U.S. patent application number 12/953731 was filed with the patent office on 2011-03-24 for hermetic type compressor and refrigeration cycle apparatus.
This patent application is currently assigned to TOSHIBA CARRIER CORPORATION. Invention is credited to Toshihiko FUTAMI, Kazuhiko MIURA, Koji SATODATE.
Application Number | 20110067434 12/953731 |
Document ID | / |
Family ID | 41377103 |
Filed Date | 2011-03-24 |
United States Patent
Application |
20110067434 |
Kind Code |
A1 |
MIURA; Kazuhiko ; et
al. |
March 24, 2011 |
HERMETIC TYPE COMPRESSOR AND REFRIGERATION CYCLE APPARATUS
Abstract
As a hermetic type compressor, a motor portion and a compression
mechanism portion that are coupled to the motor portion with a
rotating shaft interposed therebetween are accommodated in a closed
vessel. The compression mechanism portion comprises a cylinder that
comprises an internal diameter hole, and a main bearing and a
sub-bearing in which a bearing hole that journals the rotating
shaft is provided and the internal diameter hole of the cylinder is
closed to form a compression chamber in the compression mechanism
portion. The main bearing and the sub-bearing have a circular
groove that is opened toward the compression chamber side, an inner
circumferential surface of the circular groove is tapered such that
a diameter increases gradually from the compression chamber side
toward an opposite side of the compression chamber side, and a
depth of the circular groove is set to 40% of a diameter of the
bearing hole.
Inventors: |
MIURA; Kazuhiko; (Fuji-shi,
JP) ; SATODATE; Koji; (Fuji-shi, JP) ; FUTAMI;
Toshihiko; (Fuji-shi, JP) |
Assignee: |
TOSHIBA CARRIER CORPORATION
Tokyo
JP
|
Family ID: |
41377103 |
Appl. No.: |
12/953731 |
Filed: |
November 24, 2010 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
|
PCT/JP2009/059719 |
May 27, 2009 |
|
|
|
12953731 |
|
|
|
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Current U.S.
Class: |
62/498 ;
418/157 |
Current CPC
Class: |
F04C 23/001 20130101;
F04C 29/12 20130101; F01C 21/02 20130101; F04C 18/3562 20130101;
F04C 2240/601 20130101; F04C 23/008 20130101; F04C 2240/52
20130101 |
Class at
Publication: |
62/498 ;
418/157 |
International
Class: |
F25B 1/00 20060101
F25B001/00; F04C 2/067 20060101 F04C002/067 |
Foreign Application Data
Date |
Code |
Application Number |
May 28, 2008 |
JP |
2008-139682 |
Claims
1. A hermetic type compressor in which a motor portion and a
compression mechanism portion that are coupled to the motor portion
with a rotating shaft interposed therebetween are accommodated in a
closed vessel, wherein the compression mechanism portion comprises:
a cylinder that comprises an internal diameter hole; and a main
bearing and a sub-bearing, in which a bearing hole that journals
the rotating shaft is provided and the internal diameter hole of
the cylinder is closed to form a compression chamber in the
compression mechanism portion, at least one of the main bearing and
the sub-bearing has a circular groove that is opened toward the
compression chamber side, an inner circumferential surface of the
circular groove is tapered such that a diameter increases gradually
from the compression chamber side toward an opposite side of the
compression chamber side, and a depth L of the circular groove is
set to at least 40% of a diameter D of the bearing hole.
2. The hermetic type compressor according to claim 1, wherein, in
the main bearing or sub-bearing that comprises the circular groove,
a minimum wall thickness b between an inner circumferential surface
of the circular groove and a circumferential surface of the bearing
hole is set so as to satisfy a relationship of an equation (1):
0.09.times.diameter D of bearing hole.gtoreq.minimum wall thickness
b.gtoreq.0.04.times.diameter D of bearing hole (1).
