1
FORM 2
THE PATENTS ACT, 1970
(39 of 1970)
&
THE PATENTS RULES, 2003
COMPLETE SPECIFICATION
[See section 10, Rule 13]
PLAIN BEARING STRUCTURE AND SCROLL COMPRESSOR;
MITSUBISHI ELECTRIC CORPORATION, A CORPORATION ORGANISED AND
EXISTING UNDER THE LAWS OF JAPAN, WHOSE ADDRESS IS 7-3,
MARUNOUCHI 2-CHOME, CHIYODA-KU, TOKYO 1008310, JAPAN
THE FOLLOWING SPECIFICATION PARTICULARLY DESCRIBES THE INVENTION
AND THE MANNER IN WHICH IT IS TO BE PERFORMED
2
DESCRIPTION
Technical Field
[0001]
The present disclosure relates to a plain bearing structure that supports a
rotation shaft provided to drive a scroll portion of a scroll compressor, 5 and to the scroll
compressor including the plain bearing structure.
Background Art
[0002]
In the past, scroll compressors have been made to include a rotation shaft that
10 drives an orbiting scroll and a plain bearing that supports a variable radial load acting
on the rotation shaft. The plain bearing of the scroll compressor is lubricated with a
fluid such as a refrigerating machine oil, and a fluid film called an oil film is formed in
a gap between the plain bearing and the rotation shaft. During operation of the scroll
compressor, the rotation shaft is rotated such that the center of the rotation shaft is
15 displaced from the center of the plain bearing by a radial load acting on the rotation
shaft. Therefore, in the gap between the rotation shaft and the plain bearing, a socalled
wedge oil film region is formed in such a manner as to gradually decrease a
circumferential gap in a rotation direction. In that region, since a refrigerating
machine oil is drawn into a narrow space as the rotation shaft is rotated, a pressure
20 (oil film pressure) is caused by the oil film, and the radial load acting on the rotation
shaft can be supported. This is referred to as a wedge effect.
[0003]
When the rotation speed of the rotation shaft in the scroll compressor is
sufficiently high, a high oil film pressure generates in the wedge oil film region in the
25 gap between the plain bearing and the rotation shaft. Thus, the rotation shaft and
the plain bearing are separated from each other by the oil film having a sufficient
thickness, and do not directly contact each other. When the rotation speed of the
rotation shaft decreases, the oil film pressure in the wedge oil film region decreases,
and in the gap between the rotation shaft and the plain bearing, the oil film is easily
30 lost. Therefore, the rotation shaft and the plain bearing may come into contact (solid
3
contact) with each other, thus causing friction and abrasion, and seizure may occur at
the plain bearing.
[0004]
In recent years, because of high thermal insulation of buildings or other factors,
scroll compressors for use in air-conditioning apparatuses have often 5 been required
to operate with a low load and in a low rotation speed range. Also, in the case
where the outside air temperature is low, it may be necessary to perform a cooling
operation with a low load. Therefore, it is necessary for the plain bearing in the scroll
compressor to ensure a sufficient oil film pressure even when the rotation speed is
10 low and to reduce occurrence of seizure.
[0005]
In a plain bearing structure of an existing scroll compressor, a plurality of Vshaped
grooves are formed to extend in an axial direction and uniformly distributed
over the entire surface of a sliding portion of a rotation shaft in order to prevent
15 depletion of an oil film and to reduce occurrence of seizure that is caused by contact
with the rotation shaft. The plain bearing structure is configured to generate a
pressure (dynamic pressure) by a pumping action of drawing the refrigerating
machine oil along the grooves (for example, Patent Literature 1).
Citation List
20 Patent Literature
[0006]
Patent Literature 1: Japanese Unexamined Patent Application Publication No.
4-370388
Summary of Invention
25 Technical Problem
[0007]
In the plain bearing structure of the scroll compressor disclosed in Patent
Literature 1, the grooves are uniformly provided over the entire perimeter of the outer
circumferential surface of the rotation shaft. Thus, by the pumping action, a dynamic
30 pressure is generated at the entire perimeter of the rotation shaft. Therefore,
4
pressure generation effects by the grooves in the entire perimeter of the outer
circumferential surface cancel out each other. Thus, it is hard to increase the oil film
pressure in the gap between the rotation shaft and the plain bearing, and it is not
possible to reduce occurrence of seizure.
5 [0008]
The present disclosure is applied to solve the above problem, and relates to a
plain bearing structure and a scroll compressor that effectively increase an oil film
pressure in a gap between a rotation shaft and a plain bearing of the scroll
compressor, and that reduce occurrence of seizure at the rotation shaft and the plain
10 bearing.
Solution to Problem
[0009]
A plain bearing structure according to an embodiment of the present disclosure
includes a rotation shaft configured to drive an orbiting scroll in a scroll compressor;
15 and a plain bearing that supports a radial load on the rotation shaft. The rotation
shaft includes a groove formed in an outer circumferential surface of the rotation shaft
and extending in an axial direction of the rotation shaft, the outer circumferential
surface facing the plain bearing. The groove is configured to cause a pumping
action to occur. The outer circumferential surface of the rotation shaft includes a
20 region P and a region Q. The region P is defined in a predetermined angular range
in a circumferential direction of the rotation shaft. The region Q being a region of the
outer circumferential surface that is other than the region P. The ratio of the area of
part of the groove that is located in the region P in the outer circumferential surface to
the area of the region P is higher than the ratio of the area of part of the groove that is
25 located in the region Q in the outer circumferential surface to the area of the region Q.
[0010]
A scroll compressor according to another embodiment of the present disclosure
includes the above plain bearing structure.
Advantageous Effects of Invention
30 [0011]
5
According to the embodiments of the present disclosure, dynamic pressures
are generated by a pumping action at the groove or groooves formed in a region
where an oil film pressure is generated by a wedge effect, and act as an additional
resistance to a variable radial load on the rotation shaft, and the thickness of an oil
film between the rotation shaft and the plain bearing is increased. 5 Moreover, in a
region where the oil film thickness is great, axial-direction grooves can be distributed
such that the number of the axial grooves per unit area is further reduced or is zero,
thereby solving the problem in which dynamic pressures generated by the pumping
action cancel out each other. Therefore, in the plain bearing structure and the scroll
10 compressor, it is possible to further increase the oil film pressure and more effectively
reduce occurrence of seizure than an existing plain bearing structure and an existing
scroll compressor.
Brief Description of Drawings
[0012]
15 [Fig. 1] Fig. 1 is an explanatory view of a sectional configuration of a scroll
compressor 100 according to Embodiment 1.
