GB2604254A

GB2604254A - Bearing for ship propulsion shaft

Info

Publication number: GB2604254A
Application number: GB2205429.0A
Authority: GB
Inventors: Kachu Yoshimasa; Yomo Masataka; Harada Kouhei; Yokogaki Kenji
Original assignee: Mikasa Corp
Current assignee: Mikasa Corp
Priority date: 2020-06-25
Filing date: 2020-11-20
Publication date: 2022-08-31
Anticipated expiration: 2040-11-20
Also published as: GB2604254B; GB2604254A9; CN114630969B; CN114630969A; JPWO2021260965A1; KR102400231B9; KR102400231B1; DE112020003968T5; GB202205429D0; WO2021260965A1; KR20220044618A; JP7343940B2

Abstract

Provided is a bearing for supporting the propulsion shaft of a ship, with which it is possible to form a water coating with excellent lubricating characteristics on a sliding surface between the bearing and the propulsion shaft, and which is excellent in wear resistance and is economical.　A bearing according to the present invention supports the propulsion shaft of a ship and includes, on an inner peripheral surface of a cylindrical shell, a sliding layer and an elastic layer in this order in the radial direction from the cylinder center, wherein the sliding layer has: a load action portion which is on the lower surface side of the inner peripheral surface of the shell and on which a load of the propulsion shaft acts, the center angle θ thereof being in a range of θ=60°-180° with the axis of symmetry being the vertical axis; and a cooling action portion which opposes the load action portion and is provided with a plurality of required water cooling grooves on both ends being the first and last. The vertical axis refers to the axis of symmetry of a lateral cross section of the shell parallel to the center of gravity direction of the propulsion shaft.

Description

[Title of Invention] BEARING FOR PROPULSION SHAFT OF SHIP

[Technical Field]

[0001] The present invention relates to a bearing that supports a propulsion shaft of a ship and particularly to a bearing that utilizes the lubrication action of a water film formed on a sliding surface with the propulsion shaft.

[Background Art]

[0002] Bearings such as shaft bracket bearings and stern tube bearings for supporting the propulsion shaft of a ship are exposed to water such as seawater, and the corrosion protection and tight seal structures are therefore important. Considering the environmental suitability and economic efficiency, rather than metal bearings or bearings that use lubricant oil, bearings that utilize the lubrication action of a water film formed on the sliding surface with the propulsion shaft are preferred. Various bearings use rubber, fluorine resin, etc. for the sliding surface with the propulsion shaft. Various bearings that utilize the lubrication action of such a water film have been proposed.

[0003] For example, Patent Document 1 proposes a shaft bracket bearing for the propulsion shaft of a ship. In this shaft bracket bearing, the most loaded portions of the propulsion shaft are supported by smooth bearing surfaces at both end portions of the bearing, while for sufficient cooling, the central portion which receives a reduced load is provided with water supply grooves that are formed on the upper inner surface of the bearing main body and the lower inner surface of the central portion, one or more circumferential grooves that connect the water supply grooves in the circumferential direction, and second circumferential grooves that connect both opening portions of the water supply grooves in the circumferential direction. It is said that this shaft bracket bearing can reduce the bearing wear and prevent the seizing because the most loaded portions of the propulsion shaft are supported by the smooth bearing surfaces at both end portions of the bearing, the central portion which receives a reduced load is sufficiently cooled, and the circumferential grooves and the second circumferential grooves allow a thin film of seawater to be formed.

[0004] Patent Document 2 proposes an electrolytic protection bearing used for the propulsion shaft of a ship or the like. Proposed method for manufacturing the electrolytic protection bearing includes inserting a rubber material between a metal base and a pad material composed of a synthetic resin excellent in the slidability, subjecting them to vulcanized adhesion to form a segment-like rubber bearing material, fitting a plurality of the segment-like rubber bearing materials into grooves that are carved side by side in the inner surface axial direction of a cylindrical shell metal or arranging a plurality of the segment-like rubber bearing materials in a barrel-like shape along the circumferential direction of the inner surface of a metal, and adhering the bottom surface of the metal base to the inner surface.

In the above rubber bearing material, ethylene tetratluoride, polyamide, high-density polyethylene, or the like can be used as the pad material, and it is said that an underwater rubber bearing having high durability and reliability can be obtained.

[0005] Patent Document 3 proposes a split bearing for supporting the propulsion shaft of a ship. The split bearing includes a plurality of bearing members that are composed of reinforcing materials fixed along the inner surface of a cylindrical shell and sliding materials on which the propulsion shaft slides The sliding materials on the approximately upper half of the shell are composed of a rubber material, and the sliding materials on the approximately lower half of the shell are composed of a fluorine resin material. It is said that this split bearing has a small sliding resistance with the propulsion shaft and can reduce the fuel consumption of the ship.

[0006] Patent Document 4 proposes a water lubrication-type bearing material used for the bearing material in a ship stern tube. More specifically, the proposed water lubrication-type bearing material contains 12 wt% to 25 wt9/0 of a tetrafluoroethylene-perfluoroalkyl vinyl ether copolymer resin (PFA resin), 18 wt% to 33 wt% of a carbon fiber, and the balance including a polytetrafluoroethylene (PTFE) resin and/or a modified PTFE resin. It is said that this water lubrication-type bearing material is excellent in the water resistance as well as the wear resistance and suitable as a bearing material for water lubrication-type bearings.

