Plain bearings for water pumps

Abstract

To provide a slide bearing device for a water pump which imparts a high dynamic pressure effect to a lubrication groove formed at the slide bearing or an end face of a thrust receiver, and is reduced in a friction coefficient between the slide bearing and the thrust receiver.SOLUTION: A slide bearing device is composed of a slide bearing 8 and a thrust receiver, and the slide bearing 8 comprises a land part 13 being a slide face, and a lubrication groove 14 for discharging lubrication water to an outside diameter side from an inside diameter side at a bearing end face. The lubrication groove 14 has an inclination face which is inclined to the land part 13, and in a projection figure when viewing the end face having the lubrication groove 14 from a front side, the lubrication groove 14 is formed of a region which is surrounded by a line component A and a line component B which are connected to the outside diameter side from the inside diameter side of the end face, a circular arc C progressing along an inside diameter face, and a circular arc D progressing along an outside diameter face. A length of the circular arc C is equal to or longer than a length of the circular arc D.SELECTED DRAWING: Figure 2

Description

[Technical Field]

[0001] The present invention relates to a sliding bearing device for water pumps used for circulating coolant in automobile engines, inverters, batteries, fuel cells, etc., and for circulating hot water in water heaters, floor heating equipment, etc. [Background technology]

[0002] Water pumps are used to circulate coolant for automobile engines, inverters, batteries, or fuel cells, and to circulate hot water for water heaters and floor heating equipment. Typical examples of water pumps used for these purposes include a magnetic pump as described in Patent Document 1 and a DC brushless pump as described in Patent Document 2. A conventional water pump will be described with reference to FIG. 12. FIG. 12 is a cross-sectional view of a DC brushless pump. In this pump 21, a motor 32 has a winding 22 with a coil arranged thereon to generate a magnetic field, and the generation of the magnetic field is controlled by a control unit. An impeller 24, to which a permanent magnet 23 is fixed, is rotatably supported by a shaft 25 so as to follow the generated magnetic field. Circulating water is drawn in and discharged as the impeller 24 rotates in response to the rotating magnetic field. The shaft 25 is fixed to a casing 26 and supported by a shaft support 27a of a cover 27. Impeller 24 is rotatably supported on shaft 25 via slide bearing 28, and shaft 25 and slide bearing 28 rotate and slide. Furthermore, both end faces of slide bearing 28 rotate and slide in the thrust direction with thrust plates 29, 30 provided between shaft support 27a of cover 27 and casing 26, respectively. A small gap is provided between both end faces of slide bearing 28 and thrust plates 29, 30, respectively.

[0003] As the rotating magnetic field generated by the winding 22 rotates, the impeller 24 rotates due to the attraction and repulsion of the fixed permanent magnet 23. This creates a pumping action, sucking in circulating water in the direction of arrow X and expelling it in the direction of arrow Y. The pressure difference at this time presses the impeller 24 toward the cover 27, causing the end face of the plain bearing 28 and thrust plate 29 to rotate and slide. There is almost no sliding between the plain bearing 28 and the thrust plate 30 on the casing 26 side, except for brief moments during start-up and shutdown, or during abnormal operation such as dry running when the pump is operating without circulating water. For this reason, the thrust plate 30 is not used on the casing 26 side, and the plain bearing 28 may slide directly on the casing 26.

[0004] In the water pump described above, the coefficient of friction can be reduced by providing lubrication grooves in either the sliding bearing or the thrust bearing. Patent Document 3 discloses a thrust support device comprising a collar fixed to a rotating shaft and having a sliding surface perpendicular to the rotating shaft, and a thrust bearing fixed to a fixed member and rotating relatively along the sliding surface. This thrust bearing is provided with a land portion parallel to the sliding surface of the collar, a tapered portion inclined relative to the sliding surface and generating dynamic pressure in the lubricating fluid between the collar and the land portion and the collar due to relative rotation, and a groove. The generation of dynamic pressure by the tapered portion reduces the coefficient of friction.

[0005] The inventors have also proposed a water pump having at least one selected from suction means that draws circulating water from one end face side of the sliding bearing to the bearing bore surface side, and discharge means that discharges circulating water from the bearing bore surface side to the other end face side of the sliding bearing (Patent Document 4).This water pump aims to reduce friction by improving the supply of circulating water to the sliding surfaces (discharge capacity). [Prior art documents] [Patent documents]

[0006] [Patent Document 1] Patent No. 3099434 [Patent Document 2] Japanese Patent Application Laid-Open No. 2006-200427 [Patent Document 3] Patent No. 5761560 [Patent Document 4] Japanese Patent Application Laid-Open No. 2015-183650 Summary of the Invention [Problem to be solved by the invention]

[0007] In recent years, there has been a demand for water pumps with reduced friction coefficients to save energy and reduced vibrations to improve quietness. The thrust bearing in Patent Document 3 is provided with a tapered portion that generates dynamic pressure in the lubricating fluid between the collar and the bearing due to relative rotation, but the angle and length of the tapered portion are not considered. Furthermore, in the thrust bearing illustrated in Patent Document 3, the boundary between the tapered portion and the land portion is formed along the radial direction of the thrust bearing, and it is believed that there is room for further improvement in the dynamic pressure effect.