[Fig. 2] Fig. 2 is a sectional view of a plain bearing structure 50 and the vicinity
of the plan bearing structure 50 in the scroll compressor 100 according to
Embodiment 1.
20 [Fig. 3] Fig. 3 is an explanatory view of a section of the plain bearing structure
50 in the scroll compressor 100 according to Embodiment 1.
[Fig. 4] Fig. 4 is a schematic diagram illustrating the plain bearing structure 50
in the scroll compressor 100 according to Embodiment 1.
[Fig. 5] Fig. 5 is an explanatory view of a sectional configuration perpendicular
25 to an axial direction of the plain bearing structure 50 as illustrated in Fig. 4.
[Fig. 6] Fig. 6 illustrates the result of analysis of hydrodynamic lubrication in the
case where a region P where grooves 20 are provided and the shape of the grooves
20 in the plain bearing structure 50 according to Embodiment 1 are changed.
[Fig. 7] Fig. 7 indicates analysis conditions for use in the analysis the result of
30 which is indicated in Fig. 6.
6
[Fig. 8] Fig. 8 illustrates a modification of the grooves 20 provided in an outer
circumferential surface of a rotation shaft 6 in the plain bearing structure 50 according
to Embodiment 1.
[Fig. 9] Fig. 9 illustrates another modification of the grooves 20 in the outer
circumferential surface of the rotation shaft 6 in the plain bearing 5 structure 50
according to Embodiment 1.
[Fig. 10] Fig. 10 illustrates still another modification of the grooves 20 disposed
in the outer circumferential surface of the rotation shaft 6 in the plain bearing structure
50 according to Embodiment 1.
10 [Fig. 11] Fig. 11 illustrates a section of the outer circumference of the rotation
shaft 6 according to Embodiment 1 as the outer circumference is developed.
[Fig. 12] Fig. 12 illustrates another section of the outer circumference of the
rotation shaft 6 according to Embodiment 1 as the outer circumference is developed.
[Fig. 13] Fig. 13 is a schematic diagram for describing a plain bearing structure
15 250 of a scroll compressor 100 according to Embodiment 2.
[Fig. 14] Fig. 14 is a schematic diagram for describing a plain bearing structure
350 of a scroll compressor 100 according to Embodiment 3.
Description of Embodiments
[0013]
20 Embodiments of a plain bearing structure in a scroll compressor will be
described below. Configurations illustrated in the figures are examples and their
illustrations are not limiting. In each of the figures, components that are the same as
or equivalent to those in a previous figure or figures are denoted by the same
reference signs. The same is true of the entire text of the specification.
25 Configurations of the components as descried in the entire text of the specification
are merely examples, and the configurations of the components are not limited to the
configurations described in the present disclosure. In particular, the way of
combining components is not limited to combining of components according to the
same embodiment. That is, it is possible to apply a component according to an
30 embodiment to another embodiment. Regarding a plurality of components that are
7
of the same kind and distinguished from each other by suffixes, in the case where the
components do not particularly need to be distinguished from each other or to be
specified, the suffixes may be omitted. In the figures, the relationship in size
between components may be different from that between actual components.
5 [0014]
Embodiment 1
Fig. 1 is an explanatory view of a sectional configuration of a scroll compressor
100 according to Embodiment 1. As illustrated in Fig. 1, the scroll compressor 100
includes a scroll compression mechanism 60 in which a fixed scroll 1 and an orbiting
10 scroll are formed in combination. In the fixed scroll 1, a spiral scroll lap is formed at
a lower surface of a base plate 1a, and in the orbiting scroll 2, a spiral scroll lap is
formed at an upper surface of a base plate 2a. The scroll lap of the fixed scroll 1
and the scroll lap of the orbiting scroll 2 are combined with each other, thereby
forming a compression chamber 5.
15 [0015]
The fixed scroll 1 and the orbiting scroll 2 are provided in an upper region in a
pressure container 11. The pressure container 11 includes an oil sump 12 at a
bottom portion of the pressure container 11. To side walls of the pressure container
11, a side a refrigerant suction pipe 13 and a refrigerant discharge pipe 14 are
20 connected. An end of the refrigerant suction pipe 13 communicates with a suction
port 3 located within the pressure container 11, and an end of the refrigerant
discharge pipe 14 communicates with a discharge port 4 located within the pressure
container 11. Refrigerant from a refrigerant circuit flows from the refrigerant suction
pipe 13, through the suction port 3, into the compression chamber 5 in the scroll
25 compression mechanism 60, and the refrigerant is compressed and then discharged
from the discharge port 4 into the refrigerant discharge pipe 14 to flow to the
refrigerant circuit.
[0016]
In a lower surface of the base plate 2a of the orbiting scroll 2, an eccentric hole
30 2b is provided. In the eccentric hole 2b, a rotation shaft 6 is fit. The rotation shaft 6
8
is connected to an electric motor 9 that is provided at a central portion of the pressure
container 11 in a height direction thereof. The rotation shaft 6 transmits a driving
force of the electric motor 9 to the orbiting scroll 2. The electric motor 9 is located
between an upper housing 8a and a lower housing 8b of a housing 8. The electric
motor 9 includes a rotor 9a and a stator 9b. The rotor 9a is 5 fixed to an outer
circumferential surface of a main shaft 6a. The stator 9b is fixed to an inner
circumferential surface of the pressure container 12, and surrounds the rotor 9a, with
a predetermined gap provided between the stator 9b and the rotor 9a.
[0017]
10 The rotation shaft 6 includes an eccentric shaft 6b on an upper end of the main
shaft 6a. The eccentric shaft 6b is eccentric to the main shaft 6a, that is, the center
of the eccentric shaft 6b is offset from that of the main shaft 6a. The main shaft 6a is
supported by a main bearing 15a provided in the upper housing 8a, which is fixed to
an inner surface of the side wall of the pressure container 11. The main bearing 15a
15 supports the main shaft 6a such that the main shaft 6a is freely rotatable. To an
upper end portion of the upper housing 8a, the fixed scroll 1 is fixed. The orbiting
scroll 2 is located between the fixed scroll 1 and the upper housing 8a. Between the
orbiting scroll 2 and the upper housing 8a, an Oldham ring 10 is provided. The
Oldham ring 10 causes the orbiting scroll 2 to be turned, while preventing the orbiting
20 scroll 2 from rotating on the axis thereof.
[0018]
A minor bearing 16 is provided in the lower housing 8b, which is fixed to the
inner surface of the side wall of the pressure container 11. The minor bearing 16 is
located below the electric motor 9, and supports the main shaft 6a such that the main
25 shaft 6a is freely rotatable. That is, in a region located above the electric motor 9,
the main shaft 6a of the rotation shaft 6 is rotatably supported by the upper housing
8a, and in a region located below the electric motor 9, the main shaft 6a is supported
by the lower housing 8b.
