Sliding bearing and method for controlling the variation of the radial clearance due to thermal expansion

Abstract

The present application relates to a kind of sliding bearing and the control method of its radial clearance variation due to thermal expansion.The sliding bearing includes bearing bush;The bearing bush includes more than three first shell;The first shell is uniformly distributed along the inner circumferential wall of bearing seat or bearing outer ring;The first shell and its corresponding bearing seat or bearing outer ring are matched or fixed with plane.In the use of sliding bearing at high temperature and low temperature, the material can be selected according to the linear expansion coefficient and the corresponding structure and size are adjusted, the purpose of effectively controlling the radial clearance variation without gap compensation structure is realized, and the technical problems in the field are solved.The present application has the outstanding characteristics of simple and reasonable structure, cost saving and safe use, especially meets the performance requirements of sliding bearing in high temperature or low temperature and large temperature change occasions.

Description

Technical Field

[0001] This invention relates to sliding bearings, and also to a method for controlling radial clearance changes caused by thermal expansion. Thermal expansion generally refers to the increase in volume of most substances as temperature rises and decreases as temperature falls, under constant external pressure. The radial clearance change of sliding bearings due to thermal expansion described in this invention includes the change in radial clearance of the sliding bearing during the process of increasing or decreasing service temperature. Background Technology

[0002] Sliding bearings have a wide range of applications. Existing sliding bearings are oil-lubricated and cooled by the lubricating oil. For ease of installation, some bearings are structured by dividing the bearing shell into several shells. When there are three or more shells evenly distributed circumferentially, the outer diameter of each shell is equal to the inner diameter of the bearing housing. Each shell has radially drilled holes for fixing it to the bearing housing, and bolts are passed through these holes to secure the shell to the bearing housing. This leads to several problems, including:

[0003] (1) In order to form an oil film, the sliding clearance between the shaft and the bearing bush is very small. If the shaft temperature is high and the bearing housing temperature is low, the shaft will seize up after thermal expansion. This means that the shaft and bearing bush cannot be too hot or the parts cannot expand too much. Therefore, the method of cooling with lubricating oil is adopted and the temperature of the lubricating oil is strictly controlled. If this expansion clearance is reserved in advance, vibration or oil film oscillation will easily occur in the early stage of operation, which is not conducive to lubrication and can easily cause damage to the shaft and bearing bush.

[0004] (2) In many cases, for the purpose of high and / or low temperature resistance, wear resistance, or corrosion resistance, the bearing bush is made of ceramic material with a small coefficient of expansion, while the bearing housing is made of metal material with a large coefficient of expansion. When both are heated and expand simultaneously, a large difference in expansion will occur, such as... Figure 13 As shown, after expansion, the inner diameter of the bearing housing is larger than the outer diameter of the split shell of the bearing bush. This means that, on the one hand, due to vibration, manufacturing, and assembly precision issues, when the radial force of the shaft acts off-center from the circumferential center of the split shell, the bearing bush is easily damaged. On the other hand, when the inner diameter of the bearing housing is larger than the outer diameter of the split shell of the bearing bush, the bearing bush may experience rotational displacement or possible self-rotational wobbling. Similarly, when both are simultaneously subjected to cooling and contraction, a large difference in contraction will occur, such as... Figure 14 As shown, after shrinkage, the inner diameter of the bearing housing is smaller than the outer diameter of the split shell of the bearing bush. This causes the outer walls of the split shell at both circumferential ends to contact the inner wall of the bearing housing, leaving a gap in the middle. This also makes the bearing bush more susceptible to crushing. To avoid this gap, the outer diameter of the split shell of the bearing bush after shrinkage should not be greater than the inner diameter of the bearing housing. A common solution is to make the initial inner diameter of the bearing housing larger. However, in the un-shrunk state, the radial force acting on the split shell, which is off-center from the circumferential center, can easily crush the bearing bush.

[0005] (3) After thermal expansion or contraction, solid particles in the medium are likely to accumulate in the gap between the bearing shell and the bearing seat. The ceramic shell is hard and has almost no toughness. When it returns to room temperature from a hot or cold state, the shell is prone to crushing.

[0006] Furthermore, the sliding bearings widely used in magnetic pumps, because their internal lubrication is oil-free, must be designed for wear resistance, high temperature resistance, and corrosion resistance. The vast majority of sliding bearings are made of ceramic materials, such as silicon carbide. Figure 12 As shown, the bearing bushes of the two sliding bearings are usually designed in a cylindrical shape, which brings a serious problem: the shaft is usually made of metal, the bearing bushes are made of ceramic materials such as silicon carbide, and the bearing housing is also made of metal. For magnetic pumps, the shaft, bearing bushes, and bearing housing are all located in the medium. When the medium is at high or low temperature, due to the difference in thermal expansion coefficients, an expansion gap must be reserved during manufacturing to accommodate the difference in thermal expansion or contraction; otherwise, the cylindrical bearing bushes will break during thermal expansion and crush during contraction. For example, a shaft with a diameter of Φ50 is made of 304 steel (at 300℃, its linear expansion coefficient α1=17.2×10). -6 / ℃); Both bearing shells are made of silicon carbide (at 300℃, its linear expansion coefficient α2=(4~5)×10 -6 / ℃, here we take α2=4×10 -6 / ℃), the thickness of a single-layer bearing bush is 20 (the units of measurement not specified in this invention are all in millimeters), the initial inner diameter Φ1 of the bearing bush can be taken as 50, then the outer diameter of the outer layer bearing bush Φ2≈90; the bearing seat is also made of 304 steel, and its inner diameter Φ3≈Φ2≈90. When the medium temperature rises from 0℃ to 300℃, calculations show that the thermally expanded shaft diameter Φ′ = 50.258 mm, and the inner diameter of the bearing bush Φ1′ = 50.06 mm. The expansion difference between the two is Φ′ - Φ1′ = 50.258 - 50.06 = 0.198 mm. This means that to prevent the cylindrical bearing bush from breaking due to thermal expansion, an additional radial clearance of approximately 0.2 mm must be reserved between the shaft and the bearing bush during manufacturing to overcome this expansion difference. Similarly, calculations show that the thermal expansion difference between the bearing housing and the bearing bush reaches 0.3429 mm. To avoid damage to the bearing bush due to thermal expansion, an additional expansion clearance is also required. However, a clearance of 0.3429 mm usually far exceeds the clearance range required for sliding fits, resulting in significant vibration and noise, easily damaging the bearing, and affecting its reliability or rendering it unusable. Furthermore, as the temperature range increases, the clearance also increases, and the same problem exists in low-temperature applications.

[0007] To address the issues arising from increased clearance due to thermal expansion and contraction, and to accommodate these gaps, Chinese patent CN201818536U discloses a gap compensation structure, which involves installing an interference fit gap compensation ring between the bearing bush and the bearing housing. Similarly, Chinese patent CN203879788U discloses a bearing housing sleeve structure that also utilizes a T-shaped sleeve element to compensate for the gap between the bearing bush and the bearing housing.