[Prior Art Document]

[Patent Document] [0007] [Patent Document I] JP2004-66977A [Patent Document 21 JP2009-103307A [Patent Document 31 Registered utility model No. 3183964 [Patent Document 4] W02016/114244 [Problems to be solved by Invention] [0008] Patent Document 1 describes a shaft bracket bearing in which the function of the bearing is divided into load supporting portions and a seizing prevention portion with consideration for the fact that the load applied to the shaft bracket bearing is not always uniform, and the cooling water grooves are arranged in a skilled way so as to form a water film. Unfortunately, this shaft bracket bearing has a problem in that the configuration of the bearing is complicated because the bearing may have to be configured in accordance with the load. On the other hand, Patent Document 2 describes a method for manufacturing a rubber bearing material of an electrolytic protection bearing and also describes rubber bearing materials having various structures. However, there is no specific description about the characteristics of the rubber bearing materials having these various structures.

[0009] Patent Document 3 describes a split bearing having a specific configuration of the rubber bearing materials and a specific arrangement form of the materials on the inner surface of the cylindrical shell metal, which are similar to those described in Patent Document 2. As for the sliding surfaces of the bearing members arranged on the inner surface of the cylindrical shell, those on the inner surface of the approximately upper half of the cylindrical shell are made of a rubber material, and those on the inner surface of the approximately lower half are made of a fluorine resin material. Patent Document 4 describes the composition of a fluorine resin suitable as a bearing material for a stern tube of a ship. However, even in such a bearing in which the bearing materials on the inner surface of the approximately lower half of the cylindrical shell are made of a fluorine resin, there is a problem in that the wear progresses under uneven surface pressure.

[0010] In view of such conventional problems and requirements, an object of the present invention is to provide, in bearings for supporting the propulsion shaft of a ship, an economical bearing that can form a water film having excellent lubricity on the sliding surface with the propulsion shaft and has low frictional properties and excellent wear resistance and durability.

[Means for solving problems] [0011] The present invention provides a bearing comprising: a cylindrical shell having an inner surface and supporting a propulsion shaft of a ship; a pair of positioning plates arranged on the inner surface of the shell; and a plurality of closed arc pieces and a plurality of gap forming arc pieces. The closed arc pieces and the gap forming arc pieces are arranged on the inner surface of the shell. The positioning plates are fixed at positions facing each other on a horizontal axis of the shell. The horizontal axis is a symmetry axis of a cross section of the shell orthogonal to a direction of a gravity center of the propulsion shaft. The closed arc pieces are arranged on a lower surface side of the shell and subjected to a load of the propulsion shaft. Each of the closed arc pieces comprises a three-layer structure. The three-layer structure comprises: a sliding layer; an intermediate layer made of an elastic body; and a base in close contact with the inner surface of the shell. The sliding layer, the intermediate layer, and the base are pressed against adjacent ones between the positioning plates so as to be fastened on the inner surface of the shell. The gap forming arc pieces are arranged opposite to the closed arc pieces. Each of the gap forming arc pieces comprises a three-layer structure and has groove forming portions on both side edge portions. The three-layer structure comprises: a sliding layer; an intermediate layer made of an elastic body; and a base in close contact with the inner surface of the shell. The intermediate layer and the base are pressed against adjacent ones between the positioning plates so as to be fastened on the inner surface of the shell and so that the groove forming portions form a groove through which cooling water flows.

[0012] In the above invention, the closed arc pieces may cover at least the lower surface of the shell corresponding to a central angle of 600 or more and not exceeding 1750 [0013] Additionally or alternatively, the sliding layers of the closed arc pieces and gap forming arc pieces may each be made of a synthetic high polymer compound containing fluorine atoms in a molecule, a polyamide resin, or a phenol resin.

[0014] Additionally or alternatively, the intermediate layers of the closed arc pieces and gap forming arc pieces made of an elastic body having specified durometer hardness from A50 to A90 may be preferably adhered to the base.

[0015] Another aspect of the present invention provides a bearing for supporting a propulsion shaft of a ship. The bearing comprises: a cylindrical shell having an inner surface; and two elements of an elastic body layer and a sliding layer fixed firmly in this order on the inner surface of the shell. The sliding layer has a load acting portion without cooling water groove and a cooling action portion with cooling water grooves. The load acting portion is symmetrical about the vertical axis. The vertical axis is a symmetry axis of a cross section of the shell parallel to a direction of a gravity center of the propulsion shaft. The load acting portion is located on a lower surface side of the inner surface of the shell and subjected to a load of the propulsion shaft. The load covers at least the inner surface of the shell corresponding to a central angle of 800 or more and not exceeding 180°. The cooling action portion arranged opposite to the load acting portion is provided with a plurality of required cooling water grooves including a first and last cooling water grooves provided at both end portions of the load acting portion.

[Effect of Invention] [0016] According to the present invention, it is possible to provide, in bearings for supporting the propulsion shaft of a ship, an economical bearing that can form a water film having excellent lubricity on the sliding surface with the propulsion shaft and has low frictional properties and excellent wear resistance and durability.

[Brief Description of Drawings]

[0017] FIG. 1 is a set of cross-sectional views each illustrating the configuration of a lining-type bearing according to the present invention.

FIG. 2 is a set of explanatory diagrams illustrating the configurations of a gap forming arc piece, a closed arc piece, and a closed arc piece having a notch, which are arranged in the lining-type bearing illustrated in FIG. I. FIG. 3 is a cross-sectional view illustrating the configuration of an integrated-type bearing according to the present invention.

FIG. 4 is a graph illustrating the amount of change in the bearing inner diameter after a wear test.

FIG. 5 is a set of explanatory diagrams each illustrating the cross-sectional shape of a bearing subjected to the wear test.

FIG. 6 is an explanatory diagram illustrating the configuration of a wear tester.

FIG. 7 is a graph illustrating surface pressure distributions on sliding surfaces of bearings having various bearing members.

FIG. 8 is a graph illustrating the relationship between a circumferential speed and a friction coefficient in the wear test.

FIG. 9 is a graph illustrating the relationship between the circumferential speed and the friction coefficient when the surface pressure in the wear test is changed variously.

FIG. 10 is a set of graphs each illustrating a temperature change of a bearing material in the wear test.