[0008] The present invention has been made in order to address these problems, and its object is to provide a sliding bearing device for a water pump that imparts a high dynamic pressure effect to the lubrication grooves provided on the end face of the sliding bearing or thrust bearing, thereby reducing the coefficient of friction between the two. [Means for solving the problem]

[0009] The sliding bearing device for a water pump of the present invention comprises an impeller, a shaft for supporting the impeller, a cylindrical sliding bearing fixed to the impeller for rotatably supporting the impeller relative to the shaft, a thrust receiver that slides against each end face of the sliding bearing, and a casing and cover that house the impeller and form a pump chamber, and is used in a water pump that draws in and discharges circulating water via the pump chamber by the rotation of the impeller, and the sliding bearing device for a water pump comprises the sliding bearing and the thrust receiver, and At least one end face or an end face of at least one member of the thrust bearing is provided with a land portion which serves as a sliding surface and a lubrication groove which discharges the circulating water from the inner diameter side to the outer diameter side, the lubrication groove has an inclined surface which is inclined with respect to the land portion, and in a projection view of the end face on which the lubrication groove is provided, viewed from the front (axial direction), the lubrication groove consists of an area surrounded by line segment A and line segment B which connect the inner diameter side to the outer diameter side of the end face, arc C along the inner diameter surface, and arc D along the outer diameter surface, and the length of arc C is equal to or longer than the length of arc D.

[0010] In the present invention, "thrust bearing" refers not only to a dedicated thrust plate provided to bear the thrust load of a sliding bearing, but also to other components such as a casing when the thrust load is received by the other components. The dedicated thrust plate may be, for example, annular. Furthermore, "circulating water" includes not only water, but also antifreeze, chemical liquids, and the like.

[0011] The angle formed by the line segment A and the line segment B is 0° to 15°.

[0012] The line segment A is located upstream of the line segment B in the relative rotation direction, and the cross-sectional shape of the lubrication groove taken along any cut surface perpendicular to the line segment A is a substantially right-angled triangle with the inclined surface as its hypotenuse, and the interior angle formed by the intersection of the line segment A and the cut surface as its vertex is 3° to 30°. Here, the term "relative rotation direction" refers to the direction of rotation of either the sliding bearing or the thrust bearing when the sliding bearing rotates, or the opposite direction to the rotation of the counterpart material when the counterpart material rotates.

[0013] The lubrication groove has a maximum depth of 0.1 mm to 1.0 mm.

[0014] The lubrication grooves are arranged in a plurality at intervals in the circumferential direction on the end surface.

[0015] The member in which the lubrication grooves are formed has an axial center on an extension of the line segment A, and the plurality of lubrication grooves are provided at a position offset downstream in the relative rotation direction from the center line passing through the line segment A of the member.

[0016] The sliding bearing is an injection-molded article made of a resin composition, has the lubrication groove on at least one end face thereof, and has gate marks formed on the outer diameter surface of the sliding bearing, and is characterized in that a weld is formed in the lubrication groove and no weld is formed in the land portion in the sliding bearing.

[0017] The thrust bearing is characterized in that the direction of the grooves formed by machining on the surface that slides against the plain bearing is not concentric.

[0018] The direction of the streaks is random. [Effects of the Invention]

[0019] The sliding bearing device of the present invention has a land portion that serves as a sliding surface and a lubrication groove that discharges circulating water on the end face of at least one of the sliding bearing or thrust bearing. The lubrication groove has an inclined surface that is inclined relative to the land portion, and furthermore, the lubrication groove has a predetermined configuration, thereby achieving excellent low-friction characteristics. Excellent low-friction characteristics are achieved because the lubrication groove provided on the end face of the sliding bearing or thrust bearing can exert an excellent dynamic pressure effect due to its groove shape, and because the lubrication groove is connected from the inner diameter side to the outer diameter side, circulating water is easily supplied to the sliding surfaces between the sliding bearing and the thrust bearing, improving the lubrication state. In other words, the lubrication groove alone combines a dynamic pressure effect and the effect of improving the supply of circulating water to the sliding surfaces. Therefore, compared to conventional thrust bearings that have both a tapered portion that generates dynamic pressure and a lubrication groove, the configuration is simpler and leads to lower costs.

[0020] In particular, in the present invention, the length of arc C along the inner diameter surface of the lubrication groove is equal to or longer than the length of arc D along the outer diameter surface, and the outlet is the same as or narrower than the inlet in the flow direction of the circulating water, making it easier to generate dynamic pressure compared to a configuration in which the outlet is wider than the inlet, such as in conventional thrust bearings.

[0021] Furthermore, the sliding bearing is an article injection-molded from a resin composition and has a lubrication groove on at least one end surface thereof, and since weld lines are formed in the lubrication groove but not in the land portion of the sliding bearing, the flatness of the land portion can be improved. Improved flatness can reduce vibrations generated during sliding, achieving excellent quietness. [Brief explanation of the drawings]

[0022] [Figure 1] FIG. 1 is a cross-sectional view of a water pump using a sliding bearing device according to the present invention. [Figure 2] FIG. 2 is a front projection view of the end face of a sliding bearing according to the present invention. [Figure 3] FIG. 3 is a cross-sectional view showing the cross-sectional shape of a lubrication groove. [Figure 4] FIG. 1 is a cross-sectional view showing another example of a sliding bearing according to the present invention. [Figure 5] FIG. 10 is a diagram showing changes due to wear of a sliding surface. [Figure 6] FIG. 1 is a front projection of the end faces of the test pieces of Examples 1 to 4. [Figure 7] FIG. 1 is a front projection of the end faces of the test pieces of Examples 5 to 11. [Figure 8] FIG. 1 is a front projection of the end face of the test piece of Comparative Example 1. [Figure 9] FIG. 10 is a front projection view of the end face of the test piece of Comparative Example 2. [Figure 10] 1 shows the change in dynamic friction coefficient over time in Example 12. [Figure 11] 1 shows the change in dynamic friction coefficient over time in Example 13. [Figure 12] FIG. 1 is a cross-sectional view of a conventional water pump. DETAILED DESCRIPTION OF THE INVENTION