[0008] The above measures typically compensate for expansion gaps by installing various gap compensation structures. However, this leads to the following problems:

[0009] (1) Due to the limited wall thickness, expansion sleeves are often made of thin plates. This limits the load-bearing capacity of a single expansion sleeve. The solution is usually to use multiple expansion sleeves for a single sliding bearing, which leads to an increase in the axial length of the bearing and the corresponding number of parts, thus reducing the reliability of the bearing and / or equipment. At the same time, it inevitably increases manufacturing and maintenance costs. Furthermore, the large gap between the bearing housing and the bearing bush can only be compensated by the expansion sleeve alone. On the other hand, adding an elastic expansion sleeve is equivalent to reducing the stiffness of the shaft, thereby reducing the critical speed of the shaft. This results in impacts and vibrations that are particularly detrimental to the bearing bush, thus reducing the safety and lifespan of the bearing.

[0010] (2) The expansion sleeve is an elastic element. At high temperatures, the elastic force is greatly reduced or lost and yielding occurs, which increases the risk of use and may even lead to unusability. Therefore, cooling measures must be taken, which brings a series of problems such as medium crystallization, scaling, heat preservation of cooling devices, antifreeze, and scaling. On the other hand, when used in low-temperature media, the bearing and / or expansion sleeve may be damaged due to the increased elastic modulus.

[0011] (3) Expansion sleeves corrode in highly corrosive media, causing the wall thickness of the expansion sleeve to gradually decrease. This leads to a gradual reduction in the elasticity and load-bearing capacity of the expansion sleeve, ultimately affecting the service life of the equipment. For example, in some cases, if the corrosion rate of the metal is 0.1 mm / year, the corrosion rate of the expansion sleeve can reach 0.2 mm / year due to simultaneous corrosion on both the inner and outer walls. In such cases, the expansion sleeve needs to be replaced regularly, increasing the corresponding maintenance costs.

[0012] (4) Regarding the cylindrical bearings used in existing magnetic pumps, such as Figure 12 As shown, after adding the expansion sleeve, the diameter of the isolation sleeve must be increased. When the isolation sleeve is made of metal, the eddy current loss during operation will increase.

[0013] (5) Since the bearing bush and the expansion sleeve are interference fit, the risks and difficulties in the manufacturing, maintenance and disassembly process are increased, thereby increasing the corresponding costs.

[0014] The aforementioned problems are prevalent in existing technologies. For example, in the metallurgical industry, many idlers on rolling mill equipment and conveyors transporting rolled high-temperature steel plates have extremely high temperatures. These temperatures are transferred to the bearings on the rolling mill equipment and idlers. The temperature of the idlers or shafts is high, while the temperature of the bearing housings is low, creating a temperature gradient. When heated, the different parts expand unevenly, which can cause the bearings to seize. To prevent bearing seizure, a large gap is initially provided. However, for a period of time after the equipment starts operating, the gap remains large because the maximum temperature has not yet been fully reached. This causes vibration in the equipment, which is detrimental to the bearings, generates noise, affects human health, and also affects the flatness of the steel plates.

[0015] For example, media such as lead-bismuth alloys or molten salts have liquid temperatures of 500–750°C. Pumps used in such high-temperature applications would require expansion sleeves or clearance compensation rings if the bearings were cylindrical, made of ceramic materials similar to those used in magnetic pumps. However, at such high temperatures, these expansion sleeves or clearance compensation rings would lose their elasticity and become unusable. This necessitates the use of metal bearings. Currently, ensuring the service life of metal bearings without cooling is a difficult task.

[0016] For example, many manufacturers of submersible molten salt pumps, including those manufactured by the globally renowned Flowsche Corporation in the United States, employ a method of cooling the bearings internally with the medium and then using the cooled medium to cool the bearings to prevent them from overheating. However, adding cooling measures introduces a series of problems and complications, such as increasing the number of components, and carries the risk of pump damage if the cooling system malfunctions. Furthermore, in many situations, the cooling of the medium is undesirable.

[0017] To address a series of challenges, such as the large differences in the coefficients of thermal expansion between the shaft, bearing shell, and bearing housing or outer ring, leading to bearing shell breakage or crushing, and the difficulty in increasing the operating temperature of the bearing, many research institutions, universities, and professionals are researching and developing tough ceramic materials. Some progress has been made, but there is still a long way to go and problems remain. For example, in the submersible molten salt pump of a concentrated solar power generation system, it is desirable for the bearing to have the highest possible temperature resistance, such as reaching 1000℃ and a service life of one year. Currently, there is clearly no bearing that can meet this requirement. Summary of the Invention

[0018] In view of the problems existing in the prior art and the actual needs in the field, the present invention specifically discloses a sliding bearing and a method for controlling the change of radial clearance due to thermal expansion, in order to achieve a constant radial clearance or a controllable change in the sliding bearing or sliding bearing assembly under high temperature or low temperature environment or service conditions with large temperature difference, without using various expansion clearance compensation structures.

[0019] In this invention, the sliding bearing assembly is based on the concept of a rolling bearing assembly, including a bearing inner ring, a bearing outer ring, and a bearing bush or a sliding body structure similar to a bearing bush located between the bearing inner ring and the bearing outer ring.

[0020] The clearance of a sliding bearing is typically 0.01 to 0.02 millimeters. The range of clearance values ​​for a sliding bearing is determined by the precision of the component it is used in. For example, the precision of the main bearing fitting between the indexing head of a grinding machine and a lathe is H7 / g6, while the precision of the bearing fitting for agricultural machinery is H11 / d11.

[0021] There are two types of clearance in sliding bearings: radial clearance and axial clearance. A certain radial clearance is necessary for sliding bearings for the following reasons: First, radial clearance is the minimum requirement for a smooth connection between the shaft and the bearing. Second, radial clearance ensures the accuracy of shaft operation. Generally, a smaller radial clearance results in higher shaft accuracy, but it cannot be arbitrarily reduced. Excessive radial clearance not only makes it difficult to form a lubricating oil film but also generates excessive heat and may lead to bearing failure and seizure. Third, radial clearance is crucial for fluid lubrication, especially in hydrodynamic bearings, where the lubricating oil film is primarily formed by the wedge-shaped clearance between the shaft and the bearing. The size of the clearance is related to the machine's accuracy during operation; a smaller radial clearance results in higher accuracy. However, if the clearance is reduced to a certain extent, fluid lubrication cannot be guaranteed. Excessive clearance will cause vibration during rotation and will also prevent the formation of a stable oil film. Therefore, the radial clearance value cannot be arbitrarily specified and must comply with the requirements of the machine's technical documents. The radial clearance of sliding bearings is extremely important; both excessive and insufficient clearance are highly detrimental. Excessive clearance makes it difficult for a lubricating oil film to form in the bearing, failing to guarantee liquid lubrication. This can also reduce the machine's operating accuracy and even cause severe vibration and noise, potentially leading to accidents.