FIG. ii is a set of graphs each illustrating the results of a vibration analysis test using an FFT analyzer for an invention example (FRB2) and a comparative example 10 (FRB16).

FIG. 12 is a set of graphs each illustrating the results of a vibration analysis test using an FFT analyzer for an invention example (FRB2) and a comparative example (FB2).

FIG. 13 is a graph illustrating the relationship between a water film pressure strength and a contact surface angle.

FIG. 14 is a graph illustrating the water film pressure strength curve that exhibits the relationship between a water film pressure strength and a circumferential speed at the cylindrical shape of the load acting portion.

[Mode(s) for Carrying out the Invention] [0018] Hereinafter, one or more embodiments of the present invention will be described with reference to the drawings. FIG. 1 illustrates the configuration of a bearing according to the present invention. A bearing 10 according to the present invention is a bearing that supports a propulsion shaft of a ship. The bearing 10 has a pair of positioning plates 12 that are arranged on the inner surface of a cylindrical shell 11 and fixed at positions facing each other on the horizontal axis of the shell, closed arc pieces 16 that are arranged on the lower surface side of the shell inner surface, and gap forming arc pieces 15 that are arranged opposite to the closed arc pieces 16. Adjacent gap forming arc pieces 15 or adjacent closed arc pieces 16 are held on the inner surface of the shell 11 between the positioning plates 12.

Such a bearing 10 is called a lining-type bearing because of its configuration. The horizontal axis of the shell 11 refers to a symmetry axis AA of the cross section of the shell 11 orthogonal to a direction of the gravity center of the propulsion shaft. As illustrated in FIG. 1(a), the bearing 10 can be configured such that the gap forming arc pieces 15 and the closed arc pieces 16 are arranged symmetrically with respect to the positioning plates 12, respectively, on the upper surface side and the lower surface side of the inner surface of the shell 11 Alternatively, as illustrated in FIG. 1(b), the bearing 10 can be configured such that the gap forming arc pieces 15 exceed the positioning plates 12 and are in contact with both end portions of the closed arc pieces 16 arranged on the lower surface side of the inner surface of the shell 11. In the latter case, the cooling water flowing through the bearing 10 increases.

[0019] The positioning plates 12 are fixed to the inner surface of the shell 11, for example, by using bolts or pins. The positioning plates 12 hold the closed arc pieces 16 or the gap forming arc pieces 15, which are arranged on the inner surface of the shell 11, so that the closed arc pieces 16 or the gap forming arc pieces 15 are interposed between the positioning plates 12 and in close contact with the inner surface of the shell 11. To reliably hold the closed arc piece 16 or the gap forming arc piece 15 on the inner surface of the shell 11, for example, it is preferred to fit the closed arc pieces 16 or the gap forming arc pieces 15 between the positioning plates 12 after chamfering the side edge surfaces of the closed arc pieces 16 or the gap forming arc pieces 15 which are in contact with the positioning plates 12. This allows each of the closed arc pieces 16 or gap forming arc pieces 15 to be held on the inner surface of the shell 11 in a predetermined pressing state.

[0020] The back surface side of each positioning plate 12 (the inner surface side of the shell 11) may be in an arc-like shape or a flat shape, and in the case of the flat shape, the shape forming is easy. The material of the positioning plates 12 may be metal or resin. For example, in the case of metal, a copper alloy is used, and in the case of resin, for example, a carbon fiber reinforced phenol resin is used. In the case of metal, it is preferred to cover the surface with a corrosion-resistant rubber.

[0021] As illustrated in FIG. 2(a), each gap forming arc piece 15 has a three-layer structure of a sliding layer 153, an intermediate layer 152, and a base 151 and is provided with groove forming portions 155 on both side edge portions. The intermediate layer 152 and the base 151 are pressed against corresponding ones between the positioning plates 12 and held on the inner surface of the shell 11, and grooves 18 (FIG. 1) are formed between adjacent groove forming portions 155 to flow water or seawater (cooling water). Half grooves 19 are formed at portions at which the gap forming arc pieces 15 are in contact with the closed arc pieces 16. The groove forming portions 155 are provided such that the depth thereof extends into the sliding layer 153 or from the sliding layer 153 to the intermediate layer 152. The depth and width dimensions of the groove forming portions 155 are selected such that the amount of cooling water Q required for cooling the bearing 10 is supplied. The above gap forming arc pieces 15 can be formed by using the same method and material as those for the closed arc pieces 16, which will be described below.

[0022] As illustrated in FIG. 2(b), each closed arc piece 16 has a three-layer structure of a sliding layer 163, an intermediate layer 162, and a base 161. This three-layer structure is formed, for example, as follows. That is, first, a flat plate-shaped three-layer structure is formed to be a predetermined width, then inclined surfaces at both side edge portions are formed, and the back surface of the base 161 and the front surface of the sliding layer 163 are formed into respective arc shapes so as to form a predetermined arc piece shape. The surface (sliding surface) of the sliding layer 163 is preferably smooth because it is a portion on which a water film is formed for water lubrication. From this point of view, even fine scratches, irregularities, etc. that do not adversely affect the formation of the water film are permissible. The three-layer structure can be formed by adhesion of the sliding layer 163, the intermediate layer 162, and the base 161. The closed arc pieces in contact with the positioning plates 12 preferably have respective notches 175 at the side edge portions in contact with the positioning plates 12 as in a notched closed arc piece 17 (FIG. 2(c)). Such a notch can prevent the end surface portion of the corresponding sliding layer 163 from rising. Moreover, such a notch is preferred for forming a water film and also preferred for preventing the stress concentration and wear of the sliding layer 163.

[0023] The base 161 can be made of metal or resin. As the base 161 made of metal, for example, a copper alloy having good machinability and corrosion resistance can be used. As the base 161 made of resin, a fiber-reinforced thermosetting resin, for example, a carbon fiber-reinforced phenol resin, can be used.

