JP2011236923A - Bearing structure of rotary shaft - Google Patents

Abstract

PROBLEM TO BE SOLVED: To make both compatible in reduction in a viscous friction loss and an increase in a minimum oil film thickness, when supporting a rotary shaft by a radial sliding bearing via lubricating oil.SOLUTION: In a bearing inner peripheral surface 41 of a large end part bearing 30, a distance to the bearing axis 30C gradually monotonously reduces to a position for advancing to an angle θ in the rotational direction of a crank pin 18 from a position corresponding to the reference direction (an angle 0°) along one side of a reciprocating load, and the distance to the bearing axis 30C suddenly increases by δ in a position corresponding to the angle θ. The distance to the bearing axis 30C gradually monotonously reduces to a position for advancing to an angle 180°+θ in the rotational direction of the crank pin 18 from a position corresponding to an angle 180°-θ, the distance to the bearing axis 30C suddenly increases by δ in a position corresponding to the angle 180°+θ, and the distance to the bearing axis 30C gradually monotonously reduces to a position for advancing in the rotational direction in the rotational direction of the crank pin 18 from a position corresponding to an angle 360°-θ.

Description

  The present invention relates to a bearing structure of a rotating shaft, and more particularly to a bearing structure of a rotating shaft that supports the rotating shaft through lubricating oil filled in a gap between the bearing inner peripheral surface of the radial slide bearing and the rotating shaft. .

  Patent Documents 1 and 2 below disclose related arts of bearing structures in which a rotating shaft is supported by a bearing inner peripheral surface of a radial slide bearing via a lubricating oil. In Patent Document 1, the cross-sectional shape of the crank pin is non-circular, and the minimum position of the bearing gap between the cross-sectional shape and the slide bearing is an axis connecting the rotation center of the crank journal and the center of the slide bearing. A crankshaft that is not on line is disclosed. In Patent Document 1, the cross-sectional shape of the crank pin is processed into an elliptical shape, and an axis line connecting the minimum position of the bearing gap and the center of the slide bearing and an axis line connecting the rotation center of the crank journal and the center of the slide bearing are formed. The angle is set to 90 ° or less in the direction opposite to the rotation direction of the crank journal. Further, Patent Document 2 discloses a journal slide bearing in which a depression or a hole communicating with the outside of the bearing body is provided in a portion where a negative pressure of the bearing body is predicted to be generated.

JP 2003-278739 A JP 61-1557819 A

  When the rotating shaft is supported by the bearing inner peripheral surface of the radial slide bearing through the lubricating oil, if a load along the radial direction of the rotating shaft acts on the bearing inner peripheral surface from the rotating shaft through the lubricating oil, Oil film pressure (wedge) that balances the load due to the wedge effect in which the lubricating oil is dragged in due to viscosity in a gap tapering in the direction of rotation of the rotating shaft, which is formed between the portion receiving the load on the peripheral surface and the rotating shaft. Oil film pressure). At that time, the load and the oil film pressure are balanced in a state in which the central axis of the rotating shaft is eccentric with respect to the bearing central axis of the radial slide bearing in a direction advanced by an angle in the rotational direction of the rotating shaft from the load acting direction. Therefore, the oil film thickness formed between the rotating shaft and the bearing inner peripheral surface also becomes the minimum film thickness at a position advanced by an angle in the rotating direction of the rotating shaft from the load acting direction. If the amount of eccentricity of the rotating shaft increases and the minimum oil film thickness decreases, the risk of seizure and wear at high loads increases.

  If the radial plain bearing is a normal round bearing, the smaller the clearance between the rotating shaft and the inner peripheral surface of the bearing, the easier the wedge effect will be and the smaller the amount of eccentricity of the rotating shaft. This is advantageous for increasing the thickness. However, if the clearance between the rotating shaft and the inner peripheral surface of the bearing is reduced, the viscous friction loss when the rotating shaft rotates with respect to the radial slide bearing increases, and in particular, the viscous friction at low temperatures where the oil viscosity is high. Loss increases. Therefore, in the case of an ordinary perfect circle bearing, it is difficult to achieve both reduction of viscous friction loss and increase of the minimum oil film thickness.

  In Patent Document 1, a tapered gap is formed in a direction opposite to the rotation direction of the crankpin between a portion that receives a load in a plain bearing (round bearing) and a crankpin having an elliptical cross section. Therefore, it is difficult to obtain a sufficient wedge effect, and it is difficult to reduce the eccentric amount of the crankpin and increase the minimum oil film thickness. Further, in Patent Document 1, since a special technique is required for non-round processing (elliptical processing) of the crankpin, the cost is increased. Patent Document 2 is a technique aimed at reducing vibration by preventing negative pressure from being generated inside the bearing body, and shows that both reducing the viscous friction loss and increasing the minimum oil film thickness are achieved. It has not been.

  An object of the present invention is to achieve both a reduction in viscous friction loss and an increase in minimum oil film thickness when a rotating shaft is supported by a radial slide bearing via lubricating oil.

  The bearing structure of the rotating shaft according to the present invention employs the following means in order to achieve the above-described object.

  The bearing structure of the rotating shaft according to the present invention is a rotating shaft that receives a reciprocating load along the radial direction of the rotating shaft through lubricating oil filled in a clearance between the bearing inner peripheral surface of the radial slide bearing and the rotating shaft. In the bearing structure of the shaft, if θ is a predetermined angle larger than 0 ° and smaller than 90 ° on the inner peripheral surface of the shaft, a direction perpendicular to the bearing central shaft and along one side of the reciprocating load is used as a reference. In addition, the distance to the bearing center axis gradually decreases to a position that advances the angle θ in the rotation direction of the rotating shaft, and the distance to the bearing center axis rapidly increases at a position corresponding to the angle θ, from a position corresponding to an angle of 180 ° −θ. The distance to the bearing center axis gradually decreases to a position that advances to an angle of 180 ° + θ in the rotation direction of the rotating shaft, and the distance to the bearing center axis rapidly increases at a position corresponding to the angle of 180 ° + θ, corresponding to an angle of 360 ° −θ. Rotating shaft from position to rotate Distance to the bearing axis to a position where the process proceeds to the reference countercurrent to the gist that gradually decreases.