[0022] The technical solution of this invention is implemented as follows:

[0023] A sliding bearing includes a bearing shell: the bearing shell includes three or more first shell shells; the first shell shells are evenly distributed circumferentially on the inner peripheral wall of the bearing housing or the outer ring of the bearing;

[0024] The first bearing shell and its corresponding bearing seat or bearing outer ring are mated in a planar manner; that is, the outer ring surface of each first bearing shell and the inner peripheral wall of its corresponding bearing seat or bearing outer ring are mating planes, or the two are fixedly connected on this plane.

[0025] The three or more first shells constitute the first bearing shell, or form a single-layer bearing shell structure.

[0026] Furthermore, a first groove is excavated on the inner circumferential wall of the bearing housing or the outer ring of the bearing for each of the first bearing shells, and a first boss is formed between two adjacent first grooves; the bottom of the first groove is a plane that mates with the outer ring surface of the corresponding first bearing shell; the radial depth of the first groove is less than the thickness of the corresponding first bearing shell, and the first bearing shell is partially embedded in the first groove radially.

[0027] Specifically, the first groove can be wide at the bottom and narrow at the top, in the shape of a dovetail groove; correspondingly, the first tile shell is wide at the bottom and narrow at the top in the radial direction.

[0028] Furthermore, the radial structure or dimensions of the first boss can have the following characteristics: its circumferential dimension near the bottom of the first groove is smaller than its circumferential dimension far from the bottom of the first groove, so the cross-section of the first boss is narrow at the bottom and wide at the top (here, "bottom" and "top" correspond to the bottom and opening of the first groove, respectively). Specifically, for example, based on the above characteristics, the cross-section of the first boss can be "T-shaped," and the horizontal side of the "T" is farther from the bottom of the first groove than its vertical side. Considering that under conditions of large temperature variations, due to thermal expansion and contraction, a gap is formed between the first tile shell embedded in the first groove and the corresponding first groove, and solid impurities will enter the gap, thus depriving the first tile shell of thermal expansion space and making it easy for the first tile shell to be crushed. To prevent this from happening, after the first boss between two adjacent circumferentially adjacent tile shells is made into the "T-shaped" structure, a certain redundancy can be reserved in the first groove, which can both avoid crushing the corresponding first tile shell and ensure the stability of the first tile shell in the first groove.

[0029] In addition to using the above-mentioned embedded groove method to connect the first bearing shell to the bearing seat or bearing outer ring, fastener fixing, bonding and snap ring connection can also be used, or a combination of two or more of the above connection methods can be used.

[0030] Specifically, the inner annular surface of the first bearing shell forms a sliding fit with the outer peripheral wall of the journal or the inner ring of the bearing;

[0031] The inner ring of the shaft or bearing has a composite layer on the surface that mates with the first bearing shell. The composite layer can be made of ceramic materials (including but not limited to silicon carbide, boron carbide, tungsten carbide silicon nitride, boron nitride, alumina, zirconium oxide, titanium oxide, etc.) or hard alloy materials (including but not limited to tungsten nickel titanium alloy, tungsten cobalt titanium alloy, tungsten zirconium titanium alloy, etc.). It is formed by spraying, plating, fusion or other surface strengthening processes and has wear-resistant, high-temperature resistant or low-temperature resistant properties.

[0032] Furthermore, the sliding bearing may also include a torsion spring located between the first bearing housing and the shaft or bearing inner ring, tightly clamped to the shaft or bearing inner ring. The torsion spring may be made of ceramic or hard alloy material.

[0033] Alternatively, to further improve the service life of the shaft, the bearing bush may include three or more second bearing shells; the second bearing shells are evenly distributed circumferentially on the outer peripheral wall of the journal or the inner ring of the bearing and are mated with each other on a plane; that is, the inner ring surface of each second bearing shell and the outer peripheral wall of the corresponding journal or the inner ring of the bearing are mating planes, or the two are fixedly connected on this plane.

[0034] The outer ring surface of the second tile shell forms a sliding fit with the inner ring surface of the first tile shell.

[0035] Here, three or more second shells correspondingly constitute the second bearing, which, together with the aforementioned first bearing, form a double-layer bearing structure.

[0036] In order to ensure that the radial force on the bearing does not extend beyond the bearing, each plane on the two circumferential sides of the first and / or second bearing is located in the same plane as the axis of the shaft.

[0037] Furthermore, the mating surfaces of the first and second bearing shells collectively form a frustum-shaped side surface, with the upper base of the frustum located at the end of the shaft. The upper base of the frustum, relative to the lower base, refers to the base with the smaller radius. This frustum-shaped mating structure of the double-layer bearing shell is similar to that of a tapered roller bearing, improving the sliding bearing's ability to bear axial loads.

[0038] Furthermore, a second groove is excavated on the outer peripheral wall of the journal or bearing inner ring corresponding to each of the second bearing shells, and a second boss is formed between two adjacent second grooves; the bottom of the second groove is a plane that mates with the corresponding second bearing shell; the radial depth of the second groove is less than the thickness of its corresponding second bearing shell, and the second bearing shell is partially embedded in the second groove radially.

[0039] Specifically, the second groove is narrow at the bottom and wide at the opening, with a trapezoidal cross-section; correspondingly, the second tile shell is narrow at the bottom and wide at the top along the radial direction.

[0040] Similar to the T-shaped structure of the first boss described above, the radial structure or dimensions of the second boss can also have the following characteristics: its circumferential dimension near the bottom of the second groove is smaller than its circumferential dimension far from the bottom of the second groove, thus the cross-section of the second boss is narrow at the bottom and wide at the top (here, "bottom" and "top" correspond to the bottom and opening of the second groove, respectively). Similarly, for example, the cross-section of the second boss can be T-shaped, with the horizontal side of the "T" further away from the bottom of the second groove than its vertical side. The design intention of the T-shaped structure of the second boss can also refer to the T-shaped structure of the first boss described above.

[0041] Alternatively, grooves can be formed along the circumferential direction on the inner ring wall of the first bearing shell and the outer ring wall of the second bearing shell, with corresponding inner and outer retaining springs installed in each groove. The elastic force generated by the retaining spring structure adjusts the radial force on the bearing shell, ensuring that the second bearing shell is always in close contact with the shaft or bearing inner ring and the first bearing shell is always in close contact with the bearing seat or bearing outer ring. This prevents impurities in the medium from entering between the contact surfaces, thus avoiding affecting the fit clearance or causing equipment damage.

[0042] Similarly, in addition to using the above-mentioned embedded groove method to connect the second shell to the shaft or bearing inner ring, the above connection can also be achieved by fastener fixing, bonding and snap ring connection, or by combining two or more of the above connection methods.

[0043] Furthermore, the number of the first and second tile shells is different; more preferably, the number of the first and second tile shells differs by one; for example, the former has one more tile shell than the latter. This avoids collisions caused by the gaps between the first and second tile shells coinciding during rotation.