[0024] The intermediate layer 162 is preferably made of an elastic body having durometer hardness from A50 to A90. For example, nitrile rubber (NBR) having durometer hardness from A50 to A90 can be used. When the intermediate layer is rubber, the three-layer structure is preferably formed by vulcanized adhesion. This allows a strong adhesive structure to be formed. In the present invention, the intermediate layer 162 has an important action/function. As will be described below, the intermediate layer 162 serves to uniformize the load from the propulsion shaft, suppress heat generation on the sliding surface of the sliding layer 163 with the propulsion shaft, and improve the wear resistance and durability of the closed arc pieces 16 or the gap forming arc pieces 15.

[0025] Moreover, the closed arc pieces 16 are evenly arranged on the circumference without making gaps, and the elastic deformation of the intermediate layer is thus limited to the vertical direction when viewed from the axis of the propulsion shaft. Therefore, there is no deformation in the shearing direction with respect to the bonded surface between the sliding layer and the intermediate elastic layer. Deterioration of the adhesion portion between the sliding layer and the intermediate elastic layer is thus suppressed. [0026] The sliding layer 163 is preferably made of a synthetic high polymer compound containing fluorine atoms (F) in a molecule from the viewpoints of low frictional properties, wear resistance, and heat resistance. For example, a fluorine-based resin such as ethylene tetrafluoride (PTFE) resin, ethylene tetrafluoride/propylene hexafluoride copolymer (FEP) resin, or ethylene tetrafluoride/perfluoroalkoxy ethylene copolymer (PFA) resin can be used. Alternatively, a polyamide resin or a phenol resin can be used.

[0027] The processing applied to the arc piece shape of the closed arc pieces 16 is performed as follows. For example, the front surface of the sliding layer 163 and the back surface of the base 161 are processed into cylindrical shape surfaces based on an arc (shape forming arc) having a radius R centered on an axis 0 illustrated in FIG. 1(a). Then, both side edges of the closed arc pieces 16 are also processed based on the axis 0 so that adjacent closed arc pieces 16 are in close contact with each other. However, the surface of the sliding layer 163 can be processed so that the center of the forming arc is located at an upward-shifted axis 0' (radius R') rather than the axis 0. The upward-shifting is preferred in order to make the load on the sliding layer 163 from the propulsion shaft uniform. The upward-shifting processing may be performed by processing the cylindrical shape surface of the shell 11 itself In this case, the upward-shifting processing can be performed efficiently.

[0028] The upward-shifting is performed by shifting the center of inner circle of the shell 11 upward by a specified distance from the center of outer circle of the shell 11. For example, when the inner diameter of the bearing is 360 to 449 mm, the predetermined clearance between the propulsion shaft diameter and the bearing inner diameter is set to 0.8 to 1.3 mm, and the upward-shifting is set to A=0.53 mm (FIG. 1(a)).

[0029] The above description is directed to the bearing 10 (lining-type bearing) in which the gap forming arc pieces 15 and the closed arc pieces 16 are bedded on the inner surface of the shell 11. This bearing 10 is preferred as a bearing that supports a large-diameter propulsion shaft. It is also preferred that the gap forming arc pieces 15 or the closed arc pieces 16 can be easily replaced as needed. On the other hand, the integrated-type bearing, which will be described below, enables efficient and economical shape forming, or molding, of the bearing and is preferred particularly as the bearing of a small ship.

[0030] This integrated-type bearing is a bearing for supporting a propulsion shaft of a ship. The bearing comprises: a cylindrical shell having an inner surface; and two elements of an elastic body layer and a sliding layer fixed firmly in this order on the inner surface of the shell. The sliding layer has a load acting portion without cooling water groove and a cooling action portion with cooling water grooves. The load acting portion is symmetrical about the vertical axis. The vertical axis is a symmetry axis of a cross section of the shell parallel to a direction of a gravity center of the propulsion shaft. The load acting portion is located on a lower surface side of the inner surface of the shell and subjected to a load of the propulsion shaft. The load covers at least the inner surface of the shell corresponding to a central angle of 800 or more and not exceeding 1800. The cooling action portion arranged opposite to the load acting portion is provided with a plurality of required cooling water grooves including a first and last cooling water grooves provided at both end portions of the load acting portion.

[0031] As illustrated in FIG. 3, an integrated-type bearing 5 of this example is configured such that a bearing material is fixed firmly to the inner surface of a cylindrical shell 1. The bearing material includes two layers of an elastic body layer 2 and a sliding layer 3 in this order from the inner surface of the shell 1. First, when the symmetry axis of a cross section of the shell parallel to a direction of the gravity center of the propulsion shaft is a vertical axis BB, the bearing 5 has a load acting portion 3a located on the lower surface side of the inner surface of the shell 1 and having a surface within a range having a central angle 0 of 600 to 180° with respect to the vertical axis BB as the symmetry axis. The bearing 5 also has a cooling action portion 3b that is provided with a plurality of required cooling water grooves 4 including a first and last cooling water grooves 4 provided at both end portions of the load acting portion 3a. The load acting portion 3a is a portion on which a water film for water lubrication is formed, and is therefore preferably smooth like the sliding layer 163 of the closed arc piece 16. The required cooling water grooves 4 can be provided, for example, between the first cooling water groove 4 and the last cooling water groove 4 at equal intervals.

[0032] The bearing 5 of this example can be formed as follows. First, the shell I and a core are mounted on a mold, a cylindrical sliding layer forming material is mounted on the core, and then an elastic body forming material is injection-molded to form a bearing raw material. Then, a plurality of cooling water grooves 4 is machined in a predetermined range (360°-0=180° to 300°) to form the cooling action portion 3b, and the bearing 5 is thus formed. The sliding layer forming material can be the same material as that for the sliding layer 153 of the gap forming arc piece 15 or the sliding layer 163 of the closed arc piece 16.