  In addition, the bearing structure of the rotary shaft according to the present invention applies a reciprocating load along the radial direction of the rotary shaft through the lubricating oil filled in the gap between the bearing inner peripheral surface of the radial slide bearing and the rotary shaft. In the bearing structure of the rotating shaft to be received, when θ is a predetermined angle larger than 0 ° and smaller than 90 °, the direction of the rotating shaft is based on the direction perpendicular to the bearing central axis and along one side of the reciprocating load. A radial plain bearing is configured including two half bearings divided into a position advanced by an angle θ in the rotational direction and a position advanced by an angle of 180 ° + θ, and one of the half bearings is located from a position corresponding to the angle θ. The bearing inner peripheral surface is formed in a rotational direction of the rotary shaft to a position that advances to an angle of 180 ° + θ, and the other half bearing extends from a position corresponding to an angle of 180 ° + θ to a position that advances in the rotational direction of the rotary shaft to an angle θ. Forms the inner peripheral surface of the bearing and corresponds to the angle θ The bearing inner peripheral surface of one half bearing is shifted radially outward from the bearing inner peripheral surface of the other half bearing, and near the bearing split position corresponding to an angle of 180 ° + θ. The gist is that the bearing inner circumferential surface of the other half bearing is displaced radially outward from the bearing inner circumferential surface of the one half bearing.

  In addition, the bearing structure of the rotating shaft according to the present invention includes a lubricating oil filled in a gap between a bearing inner peripheral surface of a radial slide bearing attached to a large end portion of a connecting rod and a crank pin of a crankshaft. In the bearing structure of the rotating shaft that supports the crankpin, when θ is a predetermined angle that is larger than 0 ° and smaller than 90 ° on the inner peripheral surface of the bearing, it is perpendicular to the bearing central axis and from the bearing central axis. The distance to the bearing center axis gradually decreases to the position advanced by the angle θ in the rotation direction of the crank pin with reference to the direction toward the small end center axis of the connecting rod, and the distance to the bearing center axis at the position corresponding to the angle θ The distance to the bearing center axis gradually decreases from a position corresponding to an angle of 180 ° -θ to a position that advances to an angle of 180 ° + θ in the rotation direction of the crankpin, and an angle of 180 ° The distance to the bearing center axis increases rapidly at a position corresponding to θ, and the distance to the bearing center axis gradually decreases from a position corresponding to an angle of 360 ° −θ to a position that advances to the reference in the rotation direction of the crankpin. And

  In addition, the bearing structure of the rotating shaft according to the present invention includes a lubricating oil filled in a gap between a bearing inner peripheral surface of a radial slide bearing attached to a large end portion of a connecting rod and a crank pin of a crankshaft. And a rotation shaft bearing structure for supporting the crankpin, where θ is a predetermined angle larger than 0 ° and smaller than 90 °, and perpendicular to the bearing central axis and from the bearing central axis to the center of the small end of the connecting rod A radial plain bearing is configured by including two half bearings divided into a position advanced by an angle θ in the rotation direction of the crankpin and a position advanced by an angle of 180 ° + θ with respect to the direction toward the shaft. The split bearing forms a bearing inner peripheral surface from a position corresponding to the angle θ to a position that advances to an angle of 180 ° + θ in the rotation direction of the crankpin, and the other half bearing has an angle of 180 ° + θ. A bearing inner peripheral surface is formed from a corresponding position to a position that advances to an angle θ in the rotation direction of the crankpin, and the bearing inner peripheral surface of one half bearing is in the vicinity of the bearing split position corresponding to the angle θ. In the vicinity of the bearing split position corresponding to an angle of 180 ° + θ, the bearing inner peripheral surface of the other half bearing is shifted from the bearing inner peripheral surface of the split bearing to the radially outer side. The gist is that it is displaced radially outward from the surface.

  In one aspect of the present invention, the curvatures of the bearing inner peripheral surfaces of the two half bearings are equal, and the center of curvature of the bearing inner peripheral surface of one half bearing is the center of curvature of the bearing inner peripheral surface of the other half bearing. On the other hand, it is preferable to shift to the bearing split position corresponding to the angle θ.

  Further, the bearing structure of the rotating shaft according to the present invention receives a load along the radial direction of the rotating shaft through the lubricating oil filled in the clearance between the bearing inner peripheral surface of the radial sliding bearing and the rotating shaft. In the bearing structure of the rotating shaft, on the inner peripheral surface of the bearing, when θ is a predetermined angle larger than 0 ° and smaller than 90 °, the shaft rotates on the basis of the direction perpendicular to the bearing central axis and along the load. The distance to the bearing center axis gradually decreases to a position that advances the angle θ in the rotation direction of the shaft, and the distance to the bearing center axis suddenly increases at a position corresponding to the angle θ, and from the position corresponding to the angle 360 ° −θ The gist is that the distance to the bearing center axis gradually decreases to a position that advances to the reference in the rotational direction.

  Further, the bearing structure of the rotating shaft according to the present invention receives a load along the radial direction of the rotating shaft through the lubricating oil filled in the clearance between the bearing inner peripheral surface of the radial sliding bearing and the rotating shaft. In the bearing structure of the rotating shaft, when θ is a predetermined angle larger than 0 ° and smaller than 90 °, the angle θ in the rotating direction of the rotating shaft is based on the direction orthogonal to the bearing central axis and along the load. A radial plain bearing is configured by including two half bearings divided into an advanced position and an angle 180 ° + θ advanced position, and one half bearing has a rotational direction of the rotary shaft from a position corresponding to the angle θ. The bearing inner peripheral surface is formed at a position that advances to an angle of 180 ° + θ, and the other half bearing has a bearing inner peripheral surface that extends from a position corresponding to the angle of 180 ° + θ to a position that advances to an angle θ in the rotation direction of the rotary shaft. Forming and bearing split position corresponding to angle θ In short, the bearing inner peripheral surface of one of the half bearing, can be summarized as are shifted radially outward from the bearing peripheral surface of the other half bearing.

  In one aspect of the present invention, the center of curvature of the bearing inner peripheral surface of one half bearing is shifted to the bearing split position corresponding to the angle θ with respect to the center of curvature of the bearing inner peripheral surface of the other half bearing. It is preferable that

  In one aspect of the present invention, it is preferable that the curvature of the bearing inner peripheral surface of one half bearing is smaller than the curvature of the bearing inner peripheral surface of the other half bearing.

  In one embodiment of the present invention, θ is preferably a predetermined angle of 20 ° or more and 70 ° or less.

  According to the present invention, when a load along the radial direction of the rotating shaft acts on the bearing inner peripheral surface of the radial slide bearing from the rotating shaft via the lubricating oil, the region contributes to generation of an oil film pressure that balances this load. Therefore, the oil film pressure that balances the load can be generated in a state where the eccentric amount of the rotating shaft is small, and the minimum oil film thickness can be increased. Furthermore, since the clearance between the rotary shaft and the bearing inner peripheral surface of the radial slide bearing can be increased in a region that does not contribute to the generation of oil film pressure that balances the load, the rotary shaft rotates with respect to the radial slide bearing. The viscous friction loss at the time can be reduced. Therefore, both reduction of viscous friction loss and increase of minimum oil film thickness can be achieved.