[0044] More preferably, to avoid collisions between the first and second bearing shells during relative rotation, their sliding mating surfaces, near their respective circumferential end faces, are slightly warped or gradually separated radially away from each other, forming radially thinning regions. The radially thinning regions on the mating surfaces of the two bearing shells, viewed in cross-section, form a straight line, an involute, or an Archimedean spiral.

[0045] To prevent hard solid particles from entering the bearing, end caps can be provided at both ends of the bearing housing or the outer ring of the bearing, and the end caps are provided with shaft holes, which form a clearance fit with the shaft.

[0046] Specifically, in order to cool the bearing bush and / or the isolation sleeve of a magnetic pump that uses the sliding bearing, the end caps at both ends are respectively provided with inlets and outlets for cooling and / or lubricating media, and are respectively connected to the flow gaps or passages of cooling and / or lubricating media inside the sliding bearing and related structures.

[0047] The flow gap or passage of the cooling and / or lubricating medium may include a through groove provided axially between the mating surfaces of a single layer of bearing and the shaft or between the mating surfaces of two layers of bearing, and / or a circumferential gap or groove between adjacent shells of each layer of bearing.

[0048] This invention also protects sliding bearing assemblies that include the aforementioned sliding bearings.

[0049] This invention also discloses a method for controlling the change in radial clearance of a sliding bearing due to thermal expansion. The method uses the sliding bearing disclosed in this invention and is divided into two cases according to different working conditions.

[0050] The first scenario refers to the condition where the service temperatures of the shaft or bearing inner ring, bearing bush, bearing housing, or bearing outer ring are the same or close:

[0051] (1) The material selection conforms to α1>α3>α2;

[0052] (2) The structural dimensions satisfy the following: t is positively correlated with D, positively correlated with the difference between α1 and α3, and negatively correlated with the difference between α3 and α2, and D... i =D + 2t + δ, and D ≥ D0;

[0053] The second scenario occurs when the service temperatures of the shaft or bearing inner ring, bearing bush, bearing housing, or bearing outer ring decrease sequentially with significant temperature differences:

[0054] (1) Its material selection meets the criteria of α3≥α1>α2;

[0055] (2) The structural dimensions satisfy the following: t is positively correlated with D, positively correlated with the difference between α1ΔT1 and α3ΔT3, and negatively correlated with the difference between α3ΔT3 and α2ΔT2, and D... i =D + 2t + δ, and D ≥ D0;

[0056] in,

[0057] α1 is the coefficient of linear expansion of the inner ring of the shaft or bearing.

[0058] α2 is the coefficient of linear expansion of the bearing bush.

[0059] α3 is the coefficient of linear expansion of the bearing housing or the outer ring of the bearing.

[0060] t is the thickness of the bearing bush.

[0061] D is the shaft diameter or the outer diameter of the bearing inner ring.

[0062] D i This refers to the inner diameter of the bearing housing or the outer ring of the bearing.

[0063] D0 is the theoretical value of the shaft diameter or the outer diameter of the bearing inner ring that meets the product design requirements.

[0064] δ is the radial clearance allowance reserved during the manufacturing of the sliding bearing.

[0065] ΔT1 is the difference between the service temperature of the shaft or bearing inner ring and the room temperature.

[0066] ΔT2 is the difference between the service temperature of the bearing and the room temperature.

[0067] ΔT3 is the difference between the service temperature of the bearing housing or outer ring and the room temperature.

[0068] α1ΔT1, α2ΔT2, and α3ΔT3 represent the products of α1 and ΔT1, α2 and ΔT2, and α3 and ΔT3, respectively.

[0069] According to the formula for calculating linear expansion (ΔL=α·L·ΔT, where ΔL is the linear expansion, α is the coefficient of linear expansion, L is the original length, and ΔT is the temperature change), the radial clearance of the sliding bearing is not only affected by the coefficient of linear expansion of the material and the service temperature or temperature change, but also positively correlated with the shaft diameter (or the outer diameter of the bearing inner ring), the thickness of the bearing bush, and the inner diameter of the bearing housing (or the outer ring of the bearing). Furthermore, the radial dimensions of the shaft (or the inner ring of the bearing), the bearing bush, and the bearing housing (or the outer ring of the bearing) are interconnected. Therefore, the corresponding method for controlling the change in the radial clearance of the sliding bearing should also include, based on the selection of materials (including corresponding mechanical properties, performance characteristics, and coefficient of linear expansion factors), adjusting the structure and / or dimensions of each component according to the requirements of the radial clearance of the sliding bearing, combined with the stress and mechanical design requirements of each part.

[0070] In principle, the expansion of the shaft (or bearing inner ring) (the expansion mentioned in this invention is a collective term including thermal expansion and thermal contraction) plus the expansion of the bearing bush equals the expansion of the bearing housing (or bearing outer ring), thus the radial clearance of the sliding bearing (or sliding bearing assembly) remains unchanged. According to this invention, the complementary radial expansion of the shaft (or bearing inner ring), bearing bush, and bearing housing (or bearing outer ring) is utilized to control the variation of the radial clearance of the sliding bearing under different service temperatures.

[0071] Meanwhile, combining theory and practice, it is known that after thermal expansion and contraction, taking a single-layer bearing as an example, the inner diameter of the first bearing shell should not be less than the outer diameter of the shaft or bearing inner ring that it contacts or mates with. The inner diameter of the first bearing shell is determined according to this principle to avoid the phenomenon in practice where the first bearing shell is damaged because it only contacts the shaft or bearing inner ring at both ends along the circumferential direction on its inner ring surface due to thermal expansion and contraction.

[0072] In this invention, "the inner diameter of the first bearing shell" and "the inner diameter of the first bearing bush" are two different concepts, and they are usually not equal. The "inner diameter of the first bearing shell" refers to the inner diameter of the curved surface on which the inner ring surface of the first bearing shell is located; the "inner diameter of the first bearing bush" refers to the diameter of the circle on which the geometric center of the inner ring surface of each first bearing shell that constitutes the first bearing bush lies.

[0073] Similarly, the "outer diameter of the first bearing" refers to the fact that the geometric center of the outer ring surface of each first bearing shell that constitutes the first bearing is located on the same circle, and the diameter of this circle is the "outer diameter of the first bearing".

[0074] Taking the single-layer bearing sliding bearing of this invention as an example, it satisfies the following relationship:

[0075] The inner diameter of the first bearing bush = shaft diameter (or outer diameter of the inner ring of the bearing) D + radial clearance δ reserved during the manufacturing of the sliding bearing;

[0076] The inner diameter of the bearing housing (or the outer ring of the bearing) = the outer diameter of the first bearing shell = the inner diameter of the first bearing shell + 2 times the bearing shell thickness t. The first bearing shell thickness t mentioned here refers to the minimum distance between the inner and outer radial surfaces of each bearing shell that constitutes the bearing shell, that is, the thickness at the thinnest radial position of the bearing shell.