The elastic body forming material can be the same material as that for the intermediate layer 152 of the gap forming arc piece b or the intermediate layer 162 of the closed arc piece 16. [0033] The bearing 5 is characterized in that the bearing material has a two-layer structure of the sliding layer and the elastic body layer and has the load acting portion 3a in a range of a central angle 0 of 60° to 180°. In general, the bearing receives the load of the propulsion shaft in a portion below the axis of the shaft, and the wear of the sliding layer occurs in that portion. Therefore, it has been tested how large the load acting portion 3a of the bearing 5 receiving the load of the propulsion shaft is necessary. The wear test results are illustrated in FIG. 4. FIG. 4 is a graph illustrating the results obtained through performing a wear test on a bearing having a bearing inner diameter 60tpx length 60 (mm) illustrated in FIG. 5 using a wear tester illustrated in FIG. 6 and measuring the bearing inner diameter after the test. The numbers added to the graph of FIG. 4 each indicate the number of the cooling water grooves 4. In FIG. 5, the number of cooling water grooves is 2 for 0=180° (contact surface angle 157.7°), 3 for 0=120° (contact surface angle 100°), and 4 for 0=90° (contact surface angle 72.2°). As additional information, the number of cooling water grooves is 6 for 0=60° (contact surface angle 41.6°) and 8 for 0=45° (contact surface angle 24.4°). The size of the cooling water grooves 4 was depth 3 mmxwidth 6 mm.

[0034] In the wear tester illustrated in FIG. 6, the temperature of the water bath is maintained at 20°C, and forced convection of water is not performed. The bearing is lifted by a pneumatic cylinder, and a predetermined load is applied to the lower side of the bearing via the sleeve of the propulsion shaft. As for the sleeve, a required sleeve can be attached to the propulsion shaft. The amount of change in the bearing inner diameter was measured after performing a break-in operation at a constant circumferential speed (a wear test at a constant circumferential speed of 1 m/s for a total of 72 hours: 24 hours with a surface pressure of 0.25 MPa, 24 hours with a surface pressure of 0.50 MPa, and 24 hours with a surface pressure of 1.00 Tv1Pa). The bearing inner diameter was measured after the bearing was held at room temperature (20°C) for 6 hours after the wear test was completed.

[0035] In the graph illustrated in FIG. 4, the horizontal axis represents the contact surface angle (corresponding to the central angle 0) of the bearing and the vertical axis represents the amount of change in the inner diameter of the bearing. According to FIG. 4, the wear progresses slowly in a range of the contact surface angle of 160° to 80°. However, it is understood that there is a contact surface angle at which the wear progresses rapidly between the contact surface angles of 80° and 40°. It is understood that a water film is formed when the contact surface angle is within the range of 160° to 80° and the bearing is in an appropriate water lubrication state. On the other hand, it is understood that the water film is in an unstable state or a broken state when the contact surface angle is not less than a certain contact surface angle within the range of 80° to 40°, and proper water lubrication of the bearing is not performed. That is, it is preferred to provide the load acting portion 3a in a range of the central angle 0 of 80° to 180°. Such characteristics can be similarly applied to the lining-type bearing (bearing 10). The bearing 10 has the positioning plates 12, and the range of the preferred contact surface angle is therefore 0=80° to 175° in consideration of the arrangement and size of the positioning plates 12. Comprehensively determining the test/examination results, which will be described below, 0=60° to 175° is preferred and 0=80° to 175° is more preferred.

[0036] As described above, the bearing 5 is characterized in that the bearing material has a two-layer structure of the sliding layer and the elastic body layer. FIGS. 7 to 10 illustrate the action/effect of the elastic body. FIG. 7 illustrates structural analysis values of the surface pressure applied to the sliding surface portion of the bearing having a bearing inner diameter of 495 mm and a bearing length of 2000 mm. The horizontal axis represents each cross-sectional position where the cross-sectional position of the bearing end on the propeller side is 0 and the cross-sectional position of the bearing end on the bow side is 2 m. The vertical axis represents the surface pressure. Parameter FRB (solid line) represents the case of a bearing whose bearing material has a two-layer structure in which the sliding layer is a fluorine resin (PTFE) and the elastic body layer is a nitrile rubber (NBR) having durometer hardness of A70. RB (broken line) represents a bearing whose bearing material has a single-layer structure of an elastic body layer (nitrile rubber). LB (dashed-dotted line) represents a bearing whose bearing material has a single-layer structure of lignum vitae. The longitudinal elastic moduli of the bearing materials are 50 MPa for FRB, ID MPa for RB, and 2000 MPa for LB. The average surface pressure (load/bearing projected area) was 0.18 MPa.

[0037] According to FIG. 7, in the case of LB, a very high surface pressure is applied to both end portions of the bearing, and the surface pressure at the cross section 0 position is the highest at 0.8 MPa. The surface pressure in the central portion is zero over a wide area. In the case of RB, the surface pressure is most uniform, and the surface pressure is within a range of 0.08 MPa to 0.28 MPa. In the case of FRB, the surface pressure is as uniform as in the case of RB. In the case of FRB, however, the surface pressure curve has a downwardly convex shape, and the surface pressure is within a range of 0.02 MPa to 0.4 MPa. Comparing the case of FRB with the case of RB, the FRB differs from the RB in that the surface pressure is higher than that of RB at both end portions of the bearing and lower than that of RB at the central portion. In the case of FRB, the hardness of the elastic body layer (durometer hardness A) is preferably 50 to 90 from the viewpoint of making the surface pressure uniform. The bearing, which will be described below, for example, FRB16, has a two-layer structure as in the above FRB and has 16 cooling water grooves. FBI 6 represents a bearing having a single-layer structure of a fluorine resin only. The above LB example can be applied to the bearing FBI 6.