It is a figure which shows schematic structure of the connecting rod of the internal combustion engine to which the bearing structure of the rotating shaft which concerns on embodiment of this invention is applied. It is a figure which shows the outline of the bearing structure of the rotating shaft using the radial slide bearing which concerns on embodiment of this invention. It is a figure explaining the effect | action of the bearing structure of the rotating shaft which concerns on embodiment of this invention. It is a figure explaining the effect | action of the bearing structure of the rotating shaft which concerns on embodiment of this invention. It is a figure explaining the effect | action of the bearing structure of the rotating shaft using a perfect circle bearing. It is a figure explaining the effect | action of the bearing structure of the rotating shaft using a perfect circle bearing. It is a figure which shows the result of having investigated the change of the minimum oil film thickness and viscous friction torque when changing the predetermined angle (theta) by fluid lubrication calculation. It is a figure which shows the experimental result which measured viscous friction torque. It is a figure which shows the other example of the bearing structure of the rotating shaft using the radial slide bearing which concerns on embodiment of this invention. It is a figure which shows the other example of the bearing structure of the rotating shaft using the radial slide bearing which concerns on embodiment of this invention. It is a figure which shows the other example of the bearing structure of the rotating shaft using the radial slide bearing which concerns on embodiment of this invention. It is a figure which shows the other example of the bearing structure of the rotating shaft using the radial slide bearing which concerns on embodiment of this invention.

  DESCRIPTION OF EMBODIMENTS Hereinafter, embodiments for carrying out the present invention (hereinafter referred to as embodiments) will be described with reference to the drawings.

  FIG. 1 is a diagram showing a schematic configuration of a connecting rod of an internal combustion engine to which a bearing structure of a rotating shaft according to an embodiment of the present invention is applied, and FIG. 2 uses a radial slide bearing according to an embodiment of the present invention. FIG. 2 is a view showing an outline of a bearing structure of a rotating shaft (crank pin), and both are views seen from the rotating shaft direction. However, in each figure including FIGS. 1 and 2, the thickness of the radial slide bearing (large end bearing) and the size of the clearance between the rotary shaft (crank pin) and the radial slide bearing (large end bearing), etc. These are shown larger than the actual size for convenience of explanation.

  The connecting rod 10 of the internal combustion engine is a component that converts the reciprocating linear motion of the piston 11 to a crankshaft (not shown) and converts it into a rotational motion. As shown in FIG. 1, the connecting rod 10 includes a small end portion 12 on the piston 11 side, a large end portion 13 on the crankshaft side, and a column portion 14 that connects between the small end portion 12 and the large end portion 13. It is composed of The connecting rod 10 is made of a material such as nickel / chromium steel, chromium / molybdenum steel, or titanium alloy, and is a member that requires high mechanical strength.

  The small end portion 12 is a connection portion with the piston 11, and a small end portion through-hole 16 through which the piston pin 15 coupled to the piston 11 is inserted is formed. The small end through hole 16 is provided with a small end bearing 17 for supporting the piston pin 15. Since the small end bearing 17 receives, for example, a high load in the explosion stroke via the piston 11, a high load capacity is required. The movement form of the small end 12 is a swinging movement from the combination of the reciprocating movement of the entire connecting rod 10 and the rotational movement of the large end 13, so that the small end bearing 17 has a wide range of bearing inner peripheral surfaces. A high load acts over the range. Therefore, for example, lubricating oil is supplied to the inner peripheral surface of the bearing from a jet hole or a communication hole (not shown) formed in the large end portion 13 to support the axial load via the lubricating oil.

  The large end portion 13 is a connecting portion with the crankshaft, and a large end through hole 19 into which the crankpin 18 of the crankshaft is inserted is formed. The large end through hole 19 is provided with a large end bearing 30 for supporting the crank pin 18. The large end portion 13 includes a large end body 20 integrally formed with the column portion 14 and a cap 21 fastened to the large end portion body 20. The large end portion main body 20 is fastened with the cap 21. A large end bearing 30 having a half structure is attached to the large end through-hole 19 to be formed.

  As described above, the column portion 14 is a portion that connects between the small end portion 12 and the large end portion 13. Generally, the column portion 14, the small end portion 12, and the large end portion main body 20 are integrated. Molded. For example, a communication hole (not shown) is formed in the column portion 14 in order to supply the lubricating oil from the large end portion 13 to the small end portion 12. Further, as described above, the crankpin 18 is a shaft that connects the crankshaft and the large end portion 13 and is rotated by receiving the force due to the reciprocating motion of the piston 11 and transmits the force to the crankshaft. It is a member for.

  As described above, when the connecting rod 10 connects the piston 11 and the crank pin 18 of the crankshaft, when the piston 11 reciprocates in a cylinder (not shown) during operation of the engine, the reciprocating motion is transmitted through the connecting rod 10. It is converted into the rotational motion of the crankshaft, and the rotational power of the crankshaft is obtained as engine output.

  The large end bearing 30 is a radial slide bearing (also referred to as a journal slide bearing) that supports a cylindrical crankpin 18 that is a rotating shaft, and as shown in FIGS. It is configured by two substantially divided half-cylindrical bearings 31A and 31B. One half bearing 31A is attached to the cap 21 as a bearing support member, and the other half bearing 31B is attached to the large end body 20 as a bearing support member, and the circumferential direction of the two half bearings 31A and 31B. The large end bearing 30 is configured by aligning the both end portions with respect to each other. Each half bearing 31A, 31B includes a back metal and a bearing alloy layer as a lining layer formed on the inner peripheral side of the back metal. Examples of the back metal include steel, and examples of the bearing alloy layer include a copper-lead alloy and an aluminum alloy. The surface of this bearing alloy layer is a bearing inner peripheral surface 41 which is a bearing sliding surface.

  An oil supply passage is formed in the crankshaft, and the lubricating oil is filled into a gap between the outer peripheral surface of the crankpin 18 and the bearing inner peripheral surface 41 of the large end bearing 30 through the oil supply passage. The large-end bearing 30 having a half-split structure supports the crank pin 18 along the radial direction of the crank pin 18 by rotatably supporting the crank pin 18 on the bearing inner peripheral surface 41 (surface of the bearing alloy layer) via lubricating oil. The load is received through the lubricating oil. The main role of the lubricating oil here is to reduce the friction loss and wear between the shaft and the inner circumferential surface of the bearing by operating the engine without forming the oil film by forming an oil film. Also plays a role of cooling, cleaning, rust prevention and so on.