[0077] Preferably, the bearing bush is made of ceramic material or hard alloy material; including but not limited to silicon carbide, boron carbide, tungsten carbide silicon nitride, boron nitride, alumina, zirconium oxide, titanium oxide, etc., or tungsten nickel titanium alloy, tungsten cobalt titanium alloy, tungsten zirconium titanium alloy, etc.

[0078] The method for controlling the change in radial clearance of sliding bearings due to thermal expansion described in this invention, through theoretical and practical analysis and derivation, and through a large number of simulation experiments and test verifications, achieves the goal of ensuring that the change in radial clearance of the sliding bearing is close to zero or within the allowable range under high temperature or low temperature conditions without the need for a special expansion compensation structure.

[0079] Compared with the prior art, the outstanding features of this invention are as follows:

[0080] 1. The sliding bearing described in this invention avoids the problem of mismatch in shape and / or size caused by different materials and different coefficients of thermal expansion of the various parts of the sliding bearing in the prior art, which causes different degrees of deformation of the mating surfaces at room temperature when the temperature changes. This avoids the problem of bearing damage caused by such mismatch.

[0081] Bearing failure accounts for a very high percentage of equipment failures. The bearing bush of this invention consists of multiple shells. The small size of the shells allows for a thinner bearing bush. The presence of a flat surface on the upper surface prevents the bearing bush from being subjected to bending moments under stress. There are no interference fit forces during assembly or disassembly, making the bearing bush less prone to damage. The structure and shape of the bearing bush in this invention result in good stress distribution. The planes on both circumferential sides of the bearing bush are on the same plane as the axis of the shaft, ensuring that the radial force on the bearing bush does not extend beyond its surface, further enhancing stress distribution. Furthermore, the split design of the bearing bush reduces the likelihood of catastrophic brittle fracture during impacts or vibrations. More importantly, in this invention, the bearing bush and connected parts have a planar fit or connection, preventing the bearing bush from swaying around the shaft centerline, avoiding collisions, and ensuring safe and reliable operation.

[0082] 2. By employing the sliding bearing and the method for controlling the radial clearance variation of the sliding bearing described in this invention, and through the rational selection and design of its components, including materials and dimensions, the goal of maintaining a constant radial clearance or keeping its variation within allowable ranges at both high and low temperatures is achieved. Without requiring a dedicated clearance compensation structure, the structure is simpler and more rational, less prone to damage, and ensures the effective and safe use of the sliding bearing (assembly), while also guaranteeing its service life. This significantly reduces costs associated with production, manufacturing, assembly, use, and maintenance. It is particularly suitable for replacing cylindrical bearing bushes or sliding bearing assemblies used in magnetic pumps, helping to reduce eddy current losses and improve energy efficiency.

[0083] This invention, applied to magnetic drive pumps, can effectively replace energy-intensive and difficult-to-maintain canned motor pumps. When the sliding bearing described in this invention is used in a magnetic drive pump, when the medium temperature is below 700℃, the medium returned to the pump outlet for lubrication or cooling does not need to be cooled, and the return flow rate can be reduced. This helps improve pump efficiency, reduce energy consumption, and avoids problems such as energy consumption and medium crystallization caused by cooling. It meets the requirements of some applications where cooling is not permissible, and eliminates the problems associated with heat exchangers, such as insulation, heat tracing, and scaling. When this invention is used with high-temperature media such as lead-bismuth alloys or molten salts at 500–750℃, the bearing's service life can be guaranteed solely by internal medium cooling without the need for external cooling methods or devices, thus avoiding a series of problems associated with external cooling measures.

[0084] In fact, the sliding bearing described in this invention can be widely used in various high-temperature or low-temperature applications, including engines, gas turbines, magnetic pumps, lead-bismuth alloy pumps, and submersible molten salt pumps. It eliminates the need for excessive expansion clearances and gap compensation structures such as expansion sleeves. Crucially, it prevents "seizing" at high temperatures, significantly improving its temperature range. It also solves the problem of bearings in high-temperature applications requiring external cooling systems and thus facing various related issues. For example, when used as a bearing in steel rolling equipment in the metallurgical industry, this invention eliminates the need for excessive clearances to prevent seizing. In particular, this invention is applicable to bearings in submersible molten salt pumps, where the bearing shell material can be silicon carbide or boron nitride, allowing for cold and thermal shock. At temperatures below 1000°C, its hardness and wear resistance are excellent, fully meeting the performance requirements of sliding bearings in this field at high temperatures. Attached Figure Description

[0085] Figure 1 This is a structural schematic diagram according to Embodiment 1 of the present invention;

[0086] Figure 2 This is a structural schematic diagram of Embodiment 2 of the present invention, and also a structural schematic diagram of a sliding bearing assembly described in the present invention;

[0087] Figure 3 This is a structural schematic diagram according to Embodiment 3 of the present invention;

[0088] Figure 4 This is a structural schematic diagram according to Embodiment 4 of the present invention;

[0089] Figure 5 yes Figure 4 Enlarged view of section C in the image;

[0090] Figure 6 This is a structural schematic diagram according to Embodiment 5 of the present invention;

[0091] Figure 7 This is a structural schematic diagram according to Embodiment 6 of the present invention;

[0092] Figure 8 This is a schematic diagram of the structure of a sliding bearing assembly according to the present invention;

[0093] Figure 9 This is a schematic diagram of the frustum-shaped mating surface of the two-layer bearing as described in an embodiment of the present invention;

[0094] Figure 10 This is a schematic diagram of the end cap and cooling circuit described in the embodiments of the present invention;

[0095] Figure 11 This is a schematic diagram of the snap ring connection structure described in the embodiments of the present invention;

[0096] Figure 12 This is a schematic diagram of the cylindrical bearing used in existing magnetic pumps;

[0097] Figure 13 This is a schematic diagram showing the contact shape between the inner circumferential wall of the bearing housing or the outer ring of the bearing and the outer ring wall of the bearing shell after thermal expansion, due to the small coefficient of thermal expansion of the bearing shell in the existing technology.

[0098] Figure 14 This is a schematic diagram showing the contact shape between the inner circumferential wall of the bearing housing or outer ring of the bearing and the outer ring wall of the bearing shell after cooling and contraction due to the small coefficient of thermal expansion of the bearing shell in the existing technology.

[0099] Figure 15 This is a diagram illustrating the simulation results of the bearing outer ring in a simulation based on an embodiment of the present invention.

[0100] Figure 16 This is a diagram illustrating the simulation results of the central axis based on an embodiment of the present invention.

[0101] Figure 17 This is a diagram illustrating the simulation results of the bearing bush in a simulation based on an embodiment of the present invention.

[0102] In the picture:

[0103] 1. Bearing housing or outer ring 2. First bearing housing 3. Shaft 4. Fastener 5. Inner bearing ring 6. Second bearing housing 7. Second boss 8. First boss 9. First groove 10. Second groove 11, 11'. Inner retaining ring 12, 12'. Outer retaining ring 13, 13'. End caps R1, R2. Thinning zone I. Cooling medium inlet O. Cooling medium outlet

[0104] 21. Split bearing shell 22. Cylindrical bearing A 23. Cylindrical bearing B 24. Expansion sleeve Detailed Implementation

[0105] The present invention will now be further described in detail with reference to the embodiments.

