obtain:

[0145] Q b =Qm*(sina*sinc+cosa*cosc) / (sinb*sinc+cosc*cosb)=Qm

[0146] Q z =Qm(sina*cosb-cosa*sinb) / (sinb*sinc+cosc*cosb)=0.014Qm

[0147] The radial force on the inner ring is:

[0148] Q e =Qa+Qx=Qa+Qz*sinc=Qa+0.014Qm*sinc=Qa+0.00077Qm

[0149] =Qa(1+0.00077 / cosa)=Qa

[0150] The axial force on the inner ring is:

[0151] Qn-Qy=Qm*sina-Qz*cosc=Qm*(sina-0.014cosc)

[0152] =0.042Qm=0.042Qa / cosa=0.042Qa

[0153] The axial force on the outer ring is Qn = Qa*taga = 0.056Qa

[0154] The axial force on the outer flange is Qy = Qz*cosc = 0.014Qm*cosc = 0.014Qa

[0155] The axial force on the roller is Qn+Qi+Qy=Qa*taga+Qb*sinb+Qz*cosc

[0156] =Qa(taga+sinb / cosa+0.014cosc / cosa)=0.11Qa

[0157] (3) The following conclusions can be drawn from the calculations in (1) and (2):

[0158] A. When the outer rings of the two types of axle box bearings are subjected to pure radial force, the outer rings transfer the same magnitude of radial force to the inner rings through the rollers, that is, Qe = Qa.

[0159] B. From the additional invalid derived axial forces borne by each part in Table 6, it can be seen that through innovative design, the various parts 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.

[0160] C. Under the same radial load, the derived axial force generated by the outer rib of the axle box bearing of the present invention is 0.41 times that of the large rib of the existing bearing. In general, 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 larger 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 the overheating of the bearing is attributed to unqualified grease; the substantial reduction in derived axial force improves the utilization rate of energy and reduces the operating cost of the high-speed rail. The above theoretical analysis and calculation results have been verified in practice. According to relevant literature and data, 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 large inner ring rib.

[0161] Table 6

[0162]

[0163] As is known to all, the two inner rings of the currently imported axle box bearings, i.e., double-row tapered roller bearings, in addition to bearing radial forces, also transmit axial forces in two directions respectively; the axle box bearing of the present invention, although its internal structure is a set of single-row tapered roller bearings, 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, its 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 ).

[0164] This embodiment innovates and optimizes the currently imported double-row tapered roller bearing into a single-row tapered roller bearing, and optimizes the two-row roller into a single-row roller. 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 100 mm, the increase in the radial basic rated dynamic load and fatigue life is achieved 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 the lack of raw materials and processing accuracy. The inventor is a certain high Take an axle box bearing designed by Tie as an example. Its inner and outer diameters are 130 and 240 respectively, which is 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, and the above parameters are brought into the calculation formula of the radial basic rated dynamic load Cr of the tapered roller bearing:

[0165] C r =b m *f c *(i*L we *cosa) 0.778 *Z 0.75 *D we 1.074 =1.1*87.7*(119*cos3.2°) 0.778 *20 0.75 *27.1 1.074

[0166] =1298(KN)

[0167] Under the same condition as the imported axle box bearings with the same inner and outer diameters, the number of parts of the axle box bearings 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 bearings also adopt the design of this patent, the existing double-row tapered roller bearings are changed into single-row tapered roller bearings, and the above model is used as an example, the dimensions of the various parts of the imported bearings are the same as those in the example above of this patent, except that the contact angle is different (see Table 2), that is, a of the imported bearing is 10°, the semi-cone angle of the roller is ф=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:

[0168] C r =b m *f c *(i*L we *cosa) 0.778 *Z 0.75 *D we 1.074 =1.1*85.4*(119*cos10°) 0.778 *20 0.75 *25.86 1.074

[0169] =1190(KN),

[0170] The radial basic dynamic load rating C of a certain model of this patent is r =1298KN, the imported bearing only has the same contact angle, or the existing a = 10°, and its radial basic rated dynamic load becomes C r =1190KN, a=3.2° of the patented bearing, 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 is 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.

[0171] 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 change from "planned maintenance" to "preventive maintenance" mode, and then to "maintenance-free mode". From loading to scrapping and replacement, bearings do not require any form of maintenance in between, truly crushing imported products in all aspects, throughout the entire process, and throughout the entire life cycle.

Claims

1. A maintenance-free rail transit axle box bearing, characterized in that: It comprises 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) together enclose a friction working portion; the retainer (4) is arranged in the friction working portion; a tapered roller (7) is arranged in a 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 ball 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 seal 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. A 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 claims 1 to 5, characterized in that: Maintenance-free axle box 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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