[0023] An example of a water pump using a sliding bearing device of the present invention will be described with reference to FIG. 1. As shown in FIG. 1, in water pump 1, casing 6 and cover 7 are fixed together to form a pump chamber that houses impeller 4. Casing 6 and cover 7 are sealed together with packing 11, preventing leakage of circulating water from within the pump chamber. Motor 12 has windings 2 with coils arranged therein, which generate a magnetic field, and the generation of this magnetic field is controlled by a control unit. Impeller 4, to which permanent magnets 3 are fixed in order to follow this generated magnetic field, is rotatably supported by shaft 5 within the pump chamber. As impeller 4 follows the rotating magnetic field and rotates within the pump chamber, circulating water is drawn in and discharged. More specifically, the rotating magnetic field generated by winding 2 causes impeller 4 to rotate due to the attraction and repulsion of fixed permanent magnets 3, which generates a pumping action that draws in circulating water in the direction of arrow X and discharges it in the direction of arrow Y.

[0024] Shaft 5 is fixed to approximately the center of casing 6 and is supported by shaft support 7a of cover 7. Impeller 4 is rotatably supported relative to shaft 5 via cylindrical plain bearing 8 fixed to its center. Shaft 5 is a fixed shaft (it does not rotate), and the outer diameter surface of shaft 5 and the inner diameter surface of plain bearing 8 rotate and slide. Both end surfaces of plain bearing 8 rotate and slide in the thrust direction with thrust plates 9, 10, which serve as thrust bearings and are respectively provided between shaft support 7a of cover 7 and casing 6. Note that a "plain bearing" is a part that receives a load and slides on its inner diameter and end surfaces, and is not necessarily limited to a single part; it may be divided into two or more parts and may be made of different materials.

[0025] In the water pump of the present invention, a lubrication groove is provided that discharges circulating water from the inner diameter side to the outer diameter side of the plain bearing 8 due to the relative rotation between the thrust receiver (thrust plates 9, 10) and the plain bearing 8 when the impeller 4 rotates. This lubrication groove is formed on at least one thrust sliding surface selected from the end face on the circulating water discharge side of the plain bearing 8 and the thrust receiver (thrust plates 9, 10) that slides against said end face.

[0026] In the configuration shown in Figure 1, when rotating, the impeller 4 is pressed against the cover 7 due to the pressure difference, and one end face of the plain bearing 8 rotates and slides against the thrust plate 9. There is almost no sliding between the other end face of the plain bearing 8 and the thrust plate 10 on the casing 6 side. The direction of the water flow caused by the pressure difference is from the thrust plate 10 side to the thrust plate 9 side. In Figure 1, a lubrication groove is formed on the end face of the plain bearing 8 that slides against the thrust plate 9, and a projection of this end face viewed from the front is shown in Figure 2.

[0027] As shown in the projection view (plan view) of Figure 2, the sliding bearing 8 has, on its annular bearing end face, a land portion 13 that forms the sliding surface, and a lubrication groove 14 that connects the inner diameter surface 8a and the outer diameter surface 8b. The only part of the sliding bearing 8 other than the lubrication groove 14 is the land portion 13, and no other grooves or recesses are formed on the bearing end face. Additionally, arrow Z in Figure 2 indicates the direction of rotation of the sliding bearing 8.

[0028] As shown in Figure 2, the lubrication groove 14 has a groove bottom surface 14a (see Figure 3) that is inclined with respect to the land portion 13, and is a groove that generates dynamic pressure due to the relative rotation between the plain bearing 8 and the thrust bearing. Three lubrication grooves 14 are provided at intervals in the circumferential direction on the bearing end face. The multiple lubrication grooves 14 are preferably provided at equal intervals in the circumferential direction, and in Figure 2, the angular interval between adjacent grooves is 120°. This interval is the angle formed by the line segments A of each lubrication groove 14. The number of lubrication grooves 14 is not particularly limited. The greater the number of lubrication grooves 14, the greater the dynamic pressure effect, but since the surface pressure on the sliding surfaces between the plain bearing 8 and the thrust bearing also increases, the number is set taking into account the usage conditions, etc.

[0029] Furthermore, the extension line of line segment A of lubrication groove 14 passes through the axial center O of sliding bearing 8. In this case, the multiple lubrication grooves 14 are provided at positions that are offset downstream in the direction of rotation from center line OA that passes through line segment A of sliding bearing 8.

[0030] As shown in FIG. 2, each lubrication groove 14 is formed by a region enclosed by line segment A, which connects the inner diameter side to the outer diameter side of the bearing end surface; line segment B, which is located downstream of line segment A in the direction of rotation of the sliding bearing 8 and connects the inner diameter side to the outer diameter side of the bearing end surface; arc C, which runs along the inner diameter surface 8a; and arc D, which runs along the outer diameter surface 8b. The sliding bearing 8 in FIG. 2 is characterized in that the length of arc C is longer than the length of arc D. During rotation of the sliding bearing 8, circulating water is discharged from the inner diameter side to the outer diameter side. The tapered shape of the lubrication groove 14 in this flow direction enhances the dynamic pressure effect. The ratio of the length of arc C to the length of arc D is preferably 1 / 3<(length of arc D) / (length of arc C)<1. Note that the length of arc C may be equal to the length of arc D.

[0031] Furthermore, in the lubrication groove 14, it is preferable that the area of ​​the opening on the inner diameter side is equal to or larger than the area of ​​the opening on the outer diameter side. Note that the "area of ​​the opening on the inner diameter side" can be calculated from an arrow view when looking from the center line OA of the sliding bearing along line segment A to the inner diameter side. Furthermore, the "area of ​​the opening on the outer diameter side" can be calculated from an arrow view when looking from the center line OA along line segment A from the outer diameter side.