[0106] Example 1

[0107] A type of sliding bearing, such as Figure 1 As shown, it includes a first bearing shell, a bearing housing or a bearing outer ring 1. The first bearing shell includes eleven first bearing shells 2, which are evenly distributed along the circumferential direction on the inner peripheral wall of the bearing housing or the bearing outer ring 1.

[0108] The outer ring surface of each first bearing shell 2 is in planar cooperation with the inner circumferential wall of its corresponding bearing seat or bearing outer ring 1, and is fixed by fasteners 4; the inner ring surface of the first bearing shell 2 is in sliding cooperation with the shaft 3 and the journal.

[0109] Example 2

[0110] A type of sliding bearing, such as Figure 2 As shown, the main difference between this embodiment and Embodiment 1 is that the inner ring surface of the first bearing housing 2 is in sliding fit with the outer peripheral wall of the bearing inner ring 5.

[0111] The inner ring 5 of the bearing, the first bearing shell (including eleven first bearing shells 2), and the outer ring 1 of the bearing are assembled to form a sliding bearing assembly.

[0112] Example 3

[0113] A type of sliding bearing, such as Figure 3 As shown, its main difference from Example 1 is:

[0114] On the inner circumferential wall of the bearing housing or outer ring 1, a first groove 9 is respectively dug for each of the first bearing shells 2 (ten first bearing shells in total). (The number is not shown in this figure; its location can be found in...) Figure 7 A first protrusion 8 is formed between two adjacent first grooves 9; the bottom of the first groove 9 is a plane that mates with the outer ring surface of the corresponding first tile shell 2; the radial depth of the first groove 9 is less than the thickness of the corresponding first tile shell 2, and the first tile shell 2 is partially embedded in the first groove 9 radially.

[0115] Better still, the first groove 9 can be wider at the bottom and narrower at the opening, such as... Figure 3 As shown, it has a dovetail groove shape; the corresponding first tile shell 2 is wider at the bottom and narrower at the top along the radial direction.

[0116] In the above embodiments, the rotating surface of the shaft 3 or the inner ring 5 of the bearing that mates with the first bearing shell 2 is formed by spraying, plating, or fusion fixing to form a composite layer with surface hardening effect and other special properties. The composite layer can be a ceramic material layer, such as silicon carbide, boron carbide, silicon nitride, boron nitride, alumina, zirconium oxide, titanium oxide, etc., or it can be a hard alloy layer, such as tungsten nickel titanium alloy, tungsten cobalt titanium alloy, tungsten zirconium titanium alloy, etc., which has wear resistance, high temperature resistance or low temperature resistance.

[0117] Example 4

[0118] A type of sliding bearing, such as Figure 4 As shown, its main difference from Embodiment 1 is that, based on Embodiment 1, it also includes a second bearing bush, which is located between the first bearing bush and the shaft, forming a double-layer bearing bush structure to improve the service life of the shaft.

[0119] Specifically, the first bearing bush includes nine first bearing shells 2, which are evenly distributed along the circumferential direction on the inner circumferential wall of the bearing seat or the outer ring 1 of the bearing, and are mated with each other on a plane and fixed thereon by fasteners 4; the second bearing bush includes ten second bearing shells 6, which are evenly distributed along the circumferential direction on the journal of the shaft 3 or the outer circumferential wall of the inner ring of the bearing, and are mated with each other on a plane and fixed thereon by fasteners 4.

[0120] The outer ring surface of the second bearing shell 6 forms a sliding fit with the inner ring surface of the first bearing shell 2. Furthermore, the number of bearing shells of the first bearing shell and the second bearing shell differs by one. This design aims to avoid collisions or obstructions when the gaps between the first bearing shell 2 and the second bearing shell 6 overlap during rotation.

[0121] Furthermore, to more effectively avoid collisions or obstruction between the two layers of the shell, the two layers are positioned near their respective circumferential end faces on their sliding mating surfaces, such as... Figure 4 As shown in section C, the regions R1 and R2 are radially thinned, with each region slightly raised or gradually separating from the others. Section C is magnified as follows: Figure 5 As shown in the figure, the arrows indicate the direction of thinning. The thinned regions R1 and R2, when viewed in cross-section, can form a straight line, an involute, or an Archimedean spiral.

[0122] This design of two-layer bearings that gradually thins radially on their sliding mating surfaces can also be used in other two-layer bearing designs of this invention.

[0123] Example 5

[0124] A type of sliding bearing, such as Figure 6 As shown, based on Embodiment 3, it further includes a second bearing bush. Specifically, the sliding bearing includes a first bearing bush and a second bearing bush.

[0125] The first bearing bush includes nine first bearing shells 2. A first groove 9 (not shown in the figure, but its location can be found in the inner circumferential wall of the bearing seat or outer ring 1) is respectively carved into the inner circumferential wall of each first bearing shell 2. Figure 7 A first protrusion 8 is formed between two adjacent first grooves 9; the bottom of the first groove 9 is a plane that mates with the outer ring surface of the corresponding first tile shell 2; the radial depth of the first groove 9 is less than the thickness of the corresponding first tile shell 2, and the first tile shell 2 is partially embedded in the first groove 9 radially; similar to the structure of the first groove in embodiment 3, the first groove 9 is wide at the bottom and narrow at the opening, and is in the shape of a dovetail groove; the corresponding first tile shell 2 is wide at the bottom and narrow at the top radially.

[0126] The second bearing bush includes ten second bearing shells 6, and a second groove 10 is respectively carved on the journal of the shaft 3 corresponding to each second bearing shell 6 (not marked in this figure, but its position can also be seen in...). Figure 7 A second protrusion 7 is formed between two adjacent second grooves 10; the bottom of each second groove 10 is a plane that mates with the corresponding second tile shell 6; the radial depth of the second groove 10 is less than the thickness of its corresponding second tile shell 6, and the second tile shell 6 is partially embedded radially into the second groove 10. The second groove 10 is narrow at the bottom and wide at the opening, with a trapezoidal cross-section; the corresponding second tile shell 6 is narrow at the bottom and wide at the top radially.

[0127] The outer ring surface of the second shell 6 forms a sliding fit with the inner ring surface of the first shell 2.

[0128] In the sliding bearing, the mating planes of the first bearing shell 2 and the bearing seat or bearing outer ring 1, and the mating planes of the second bearing shell 6 and the shaft 3, ensure that the bearing shells themselves do not rotate within their respective grooves, and that the contact area is large, the load-bearing capacity is strong, and they are not easily broken. Furthermore, the first bearing shell and the second bearing shell always maintain a state of near-total contact, including, for example, one first bearing shell 2 of the first bearing shell and one second bearing shell 6 of the second bearing shell maintaining near-total contact and mating, or one first bearing shell 2 of the first bearing shell simultaneously contacting and mating with two second bearing shells 6 of the second bearing shell, or two first bearing shells 2 of the first bearing shell simultaneously contacting and mating with one second bearing shell 6 of the second bearing shell. Thus, the large contact area between the two bearing shells prevents them from falling off, and the probability of rotational interference between the two bearing shells is very small, thereby avoiding collisions between them.