  In the cycle of the internal combustion engine (four-stroke engine), at the compression top dead center, a high load due to the combustion pressure in the cylinder acts on the bearing inner peripheral surface 41 of the large end bearing 30 from the crankpin 18. Here, the direction of application of a high load from the crank pin 18 to the bearing inner peripheral surface 41 is the bearing central axis 30C (large end central axis) of the large end bearing 30 and the bearing central axis 17C (large end central axis) of the small end bearing 17. It is a direction (the direction indicated by the arrow F1 in FIGS. 1 and 2) that is orthogonal to the small end portion central axis and is directed from the bearing central axis 30C to the bearing central axis 17C. Further, at the intake bottom dead center and the exhaust bottom dead center, a high load due to the inertia force of the piston system acts on the bearing inner peripheral surface 41 of the large end bearing 30 from the crank pin 18. Here, the direction in which a high load is applied from the crank pin 18 to the bearing inner peripheral surface 41 is also orthogonal to the bearing central shaft 30C and the bearing central shaft 17C and is directed from the bearing central shaft 30C to the bearing central shaft 17C (FIG. 1). , 2 (direction indicated by arrow F1). Even at the intake top dead center, a high load due to the inertia force of the piston system acts on the bearing inner peripheral surface 41 of the large end bearing 30 from the crank pin 18. However, the direction of the high load applied from the crank pin 18 to the bearing inner peripheral surface 41 here is orthogonal to the bearing central shaft 30C and the bearing central shaft 17C and is a direction from the bearing central shaft 17C to the bearing central shaft 30C ( (Direction shown by arrow F2 in FIGS. 1 and 2). As described above, the high load acting on the bearing inner peripheral surface 41 of the large end bearing 30 from the crank pin 18 is mainly a load due to the combustion pressure in the cylinder and the inertia force of the piston system. The reciprocating load is along the direction of the connecting rod main shaft 10A perpendicular to the central shaft 30C and the bearing central shaft 17C of the small end bearing 17 (directions indicated by arrows F1 and F2 in FIGS. 1 and 2). The large end bearing 30 receives this reciprocating load on the bearing inner peripheral surface 41 via the lubricating oil.

  In a normal connecting rod, a split surface (a split surface of the large end bearing 30) between the large end main body 20 and the cap 21 at the large end 13 is perpendicular to the connecting rod main shaft 10A (direction of reciprocating load). On the other hand, in the present embodiment, when θ is a predetermined angle larger than 0 ° and smaller than 90 °, split surfaces 13C and 13D (large end bearings) of the large end main body 20 and the cap 21 at the large end 13 are used. 30 split surfaces 31C, 31D) are inclined at a predetermined angle θ in the rotation direction of the crank pin 18 (the direction indicated by the arrow R in FIGS. 1 and 2) with respect to the connecting rod main shaft 10A (direction of reciprocating load). . Here, a direction perpendicular to the bearing center axis 30C and directed from the bearing center axis 30C to the bearing center axis 17C (small end center axis), that is, a direction along one side of the reciprocating load (indicated by an arrow F1 in FIGS. 1 and 2). When the reference direction (angle 0 °) is the reference direction (angle 0 °), the position (divided surface 13C) advanced by the angle θ along the rotation direction of the crankpin 18 around the bearing center axis 30C with respect to the reference direction and the angle 180 ° + θ. The large end portion 13 is divided into the large end portion main body 20 and the cap 21 at the advanced position (the dividing surface 13D).

  In the cap 21, a bearing mounting surface 13 </ b> A that is a substantially semi-cylindrical concave curved surface is formed from a position corresponding to the angle θ to a position that advances to an angle of 180 ° + θ in the rotation direction of the crankpin 18. One half bearing 31A is mounted on the surface 13A. In the large-end main body 20, a bearing mounting surface 13B that is a substantially semi-cylindrical concave curved surface is formed from a position corresponding to an angle of 180 ° + θ to a position that advances to an angle θ in the rotation direction of the crankpin 18. The other half bearing 31B is mounted on the bearing mounting surface 13B. Therefore, a large angle position (divided surface 31C) advances along the rotational direction of the crankpin 18 around the bearing center axis 30C with respect to the reference direction (divided surface 31C) and a position advanced 180 ° + θ (divided surface 31D). The end bearing 30 is divided into two half bearings 31A and 31B. Then, one half bearing 31A forms a substantially semi-cylindrical bearing inner peripheral surface 41A from a position corresponding to the angle θ to a position that advances to an angle of 180 ° + θ in the rotation direction of the crank pin 18, and the other half bearing The bearing 31B forms a substantially semi-cylindrical bearing inner peripheral surface 41B from a position corresponding to an angle of 180 ° + θ to a position proceeding to the angle θ in the rotation direction of the crankpin 18, and by these bearing inner peripheral surfaces 41A, 41B. The inner peripheral surface 41 of the large end bearing 30 is formed. The curvatures of the two bearing mounting surfaces 13A and 13B are equal to each other, and the curvatures of the bearing inner peripheral surfaces 41A and 41B in the two half bearings 31A and 31B are equal to each other, both of which are smaller than the curvature of the outer peripheral surface of the crankpin 18.

  Further, in the present embodiment, the center of curvature 30A of the bearing mounting surface 13A of the cap 21 is shifted from the bearing center shaft 30C by δ / 2 toward the dividing surface 13C corresponding to the angle θ, and the large end body The center of curvature 30B of the 20 bearing mounting surfaces 13B is deviated by δ / 2 from the bearing center shaft 30C toward the divided surface 13D corresponding to an angle of 180 ° + θ. That is, the center of curvature 30A of the bearing mounting surface 13A is deviated from the center of curvature 30B of the bearing mounting surface 13B by δ toward the dividing surface 13C corresponding to the angle θ. Accordingly, the center of curvature 30A of the bearing inner peripheral surface 41A of the half bearing 31A is shifted by δ / 2 toward the dividing surface 31C (bearing dividing position) corresponding to the angle θ with respect to the bearing center shaft 30C. The center of curvature 30B of the bearing inner peripheral surface 41B of the split bearing 31B is shifted from the bearing center shaft 30C by δ / 2 toward the dividing surface 31D (bearing dividing position) corresponding to an angle of 180 ° + θ. That is, the center of curvature 30A of the bearing inner peripheral surface 41A of the half bearing 31A is on the side of the split surface 31C (bearing split position) corresponding to the angle θ with respect to the center of curvature 30B of the bearing inner peripheral surface 41B of the half bearing 31B. Is shifted by δ. As a result, a δ discrepancy occurs between the half bearings 31A and 31B mounted on the bearing mounting surfaces 13A and 13B. In the vicinity of the split surface 31C corresponding to the angle θ, the bearing inner peripheral surface 41A of the half bearing 31A. However, in the vicinity of the split surface 31D corresponding to an angle of 180 ° + θ, the bearing inner peripheral surface 41B of the half bearing 31B is half-slipted by δ radially outward from the bearing inner peripheral surface 41B of the half bearing 31B. The split bearing 31A is displaced by δ radially outward from the bearing inner peripheral surface 41A. In each of the drawings including FIG. 2, δ is shown larger than the actual size for convenience of explanation.