[0019]
9
Fig. 2 is a sectional view of a plain bearing structure 50 and the vicinity of the
plain bearing structure 50 in the scroll compressor 100 according to Embodiment 1.
In Embodiment 1, the scroll compressor 100 includes a plain bearing structure 50a
and a plain bearing structure 50b. The plain bearing structure 50a is provided at a
location where the upper housing 8a supports the rotation shaft 5 6, and the plain
bearing structure 50b is provided at a location where the orbiting scroll 2 and the
rotation shaft 6 are connected to each other. In the plain bearing structure 50a, the
main bearing 15a, which is fixed to a central hole of the upper housing 8a, and the
main shaft 6a of the rotation shaft 6 are fit. In the plain bearing structure 50b, an
10 orbiting bearing 15b fixed to the eccentric hole 2b of the orbiting scroll 2 and the
eccentric shaft 6b of the rotation shaft 6 are fit. In the following description, the main
bearing 15a and the orbiting bearing 15b may be collectively referred to as a plain
bearing 15.
[0020]
15 The main bearing 15a is fixed to the hole formed in the central portion of the
upper housing 8a by press-fitting, and supports the main shaft 6a such that the main
shaft 6a is freely slidable. The structure of the main bearing 15a is not limited to the
structure as illustrated in Fig. 1. For example, the hole of the upper housing 8a itself
may serve as the main bearing 15a, or even if the main bearing 15a is not provided in
20 the upper housing 8a, the main bearing 15a may be fixed by a method other than
press-fitting.
[0021]
The orbiting bearing 15b is fixed inside the eccentric hole 2b of the orbiting
scroll 2 by press-fitting and is connected to the eccentric shaft 6b such that the
25 eccentric shaft 6b can slide freely. The configuration of the orbiting bearing 15b is
not limited to the configuration illustrated in Fig. 1. For example, the eccentric hole
2b itself may serve as the orbiting bearing 15b, or even in the case where the orbiting
bearing 15b and the eccentric hole 2b are separate elements, the orbiting bearing 15b
may be fixed by a method other than press-fitting.
30 [0022]
10
An axial-direction oil supply hole 7a is formed to extend through the rotation
shaft 6 in the axial direction of the rotation shaft 6. A lower end of the rotation shaft 6
is fit to a pump 7b located below the electric motor 9. A lower opening port 7e of the
pump 7b is immersed in refrigerating machine oil in the oil sump 12 in the lower
portion of the pressure container 11. The axial-direction oil supply 5 hole 7a, which
extends through the rotation shaft 6 in the axial direction, communicates with a radial
oil supply outlet 7c that is formed as an opening in a surface of the main shaft 6a that
faces the main bearing 15a. The axial-direction oil supply hole 7a also
communicates with an oil supply outlet 7d that is formed as an opening in an upper
10 end face of the rotation shaft 6. The refrigerating machine oil drawn up from the oil
sump 12 by the pump 7b flows through the axial-direction oil supply hole 7a and
branches into refrigerating machine oils that flow into the radial oil supply outlet 7c
and the oil supply outlet 7d, and these refrigerating machine oils are supplied to the
plain bearing structures 50a and 50b. In such a manner, the axial-direction oil
15 supply hole 7a, the pump 7b, the radial oil supply outlet 7c, and the oil supply outlet
7d form a mechanism that supplies oil to the plain bearing 15 in the scroll compressor
100.
[0023]
(Operation of Scroll Compressor 100)
20 Next, the operation of the scroll compressor 100 will be described. When the
electric motor 9 is driven, the rotation shaft 6 is rotated along with the rotor 9a. The
driving force of the electric motor 9 is transmitted to the orbiting scroll 2 by the rotation
shaft 6. The orbiting scroll 2 is turned while being prevented by the Oldham ring 10
from rotating on the axis of the orbiting scroll 2. As a result, the compression
25 chamber 5, which is formed by combing the scroll lap of the orbiting scroll 2 and the
scroll lap of the fixed scroll 1, is moved toward the center while gradually decreasing
the volume of the compression chamber 5. The pressure of the refrigerant that has
been sucked from the suction port 3 into the compression chamber 5 after having
passed through the refrigerant suction pipe 13 gradually increases, and then, after
30 passing though the discharge port 4, the refrigerant is pumped from the compression
11
chamber 5 to flow out from the refrigerant discharge pipe 14 to a refrigerant pipe
located outside the scroll compressor 100.
[0024]
In the above sequential operations, a gas load due to compression of
refrigerant gas and a centrifugal force due to the turning of the 5 orbiting scroll 2
primarily act on the rotation shaft 6. The resultant of the gas load and the centrifugal
force acts as a variable radial load W that rotates in synchronization with the rotation
of the rotation shaft 6.
[0025]
10 (Functions of Plain Bearing Structure 50 According to Embodiment 1)
Fig. 3 is an explanatory view of a section of the plain bearing structure 50 in the
scroll compressor 100 according to Embodiment 1. In Fig. 3, grooves 20 formed in
the plain bearing structure 50 are omitted. In other words, Fig. 3 illustrates the plain
bearing structure 50 in which grooves 20 are not formed in an outer circumferential
15 surface of the rotation shaft 6 that faces the plain bearing 15. When the variable
radial load W acts on the rotation shaft 6, the rotation shaft 6 is rotated by the variable
radial load W, with the center of the rotation shaft 6 displaced from the axial center of
the plain bearing 15. A region interposed between the rotation shaft 6 and the plain
bearing 15 is filled with the refrigerating machine oil that is supplied through the radial
20 oil supply outlet 7c or the oil supply outlet 7d to the plain bearing 15, and an oil film 18
is formed in the region. It should be noted that a region of a circumferential flow
passage that gradually narrows in the rotation direction in such a manner as to be
shaped into a wedge because of eccentricity of the rotation shaft 6 will be referred to
as a wedge region 18a, and a region of the circumferential flow passage that
25 gradually broadens in the rotation direction to be shaped into an inverted wedge, in
contrast to the wedge region 18a, will be referred to as an inverted wedge region 18b.
In the wedge region 18a, because the refrigerating machine oil is drawn into a narrow
region by the rotation of the rotation shaft 6, an oil film pressure 19 is generated.
When the oil film pressure 19 corresponding to the variable radial load W is
30 generated, the rotation shaft 6 and the plain bearing 15 are separated from each
12
other by the oil film 18, and the rotation shaft 6 and the plain bearing 15 are rotated
without directly contacting each other.