[0032] In FIG. 2, angle θ1 indicates the angle between line segment A and line segment B of lubrication groove 14. More specifically, it refers to the acute angle (0° or more and 90° or less) between the extension of line segment A and the extension of line segment B. As angle θ1 increases, the dynamic pressure effect improves, but the surface pressure also increases, so angle θ1 is preferably 0° to 15°. In FIG. 2, line segment A and line segment B are not parallel, and angle θ1 is greater than 0°. In this embodiment, angle θ1 is preferably 1° to 15°, more preferably 5° to 15°, and even more preferably 10° to 15°.

[0033] As will be shown in the examples below, the line segments A and B may be parallel to each other (angle θ1=0°). In this case, the length of the arc C will be longer than the length of the arc D.

[0034] Next, FIG. 3 shows a cross-sectional view taken along line aa in FIG. 2. FIG. 3 is a cross-sectional view of the lubrication groove taken along a plane perpendicular to line A. As shown in FIG. 3, the cross-sectional shape of the lubrication groove 14 is a right-angled triangle. This right-angled triangle has three vertices: an intersection v1 between line A and the cut surface, an intersection v2 between line B and the cut surface, and an intersection v3 between the groove bottom surface 14a and the groove side surface 14b. The hypotenuse of the right-angled triangle corresponds to the groove bottom surface 14a of the lubrication groove 14. The groove bottom surface 14a is an inclined surface that is inclined relative to the land portion 13, and the groove depth decreases toward the upstream side in the rotation direction Z. When the sliding bearing 8 rotates, the circulating water supplied to the lubrication groove 14 acts to be pushed toward the intersection v1. Therefore, by providing such a lubrication groove 14, dynamic pressure is generated by the circulating water.

[0035] The inclination angle θ2 of the groove bottom surface 14a with respect to the land portion 13 (which is also the interior angle with the intersection point v1 as the vertex) is preferably 3° to 30°. If the angle θ2 is less than 3°, the flow rate of circulating water passing through the lubrication groove decreases, which may result in an insufficient supply of circulating water to the sliding surfaces between the sliding bearing and the thrust bearing. If the angle θ2 exceeds 30°, the dynamic pressure effect may be insufficient. Furthermore, the angle θ2 is more preferably 5° to 20°. The angle θ2 may be constant from the inner diameter side to the outer diameter side, or may vary continuously. If the angle θ2 varies, each angle θ2 in any cross section is preferably within the range of 3° to 30°, and more preferably within the range of 5° to 20°.

[0036] The maximum depth H of the lubrication groove 14 in the axial direction of the sliding bearing 8 is preferably 0.1 to 1.0 mm. The maximum depth H is the depth from the land portion 13 to the deepest part of the lubrication groove 14. If the maximum depth H is less than 0.1 mm, the flow rate of circulating water passing through the lubrication groove 14 decreases, which may result in an insufficient supply of circulating water to the sliding surface between the land portion 13 on the end face and the thrust bearing. Furthermore, if the maximum depth H exceeds 1.0 mm, the dynamic pressure effect caused by the circulating water being forced toward the intersection point v1 may be insufficient. In FIG. 3, the deepest part of the lubrication groove 14 is formed linearly from the inner diameter side to the outer diameter side. The groove depth of the lubrication groove 14 may vary from the inner diameter side to the outer diameter side, in which case the maximum depth H is preferably in the range of 0.1 to 1.0 mm.

[0037] For example, if the maximum depth H of the lubrication groove 14 is constant from the inner diameter side to the outer diameter side, the cross-sectional shape of the lubrication groove 14 is a right triangle, and the angle θ1 (see Figure 2) is not 0°, the angle θ2 increases continuously from the inner diameter side to the outer diameter side.

[0038] While FIG. 3 shows a case where the cross-sectional shape of the lubrication groove 14 is a right-angled triangle, it may also be a substantially triangular right-angled shape. For example, the hypotenuse constituting the groove bottom surface 14a and the sides constituting the groove side surface 14b may be slightly curved as long as the effects of the present invention are not impaired. Furthermore, the interior angle at the intersection point v2 between the line segment B and the cut surface may be 70° to 90° (preferably 80° to 90°). It is preferable that the groove side surface 14b be nearly perpendicular to the land portion 13. Furthermore, the intersection point v2 may be chamfered or rounded, and the intersection point v3 may have a corner radius. The chamfering and radius of the intersection point v2 are not limited, but may be, for example, about 0.1 to 0.2° C. The radius of the corner radius of the intersection point v3 is not limited, but may be, for example, about 0.1 to 0.2° R.

[0039] Another example of a sliding bearing according to the present invention is shown in Figure 4. Similar to Figure 3, Figure 4 is a cross-sectional view of a lubrication groove taken along a plane perpendicular to line segment A. As shown in Figure 4, an inclined surface 16c with a steeper slope than groove bottom surface 16a is provided at the boundary between the land portion 15 and the upstream end of groove bottom surface 16a in the direction of rotation Z. In this case, the path of the upstream end of inclined surface 16c in the direction of rotation Z forms line segment A. In this example, the inclination angle θ3 of inclined surface 16c with respect to land portion 15 may be any angle greater than angle θ2, and is, for example, 20° to 90°, and preferably 40° to 60°. If angle θ3 is less than 20 degrees, the change in area of ​​land portion 15 will be significant when wear of land portion 15 occurs, which may prevent the desired effect from being achieved. Furthermore, R may be provided at the intersection v1' of inclined surface 16c and line segment A. The intersection v2' between the line segment B and the cutting surface may be chamfered or rounded, and the intersection v3' between the groove bottom surface 16a and the groove side surface 16b and the intersection v4' between the groove bottom surface 16a and the inclined surface 16c may have a rounded corner. The size of the rounded corner at the intersection v1' is not limited, but may be, for example, about R0.1 to 0.2. The size of the chamfered corner or rounded corner at the intersection v2' is not limited, but may be, for example, about C0.1 to 0.2, for chamfering, and about R0.1 to 0.2, for rounding. The size of the rounded corners at the intersections v3' and v4' is not limited, but may be, for example, about R0.1 to 0.2.