  With the above configuration, in the bearing inner peripheral surface 41 of the large end bearing 30, the bearing center extends from a position corresponding to the reference direction (angle 0 °) to a position (divided surface 31 </ b> C) that advances to the angle θ in the rotation direction of the crankpin 18. The distance to the shaft 30C gradually decreases monotonously, and the distance to the bearing center shaft 30C rapidly increases by δ at a position corresponding to the angle θ. The distance from the bearing inner peripheral surface 41 to the bearing center axis 30C gradually decreases monotonically from the position corresponding to the angle θ to the position (divided surface 31D) that advances to the angle 180 ° + θ in the rotation direction of the crankpin 18 and the angle 180 The distance between the bearing inner peripheral surface 41 and the bearing central shaft 30C increases rapidly by δ at a position corresponding to ° + θ. The distance from the bearing inner peripheral surface 41 to the bearing center shaft 30C gradually and monotonously decreases from a position corresponding to an angle of 180 ° + θ to a position that proceeds in the reference direction (angle 0 °) in the rotation direction of the crankpin 18.

  As a result, the clearance (clearance) between the bearing inner peripheral surface 41 and the crank pin 18 in a no-load state in which the center shaft 18C of the crank pin 18 coincides with the bearing center shaft 30C of the large end bearing 30 is a reference. It gradually decreases monotonically from a position corresponding to the direction (angle 0 °) to a position proceeding to the angle θ in the rotation direction of the crankpin 18 and rapidly expands by δ at the position corresponding to the angle θ. Then, the clearance gradually decreases monotonically from the position corresponding to the angle θ to the position that advances to the angle 180 ° + θ in the rotation direction of the crankpin 18, and the clearance rapidly increases by δ at the position corresponding to the angle 180 ° + θ. Then, the clearance gradually decreases monotonously from a position corresponding to the angle 180 ° + θ to a position where the crank pin 18 rotates in the reference direction (angle 0 °).

  When the crank pin 18 rotates with respect to the large end bearing 30, the combustion pressure in the cylinder at the compression top dead center or the inertial force of the piston system at the intake bottom dead center or exhaust bottom dead center along the connecting rod main shaft 10A. When a high load in the direction from the bearing center shaft 30C toward the bearing center shaft 17C (the direction indicated by the arrow F1 in FIGS. 1 and 2) acts on the bearing inner peripheral surface 41 from the crank pin 18 through the lubricating oil, Oil film pressure that balances the high load by a wedge effect in which the lubricating oil is dragged in due to viscosity in a gap tapered between the portion receiving the high load in 41 and the crankpin 18 in the rotation direction of the rotary shaft. (Wedge oil film pressure) is generated. In this case, as shown in FIG. 3, the center axis 18C of the crankpin 18 advances from the bearing center axis 30C of the large end bearing 30 by an angle in the rotational direction of the crankpin 18 from the high load acting direction. The high load and the oil film pressure are balanced in the state of being eccentric in the vertical direction. Further, due to the inertial force of the piston system at the intake top dead center, a high load in the direction from the bearing central shaft 17C to the bearing central shaft 30C (the direction indicated by the arrow F2 in FIGS. 1 and 2) along the connecting rod main shaft 10A is cranked. When acting on the bearing inner peripheral surface 41 from the pin 18 via the lubricating oil, the pin 18 is tapered in the rotational direction of the rotary shaft formed between the portion receiving the high load on the bearing inner peripheral surface 41 and the crank pin 18. An oil film pressure that balances a high load is generated in the gap by a wedge effect in which the lubricating oil is dragged due to viscosity. Also in this case, as shown in FIG. 4, the center axis 18C of the crank pin 18 advances from the bearing center axis 30C of the large end bearing 30 by an angle in the rotational direction of the crank pin 18 from the high load acting direction. The high load and the oil film pressure are balanced in the state of being eccentric in the vertical direction.

  5 and 6, the center shaft 18 </ b> C of the crank pin 18 (rotating shaft) is the large end bearing 30 so that an oil film pressure that is balanced with the load is generated as the load is applied. Therefore, the oil film thickness between the crank pin 18 and the large end bearing 30 is the minimum film at a position advanced by an angle in the rotation direction of the crank pin 18 with respect to the load direction. Thick. When the average clearance between the crankpin 18 and the large end bearing 30 in the no-load state as shown in FIG. 5 is large, the crankpin 18 and the large end in the no-load state as shown in FIG. Compared to the case where the average clearance between the bearings 30 is small, the viscous friction loss when the crankpin 18 rotates with respect to the large-end bearing 30 is reduced, but the oil film pressure generation region due to the wedge effect is narrow. Thus, the amount of eccentricity of the crank pin 18 with respect to the same load is large, and the minimum oil film thickness hmin is small. As a result, there is an increased risk of seizure and wear at high loads. As described above, in the case of a normal perfect circle bearing, although it is advantageous for the viscous friction reduction that the average clearance between the crankpin 18 and the large end bearing 30 is large, the minimum oil film thickness hmin For the increase, it is advantageous that the average clearance between the crankpin 18 and the large end bearing 30 is small, and it is difficult to achieve both reduction of viscous friction and increase of the minimum oil film thickness hmin.

  On the other hand, in the present embodiment, the bearing inner periphery in a no-load state extends from a position corresponding to an angle of 360 ° −θ to a position that advances to the angle θ through the reference direction (angle 0 °) in the rotation direction of the crankpin 18. The clearance between the surface 41 and the crank pin 18 (the distance of the bearing inner peripheral surface 41 relative to the bearing center axis 30C) gradually and monotonously decreases, so that the rotation direction of the crank pin 18 increases from an angle 360 ° -θ to an angle θ. A tapered gap is formed even in a no-load state. Accordingly, when a high load along the reference direction (the direction indicated by the arrow F1 in FIGS. 1 and 2) acts on the bearing inner peripheral surface 41 from the crank pin 18 via the lubricating oil, this tapered gap (FIG. 3). In the range A), the wedge effect in which the lubricating oil is dragged by viscosity can be increased, and the wedge effect in the region contributing to oil film pressure generation can be increased. Therefore, as shown in FIG. 3, in the state where the eccentric amount of the crankpin 18 is small, it is possible to generate an oil film pressure that balances a high load along the reference direction, and it is possible to increase the minimum oil film thickness hmin. In the present embodiment, between the bearing inner peripheral surface 41 and the crank pin 18 in the no-load state from the position corresponding to the angle 180 ° −θ to the position that advances the angle 180 ° + θ in the rotation direction of the crank pin 18. The clearance (the distance from the bearing inner peripheral surface 41 to the bearing center axis 30C) gradually decreases monotonously, so that the taper gap in the rotational direction of the crankpin 18 from the angle 180 ° −θ to the angle 180 ° + θ is in an unloaded state. But formed. Accordingly, when a high load along the direction opposite to the reference direction (the direction indicated by the arrow F2 in FIGS. 1 and 2) acts on the bearing inner peripheral surface 41 from the crank pin 18 via the lubricating oil, this tapered gap It is possible to increase the wedge effect in which the lubricating oil is dragged into (range B in FIG. 4) due to the viscosity, and it is possible to increase the wedge effect in the region contributing to oil film pressure generation. Therefore, as shown in FIG. 4, in the state where the eccentric amount of the crankpin 18 is small, it is possible to generate an oil film pressure that balances a high load along the direction opposite to the reference direction, and to increase the minimum oil film thickness hmin. Can do.