[0026]
A phase difference between a direction in which the variable radial load W acts
on the rotation shaft 6 in the scroll compressor 100 and a circumferential 5 position
where the surface of the rotation shaft 6 is located is kept at a predetermined value.
It should be noted that a circumferential position where a line L0 of action of the
variable radial load that starts at the axis center of the rotation shaft 6 intersects with
the outer circumferential surface of the rotation shaft 6 is defined as the origin point
10 for the circumferential angular position at the outer circumferential surface of the
rotation shaft. For a direction angle , the opposite direction to the rotation direction
of the rotation shaft is positive, and the origin point is defined as 0 degrees.
Therefore,  falls within the range of 0 degrees to 360 degrees.
[0027]
15 Fig. 4 is a schematic diagram illustrating the plain bearing structure 50 in the
scroll compressor 100 according to Embodiment 1. Fig. 5 is an explanatory view of
a sectional configuration perpendicular to the axial direction of the plain bearing
structure 50 as illustrated in Fig. 4. The plain bearing structure 50 in the scroll
compressor 100 according to Embodiment 1 has the grooves 20 that are formed in
20 the surface of the rotation shaft 6 that faces the plain bearing 15, in such a manner as
to extend in the axial direction of the rotation shaft 6.
[0028]
In Embodiment 1, in a circumferential angular range of 0 degrees to  of the
outer circumferential surface of the rotation shaft 6, the ratio of the area of grooves 20
25 extending in the axial direction to the surface area in the circumferential angular
range is higher than that of the grooves 20 formed in the other portion of the outer
circumferential surface of the rotation shaft 6 to the surface area of the other portion.
A region located in the circumferential angular range of 0 degrees to  in the outer
circumferential surface of the rotation shaft 6 will referred to as a region P. A region
30 of the outer circumferential surface of the rotation shaft 6 that is other than the region
13
P will be referred to as a region Q. Fig. 4 illustrates by way of example the case
where the ratio of the area of the grooves 20 in the region Q of the outer
circumferential surface of the rotation shaft 6 to the surface area is 0. That is,
referring to Fig. 4, the grooves 20 are provided only in the region P of the outer
circumferential surface of the rotation shaft 6 and are not provided 5 in the region Q,
which is other than the region P. Fig. 5 illustrates by way of example the plain
bearing structure 50 in which  is 90 degrees and the ratio of the area of grooves 20
in a region of the outer circumference surface that is other than the circumferential
angular range where 0 degrees    90 degrees to the surface area is 0.
10 [0029]
As illustrated in Fig. 4, the grooves 20 extending in the axial direction are bent
in the outer circumferential surface of the rotation shaft 6. To be more specific, each
of the grooves 20 includes two flow passages that extend from their opposite ends in
the axial direction to a center portion of the groove 20 such that the closer portions of
15 the flow passages to the central portion of the groove 20, the more greatly the
portions are retreated in the opposite direction to the rotation direction of the rotation
shaft 6, and the flow passages join each other at a bent portion 21. In other words,
the groove 20 is bent and V-shaped in the rotation direction of the rotation shaft 6.
Because the groove 20 extending in the axial direction is bent, when the rotation shaft
20 6 is rotated, the refrigerating machine oil flow from the opposite ends of the groove 20
toward the bent portion 21 along an inner wall surface of the groove 20. To be more
specific, in the plain bearing structure 50, on the surface of the rotation shaft 6 that
faces the plain bearing 15, the refrigerating machine oil flows from the opposite ends
toward the central portion. The flows of the refrigerating machine oil join each other
25 at the bent portion 21 of the groove 20, and a pressure (dynamic pressure) is
generated at the bent portion 21 by compression of the refrigerating machine oil.
This phenomenon is referred to as a pumping effect due to dynamic pressure
grooves.
[0030]
14
As illustrated in Figs. 4 and 5, the rotation shaft 6 has the axial-direction oil
supply hole 7a that is formed in the central portion of the rotation shaft 6 to extend
therethrough in the axial direction. As illustrated in Fig. 5, the axial-direction oil
supply hole 7a and the outer circumferential surface of the rotation shaft 6 have the
radial oil supply outlet 7c that extends from the axis center in the 5 radial direction.
Preferably, part of the outer circumferential surface of the rotation shaft 6 where the
radial oil supply outlet 7c has an opening portion should be located in a region of the
outer circumferential surface of the rotation shaft 6 that is other than the
circumferential angular range of 0 degrees to , that is, the region Q. Alternatively,
10 as in the plain bearing structure 50b, it is preferable that the oil supply outlet 7d be
provided in an axial-direction end portion of the eccentric shaft 6b, and refrigerating
machine oil be supplied to the gap between the orbiting bearing 15b and the eccentric
shaft 6b.
[0031]
15 In the plain bearing structure 50 according to Embodiment 1, the grooves 20
can be arranged such that the ratio of the grooves 20 to the outer circumferential
surface of the rotation shaft 6 is high in the wedge region 18a of the oil film 18 and is
low in the region other than the wedge region 18a. Thus, in the wedge region 18a
where the oil film has a smaller thickness, a dynamic pressure is generated by the
20 pumping action at the grooves 20, and is added to an oil film pressure generated by
the rotation of the rotation shaft 6, and the resultant of the dynamic pressure and the
oil film pressure acts as a resistance to the variable radial load W. Because the oil
film thickness in the wedge region 18a increases, occurrence of seizure at the plain
bearing structure 50 can be reduced.
25 [0032]
In the related art, grooves 20 are provided over the whole outer circumferential
surface of the rotation shaft 6, and a dynamic pressure by the pumping action at the
grooves 20 is generated over the whole area of an oil film 18. Therefore, a dynamic
pressure generated by a pumping action at grooves 20 formed in a wedge region 18a
30 of the oil film 18 is cancelled out by a dynamic pressure generated by a pumping
15
action at grooves 20 formed in a region other than the wedge region 18a. Therefore,
in the case where the grooves 20 are arranged over the whole area of the outer
circumferential surface of the rotation shaft 6 as in the related art, in the wedge region
18a, only the oil film pressure 19 due to the rotation of the rotation shaft 6 acts as a
resistance to the variable radial load W, and unlike the plain bearing 5 structure 50
according to Embodiment 1, the oil film thickness cannot be increased.