[0038] FIGS. 8 and 9 are graphs illustrating the relationship between a circumferential speed and a friction coefficient for a bearing having a bearing inner diameter of 60 mm and a bearing length of 60 mm. FIG. 10 is a set of graphs illustrating the temperature measurement results of a bearing member (sliding layer or elastic body layer under the cooling water grooves) during a model operation test for a bearing having a bearing inner diameter of 60 mm and a bearing length of 190 mm. The cross section of the bearing (FIGS. 8 to 10) has the bearing shape illustrated in FIG. 5. The friction coefficient was obtained through performing the above break-in operation at a constant circumferential speed and then performing a wear test for a predetermined time (0.5 h or 1 h) at each surface pressure (0.25 MPa to 1.00 MPa) and each circumferential speed (0.10 m/s to 4.00 m/s). For the wear test, the wear tester illustrated in FIG. 6 was used. The temperature of the water bath was maintained at 32°C.

[0039] In FIG. 8, the horizontal axis represents the circumferential speed and the vertical axis represents the friction coefficient when the surface pressure is 0.25 MPa. Parameters FRB16, FRB8, and FRB2 represent the cases in which the bearing members have a two-layer structure and the numbers of cooling water grooves are 16 (contact surface angle is 12.7°), 8 (contact surface angle is 24.4°), and 2 (contact surface angle is 157.7°), respectively. FB16 represents a case in which the bearing member has a single-layer structure with no elastic body layer and the number of cooling water grooves is 16 (contact surface angle is 12.7°). According to FIG. 8, the friction coefficient curves of FRB16, FRB8, and FRB2 are almost the same, and the friction coefficient decreases rapidly within a range of the circumferential speed of 0.1 m/s to 1 m/s. The friction coefficient curves become 0.005 at a circumferential speed of 1 m/s and thereafter become a constant value regardless of the circumferential speed. On the other hand, in the case of FB16, the friction coefficient decreases rapidly within a range of the circumferential speed of 0.1 m/s to 1 m/s, but it is 0.03 at a circumferential speed of 1 m/s. Then, at a circumferential speed of 4 m/s, the friction coefficient is almost the same (0.007) as those of the bearings with a two-layer structure, such as FRB16. That is, it can be found that the friction coefficient of the bearing having a two-layer structure with an elastic body layer is the smallest when the number of cooling water grooves is 2, but is almost the same regardless of the number of cooling water grooves (2 to 16) It can also be found that the bearings having a two-layer structure have a smaller friction coefficient than that of the bearing having a single-layer structure.

[0040] The tendency that the friction coefficient decreases rapidly with the circumferential speed and becomes almost constant above a certain circumferential speed, as illustrated in FIG. 8, is exhibited also when the surface pressure is high. FIG. 9 illustrates the friction coefficient of a bearing (FRB8) having a two-layer structure under various surface pressures. According to FIG. 9, there is a tendency that a low friction coefficient is exhibited at a low surface pressure (0.25 1VIPa) and a high friction coefficient is exhibited at a high surface pressure (1.0 MPa), but it can be found that the variation of measurement points due to different surface pressures is small.

[0041] FIG. 10 is a set of graphs illustrating the temperature change of the bearing material in each cycle of a constant circumferential speed/surface pressure operation of the model operation. FIG. 10(a) illustrates the temperature change of the bearing material of the bearing (FRB16) having a two-layer structure. FIG. 10(b) illustrates the temperature change of the bearing material of the bearing (FB16) having a single-layer structure. The model operation refers to a wear test in which a load is applied so that the surface pressure at both end portions of the bearing becomes 0.8 MPa and the average surface pressure becomes 0.31 M1Pa by adjusting the sleeve shape in the wear test, and a circumferential speed change operation and a constant circumferential speed/surface pressure operation are performed as one cycle and repeated. The circumferential speed change operation is a wear test in which the surface pressure (0.31 M_Pa) is kept constant and the circumferential speed is increased in a stepwise manner to 0.4 m/s/30 min, 0.6 m/s/30 min, 1.0 m/s/30 min, and 1.6 m/s/30 min. The constant circumferential speed/surface pressure operation is a wear test performed at a circumferential speed (0.6 m/s) and a surface pressure (0.31 N4Pa) for 25 hours. In the figure, the time is indicated as 0 to 25 h, 25 to 50 h, and so on. The wear test of this example was performed for 4 cycles (108 hours in total), and the circumferential speed change operation was further performed for 2 hours (110 hours in total).

[0042] According to FIG. 10, in the case of FRB16, the temperature of the bearing material is almost constant within a range of 36°C to 40°C. On the other hand, in the case of FB16, it can be found that the temperature is high within a range of 37°C to 45°C and the temperature variation is large. In particular, the temperature variation of FB16 in the case of constant circumferential speed/surface pressure operation (25 to 50 h) is large. In the circumferential speed change operation, there is a tendency that when the circumferential speed is increased, the amount of water flowing into the cooling water grooves increases and the temperature of the bearing material decreases. This phenomenon was more remarkable in the case of the bearing FRB16 than in the case of the bearing FB16.

[0043] After the above test, the amount of wear was measured for the bearing FRB16 and the bearing FB16. The amount of wear of the bearing FB16 was 2.6 times the amount of wear of the bearing FRB16. That is, the bearing FRB16 having a two-layer structure has a uniform surface pressure and a small friction coefficient of the sliding layer portion as compared with the bearing FB16 having a single-layer structure. Moreover, it can be found that the bearing FRB16 has a stable low temperature of the bearing material in operation and a small amount of wear. The wear resistance of the bearing FRB16 is superior to that of the bearing FB16. Such a feature can be applied to the above lining-type bearing (bearing 10). In addition, the design values of bearings for ships are generally set to a circumferential speed of 12 m/s or less and a surface pressure of 0.6 MPa or less, and the results of the above wear test can be applied to general ships.