  Further, in the present embodiment, the clearance between the bearing inner peripheral surface 41 and the crank pin 18 is suddenly increased by δ at a position corresponding to the angle θ, so that the rotation direction of the crank pin 18 from the position corresponding to the angle θ. The clearance in the range advanced to (range C in FIG. 3) can be increased, and the clearance in the region that does not contribute to oil film pressure generation can be increased. Then, the clearance between the bearing inner peripheral surface 41 and the crank pin 18 is suddenly increased by δ at a position corresponding to an angle of 180 ° + θ, so that the crank pin 18 rotates in a rotational direction from a position corresponding to an angle of 180 ° + θ. The clearance in the advanced range (range D in FIG. 4) can be increased, and the clearance in the region that does not contribute to oil film pressure generation can be increased.

  As described above, in this embodiment, the wedge effect in the region contributing to the generation of the oil film pressure that balances the high load can be increased, and the eccentric amount of the crankpin 18 is reduced and the minimum oil film thickness hmin is increased. Therefore, seizure and wear at high temperature (low oil viscosity) and high load can be prevented. Since the clearance in the region that does not contribute to oil film pressure generation can be increased, the viscous friction loss when the crankpin 18 rotates with respect to the large end bearing 30 at low temperatures (high oil viscosity) is reduced. can do. Therefore, both viscous friction reduction and minimum oil film thickness hmin increase can be achieved. Furthermore, since the half bearings 31A and 31B themselves can use a half bearing having a general structure as it is, the cost increase is due to processing on the large end portion 13 (large end portion main body 20 and cap 21) side. Just what you need.

  FIG. 7 shows the result of examining the change in the minimum oil film thickness hmin and the viscous friction torque by the fluid lubrication calculation when the predetermined angle θ is changed in the configuration of the present embodiment. When calculating the minimum oil film thickness hmin and the viscous friction torque, the average clearance between the rotating shaft and the bearing is set to 40 μm, and not only the value of θ but also the value of δ is changed in the range of 10 μm to 60 μm. Yes. Furthermore, for comparison, the minimum oil film thickness hmin and viscous friction torque in the case of a perfect circle bearing are also calculated while changing the clearance between the rotating shaft and the bearing in the range of 20 μm to 50 μm, and the calculation results are also shown in FIG. 7 shows. When calculating the minimum oil film thickness hmin, the temperature of the lubricating oil is 120 ° C., the rotational speed of the rotating shaft is 2000 rpm, and the load acting on the bearing from the rotating shaft is 50 kN. High load conditions. When calculating the viscous friction torque, the temperature of the lubricating oil is 20 ° C., the rotational speed of the rotating shaft is 1000 rpm, and the load acting on the bearing from the rotating shaft is 10 kN. Low temperature (high oil viscosity) / low load As a condition.

  As shown in the calculation result of FIG. 7, in the case of θ = 20 ° to θ = 70 °, viscous friction torque is ensured while ensuring a minimum oil film thickness hmin equal to or greater than that of a perfect circle bearing. It can be seen that However, in the case of θ = 10 ° and θ = 80 °, the viscous friction torque is increased as compared with the case of the perfect circle bearing. Therefore, in the configuration of the present embodiment, in order to reduce the viscous friction torque, it is preferable to set the value of θ to a predetermined angle of 20 ° or more and 70 ° or less. In order to further reduce the viscous friction torque, it is preferable to set the value of θ to 20 °.

  FIG. 8 shows the experimental results of measuring the viscous friction torque when θ = 30 °, δ = 20 μm, and the minimum clearance Cmin between the rotating shaft and the bearing at no load is 43.5 μm. For comparison, FIG. 8 also shows the viscous friction torque in the case of a perfect circle bearing measured under the condition that the clearance between the rotating shaft and the bearing is 53.5 μm, and the experimental results are also shown. When measuring the viscous friction torque, the temperature of the lubricating oil is 25 ° C., the rotational speed of the rotating shaft is 1200 rpm, and the load acting on the bearing from the rotating shaft is 10 kN. As shown in the experimental results of FIG. 8, it can be seen that according to the configuration of the present embodiment, the viscous friction torque can be reduced as compared with the case of the perfect circle bearing.

  In the above description, the large-end bearing 30 of the connecting rod 10 has been described as an example of the rotating shaft bearing structure using the radial slide bearing according to the embodiment of the present invention. However, the bearing structure of the rotary shaft using the radial slide bearing according to the present invention can be applied to a bearing of a rotary machine as shown in FIGS. 9 and 10 in addition to the large end bearing 30 of the connecting rod 10. It is. As described above, the rotary shaft bearing structure using the radial slide bearing according to the present invention can be applied to various bearings as long as the rotary shaft is supported by the radial slide bearing through the lubricating oil. is there. In the following description, the same or corresponding components as those shown in FIGS. 1 to 4 are denoted by the same reference numerals, and the components that are not described are the same as those shown in FIGS. is there.

In the configuration example shown in FIGS. 9 and 10, when the following conditions (1) to (3) are satisfied, the load acting on the bearing inner peripheral surface 41 of the radial slide bearing 30 from the rotating shaft 18 through the lubricating oil. Is mainly a downward load (static load) in one direction due to the weight of the rotating portion 50. In this case, as shown in FIG. 10, the reference direction (load acting direction) is set to a vertically downward direction, and the rotation direction of the rotary shaft 18 is set around the bearing center axis 30C with respect to the reference direction (vertically downward direction). The bearing support member 13 is divided into the bearing support member main body 20 and the cap 21 at a position advanced along the angle θ (divided surface 13C) and a position advanced 180 ° + θ (divided surface 13D). Then, one half bearing 31A is mounted on the bearing mounting surface 13A formed on the cap 21, and the other half bearing 31B is mounted on the bearing mounting surface 13B formed on the bearing support member body 20.
(1) The rotating shaft 18 is a horizontal axis, and the weight of the rotating portion 50 is relatively large.
(2) The rotation speed of the rotating portion 50 is relatively low and the inertia (rotation) load component is small.
(3) The rotation direction of the rotating shaft 18 is a fixed direction.