[0040] Next, the effect of the configuration in Fig. 4 is shown in Fig. 5. As shown in Fig. 5(a), this sliding bearing has an inclined surface 16c at the boundary between groove bottom surface 16a and land portion 15 (sliding surface) that is steeper than groove bottom surface 16a, and therefore, compared to a case in which this inclined surface is not provided (Fig. 5(b)), even if the sliding surface wears to the same extent, the increase in surface area of ​​the sliding surface is smaller, and changes in torque are suppressed.

[0041] In the present invention, the material of the sliding bearing is not particularly limited, and synthetic resin, carbon material, metal, etc. can be used. Of these, it is preferable to use synthetic resin, and from the standpoint of ease of molding, it is even more preferable to use thermoplastic resin. In particular, the sliding bearing is preferably an injection-molded article made of a resin composition containing a thermoplastic resin. During injection molding, at least one gate is disposed on the outer diameter surface of the sliding bearing, and the molten resin composition flows into the cavity from this gate. If the gate is a multi-point gate, it is preferable that the gates be spaced equally apart in the circumferential direction.

[0042] Here, the gate position will be explained using Figure 2. The sliding bearing 8 in Figure 2 is an injection-molded article that was injection-molded using three gates (three-point gates) indicated by the black arrows, and has gate marks on its outer diameter surface. The interval between the gates is set at 120° in the circumferential direction. There are no particular restrictions on the gate positions, but as shown in Figure 2, it is preferable that welds W are formed within the lubrication grooves 14 and that welds W are not formed in areas other than the lubrication grooves 14 (land portions 13). Welds W are formed near the middle between adjacent gates, and so in Figure 2, the gate positions have been adjusted so that three welds W are formed within each of the three lubrication grooves 14.

[0043] Welds are formed where the molten resin flows meet, and they may be convex compared to areas without welds. Therefore, by not forming welds on the land portion 13 on the end surface, the flatness of the land portion 13 can be reduced. Note that if the convex portion caused by the weld extends over a fairly wide area centered on the weld, it is not essential that the entire convex portion is contained within the lubrication groove. The position of the weld can be confirmed using known methods such as microscopic observation. Note that the axial position of the gate in the sliding bearing 8 is not limited, but it is preferably near the center of the axial length of the sliding bearing 8.

[0044] Furthermore, the flatness of the land portion 13 is preferably 0.08 mm or less, and more preferably 0.05 mm or less. Flatness is defined in JIS B0621-1984. Flatness can be measured by either contact measurement using a dial gauge or non-contact measurement using height information obtained by irradiating the bearing with laser light. By improving the flatness of the flat portion (land portion 13) excluding the lubrication grooves 14 on the bearing end face of the sliding bearing 8, it is possible to reduce vibrations that occur during sliding.

[0045] In addition to the end faces described above, it is preferable that grooves be formed on the radial sliding surface of the sliding bearing's inner diameter surface in the sliding bearing of the present invention. For example, linear grooves parallel to the axial direction or spiral grooves can be formed. It is also preferable that these grooves be dynamic pressure grooves. The provision of dynamic pressure grooves allows water to be forced into the closed sliding surface, making it possible to supply a large amount of water, generating a load in the anti-load direction, forming a water film, and resulting in a low friction coefficient. It is also expected that an air-cooling effect can be achieved even in abnormal conditions such as a lack of water. In the above-mentioned spiral grooves, aligning the spiral rotation direction with the rotation direction of the shaft makes it easier to generate dynamic pressure. It is also preferable that these grooves be used in combination with communicating grooves (grooves that communicate from one end face of the bearing to the other end face) and non-communicating grooves.

[0046] Below, we will explain the resin composition when the sliding bearing is an injection-molded article. The base resin of the resin composition is preferably a thermoplastic resin. There are no limitations on the type of thermoplastic resin, but from the standpoints of heat resistance and chemical resistance, it is preferable that it be an engineering plastic or super engineering plastic. Specific examples include polyphenylene sulfide (PPS) resin, polyether ether ketone (PEEK) resin, polyethersulfone resin, polyetherimide resin, polyamide resin, thermoplastic polyimide resin, and polyamideimide resin. Of these resins, PPS resin and PEEK resin are preferred, as they have particularly excellent chemical resistance and low water absorption. From an economical standpoint, PPS resin is particularly preferred. By using a resin composition that uses PPS resin as the base resin, it is possible to provide an inexpensive sliding bearing.

[0047] PPS resin is a crystalline thermoplastic resin with a polymer structure in which benzene rings are linked in the para position by sulfur bonds. PPS resin has a melting point of approximately 280°C and a glass transition temperature of 93°C, and possesses extremely high rigidity and excellent heat resistance, dimensional stability, and abrasion resistance. PPS resins are classified into crosslinked, semi-crosslinked, linear, and branched types depending on their molecular structure, but the present invention can use any type of PPS resin without being limited by these molecular structures or molecular weights.

[0048] PEEK resin is a crystalline thermoplastic resin with a polymer structure in which benzene rings are linked in the para position by carbonyl groups and ether bonds. PEEK resin has a melting point of approximately 340°C and a glass transition point of 143°C, and has excellent heat resistance, creep resistance, load resistance, abrasion resistance, sliding properties, fatigue properties, and other properties, as well as excellent moldability.