[0033]
In contrast, in the plain bearing structure 50 according to Embodiment 1, in
addition to the oil film pressure 19, the dynamic pressure generated by the pumping
10 action acts, and the oil film thickness in the wedge region 18a is increased, and
occurrence of seizure can be more effectively reduced. That is, in the plain bearing
structure 50 according to Embodiment 1, in the region of the oil film 18 where the oil
film thickness is relatively great, the ratio of the grooves 20 to the outer
circumferential surface of the rotation shaft 6 can be made low or zero. In other
15 words, in the plain bearing structure 50, the generation of a dynamic pressure by the
pumping action is reduced in the region other than the wedge region 18a of the oil
film 18, and a dynamic pressure by the pumping action is generated in the wedge
region 18a. Accordingly, in the plain bearing structure 50, the dynamic pressure
generated in the wedge region 18a by the pumping action is not cancelled out, and
20 can be applied as a resistance to the variable radial load W. In addition, the oil film
thickness can be effectively increased, and the occurrence of seizure can be reduced.
[0034]
Fig. 6 illustrates the result of analysis of hydrodynamic lubrication in the case
where the region P in the grooves 20 are provided and the shape of the grooves 20 in
25 the plain bearing structure 50 according to Embodiment 1 are changed. In Fig. 6, an
angle  that is the upper limit of the circumferential angular range of the region P
where the grooves 20 are arranged, and an angle  that is an acute-angle component
of the angle of each of the flow passages of the grooves 20 to the circumferential
direction, are parameters. Fig. 6 illustrates the minimum oil film thickness for each of
30 the combinations of the angles  and  in the case where  is changed in the range of
16
0 degrees to 360 degrees and in the case where  is changed in the range of 0
degrees to 90 degrees. It should be noted that the minimum gap between the plain
bearing 15 and the rotation shaft 6 is defined as the minimum oil film thickness. The
numerical values of the minimum oil film thickness as indicated in Fig. 6 are
expressed as the ratios (minimum oil film thickness ratios) of the 5 above minimum oil
film thickness to the minimum oil film thickness in a plain bearing in which the rotation
shaft 6 includes no grooves 20. It should also be noted that the angle , which is the
acute-angle component of the angle of the groove 20 to the circumferential direction,
is  indicated in Fig. 4 and represents the angle of inclination of each of the flow
10 passages included in the groove 20 relative to an imaginary line L1 parallel to the
central axis of the rotation shaft 6.
[0035]
Fig. 7 indicates analysis conditions for use in the analysis the result of which is
indicated in Fig. 6. The analysis in Fig. 6 was conducted under the conditions in
15 which dimensions of portions of the plain bearing structure 50 are as follows: the
inside diameter of the plain bearing 15 is 30 mm; the length of the plain bearing 15 in
the axial direction is 30 m; the radius gap between the plain bearing 15 and the
rotation shaft 6 is 30 m; the depth of the groove 20 from the outer circumferential
surface of the rotation shaft 6 is 30 m. The variable radial load W applied from the
20 rotation shaft 6 is 5000 N, and the rotation speed of the rotation shaft 6 is 3000 rpm.
The result of the analysis in Fig. 6 indicates that in the combinations of the angle 
and the angle  in the hatched region, the minimum oil film thickness ratio exceeds 1.
That is, the hatched region in Fig. 6 represents the combinations of the angle  and
the angle  where the minimum oil film thickness in the plain bearing structure 50
25 according to Embodiment 1 is greater than that in the plain bearing in the related art.
In other words, it represents the combinations of  and  where the plain bearing
structure 50 can obtain an oil film pressure greater than that in the plain bearing in the
related art.
[0036]
17
As indicated in Fig. 6, in the plain bearing structure 50 according to
Embodiment 1, a region where a combination of  and  at which the minimum oil film
thickness is greater than that in the plain bearing in the related art is present can be
expressed with the following combinations of angular ranges: a combination of 40
degrees <  < 270 degrees and 40 degrees <  < 65 degrees; a 5 combination of 30
degrees <  < 250 degrees; and 25 degrees <  < 60 degrees, or a combination of 70
degrees <  < 220 degrees and 15 degrees <  < 65 degrees. The angles  and 
can be arbitrarily selected from the above angular ranges.
[0037]
10 The condition in which  is 360 degrees corresponds to the case where the
grooves 20 are uniformly provided over the entire perimeter of the rotation shaft.
This corresponds to a configuration described in Japanese Unexamined Patent
Application Publication No. 4-370388, which is a prior art document. As illustrated in
Fig. 6, in the case where  is 360 degrees, the minimum oil film thickness ratio is
15 smaller than 1 for  of all angles, and thus unlike a configuration in which the rotation
shaft 6 has no grooves 20, the oil film thickness is not increased. Far from it, the oil
film thickness is decreased. That is, in the plain bearing structure 50 according to
Embodiment 1, the oil film thickness can be further increased and the occurrence of
seizure can be reduced, by providing the grooves 20 in the region P and changing the
20 circumferential angular range of the region P, as compared with the case where the
grooves 20 are provided over the entire perimeter.
[0038]
Although the above description is made with respect to the case where the
grooves 20 are provided only in the region P at the outer circumferential surface of
25 the rotation shaft 6 and no grooves 20 are provided in the region Q other than that
region P, the grooves 20 can be provided in the region Q. However, in the case
where the grooves 20 are provided in the outer circumferential surfaces in the regions
P and Q at the same ratio, this configuration is similar to the configuration as
described in the prior art document, and the advantage of increasing the oil film
30 pressure by the pumping action at the grooves 20 in the wedge region 18a is not
18
obtained. Therefore, in the case where grooves 20 are provided in the region Q,
preferably, the ratio of the area of the grooves 20 to the area of the outer
circumferential surface in the region Q should be smaller than that in the region P. In
this case, in the plain bearing structure 50, the oil film pressure can be increased by
the pumping action at the grooves 20 in the wedge region 18a, 5 and the oil film
thickness can be increased and the occurrence of seizure can be reduced, as
compared with the case where the grooves 20 are uniformly provided over the entire
perimeter.
[0039]
10 (Modifications of Grooves 20)
Regarding the shape of each of the grooves 20 that extend in the axial
direction of the rotation shaft 6, it is descried above by way of example that the two
flow passages of each groove 20 are linearly formed to have a uniform thickness and
are symmetric with respect to the center of the plain bearing 15 in the axial direction
15 such that the two flow passages join each other at the bent portion 21. The shape of
the groove 20 is not limited to the above shape. Regarding the shape of the groove
20, it suffices that the flow passages located on the opposite sides with respect to the
bent portion 21 are provided in such a manner to be gradually retreated from the
opposite ends of the groove 20 in the circumferential direction and in the rotation
20 direction of the rotation shaft 6.
[0040]
Figs. 8 to 10 illustrate modifications of the grooves 20 in the outer
circumferential surface of the rotation shaft 6 of the plain bearing structure 50
according to Embodiment 1. For example, the grooves 20 are modified as illustrated
25 in Fig. 8 that illustrates in grooves 20a. To be more specific, in each of the grooves
20a as illustrated in Fig. 8, a bent portion 21a is displaced in the axial direction of the
rotation shaft 6. The groove 20a is V-shaped such that the flow passages of the
groove 20a are arranged asymmetric in the axial direction of the rotation shaft 6.