  Also in the configuration example shown in FIG. 10, the center of curvature 30A of the bearing inner peripheral surface 41A of the half bearing 31A is a split surface 31C corresponding to the angle θ with respect to the center of curvature 30B of the bearing inner peripheral surface 41B of the half bearing 31B. By shifting by δ toward the (bearing split position) side, in the vicinity of the split surface 31C corresponding to the angle θ, the bearing inner peripheral surface 41A of the half bearing 31A is more than the bearing inner peripheral surface 41B of the half bearing 31B. It is shifted by δ outward in the radial direction. With this configuration, on the bearing inner peripheral surface 41, the distance from the bearing central shaft 30C gradually increases from a position corresponding to the reference direction (angle 0 °) to a position (divided surface 31C) that advances to the angle θ in the rotation direction of the rotary shaft 18. The distance to the bearing center shaft 30C increases abruptly by δ at a position corresponding to the angle θ, and proceeds from the position corresponding to the angle 360 ° −θ to the reference direction (angle 0 °) in the rotational direction of the rotary shaft 18. The distance to the bearing center shaft 30C gradually and monotonously decreases to the position. As a result, the bearing inner peripheral surface 41 and the rotating shaft 18 in a no-load state from a position corresponding to an angle of 360 ° −θ to a position proceeding to the angle θ through the reference direction (angle 0 °) in the rotating direction of the rotating shaft 18. The clearance between and gradually decreases. Therefore, when a load along the reference direction (the direction indicated by the arrow F1 in FIGS. 9 and 10) acts on the bearing inner peripheral surface 41 from the rotating shaft 18 via the lubricating oil, the angle 360 ° to θ from the angle θ. The wedge effect in the region contributing to the generation of the oil film pressure can be increased. Accordingly, an oil film pressure that balances a high load along the reference direction can be generated in a state where the eccentric amount of the rotating shaft 18 is small, and the minimum oil film thickness hmin can be increased. Furthermore, the clearance between the bearing inner peripheral surface 41 and the rotary shaft 18 at the position corresponding to the angle θ rapidly increases by δ, so that the oil film advances in the rotational direction of the rotary shaft 18 from the position corresponding to the angle θ. The clearance in the region that does not contribute to pressure generation can be increased. Therefore, it is possible to reduce viscous friction loss when the rotating shaft 18 rotates with respect to the radial slide bearing 30. In that case, it is preferable to set the value of θ to a predetermined angle of 20 ° or more and 70 ° or less.

  Further, when the load acting on the bearing inner peripheral surface 41 of the radial slide bearing 30 from the rotating shaft 18 through the lubricating oil is mainly in one direction, for example, as shown in FIGS. 11 and 12, the half bearing 31A The curvature of the bearing inner peripheral surface 41A can be made smaller than the curvature of the bearing inner peripheral surface 41B of the half bearing 31B. In the configuration example shown in FIG. 11, the radius of curvature of the bearing inner peripheral surface 41A of the half bearing 31A is larger by δ / 2 than the radius of curvature of the bearing inner peripheral surface 41B of the half bearing 31B. The center of curvature 30A of the peripheral surface 41A is deviated by δ / 2 toward the split surface 31C (bearing split position) side corresponding to the angle θ with respect to the center of curvature 30B of the bearing inner peripheral surface 41B of the half bearing 31B. Thus, in the vicinity of the split surface 31C corresponding to the angle θ, the bearing inner peripheral surface 41A of the half bearing 31A is displaced radially outward from the bearing inner peripheral surface 41B of the half bearing 31B by δ. In the configuration example shown in FIG. 12, the center of curvature 30A of the bearing inner peripheral surface 41A of the half bearing 31A coincides with the center of curvature 30B of the bearing inner peripheral surface 41B of the half bearing 31B. The radius of curvature of the bearing inner peripheral surface 41A is larger than the radius of curvature of the bearing inner peripheral surface 41B of the half bearing 31B by δ, so that the bearing inner circumference of the half bearing 31A is near the split surface 31C corresponding to the angle θ. The surface 41A is displaced by δ radially outward from the bearing inner peripheral surface 41B of the half bearing 31B.

  Even in the configuration examples shown in FIGS. 11 and 12, when a load along the reference direction (the direction indicated by the arrow F1 in FIGS. 11 and 12) acts on the bearing inner peripheral surface 41 from the rotary shaft 18 via the lubricating oil, Since the wedge effect in the region contributing to the generation of the oil film pressure can be increased, the oil film pressure that balances the high load along the reference direction can be generated with the eccentric amount of the rotating shaft 18 being small, and the minimum oil film The thickness hmin can be increased. Furthermore, since the clearance between the bearing inner peripheral surface 41 and the rotary shaft 18 is suddenly increased by δ at a position corresponding to the angle θ, the clearance in a region that does not contribute to oil film pressure generation can be increased. The viscous friction loss when the rotating shaft 18 rotates with respect to the radial slide bearing 30 can be reduced. In that case, it is preferable to set the value of θ to a predetermined angle of 20 ° or more and 70 ° or less.

  As mentioned above, although the form for implementing this invention was demonstrated, this invention is not limited to such embodiment at all, and it can implement with a various form in the range which does not deviate from the summary of this invention. Of course.

  10 connecting rod, 10A connecting rod main shaft, 11 piston, 12 small end, 13 large end (bearing support member), 13A, 13B bearing mounting surface, 13C, 13D, 31C, 31D split surface, 14 column portion, 15 piston pin , 16 Small end through hole, 17 Small end bearing, 17C, 30C Bearing central axis, 18 Crank pin (Rotating shaft), 18C Central axis, 19 Large end through hole, 20 Large end main body (Bearing support member main body ), 21 cap, 30 large end bearing (radial sliding bearing), 30A, 30B curvature center, 31A, 31B half bearing, 41, 41A, 41B bearing inner peripheral surface, 50 rotating parts.