3. The hermetic type compressor according to claim 1, wherein a
compression chamber of the compression mechanism portion
accommodates an eccentric portion that is eccentrically provided
while integrated with the rotating shaft and a rolling piston that
is fitted in the eccentric portion to rotate eccentrically in the
compression chamber in association with rotation of the rotating
shaft, and assuming that e is an eccentric amount of the eccentric
portion and r is an outer circumferential radius of the rolling
piston, an outer circumferential radius g of the circular groove
satisfies relationships of equations (2) and (3): 0.5
mm.ltoreq.[outer circumferential radius r (mm) of rolling
piston-eccentric amount e (mm) of eccentric portion]-outer
circumferential radius g (mm) of circular groove (2) outer
circumferential radius g (mm) of circular groove>diameter D (mm)
of bearing hole/2+minimum wall thickness b (mm) (3).
4. The hermetic type compressor according to claim 1, wherein the
main bearing and the sub-bearing have a flange whose wall thickness
is set to a depth L of the circular groove or less.
5. A refrigeration cycle apparatus comprising: the hermetic type
compressor according to claim 1; a condenser; an expansion device;
and an evaporator.
Description
CROSS REFERENCE TO RELATED APPLICATIONS
[0001] This is a Continuation Application of PCT Application No.
PCT/JP2009/059719, filed May 27, 2009, which was published under
PCT Article 21(2) in Japanese.
[0002] This application is based upon and claims the benefit of
priority from prior Japanese Patent Application No. 2008-139682,
filed May 28, 2008, the entire contents of which are incorporated
herein by reference.
BACKGROUND OF THE INVENTION
[0003] 1. Field of the Invention
[0004] The present invention relates to a hermetic type compressor
whose bearing structure is improved and a refrigeration cycle
apparatus that includes the hermetic type compressor to form a
refrigeration cycle.
[0005] 2. Description of the Related Art
[0006] Frequently, a rotary hermetic type compressor is used in the
refrigeration cycle apparatus. In the rotary hermetic type
compressor, a motor portion and a compression mechanism portion
that is coupled to the motor portion via a rotating shaft
(crankshaft) interposed therebetween are accommodated in a closed
vessel. In this kind of compressor, a refrigerant is introduced
into a compression chamber formed in a cylinder and compressed,
whereby a compressive load acts on the rotating shaft.
[0007] Accordingly, the rotating shaft generates a flexural
deformation, and a rotating shaft portion in a flexure direction
and a bearing that journals the rotating shaft come into partial
contact with each other unless some sort of measure is taken.
Smooth rotation of the rotating shaft is spoiled, which leads to
damage of the rotating shaft and bearing. Therefore, for example,
Jpn. Pat. Appln. KOKAI Publication No. 2004-124834 proposes a
bearing structure in order to properly bear the flexural
deformation of the rotating shaft.
[0008] In the technique proposed in Jpn. Pat. Appln. KOKAI
Publication No. 2004-124834, according to the flexural deformation
of the rotating shaft due to the compressive load in the cylinder,
a groove is provided on a cylinder side of a main bearing to allow
the flexural deformation of the main bearing, and a center of an
internal diameter on the motor side of the main bearing is
eccentrically disposed by a predetermined amount with respect to a
center of an internal diameter on the cylinder side in a direction
of the flexural deformation of the rotating shaft.
BRIEF SUMMARY OF THE INVENTION
[0009] However, in the groove on the cylinder side of the main
bearing in the technique, a diameter of an inner circumferential
surface of the main bearing is kept constant over a total length,
and a thickness between the inner circumferential surface of the
groove and an inner circumference of a bearing hole is also kept
constant over the total length.
[0010] Accordingly, although the partially strong contact between
the rotating shaft and the bearing can be avoided in a certain
range of the groove by the flexure of the bearing, rigidity of the
bearing increases rapidly at an end of the groove, and a contact
load is concentrated on the end of the groove. Therefore, local
abrasion is generated, and bearing reliability cannot sufficiently
be enhanced.
[0011] In view of the foregoing, an object of the invention is to
provide a hermetic type compressor in which, according to the
flexural deformation of the rotating shaft due to the compressive
load in the cylinder, uneven contact with the rotating shaft is
prevented in at least one of the main bearing and sub-bearing,
thereby achieving the enhancement of the reliability and a longer
operation life.
[0012] Another object of the invention is to provide a
refrigeration cycle apparatus that includes the hermetic type
compressor to form the refrigeration cycle, thereby improving
refrigeration efficiency.