[0041]
19
In grooves 20b as illustrated in Fig. 9, each of the flow passages extending
from the opposite ends in the axial direction are V-shaped to have a non-uniform
width such that the width of each flow passage decreases in a direction toward a bent
portion 21c. Furthermore, in grooves 20c as illustrated in Fig. 10, each of the flow
passages extending from the opposite ends of the rotation shaft 5 6 in the axial
direction are curved and U-shaped. Since all the grooves 20a to 20c can generate a
dynamic pressure by a pumping action, in the plan bearing structure including any of
the grooves 20a to 20c, the oil film pressure can be increased and the occurrence of
seizure can be reduced as in the plain bearing structure 50 using the groove 20.
10 [0042]
Figs. 11 and 12 illustrate sections of the outer circumference of the rotation
shaft 6 according to Embodiment 1 as the outer circumference is developed. As
illustrated in Fig. 11, each of the grooves 20 in the plain bearing structure 50
according to Embodiment 1 has a rectangular sectional shape. However, the
15 sectional shape of the groove 20 is not limited to a rectangle and can be any shape
as long as an oil film pressure can be generated at a rear end portion 25, which is an
end portion located on the rear side of the groove 20 in the rotation direction of the
rotation shaft 6, because of a decrease in the gap between the rotation shaft 6 and
the plain bearing 15.
20 [0043]
Each of grooves 20d as illustrated in Fig. 12 is a modification of the groove 20
as illustrated in Fig. 11 and has a rectangular sectional shape in the circumferential
direction. The groove 20d may have a V-shaped section. Furthermore, the groove
20d is not limited to a groove having flow passages that are symmetric with respect to
25 an imaginary line extending through the center of the rotation shaft 6 in the section
that is formed as illustrated in Fig. 12, and may be formed such that flow passages of
the groove 20d are arranged asymmetrical as illustrated in Fig. 12. In the groove
20d, at a inclined surface 23 that is inclined at a small angle and starts at a front end
portion 22 of the groove 20d that is located on a front side in the rotation direction of
30 the rotation shaft 6, the refrigerating machine oil easily flows along the inclined
20
surface 23 because of the rotation of the rotation shaft 6. Therefore, the groove 20d
enables the refrigerating machine oil to be easily drawn into the groove 20d. At a
rear end portion 25 of the groove 20d that is located on a rear side in the rotation
direction of the rotation shaft 6, the groove 20d has a steep inclined surface 24,
whereby the size of the gap between the rotation shaft 6 and the 5 plain bearing 15
sharply decreases. Therefore, a great oil film pressure can be easily obtained. It
should be noted that the angle of the inclined surface 24 may be a right angle to a
tangent to the outer circumferential surface. In each of the grooves 20d, because
the refrigerating machine oil easily flows along the inclined surface 23, at the inclined
10 surface 24, the flow of the refrigerating machine oil is sharply reduced, and the oil film
pressure can be easily increased. The angles of the inclined surfaces 23 and 24 can
be changed as needed. The angle of a bottom 26 of the groove 20 as illustrated in
Fig. 11 may be changed such that the depth of the groove 20 may increase in the
opposite direction to the rotation direction of the rotation shaft 6. Furthermore, the
15 inclined surface 23, the inclined surface 24, and the bottom 26 may be flat surfaces or
curved surfaces. The inclined surfaces 23 and 24 will be referred to as first and
second inclined surfaces, respectively.
[0044]
Embodiment 2
20 In a plain bearing structure 250 in a scroll compressor 100 according to
Embodiment 2, circumferential oil supply grooves 30 are added to the opposite end
portions of the rotation shaft 6 in the axial direction in the plain bearing structure 50
according to Embodiment 1. Embodiment 2 will be described mainly regarding the
differences between Embodiments 1 and 2.
25 [0045]
Fig. 13 is a schematic diagram for describing the plain bearing structure 250 in
the scroll compressor 100 according to Embodiment 2. As illustrated in Fig. 13, in
the rotation shaft 6, the grooves 20 are provided in the region P, and the
circumferential oil supply grooves 30 are provided in the opposite end portions of the
30 grooves 20 in the axial direction such that the circumferential oil supply grooves 30
21
extend in the circumferential direction of the rotation shaft 6. The circumferential oil
supply grooves 30 do not face the plain bearing 15 for the rotation shaft 6 and are not
covered by the plain bearing 15. The circumferential oil supply grooves 30 are
continuous with the end portions of the grooves 20 that extend in the axial direction of
the rotation shaft 6. The circumferential oil supply grooves 30 allow 5 the refrigerating
machine oil to flow from the circumferential oil supply grooves 30 into the grooves 20.
Because of this configuration, the refrigerating machine oil retained in the
circumferential oil supply grooves 30 can be supplied to the gap between the plain
bearing 15 and the rotation shaft 6, thereby preventing depletion of the refrigerating
10 machine oil in the plain bearing structure 250.
[0046]
The circumferential oil supply grooves 30 are continuous with the grooves 20
extending in the axial direction of the rotation shaft 6. Thus, the refrigerating
machine oil retained in the circumferential oil supply grooves 30 flows into the
15 grooves 20, and the amount of the refrigerating machine oil drawn into the groove 20
is larger than that in Embodiment 1. Therefore, the oil film 18 in the region where
the grooves 20 are provided can have a high oil film pressure.
[0047]
Because the bottom of each of the circumferential oil supply grooves 30 does
20 not face the plain bearing 15, the area of the outer circumferential surface of the
rotation shaft 6, which can be supported by the plain bearing 15, is not reduced.
Therefore, even in the case the circumferential oil supply grooves 30 communicating
with the grooves 20 are provided, the load that can be received by the plain bearing
15 is not decreased. The bottom of each of the circumferential oil supply grooves 30
25 may be continuous with the axial-direction oil supply hole 7a via the radial oil supply
outlet 7c. In the case where the circumferential oil supply groove 30 communicates
with the radial oil supply outlets 7c, the refrigerating machine oil supplied from the
radial oil supply outlet or outlets 7c is supplied to the grooves 20 that extend in the
axial direction, thereby preventing the refrigerating machine oil in the plain bearing
30 structure 250 from being further reduced.