Claims (10)

A bearing structure of a rotating shaft that receives a reciprocating load along the radial direction of the rotating shaft via lubricating oil filled in a clearance between the bearing inner peripheral surface of the radial slide bearing and the rotating shaft,
On the bearing inner surface,
When θ is a predetermined angle larger than 0 ° and smaller than 90 °,
The distance to the bearing center axis gradually decreases to a position advanced by an angle θ in the rotational direction of the rotating shaft with reference to a direction perpendicular to the bearing center axis and along one side of the reciprocating load,
The distance to the bearing center axis suddenly increases at a position corresponding to the angle θ,
The distance to the bearing center axis gradually decreases from a position corresponding to an angle of 180 ° -θ to a position that advances to an angle of 180 ° + θ in the rotational direction of the rotary shaft,
The distance to the bearing center axis suddenly increases at a position corresponding to an angle of 180 ° + θ,
A rotating shaft bearing structure in which the distance to the bearing center axis gradually decreases from a position corresponding to an angle of 360 ° -θ to a position that proceeds to the reference in the rotation direction of the rotating shaft. A bearing structure of a rotating shaft that receives a reciprocating load along the radial direction of the rotating shaft via lubricating oil filled in a clearance between the bearing inner peripheral surface of the radial slide bearing and the rotating shaft,
When θ is a predetermined angle larger than 0 ° and smaller than 90 °,
Two halved bearings divided into a position advanced by an angle θ and a position advanced by an angle of 180 ° + θ in the rotational direction of the rotary shaft with reference to a direction perpendicular to the bearing central axis and along one side of the reciprocating load A radial plain bearing is configured including
One half bearing forms a bearing inner peripheral surface from a position corresponding to the angle θ to a position that advances to an angle of 180 ° + θ in the rotation direction of the rotary shaft,
The other half bearing forms a bearing inner peripheral surface from a position corresponding to an angle of 180 ° + θ to a position that advances to the angle θ in the rotation direction of the rotary shaft,
Near the bearing split position corresponding to the angle θ, the bearing inner peripheral surface of one half bearing is shifted radially outward from the bearing inner peripheral surface of the other half bearing,
A bearing structure of a rotating shaft, wherein a bearing inner peripheral surface of the other half bearing is displaced radially outward from a bearing inner peripheral surface of one half bearing in the vicinity of a bearing split position corresponding to an angle of 180 ° + θ. The bearing structure of the rotating shaft that supports the crank pin through the lubricating oil filled in the clearance between the bearing inner peripheral surface of the radial slide bearing attached to the large end of the connecting rod and the crank pin of the crank shaft. There,
On the bearing inner surface,
When θ is a predetermined angle larger than 0 ° and smaller than 90 °,
The distance to the bearing center axis gradually decreases to a position that advances the angle θ in the rotation direction of the crankpin with respect to the direction perpendicular to the bearing center axis and from the bearing center axis toward the small end center axis of the connecting rod.
The distance to the bearing center axis suddenly increases at a position corresponding to the angle θ,
The distance to the bearing center axis gradually decreases from a position corresponding to an angle of 180 ° -θ to a position that advances to an angle of 180 ° + θ in the rotation direction of the crankpin,
The distance to the bearing center axis suddenly increases at a position corresponding to an angle of 180 ° + θ,
A rotating shaft bearing structure in which the distance from the bearing central axis gradually decreases from a position corresponding to an angle of 360 ° -θ to a position that proceeds to the reference in the rotation direction of the crankpin. The bearing structure of the rotating shaft that supports the crank pin through the lubricating oil filled in the clearance between the bearing inner peripheral surface of the radial slide bearing attached to the large end of the connecting rod and the crank pin of the crank shaft. There,
When θ is a predetermined angle larger than 0 ° and smaller than 90 °,
Divided into a position advanced by an angle θ in the rotation direction of the crank pin and a position advanced by an angle of 180 ° + θ, with reference to the direction perpendicular to the bearing central axis and from the bearing central axis toward the central axis of the small end of the connecting rod A radial plain bearing is constructed including two half bearings,
One half bearing forms a bearing inner peripheral surface from a position corresponding to the angle θ to a position that advances to an angle of 180 ° + θ in the rotation direction of the crankpin,
The other half bearing forms a bearing inner peripheral surface from a position corresponding to an angle of 180 ° + θ to a position that advances to an angle θ in the rotation direction of the crankpin,
Near the bearing split position corresponding to the angle θ, the bearing inner peripheral surface of one half bearing is shifted radially outward from the bearing inner peripheral surface of the other half bearing,
A bearing structure of a rotating shaft, wherein a bearing inner peripheral surface of the other half bearing is displaced radially outward from a bearing inner peripheral surface of one half bearing in the vicinity of a bearing split position corresponding to an angle of 180 ° + θ. A bearing structure for a rotating shaft according to claim 2 or 4,
The curvature of the inner peripheral surface of the two half bearings is equal,
The bearing structure of the rotating shaft in which the center of curvature of the inner circumferential surface of the half bearing is shifted to the bearing split position corresponding to the angle θ with respect to the center of curvature of the inner circumferential surface of the other half bearing. . A bearing structure of a rotating shaft that receives a load along a radial direction of the rotating shaft through lubricating oil filled in a clearance between the bearing inner peripheral surface of the radial slide bearing and the rotating shaft,
On the bearing inner surface,
When θ is a predetermined angle larger than 0 ° and smaller than 90 °,
The distance to the bearing center axis gradually decreases to a position advanced by an angle θ in the rotational direction of the rotating shaft, based on the direction perpendicular to the bearing center axis and along the load,
The distance to the bearing center axis suddenly increases at a position corresponding to the angle θ,
A rotating shaft bearing structure in which the distance to the bearing center axis gradually decreases from a position corresponding to an angle of 360 ° -θ to a position that proceeds to the reference in the rotation direction of the rotating shaft. A bearing structure of a rotating shaft that receives a load along a radial direction of the rotating shaft through lubricating oil filled in a clearance between the bearing inner peripheral surface of the radial slide bearing and the rotating shaft,
When θ is a predetermined angle larger than 0 ° and smaller than 90 °,
Radial including two halved bearings divided into a position advanced by an angle θ and a position advanced by an angle of 180 ° + θ in the rotational direction of the rotating shaft with reference to the direction along the load perpendicular to the bearing center axis A plain bearing is constructed,
One half bearing forms a bearing inner peripheral surface from a position corresponding to the angle θ to a position that advances to an angle of 180 ° + θ in the rotation direction of the rotary shaft,
The other half bearing forms a bearing inner peripheral surface from a position corresponding to an angle of 180 ° + θ to a position that advances to the angle θ in the rotation direction of the rotary shaft,
A bearing structure of a rotating shaft, wherein a bearing inner peripheral surface of one half bearing is displaced radially outward from a bearing inner peripheral surface of the other half bearing in the vicinity of a bearing split position corresponding to an angle θ. The rotary shaft bearing structure according to claim 7,
The bearing structure of the rotating shaft in which the center of curvature of the inner circumferential surface of the half bearing is shifted to the bearing split position corresponding to the angle θ with respect to the center of curvature of the inner circumferential surface of the other half bearing. . A bearing structure for a rotating shaft according to claim 7 or 8,
The bearing structure of a rotating shaft, wherein the curvature of the inner circumferential surface of one half bearing is smaller than the curvature of the inner circumferential surface of the other half bearing. It is a bearing structure of a rotating shaft given in any 1 paragraph of Claims 1-9,
θ is a rotating shaft bearing structure having a predetermined angle of 20 ° or more and 70 ° or less.

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