[0013] A hermetic type compressor of the present invention
comprises, a motor portion and a compression mechanism portion that
are coupled to the motor portion with a rotating shaft interposed
therebetween are accommodated in a closed vessel, the compression
mechanism portion comprises a cylinder that comprises an internal
diameter hole; and a main bearing and a sub-bearing in which a
bearing hole that journals the rotating shaft is provided and the
internal diameter hole of the cylinder is closed to form a
compression chamber in the compression mechanism portion, at least
one of the main bearing and the sub-bearing have a circular groove
that is opened toward the compression chamber side, an inner
circumferential surface of the circular groove is tapered such that
a diameter increases gradually from the compression chamber side
toward an opposite side of the compression chamber side, and a
depth of the circular groove is set to at least 40% of a diameter
of the bearing hole.
[0014] A refrigeration cycle apparatus of the present invention
comprises, the hermetic type compressor; a condenser; an expansion
device; and an evaporator.
BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWING
[0015] FIG. 1 is a refrigeration cycle configuration diagram of a
refrigeration cycle apparatus according to a first embodiment of
the invention and a longitudinal sectional view of a hermetic type
compressor.
[0016] FIG. 2 is an enlarged longitudinal sectional view of a
compression mechanism portion of the hermetic type compressor.
[0017] FIG. 3 is an enlarged longitudinal sectional view of a
compression mechanism portion of a hermetic type compressor
according to a second embodiment of the invention.
[0018] FIG. 4 is a longitudinal sectional view of a main part of a
hermetic type compressor according to a third embodiment of the
invention.
[0019] FIG. 5 is a longitudinal sectional view of a main part of a
hermetic type compressor according to a fourth embodiment of the
invention.
[0020] FIG. 6 is a characteristic diagram of a circular groove
depth effect in the invention.
[0021] FIG. 7 is a characteristic diagram of a circular groove
minimum wall thickness effect in the invention.
[0022] FIG. 8 is a characteristic diagram of a circular groove
minimum seal width effect in the invention.
[0023] FIG. 9 is a characteristic diagram of a circular groove
slope effect in the invention.
[0024] FIG. 10 is a longitudinal sectional view of a hermetic type
compressor according to a modification of the third embodiment of
the invention.
[0025] FIG. 11 is a plan view of a discharge valve mechanism
mounted on an intermediate partition plate of the modification.
[0026] FIG. 12 is a sectional view of an intermediate partition
plate and a discharge valve mechanism of a first example of the
modification.
[0027] FIG. 13 is a sectional view of an intermediate partition
plate and a discharge valve mechanism of a first example of the
modification.
DETAILED DESCRIPTION OF THE INVENTION
[0028] Embodiments of the invention will be described below with
reference to the drawings. FIG. 1 is a longitudinal sectional view
of a hermetic type compressor 1 and a refrigeration cycle
configuration diagram of a refrigeration cycle apparatus R.
[0029] In FIG. 1, the numeral 1 designates a hermetic type rotary
compressor (hereinafter simply referred to as "compressor"), and
the compressor 1 is described later. A refrigerant pipe P is
connected to an upper end portion of the compressor 1. A condenser
2, an expansion valve (expansion device) 3, an evaporator 4, and an
accumulator 5 are sequentially provided in the refrigerant pipe P.
The refrigerant pipe P is also connected to a side portion of the
compressor 1 from the accumulator 5, thereby forming a
refrigeration cycle of the refrigeration cycle apparatus R.
[0030] The compressor 1 will be described next. The compressor 1
comprises a closed vessel 10. A motor portion 11 is accommodated on
an upper portion side in the closed vessel 10, and a compression
mechanism portion 12 is accommodated on a lower portion side. The
motor portion 11 and the compression mechanism portion 12 are
coupled to each other via a rotating shaft 13 interposed
therebetween.
[0031] A discharge portion la formed by a hole portion is provided
in an upper surface portion of the closed vessel 10, and the
refrigerant pipe P communicated with the condenser 2 is connected
to the discharge portion 1a. A suction portion 1b formed by a hole
portion is provided in a circumferential wall in a lower portion of
the closed vessel 10, and the refrigerant pipe P communicated with
the accumulator 5 is connected to the suction portion 1b.
[0032] The motor portion 11 comprises a rotor 15 and a stator 16.