[0044] As described above, the bearing 5 or 10 according to the present invention has the sliding layer 3, 153, or 163 having excellent wear resistance and heat resistance and the elastic body layer 2 or the intermediate layer 152 or 162 (elastic body layer) for uniformizing the load from the propulsion shaft as the bearing material on the inner surface of the shell. Moreover, the inner surface of the bearing 5 or 10 in contact with the propulsion shaft has the load acting portion 3a receiving a load from the propulsion shaft or a portion (load acting portion) arranged with each closed arc piece and the cooling action portion 3b having the cooling water grooves 4 opposed thereto or a portion (cooling action portion) arranged with each gap forming arc piece having the groove forming portions 155 which form the groove 18. The load acting portion is arranged at a central angle 0 of 60° to 180° or a central angle 0 of 600 to 175° on the inner surface of the shell. In this range, a water film is stably formed and water lubrication is performed. Moreover, the cooling action portion is provided with the required cooling water grooves 4 or 18 so that the required cooling of the bearing 5 or 10 is performed. When the bearing inner diameter is D (cm), an amount of cooling water Q (kg/hr) supplied to the cooling water grooves 4 or 18 is preferably set such that k=4 is obtained, provided that Q>kxD2 and D=15 to 100 (cm). Q=4xD2 is the amount of cooling water practically required for a large ship (D=15 to 100). However, the amount of cooling water to be used is determined in consideration of the characteristics and type of the ship, the cost of the required pump, etc. Then, the size and number of the cooling water grooves 4 or 18 required for supplying such an amount of cooling water are determined. The bearings or 10 according to the present invention is excellent in the wear resistance and durability without requiring circumferential grooves or water supply grooves for forcibly cooling the portion receiving the load from the propulsion shaft as in the conventional shaft bracket bearing of a ship. Moreover, the bearing 5 or 10 can be used not only for the shaft bracket bearing but also for the stern tube bearing.

[0045] In addition, the bearing 5 or 10 (the bearing according to the present invention) has a two-layer structure in which the sliding surface is smooth at the load acting portion as described above and which includes the elastic layer, and is in a uniform stress state. Therefore, as illustrated in FIGS. 11 and 12 which will be described below, the bearing can prevent the occurrence of abnormal vibration that is caused periodically as in a comparative example. Moreover, as illustrated in FIG. 13, the bearing is in a water lubrication state because of the water film, and it is possible to prevent the occurrence of a problematic stick-slip phenomenon. it is said that the stick-slip phenomenon is likely to occur under low circumferential speed rotation, and it has been a problem for marine research ships and submarines that navigate in a low speed mode and obtain information by sonar, but the bearing can be suitably used as a bearing for the propulsion shaft of such a marine research ship or submarine. This is because the bearing can significantly reduce the vibration, which occurs on the bearing portion, due to the synergistic effect of eliminating the cooling water grooves at the bearing lower portion unlike the conventional bearing and adopting a two-layer structure that includes an elastic layer unlike the conventional single-layer bearing.

[0046] FIG. 11 illustrates the results of a vibration analysis test using an FFT (fast Fourier transform) analyzer when the shaft is supported by the bearings of the invention example (FRB2) and the comparative example (FRB16). In FIG. 11, the horizontal axis represents the period (0.47 when the shaft rotates once, 0.94 when the shaft rotates twice), the vertical axis represents the vibration intensity (m/s2), and the white portions represent the vibration state of the invention example. The thorn-shaped overhanging portions represent the vibration state of the comparative example. According to FIG. 11, in the invention example, a small stable vibration state is observed. On the other hand, in the comparative example, a very strong vibration is partially observed during the vibration having a constant width stronger than the vibration of the invention example. The comparative example corresponds to the bearing described in Patent Document 2 or 3.

[0047] FIG. 12 illustrates the results of a vibration analysis test using an FFT analyzer when the shaft is supported by the bearings of the invention example (FRB2) and the comparative example (FB2). In FIG. 12, the horizontal axis represents the period, the vertical axis represents the vibration intensity, the fine thorn-shaped black line portions represent the invention example, and the gray portions represent the comparative example. FB2 represents a bearing having a single-layer structure without an elastic layer and having two cooling water grooves. According to FIG. 12, in the invention example, a small stable vibration is observed from 0 h to 100 h. On the other hand, in the comparative example, the vibration is performed with a small and stable strength equivalent to that of the invention example from 0 h to 50 h, but strong vibration is observed in the vicinity of 0.4 rotation at 75 h and in the vicinity of 0.8 rotation at 100 h The comparative example corresponds to the bearing described in Patent Document 1.

[0048] For the bearing 5, how the contact surface angle or circumferential speed affects the water film pressure strength was analyzed by using contact problem analysis software 10 TED/CPA (Tribology Engineering Dynamics/Contact Problem Analyzer). The analysis results are illustrated in FIG. 13. The analysis conditions were bearing inner diameterxlength of 5509x1100 (mm), a bearing surface pressure of 0.5 N4Pa, and a circumferential speed of 0.4, 1, 2, or 4 (m/s). Table 1 lists the relationship between the contact surface angle and the number of grooves.

[0049]

[Table]]

Contact surface angle (°) 10.4 22.5 41.7 56.7 80.3 110.5 170.9 Number of grooves 26 14 8 6 4 3 2 [0050] In FIG. 13, the horizontal axis represents the contact surface angle, the vertical axis represents the water film pressure strength, and the parameter is the circumferential speed.

According to FIG. 13, the water film pressure strength is constant when the contact surface angle is 00 to 20°, but increases rapidly when the contact surface angle is 20° to 400 However, the amount of increase in the water film pressure strength with respect to the contact surface angle decreases gradually from 40° to 60° and decreases rapidly beyond 60°.