[0129] Example 6

[0130] A type of sliding bearing, such as Figure 7 As shown, based on embodiment 5, further adjustments were made to the grooves 9 and 10 and the corresponding bosses 8 and 7 according to specific needs.

[0131] Specifically, considering thermal expansion or contraction, gaps are easily formed between the mating surfaces of the first bearing shell 2 and the bearing bracket or bearing outer ring 1, and between the second bearing shell 6 and the shaft 3 (or bearing inner ring). Solid impurities can enter these gaps, depriving the bearing shells of space for thermal expansion and contraction. After thermal expansion or contraction, the presence of these solid impurities increases the probability of the bearing shells being crushed. To avoid the crushing of the bearing shells, a space is reserved along the bottom of the first groove 9 and the second groove 10. The corresponding first protrusion 8 and second protrusion 7 in the middle of the two grooves are respectively formed with a smaller circumferential dimension near the bottom of their respective grooves and a larger circumferential dimension further away from the bottom of their respective grooves, presenting a protrusion form that is narrow at the bottom and wide at the top. For example, a T-shaped first protrusion 8 and second protrusion 7 are formed. Figure 7 As shown.

[0132] Furthermore, to address the issue of gaps easily forming between the mating surfaces of the tile shell and the corresponding groove during thermal expansion and contraction, allowing solid impurities to enter, an inner and outer retaining spring structure can be employed to adjust the radial force on the tile shell, thereby offsetting or reducing or eliminating the gap due to the force during thermal expansion and contraction. Specifically, such as... Figure 11 As shown, grooves are respectively opened along the circumferential direction on the inner wall of the first tile shell 2 and the outer wall of the second tile shell 6, wherein inner retaining springs 11 and 11' and outer retaining springs 12 and 12' are respectively installed.

[0133] In order to ensure that the radial force on the bearing does not extend beyond the bearing, each plane on both sides of the bearing is located in the same plane as the axis of shaft 3.

[0134] Furthermore, to better prevent solid impurities from entering the gap between the mating surfaces of the bearing, end caps 13 and 13' can be respectively provided at both ends of the bearing housing or the outer ring 1 of the bearing. The end caps 13 and 13' are provided with shaft holes, which form a clearance fit with the shaft 3. Figure 10 As shown. Meanwhile, in order to cool the housings 2, 6, and / or the isolation sleeve of the magnetic pump using the sliding bearing (or its assembly), cooling and / or lubricating medium inlets I and outlets O are respectively provided on the end caps 13, 13', and are respectively connected to the flow gaps or passages of the cooling and / or lubricating medium in the sliding bearing (or its assembly).

[0135] The flow gap or passage of the cooling and / or lubricating medium may include a through groove provided along the axial direction between the mating surfaces of a single layer of bearing and the shaft or between the mating surfaces of two layers of bearing, and / or a circumferential gap or groove between adjacent shells of each layer of bearing (not shown in the figure).

[0136] Furthermore, considering the axial load requirements of the bearing, the mating surfaces of the first bearing housing 2 and the second bearing housing 6 form a frustum-shaped side surface, similar to that of a tapered roller bearing. Figure 9 As shown, the upper base of the frustum shape is located at the end of the shaft 3.

[0137] Combining the two-layer bearing structure described in the above embodiments, assembling the inner bearing ring 5, the outer bearing ring 1, and the first and second bearing shells together constitutes a sliding bearing assembly, such as... Figure 8 As shown.

[0138] Based on the sliding bearing described in the above embodiments, the inventors further used the single-layer bearing described in Embodiment 3 above as an example to simulate and calculate the change in radial clearance due to thermal expansion (using Siemens NXNASTRAN simulation software, and a desktop computer with a 64-core processor, 128 threads, and 32GB of memory).

[0139] The service temperatures of the inner ring, bearing bush, and outer ring of the shaft or bearing are set to be the same or close; a shaft with a diameter D = 50 is selected and 321 steel is used (its linear expansion coefficient α1 = 17.2 × 10⁻⁶ at 300℃). -6 / ℃); the bearing shell is made of silicon carbide (at 300℃, its coefficient of linear expansion is α2=4×10). -6 / ℃), the bearing shell thickness (taken as its circumferential center thickness) is t=18.421 (calculated according to the principle of the sliding bearing radial clearance change control method due to thermal expansion described in this invention), the initial inner diameter Φ1 of the bearing shell is 50+δ; the bearing outer ring is made of 3Cr13 steel (at 300℃, its linear expansion coefficient α3=11.6×10 -6 / ℃), its inner diameter D i =86.842+δ; When the medium temperature rises from 0℃ to 300℃, the simulation calculation results of the bearing outer ring, shaft, and bearing shell (specifically represented by a shell constituting the bearing shell) are as follows: Figure 15 , Figure 16 and Figure 17 As shown, the corresponding dimensions are: shaft diameter D′=50.258, bearing shell thickness t′=18.44308, and the inner diameter D of the bearing outer ring. i =87.14406 (Since the radial clearance δ reserved during the manufacturing of sliding bearings is usually within a range of several channels, one channel is 0.01mm, and its impact on the overall calculation result does not exceed 10). -4 The change in radial clearance of the sliding bearing after thermal expansion (on the order of mm, negligible in calculations) is as follows:

[0140] D i ′-D′-2t′=87.14406-50.258-2×18.44308=-0.0001

[0141] The simulation results show that the radial clearance change due to thermal expansion of the sliding bearing is 0.0001 mm. Even without considering factors such as computer resolution limitations (which do exist), this order of magnitude is negligible in practice. The simulation results further demonstrate that the sliding bearing and its control method for radial clearance changes due to thermal expansion, as described in this embodiment of the invention, effectively achieve the goal of keeping the radial clearance of the sliding bearing (assembly) constant or within a controllable range when the service temperature changes.

[0142] Obviously, the above embodiments are only some embodiments of the present invention, and not all embodiments. Based on this, any technical solution obtained by those skilled in the art, within the scope of the technology disclosed in this invention, by making equivalent substitutions or modifications to the technical solutions and inventive concepts of the present invention without creative effort, should be covered within the protection scope of this invention.