[0049] In the above resin composition, it is preferable to blend polytetrafluoroethylene (PTFE) resin to impart frictional properties in a water-free state where no water film is formed. It is also preferable to blend graphite to impart frictional properties in circulating water. Furthermore, graphite also has the effect of improving the dimensional accuracy of the sliding bearing during injection molding. The inclusion of graphite can also reduce the flatness of the land portion.

[0050] The PTFE resin may be a molding powder produced by suspension polymerization, a fine powder produced by emulsion polymerization, or recycled PTFE. To stabilize the fluidity of the resin composition, recycled PTFE is preferred, as it is less likely to become fibrous due to shear during molding and to increase the melt viscosity. Recycled PTFE refers to powders that have undergone heat treatment (thermal history) or have been irradiated with gamma rays or electron beams. Examples of recycled PTFE include powders obtained by heat-treating molding powder or fine powder, powders obtained by further irradiating this powder with gamma rays or electron beams, powders obtained by pulverizing a molding powder or fine powder compact, powders obtained by subsequently irradiating gamma rays or electron beams, and powders obtained by irradiating molding powder or fine powder with gamma rays or electron beams.

[0051] Commercially available PTFE resins that can be used in the present invention include Kitamura's KTL-610, KTL-450, KTL-350, KTL-8N, and KTL-400H, Mitsui-Chemours Fluoroproducts' Teflon® 7-J and TLP-10, AGC's Fluon G163, L150J, L169J, L170J, L172J, and L173J, Daikin Industries, Ltd.'s Polyflon M-15 and Lubron L-5, and 3M Japan's Dyneon TF9205 and TF9207. PTFE resins modified with perfluoroalkyl ether groups, fluoroalkyl groups, or other fluoroalkyl-containing side chain groups may also be used. Among the above, examples of PTFE resins irradiated with gamma rays or electron beams include Kitamura Co., Ltd.'s KTL-610, KTL-450, KTL-350, KTL-8N, and KTL-8F, and AGC Co., Ltd.'s Fluon L169J, L170J, L172J, and L173J.

[0052] The graphite may be either natural or artificial. The particle shape may be flaky or spherical, with flaky being more preferred as it is less likely to fall off during sliding. Examples of natural graphite include ACP manufactured by Nippon Graphite Industries Co., Ltd., and examples of artificial graphite include KS-6, KS-25, and KS-44 manufactured by Imerys GC Japan.

[0053] It is preferable to blend carbon fiber into the resin composition to improve the rigidity, wear resistance, and dimensional accuracy of the sliding bearing. The carbon fiber may be either pitch-based or PAN-based, as classified based on the raw material. The average fiber diameter of the carbon fiber is 20 μm or less, preferably 5 μm to 15 μm. Carbon fibers thicker than this range are prone to generating extreme pressure, are less effective in improving load resistance, and, when the mating material, such as a rotating shaft or thrust bearing, is stainless steel, may be subject to significant wear and tear.

[0054] The carbon fibers may be either chopped fibers or milled fibers, but milled fibers with a fiber length of less than 1 mm are preferred, and more preferably, the average fiber length is 20 μm to 200 μm. If the fiber length is less than 20 μm, it is difficult to obtain sufficient rigidity and reinforcing effect, and wear resistance may be poor. If the fiber length exceeds 200 μm, extreme pressure is likely to be generated, and if the mating material, such as a rotating shaft or thrust bearing, is stainless steel, wear damage to the mating material may be significant. The average fiber diameter can be measured using an electron microscope or atomic force microscope commonly used in this field. The average fiber diameter can also be calculated as the number-average fiber diameter based on the above measurement.

[0055] Commercially available milled fibers that can be used in the present invention include pitch-based carbon fibers such as Kureha Corp.'s KUREKA M-101S, M-101F, and M-201S, Mitsubishi Chemical Corp.'s DIALEAD K223HM-200 μm and DIALEAD K223HM-50 μm, and Nippon Graphite Fiber Co., Ltd.'s HC-600-15M. Similar PAN-based carbon fibers include Toho Tenax Corp.'s BESPFITE HT M100 160MU and HT M100 40MU, and Toray Industries, Inc.'s TORAYCA MLD-30 and MLD-300. Examples of chopped fibers include pitch-based carbon fibers such as Mitsubishi Plastics' DIALEAD K223HE, and Toray Industries, Inc.'s TORAYCA T010-003.

[0056] Well-known resin additives may be added to the resin composition to the extent that they do not impair the effects of the present invention. Examples of such additives include friction property improvers such as boron nitride, molybdenum disulfide, and tungsten disulfide, and colorants such as carbon powder, iron oxide, and titanium oxide.

[0057] The blending ratio of the resin composition is preferably 3 to 30 volume %, more preferably 5 to 20 volume %, of at least PTFE resin and / or graphite, 5 to 30 volume %, more preferably 10 to 20 volume %, of carbon fiber, and the remainder is the base resin.Furthermore, it is preferable to use PTFE resin and graphite in combination.

[0058] The materials that make up the resin composition can be mixed as needed using a Henschel mixer, ball mixer, ribbon blender, or the like, and then melt-kneaded in a melt extruder such as a twin-screw kneading extruder to obtain molding pellets. Additives may be added using a side feed when melt-kneading in a twin-screw kneading extruder or the like. These molding pellets are then injection molded into a sliding bearing.

[0059] In the sliding bearing device of the present invention, the material of the thrust bearing is not particularly limited, but metal is preferably used, and stainless steel is more preferably used. Furthermore, the surface of the thrust bearing may be coated with a known coating such as diamond-like carbon (DLC) or a resin coating.