22
[0048]
In the above configuration, the refrigerating machine oil supplied from the pump
7b through the axial-direction oil supply hole 7a can be caused to directly flow into the
circumferential oil supply grooves 30. Therefore, the oil amount of refrigerating
machine oil retained in the circumferential oil supply grooves 30 5 can be effectively
increased. Furthermore, it is preferable that at the bottom of each of the
circumferential oil supply grooves 30, an opening portion of the radial oil supply outlet
7c be provided at a circumferential position where the oil amount of the refrigerating
machine oil drawn into the gap between the rotation shaft 6 and the plain bearing 15
10 is large. For example, in the case where the opening portion of each of the radial oil
supply outlets 7c is located at a position where  is 0 degrees, the grooves 20 are
close to the radial oil supply outlet 7c, and the refrigerating machine oil supplied from
the axial-direction oil supply hole 7a through the radial oil supply outlet 7c to the
circumferential oil supply grooves 30 can be supplied to the grooves 20 via the
15 shortest possible path. Therefore, the refrigerating machine oil can be more
effectively supplied to the plain bearing structure 250.
[0049]
Embodiment 3
In a plain bearing structure 350 in a scroll compressor 100 according to
20 Embodiment 3, an axial-direction oil supply groove 40 is added to the region Q of the
outer circumferential surface of the rotation shaft 6 in the plain bearing structure 250
according to Embodiment 2. Embodiment 3 will be described mainly regarding the
differences between Embodiment 3 and Embodiments 1 and 2.
[0050]
25 Fig. 14 is a schematic diagram for describing the plain bearing structure 350 in
the scroll compressor 100 according to Embodiment 3. The plain bearing structure
350 according to Embodiment 3 is the same as the plain bearing structures according
to Embodiments 1 and 2 on the following points: the rotation shaft 6 has the grooves
20 extending in the axial direction; the grooves 20 are provided in the outer
30 circumferential surface of the rotation shaft 6 in the circumferential angular range of 0
23
degrees to ; the ratio of the grooves 20 to the surface area in the other angular
range is 0; and the grooves 20 are formed as dynamic pressure grooves having bent
portions. Furthermore, the circumferential oil supply grooves 30 continuous with the
grooves 20 are provided. In this regard, the plain bearing structure 350 is the same
as the plain bearing surface 5 in Embodiment 2.
[0051]
In the rotation shaft 6 according to Embodiment 3, the radial oil supply outlet 7c
is provided in a surface of the rotation shaft 6 that faces the plain bearing 15. The
axial-direction oil supply groove 40 is provided in the region Q of the outer
10 circumferential surface of the rotation shaft 6, in which the grooves 20 are not
provided. In the bottom of the axial-direction oil supply groove 40, the radial oil
supply outlet 7c is provided to communicate with the axial-direction oil supply hole 7a.
Furthermore, opposite ends of the axial-direction oil supply groove 40 in the axial
direction are continuous with the respective circumferential oil supply grooves 30, and
15 the refrigerating machine oil can flow from the axial-direction oil supply hole 7a, flow
through each of the radial oil supply outlets 7c and the axial-direction oil supply
groove 40, and then flow to the circumferential oil supply grooves 30. The
refrigerating machine oil that has flowed into the circumferential oil supply groove 30
flows into the grooves 20, which are continuous with the circumferential oil supply
20 groove 30 and extend in the axial direction.
[0052]
In Embodiment 2, the radial oil supply outlet 7c is open to the circumferential oil
supply grooves 30, which do not face the plain bearing 15. Inevitably, supplied
refrigerating machine oil flows from the outer circumferential surface of the rotation
25 shaft 6, which faces the plain bearing 15, to the outside. However, in the plain
bearing structure 350, because the radial oil supply outlet 7c is open in the outer
circumferential surface of the rotation shaft 6, which faces the plain bearing 15, the
entire amount of refrigerating machine oil that is supplied by the pump 7b through the
axial-direction oil supply hole 7a and the radial oil supply outlet 7c is supplied to the
30 gap between the plain bearing 15 and the rotation shaft 6. Therefore, the amount of
24
the supply of the refrigerating machine oil to the gap between the plain bearing 15
and the rotation shaft 6 can be increased.
[0053]
The radial oil supply outlet 7c is continuous with the axial-direction oil supply
groove 40, whereby the refrigerating machine oil can be retained in the 5 axial-direction
oil supply groove 40. Therefore, the plain bearing structure can further effectively
prevent depletion of the refrigerating machine oil. Moreover, because the axialdirection
oil supply groove 40 is continuous with the circumferential oil supply grooves
30, as in Embodiment 2, the refrigerating machine oil can be retained also in the
10 circumferential oil supply grooves 30, and the amount of the supply of the
refrigerating machine oil to the grooves 20 through the circumferential oil supply
grooves 30 can be increased.
[0054]
The axial-direction oil supply groove 40 includes two flow passages that join
15 each other at a bent portion 41. The closer part of each of the flow passages to the
central portion of the axial-direction oil supply groove 40, the more forwardly the part
of the flow passage is located in the circumferential direction and the rotation
direction. That is, the axial-direction oil supply groove 40 is V-shaped in the opposite
direction to the rotation direction of the rotation shaft 6. The radial oil supply outlet
20 7c can be provided at the bent portion 41. That is, the flow passages that form the
axial-direction oil supply groove 40 and extend from the bent portion 41 in the axial
direction join the radial oil supply outlet 7c at the bent portion 41.
[0055]
The grooves 20 described regarding Embodiment 1 are provided to facilitate
25 the inflow of the refrigerating machine oil into the gap between the rotation shaft 6 and
the plain bearing 15 and effectively generate a dynamic pressure in the direction
where the variable radial load W is supported by the pumping action. By contrast, in
the plain bearing structure 350 according to Embodiment 3, the axial-direction oil
supply groove 40 is provided to cause the refrigerating machine oil supplied from the
30 axial-direction oil supply hole 7a to the gap between the rotation shaft 6 and the plain
25
bearing 15 through the radial oil supply outlet 7c to be efficiently supplied to the
circumferential oil supply grooves 30. In other words, the axial-direction oil supply
groove 40 has the bent portion 41, and the two flow passages extending from the
bent portion 41 are inclined in the opposite direction to the rotation direction of the
rotation shaft 6. Thus, when the rotation shaft 6 is rotated, the refrigerating 5 machine
oil that flows in the axial-direction oil supply groove 40 flows along the wall surface of
the axial-direction oil supply groove 40 toward the circumferential oil supply grooves
30. Because of this flow, the oil amount of the refrigerating machine oil supplied
from the axial-direction oil supply groove 40 to the circumferential oil supply grooves
10 30 increases, the oil amount of the refrigerating machine oil supplied from the
circumferential oil supply grooves 30 to the grooves 20 increases, and thus a high oil
film pressure can be obtained in the wedge region 18a. Furthermore, in the plain
bearing structure 350, it is possible to reduce depletion of the refrigerating machine
oil, increase the oil film thickness, and also reduce the occurrence of seizure.