Claims

1. A sliding bearing, comprising a bearing shell; the bearing shell comprising three or more first bearing shells; the first bearing shells being evenly distributed circumferentially on the inner peripheral wall of a bearing housing or an outer ring of a bearing; the first bearing shells being planarly fitted or fixed to their corresponding bearing housing or outer ring; Its features are: The bearing bush also includes three or more second bearing shells; the second bearing shells are evenly distributed circumferentially on the outer peripheral wall of the journal or the inner ring of the bearing and are mated or fixed in a planar manner; the outer ring surface of the second bearing shell forms a sliding fit with the inner ring surface of the first bearing shell; When the service temperatures of the shaft or bearing inner ring, bearing bush, bearing housing, or bearing outer ring are the same or close: (1) The materials selected meet the criteria of α1 > α3 > α2; (2) The structural dimensions satisfy the following: t is positively correlated with D, positively correlated with the difference between α1 and α3, and negatively correlated with the difference between α3 and α2, and D i =D+2t+δ, and D≥D0, that is, t=D(α1-α3) / (α3-α2); Under the condition that the service temperatures of the shaft or bearing inner ring, bearing bush, bearing housing, or bearing outer ring decrease sequentially and the temperature difference is significant: (1) Its material selection conforms to α3≧α1>α2; (2) The structural dimensions satisfy that t is positively correlated with D, positively correlated with the difference between α1ΔT1 and α3ΔT3, and negatively correlated with the difference between α3ΔT3 and α2ΔT2, and D i =D+2t+δ, and D≥D0, that is, t=D(α1ΔT1 - α3ΔT3) / (α3ΔT3 - α2ΔT2); Where α1 is the linear expansion coefficient of the shaft or bearing inner ring, α2 is the linear expansion coefficient of the bearing bush, and α3 is the linear expansion coefficient of the bearing housing or bearing outer ring; t is the thickness of the bearing bush, and D is the shaft diameter or the outer diameter of the bearing inner ring. i D0 is the inner diameter of the bearing housing or outer ring of the bearing; D0 is the theoretical value of the shaft diameter or outer diameter of the bearing inner ring that meets the product design requirements; δ is the radial clearance reserved during the manufacturing of the sliding bearing; ΔT1 is the difference between the service temperature of the shaft or bearing inner ring and the room temperature; ΔT2 is the difference between the service temperature of the bearing bush and the room temperature; ΔT3 is the difference between the service temperature of the bearing housing or outer ring of the bearing and the room temperature; α1ΔT1, α2ΔT2, and α3ΔT3 represent the products of α1 and ΔT1, α2 and ΔT2, and α3 and ΔT3, respectively.

2. The sliding bearing as described in claim 1, characterized in that: A first groove is excavated on the inner circumferential wall of the bearing housing or the outer ring of the bearing for each of the first bearing shells, and a first boss is formed between two adjacent first grooves; the bottom of the first groove is the plane that mates with the corresponding first bearing shell; the radial depth of the first groove is less than the thickness of the corresponding first bearing shell, and the first bearing shell is partially embedded in the first groove radially.

3. The sliding bearing as described in claim 2, characterized in that: The first groove is wide at the bottom and narrow at the opening, forming a dovetail shape; correspondingly, the first tile shell is wide at the bottom and narrow at the top along the radial direction.

4. The sliding bearing as described in claim 2, characterized in that: If the circumferential dimension of the first boss near the bottom of the first groove in the radial direction is smaller than the circumferential dimension of the first boss far from the bottom of the first groove in the radial direction, then the cross-section of the first boss will be narrow at the bottom and wide at the top.

5. The sliding bearing as described in claim 1, characterized in that: The first bearing shell slides in conjunction with the shaft or the inner ring of the bearing. The inner ring of the shaft or bearing has a composite layer that is wear-resistant, high-temperature resistant, or low-temperature resistant, which mates with the first bearing shell.

6. The sliding bearing as described in claim 1, characterized in that: It also includes a torsion spring, which is located between the first bearing shell and the shaft or bearing inner ring, and is tightly clamped to the shaft or bearing inner ring.

7. The sliding bearing as described in claim 1, characterized in that: The mating surfaces of the first and second tile shells together form a frustum-shaped side surface, with the upper base of the frustum located at the end of the shaft.

8. The sliding bearing as described in claim 1, characterized in that: A second groove is excavated on the outer peripheral wall of the journal or bearing inner ring for each of the second bearing shells, and a second boss is formed between two adjacent second grooves; the bottom of each second groove is a plane that mates with the corresponding second bearing shell; the radial depth of the second groove is less than the thickness of the corresponding second bearing shell, and the second bearing shell is partially embedded in the second groove radially.

9. The sliding bearing as described in claim 8, characterized in that: The second groove is narrow at the bottom and wide at the opening, with a trapezoidal cross-section; correspondingly, the second tile shell is narrow at the bottom and wide at the top along the radial direction.

10. The sliding bearing as described in claim 8, characterized in that: If the circumferential dimension of the second boss near the bottom of the second groove in the radial direction is smaller than the circumferential dimension of the second boss far from the bottom of the second groove in the radial direction, then the cross-section of the second boss will be narrow at the bottom and wide at the top.

11. The sliding bearing as claimed in claim 1, characterized in that: The number of the first tile shell is different from the number of the second tile shell.

12. The sliding bearing as claimed in claim 1, characterized in that: The first and second tile shells, on their sliding mating surfaces near their respective circumferential end faces, are slightly raised or gradually separated radially away from each other, forming radially thinned regions.

13. A sliding bearing assembly comprising the sliding bearing as described in any one of claims 1 to 12.

14. A method for controlling the change in radial clearance of a sliding bearing due to thermal expansion, wherein the sliding bearing described in any one of claims 1 to 12 is used; When the service temperatures of the shaft or bearing inner ring, bearing bush, bearing housing, or bearing outer ring are the same or close: (1) The materials selected meet the criteria of α1 > α3 > α2; (2) The structural dimensions satisfy the following: t is positively correlated with D, positively correlated with the difference between α1 and α3, and negatively correlated with the difference between α3 and α2, and D i =D+2t+δ, and D≥D0; Under the condition that the service temperatures of the shaft or bearing inner ring, bearing bush, bearing housing, or bearing outer ring decrease sequentially and the temperature difference is significant: (1) Its material selection conforms to α3≧α1>α2; (2) The structural dimensions satisfy that t is positively correlated with D, positively correlated with the difference between α1ΔT1 and α3ΔT3, and negatively correlated with the difference between α3ΔT3 and α2ΔT2, and D i =D+2t+δ, and D≥D0; in, α1 is the linear expansion coefficient of the shaft or bearing inner ring, α2 is the linear expansion coefficient of the bearing bush, and α3 is the linear expansion coefficient of the bearing housing or bearing outer ring; t is the thickness of the bearing bush, and D is the shaft diameter or the outer diameter of the bearing inner ring. i D0 is the inner diameter of the bearing housing or outer ring of the bearing; D0 is the theoretical value of the shaft diameter or outer diameter of the bearing inner ring that meets the product design requirements; δ is the radial clearance reserved during the manufacturing of the sliding bearing; ΔT1 is the difference between the service temperature of the shaft or bearing inner ring and the room temperature; ΔT2 is the difference between the service temperature of the bearing bush and the room temperature; ΔT3 is the difference between the service temperature of the bearing housing or outer ring of the bearing and the room temperature; α1ΔT1, α2ΔT2, and α3ΔT3 represent the products of α1 and ΔT1, α2 and ΔT2, and α3 and ΔT3, respectively.

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