[0060] In thrust bearings, it is preferable that the direction of the machined grooves on the surface that slides against the plain bearing is not concentric. To achieve this, it is preferable that this surface be formed by a processing method other than turning. If this surface is formed by turning, the direction along the grooves (circumferential direction) and the lubrication grooves of the plain bearing will be perpendicular, making it difficult for wear powder to be discharged to the outer diameter side of the plain bearing when the plain bearing slides against the thrust bearing. As a result, wear powder may become trapped in the sliding surface, causing large fluctuations in the friction coefficient.

[0061] A specific processing method other than turning is to manufacture a thrust bearing by pressing a metal steel plate. This method is particularly preferable because the direction of the grain is random. Here, "random grain direction" refers to a state in which the grains are not aligned in a single direction, but are oriented in various directions. Other processing methods, such as surface grinding, that result in unidirectional grains are also preferable because they are less likely to cause wear particles to get caught in the sliding surface.

[0062] 1 to 5, lubrication grooves for discharging liquid and generating dynamic pressure are provided on the bearing end surface of the sliding bearing, but instead of or in addition to this configuration, similar lubrication grooves may be provided on the sliding surface of the thrust bearing. Furthermore, the circulating water in a water pump is not limited to water, and antifreeze solution, chemical liquids, etc. can also be suitably used. [Example]

[0063] The cylindrical test pieces (inner diameter 10 mm, outer diameter 17 mm, height 13 mm) used in Examples 1 to 13 and Comparative Examples 1 and 2 were produced by injection molding using a resin composition with PPS resin as the base resin. The resin composition contained 5 volume % PTFE resin, 15 volume % graphite, 10 volume % carbon fiber, and the remainder PPS resin. Details of each material are as follows: [PPS resin] Tosoh: B-042 [PTFE resin] Kitamura: KTL-610 (recycled PTFE) [Carbon fiber] Kureha Corporation: Kureka M107T (average fiber length 0.4 mm, average fiber diameter 18 μm) [Graphite] Imerys GC Japan: KS-25 (artificial graphite, flakes)

[0064] FIG. 6 shows a projection of the end face of the cylindrical test piece used in Examples 1 to 4, viewed from the front. In the cylindrical test piece 17 in FIG. 6, the cross-sectional shape of the lubrication groove cut along any plane perpendicular to line segment A is a right-angled triangle, and the maximum depth of the lubrication groove is constant regardless of the radial position. Furthermore, the length of arc C is longer than the length of arc D. In Examples 1 to 4, since angle θ1 is not 0°, the groove width of the lubrication groove narrows as it approaches the outer diameter side, and angle θ2 (see FIG. 3) gradually increases from the inner diameter side to the outer diameter side. In Examples 1 to 4, the vertical length L of line segment A between the inner diameter side end of line segment A and the inner diameter side end of line segment B was fixed at 1.70 mm.

[0065] Fig. 7 shows a projection view of the end face of the cylindrical test piece used in Examples 5 to 13, viewed from the front. In the cylindrical test piece 18 in Fig. 7, the cross-sectional shape of the lubrication groove cut by any plane perpendicular to line segment A is a right triangle, and the maximum depth of the lubrication groove is constant regardless of the radial position. In addition, the length of arc C is longer than the length of arc D. In Examples 5 to 13, line segment A and line segment B are parallel (angle θ1 = 0°), and angle θ2 is constant from the inner diameter side to the outer diameter side. In addition, length L was changed in Examples 5 to 13.

[0066] Fig. 8 shows a projection of the end face of the cylindrical test piece used in Comparative Example 1, viewed from the front. In the cylindrical test piece 19 of Fig. 8, the cross-section of the lubrication groove cut by any plane perpendicular to line segment A is a right-angled triangle, and the maximum depth of the lubrication groove is constant regardless of the radial position. In addition, the length of arc C is shorter than the length of arc D. Therefore, the groove width of the lubrication groove increases as it approaches the outer diameter side, and angle θ2 gradually decreases.

[0067] Fig. 9 shows a projection view of the end face of the cylindrical test piece used in Comparative Example 2, viewed from the front. The cylindrical test piece 20 in Fig. 9 does not have a lubrication groove provided on the end face.

[0068] The flatness of the land portion of each cylindrical test piece was measured using a dial gauge (in accordance with JIS B0621-1984). The land portion was also observed using an optical microscope for the presence or absence of welds.

[0069] The dynamic friction coefficients in antifreeze (50% ethylene glycol, 50% water) were measured using a ring-on-disc testing machine with the cylindrical test specimens and a disk-shaped mating member (SUS304). In Examples 1 to 11 and Comparative Examples 1 and 2, the test conditions were a speed of 125 m / min, a load of 38 N, and a temperature of 30°C. The surface of the disk-shaped mating member that slides against the cylindrical test specimen was finished by flat grinding. The dimensions and test results of each cylindrical test specimen are shown in Tables 1 and 2. For the cylindrical test specimens shown in Table 1, the angle θ2 changes continuously in the radial direction, so the minimum and maximum values ​​are shown.

[0070]

Table 1

[0071]

Table 2

[0072] As shown in Table 1, in Example 1 (θ2 = 10 - 16°), Example 2 (θ2 = 10 - 23°), and Example 4 (θ2 = 10 - 43°), the minimum value of the angle θ2 is 10° and the maximum depth is 0.3 mm for all, but the setting of the angle θ1 is different. Example 1 (θ1 = 10°) and Example 2 (θ1 = 15°) had lower friction than Example 4 (θ1 = 20°). Also, Example 3 (θ2 = 16 - 36°, maximum depth = 0.5 mm) had a larger θ2 and a higher dynamic friction coefficient than Example 1 and Example 2. This is considered to be because the dynamic pressure generation effect decreased as θ2 increased. On the other hand, although θ2 is small in Comparative Example 1 (θ2 = 7 - 10°), the length of arc C is shorter than the length of arc D (C < D), and the dynamic pressure generation effect decreased due to the groove width widening toward the outer diameter side. As a result, the dynamic friction coefficient of Comparative Example 1 was higher than that of Example 1 - 2 and Example 5 - 11.