15 Reference Signs List
[0056]
1 fixed scroll, 1a base plate, 2 orbiting scroll, 2a base plate, 2b
eccentric hole, 3 suction port, 4 discharge port, 5 compression chamber, 6
rotation shaft, 6a main shaft, 6b eccentric shaft, 7a axial-direction oil supply
20 hole, 7b pump, 7c radial oil supply outlet, 7d oil supply outlet, 7e lower
opening port, 8 housing, 8a upper housing, 8b lower housing, 9 electric
motor, 9a rotor, 9b stator, 10 Oldham ring, 11 pressure container, 12
oil sump, 13 refrigerant suction pipe, 14 refrigerant discharge pipe, 15 plain
bearing, 15a main bearing, 15b orbiting bearing, 16 minor bearing, 18 oil
25 film, 18a region, 18b region, 19 oil film pressure, 20 groove, 20a
groove, 20b groove, 20c groove, 20d groove, 21 bent portion, 21a
bent portion, 21c bent portion, 22 front end portion, 23 inclined surface, 24
inclined surface, 25 rear end portion, 26 bottom, 30 circumferential oil
supply groove, 40 axial-direction oil supply groove, 41 bent portion, 50 plain
30 bearing structure, 50a plain bearing structure, 50b plain bearing structure, 60
26
scroll compression mechanism, 100 scroll compressor, 250 plain bearing
structure, 350 plain bearing structure, L0 line of action, L1 imaginary line,
P region, Q region, W variable radial load,  angle,  angle
27
We Claim:
[Claim 1]
A plain bearing structure comprising:
a rotation shaft configured to drive an orbiting scroll in a scroll 5 compressor; and
a plain bearing that supports a variable radial load on the rotation shaft,
wherein the rotation shaft includes a groove formed in an outer circumferential
surface of the rotation shaft and extending in an axial direction of the rotation shaft,
the outer circumferential surface facing the plain bearing,
10 the groove is configured to cause a pumping action to occur,
the outer circumferential surface of the rotation shaft includes a region P and a
region Q, the region P being defined in a predetermined angular range in a
circumferential direction of the rotation shaft, the region Q being a region of the outer
circumferential surface that is other than the region P, and
15 a ratio of an area of part of the groove that is located in the region P in the
outer circumferential surface to an area of the region P is higher than a ratio of an
area of part of the groove that is located in the region Q in the outer circumferential
surface to an area of the region Q.
[Claim 2]
20 The plain bearing structure of claim 1, wherein the groove is not provided in the
region Q.
[Claim 3]
The plain bearing structure of claim 1 or 2, wherein opposite end portions of the
groove in the axial direction are located forward of a central portion of the groove in
25 the circumferential direction of the rotation shaft and a rotation direction thereof.
[Claim 4]
The plain bearing structure of claim 3, wherein the groove is V-shaped in the
rotation direction of the rotation shaft.
[Claim 5]
30 The plain bearing structure of any one of claims 1 to 4, wherein
28
where a circumferential position at which a line of action of the variable radial
load that starts at a center of the rotation shaft intersects with the outer circumferential
surface of the rotation shaft is defined as a 0-degree position and an angle  is set in
an opposite direction to a rotation direction of the rotation shaft,
where the region P is provided in an angular range of 0 5 degrees to , and
where an acute-angle component of an angle of the part of the groove that is
located in the region P to an imaginary line L1 parallel to the axial direction of the
rotation shaft is an angle ,
a combination of a range of values of the angle  and a range of values of the
10 angle  is selected from the following combinations:
a combination of 40 degrees <  < 270 degrees and 40 degrees <  < 65
degrees,
a combination of 30 degrees <  < 250 degrees and 25 degrees <  < 60
degrees, and
15 a combination of 70 degrees <  < 220 degrees and 15 degrees <  < 65
degrees, and
the angle  and the angle  are set to angles that falls within the ranges of the
combination selected.
[Claim 6]
20 The plain bearing structure of any one of claims 1 to 5, wherein a bottom of the
groove is provided such that a depth of the groove increases in an opposite direction
to a rotation direction of the rotation shaft.
[Claim 7]
The plain bearing structure of claim 6, wherein the bottom of the groove is an
25 inclined surface.
[Claim 8]
The plain bearing structure of any one of claims 1 to 5, wherein the groove
includes
a front end portion located on a front side in a rotation direction of the
30 rotation shaft,
29
a rear end portion located on a rear side in the rotation direction of the
rotation shaft,
a first inclined surface whose depth increases in a direction from the
front end portion toward the rear end portion, and
a second inclined surface whose depth decreases in 5 a direction toward
the rear end portion, the second inclined surface being continuous with the first
inclined surface, and
the first inclined surface is inclined at a smaller angle than the second inclined
surface.
10 [Claim 9]
The plain bearing structure of any one of claims 1 to 8, further comprising:
an axial-direction oil supply hole extending through the rotation shaft in the
axial direction; and
a radial oil supply outlet continuous with the axial-direction oil supply hole and
15 formed in the region Q of the outer circumferential surface of the rotation shaft.
[Claim 10]
The plain bearing structure of any one of claims 1 to 8, wherein the rotation
shaft has circumferential oil supply grooves provided in respective regions of the
outer circumferential surface that are located outward of a region thereof that faces
20 the plain bearing, the circumferential oil supply grooves extending in the
circumferential direction, and
the circumferential oil supply grooves are continuous with opposite end
portions of the groove in the axial direction.
[Claim 11]
25 The plain bearing structure of claim 10, wherein the rotation shaft further
includes an axial-direction oil supply groove extending in the axial direction in the
region Q, and
the axial-direction oil supply groove is continuous with the circumferential oil
supply grooves at the opposite end portions in the axial direction.
30
30
[Claim 12]
The plain bearing structure of claim 11, further comprising:
an axial-direction oil supply hole extending through the rotation shaft in the
axial direction; and
a radial oil supply outlet continuous with the axial-direction oil 5 supply hole and
formed in in the region Q of the outer circumferential surface of the rotation shaft.
[Claim 13]
The plain bearing structure of claim 12, wherein the axial-direction oil supply
groove is V-shaped in an opposite direction to a rotation direction of the rotation shaft
10 and includes a bent portion at a central portion of the axial-direction oil supply groove
in the axial direction, and
the axial-direction oil supply groove is continuous with the radial oil supply
outlet at the central portion in the axial direction.
[Claim 14]
15 A scroll compressor comprising the plain bearing structure of any one of claims
1 to 13.