Then, when the contact surface angle becomes 80° or more, the water film pressure strength becomes almost constant with respect to the contact surface angle. That is, the water film pressure strength curve has a stepped shape including a straight portion having a contact surface angle of 0° to 20°, a curved portion of 20° to 80°, and a one level higher straight portion of 800 or more.

[0051] FIG. 14 illustrates the relationship between the water film pressure strength and the circumferential speed of the portion in which the above water film pressure strength curve exhibits a curved shape. In FIG. 14, the horizontal axis represents the circumferential speed, the vertical axis represents the water film pressure strength, and the parameter is the contact surface angle. According to FIG. 14, each of the water film pressure strength curves having a contact surface angle of 22.5° to 80.3° is approximately linear, and the water film pressure strength is proportional to the circumferential speed. The water film pressure strength curves with contact surface angles of 56.7° and 80.3° almost overlap each other The gradient of the water film pressure strength curve with a contact surface angle of 41.7° is 0.8 times that of the water film pressure strength curve with a contact surface angle of 80.3°. The gradient of the water film pressure strength curve with a contact surface angle of 22.5° is 0.25 times that of the water film pressure strength curve with a contact surface angle of 80.3°.

Considering the above results and considerations of FIGS. 13 to 14 and the above-described results and considerations of the wear test illustrated in FIG. 4, it is understood that, in the bearing according to the present invention, the central angle 0 of the load acting portion 3a is preferably within a range of 60° to 180° and more preferably within a range of 80° to 180°.

The circumferential speed of a ship is 2 to 4 m/s in normal navigation. A circumferential speed of 4 m/s corresponds to high-speed navigation, and a circumferential speed of 0.4 m/s corresponds to harbor navigation.

[0052] The bearing according to the present invention has been described above. The bearing according to the present invention has a sliding layer having a smooth surface and composed of a material excellent in the wear resistance and the heat resistance and an elastic body layer capable of uniformizing the load from a propulsion shaft and is excellent in the low frictional properties, wear resistance, and durability. Moreover, the present invention can improve the durability of the bearing through eliminating the cooling water grooves on the bearing lower surface side, which receives a large amount of physical stress due to deformation caused by the weight of the supporting propulsion shaft and rotation of the propulsion shaft, thereby to minimize the contact between water or seawater and the intermediate elastic layer and suppressing the chemical erosion of the intermediate elastic layer due to the highly alkaline environment or hydrogen sulfide environment caused by water or seawater.

[Description of Reference Numerals] [0053]

1 Shell 2 Elastic body layer 3 Sliding layer 4 Cooling water groove Bearing Bearing 11 Shell 12 Positioning plate Gap forming arc piece 151 Base 152 Intermediate layer 153 Sliding layer Groove forming portion 16 Closed arc piece 161 Base 162 Intermediate layer 163 Sliding layer 17 Notched closed arc piece 18 Groove 19 Half groove

Claims

CLAIMSA bearing comprising: a cylindrical shell having an inner surface and supporting a propulsion shaft of a ship; a pair of positioning plates arranged on the inner surface of the shell; and a plurality of closed arc pieces and a plurality of gap forming arc pieces, the closed arc pieces and the gap forming arc pieces being arranged on the inner surface of the shell, the positioning plates being fixed at positions facing each other on a horizontal axis of the shell, the horizontal axis being a symmetry axis of a cross section of the shell orthogonal to a direction of a gravity center of the propulsion shaft the closed arc pieces being arranged on a lower surface side of the shell and subjected to a load of the propulsion shaft, each of the closed arc pieces comprising a three-layer structure, the three-layer structure comprising: a sliding layer; an intermediate layer made of an elastic body; and a base in close contact with the inner surface of the shell, the sliding layer, the intermediate layer, and the base being pressed against adjacent ones between the positioning plates so as to be fastened on the inner surface of the shell, the gap forming arc pieces being arranged opposite to the closed arc pieces, each of the gap forming arc pieces comprising a three-layer structure and having groove forming portions on both side edge portions, the three-layer structure comprising: a sliding layer; an intermediate layer made of an elastic body; and a base in close contact with the inner surface of the shell, the intermediate layer and the base being pressed against adjacent ones between the positioning plates so as to be fastened on the inner surface of the shell and so that the groove forming portions form a groove through which cooling water flows.
2. The bearing according to claim 1, wherein the closed arc pieces cover at least the lower surface of the shell corresponding to a central angle of 600 or more and not exceeding 175°.
3. The bearing according to claim 1, wherein the sliding layers of the closed arc pieces and gap forming arc pieces are each made of a synthetic high polymer compound containing fluorine atoms in a molecule, a polyamide resin, or a phenol resin.
The bearing according to claim 1, wherein the intermediate layers of the closed arc pieces and gap forming arc pieces made of an elastic body having specified durometer hardness from A50 to A90 are adhered to the base.
5. A bearing for supporting a propulsion shaft of a ship, the bearing comprising: a cylindrical shell having an inner surface; and two elements of an elastic body layer and a sliding layer fixed firmly in this order on the inner surface of the shell, the sliding layer having a load acting portion without cooling water groove and a cooling action portion with cooling water grooves, the load acting portion being symmetrical about the vertical axis, the vertical axis being a symmetry axis of a cross section of the shell parallel to a direction of a gravity center of the propulsion shaft, the load acting portion being located on a lower surface side of the inner surface of the shell and subjected to a load of the propulsion shaft, the load covering at least the inner surface of the shell corresponding to a central angle of 80° or more and not exceeding 180°, the cooling action portion arranged opposite to the load acting portion, being provided with a plurality of required cooling water grooves including a first and last cooling water grooves provided at both end portions of the load acting portion