[0073] Also, as shown in Table 2, in Examples 5 - 11, line segment A and line segment B are parallel to each other, and the angle θ2 and the maximum depth are different. Examples 5 - 6 and Examples 8 - 11 with θ2 of 5 - 30° and a maximum depth of 0.1 - 0.6 mm had a dynamic friction coefficient of 0.031 - 0.043 and were low - friction. On the other hand, the dynamic friction coefficient of Example 7 with a flatness of 0.09 mm was 0.051, showing a slightly higher value. Also, Comparative Example 2 had no lubrication groove on the end face, and the dynamic friction coefficient was 0.066, which was higher than any of the examples. Also, in Examples 5 and Examples 8 - 11, there were lubrication grooves near the middle of adjacent gates, and weld lines were formed in the lubrication grooves. On the other hand, in Examples 6 and Example 7, there were lubrication grooves at positions away from the middle of adjacent gates, and weld lines were formed on the land portion.

[0074] Next, in Examples 12 and 13, cylindrical test pieces of the same shape shown in Table 3 were used to measure the variation in the dynamic friction coefficient due to differences in the processing method of the disc-shaped mating member. In Example 12, the surface of the disc-shaped mating member that slides against the cylindrical test piece was finished by flat grinding. After this processing, the streaks on the surface were generally aligned in one direction. In Example 13, the surface of the disc-shaped mating member that slides against the cylindrical test piece was finished by turning. After this processing, the streaks on the surface were formed in a concentric pattern. The test conditions were the same as in Examples 1 to 11 and Comparative Examples 1 and 2, and the variation in the dynamic friction coefficient over 600 seconds is shown in Figure 10 (Example 12) and Figure 11 (Example 13), respectively.

[0075] [Table 3]

[0076] 10 and 11, the dynamic friction coefficient fluctuates little in Fig. 10 (Example 12), where the disc-shaped mating material is surface-ground, whereas an increase in the dynamic friction coefficient due to the incorporation of wear particles is observed around 160 seconds in Fig. 11 (Example 13), where the disc-shaped mating material is turned. In both cases, the dynamic friction coefficient showed low values. [Industrial Applicability]

[0077] The sliding bearing device of the present invention has low friction characteristics and is extremely quiet, making it suitable for use as a sliding bearing device in water pumps that circulate coolant for automobile engines, inverters, batteries, fuel cells, etc., or that circulate hot water for water heaters, floor heating equipment, etc. Note that the sliding bearing device of the present invention is not limited to use in water pumps that circulate water, but is also useful as a pump that moves and supplies water. Similar effects can be expected in pumps that circulate, move, and supply liquids other than water, such as chemicals, solvents, oils, and beverages. [Explanation of symbols]

[0078] 1 water pump 2 windings 3. Permanent magnets 4 impeller 5-axis 6 Casing 7 Cover 8, 8' plain bearing 9 Thrust plate 10 Thrust plate 11 Packing 12 motors 13 Land Department 14 Lubrication groove 15 Land Department 16 Lubrication groove 17 Cylindrical test piece 18 Cylindrical test specimen W weld part

Claims

1. a cylindrical sliding bearing fixed to the impeller for supporting the impeller so as to rotatably support the impeller relative to the shaft; thrust bearings that slide against each end face of the sliding bearing; and a casing and cover that house the impeller and form a pump chamber, the sliding bearing being used in a water pump that draws in and discharges circulating water via the pump chamber as the impeller rotates, the water pump sliding bearing is provided, on at least one end face, with a land portion that serves as a sliding surface and a lubrication groove that discharges the circulating water from the inner diameter side to the outer diameter side, In a projection view of the end face on which the lubrication groove is provided, viewed from the front, the lubrication groove comprises an area surrounded by line segments A and B, which are connected by straight lines from the inner diameter side to the outer diameter side of the end face, an arc C along the inner diameter surface, and an arc D along the outer diameter surface, and line segment A is located upstream of line segment B in the direction of rotation of the sliding bearing, The length of the arc C is equal to or longer than the length of the arc D, a groove bottom surface of the lubrication groove that is an inclined surface inclined at an angle of 3° to 30° with respect to the land portion, and the groove depth becomes shallower toward the line segment A.

2. The sliding bearing for a water pump according to claim 1, characterized in that the angle formed by the line segment A and the line segment B is between 0° and 15°.

3. 3. The sliding bearing for a water pump according to claim 1, wherein the cross-sectional shape of the lubrication groove taken along any cut surface perpendicular to the line segment A is a substantially right-angled triangle with the inclined surface as its hypotenuse.

4. The sliding bearing for a water pump according to any one of claims 1 to 3, characterized in that the maximum depth of the lubrication groove is 0.1 mm to 1.0 mm.

5. 5. The sliding bearing for a water pump according to claim 1, wherein a plurality of the lubrication grooves are arranged at intervals in the circumferential direction on the end face.

6. the sliding bearing has an axial center on an extension of the line segment A, The sliding bearing for a water pump according to claim 5, characterized in that the plurality of lubrication grooves are provided at positions that are offset downstream in the direction of rotation of the sliding bearing from a center line of the sliding bearing that passes through the line segment A.

7. the sliding bearing is an injection-molded article of a resin composition, and has the lubrication groove on at least one end face thereof, 7. The sliding bearing for a water pump according to claim 1, characterized in that gate marks are formed on the outer diameter surface of the sliding bearing, and in the sliding bearing, a weld is formed within the lubrication groove, and no weld is formed in the land portion.

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