The rotor 15 is fitted in and fixed to a rotating shaft 13. An
inner circumferential surface of the stator 16 faces an outer
circumferential surface of the rotor 15 with a narrow gap, and the
stator 16 is fitted in and fixed to an inner circumferential wall
of the closed vessel 10.
[0033] The compression mechanism portion 12 will be described below
with reference to FIGS. 1 and 2. FIG. 2 is an enlarged longitudinal
sectional view illustrating the compression mechanism portion
12.
[0034] The compression mechanism portion 12 comprises a cylinder
20, a main bearing 21, and a sub-bearing 22. The cylinder 20 is
fitted in and fixed to the inner circumferential wall of the closed
vessel 10, and an internal diameter hole S is made in an axial
center of the cylinder 20. The main bearing 21 is mounted on an
upper surface of the cylinder 20. The sub-bearing 22 is mounted on
a lower surface of the cylinder 20. The cylinder internal diameter
hole S is closed by the main bearing 21 and the sub-bearing 22 to
form a space, and the space constitutes a compression chamber
(hereinafter referred to as "cylinder chamber") S.
[0035] In the rotating shaft 13, a portion between the motor
portion 11 and the upper surface of the cylinder 20 is inserted and
journaled in a bearing hole N made in the main bearing 21. In the
rotating shaft 13, a portion between the lower surface and a lower
end of the cylinder 20 is inserted and journaled in a bearing hole
N made in the sub-bearing 22.
[0036] The main bearing 21 and the sub-bearing 22 comprise flanges
21a and 22a and cylindrical pivot portions 21b and 22b,
respectively. The flanges 21a and 22a close the cylinder internal
diameter hole S. The cylindrical pivot portions 21b and 22b are
projected along axial center portions of the flanges 21a and 22a
while integrated with the flanges 21a and 22a, and the cylindrical
pivot portions 21b and 22b comprise the bearing holes N in which
the rotating shaft 13 is journaled. Circular grooves K are provided
in the main bearing 21 and the sub-bearing 22, and the circular
grooves K are described later.
[0037] An eccentric portion 13a whose center axis is eccentrically
located by an eccentric amount e is integrally provided in the
rotating shaft 13. A rolling piston (hereinafter simply referred to
as "roller") 25 is fitted in a circumferential surface of the
eccentric portion 13a. The roller 25 and the eccentric portion 13a
are accommodated in the cylinder chamber S, and part of an outer
circumferential wall of the roller 25 is designed to come into
linear contact with a circumferential wall of the cylinder chamber
S along an axis direction. Accordingly, a position at which the
outer circumferential wall of the roller 25 comes into contact with
the circumferential wall of the cylinder chamber S moves gradually
in a circumferential direction by the rotation of the rotating
shaft 13.
[0038] A blade chamber (not illustrated) is provided in the
cylinder 20. A compression spring is accommodated in the blade
chamber, and a blade that receives a back pressure from the
compression spring is movably accommodated. A leading end edge of
the blade is in contact with part of the outer circumferential wall
of the roller 25 along the axis direction, and therefore the blade
always divides the cylinder chamber S into two.
[0039] A discharge hole 26 is made in the main bearing 21. A
position at which the discharge hole 26 is made is located near a
region where the blade comes into contact with the roller 25, and
the position constitutes one side portion of the region. A
discharge valve mechanism 27 is provided in the discharge hole 26,
and the discharge valve mechanism 27 is covered with a valve cover
28 mounted on the main bearing 21. A guide hole 28c is made in the
valve cover 28 so as to be opened into the closed vessel 10.
[0040] In the cylinder 20, the hole portion constituting the
suction portion 1b is provided across the region where the blade
comes into contact with the roller 25 from the discharge hole 26.
The suction portion 1b is communicated with the closed vessel 10
while radially piercing the cylinder 20, and the suction portion 1b
is connected to the refrigerant pipe P communicated with the
accumulator 5.
[0041] The circular grooves K, provided in the main bearing 21 and
the sub-bearing 22, will be described in detail.
[0042] The circular groove K provided in the main bearing 21 and
the circular groove K provided in the sub-bearing 22 have the same
structure, shape, and dimensions. At this point, only the circular
groove K of the main bearing 21 is described. In the circular
groove K of the sub-bearing 22, the same component is designated by
the same numeral, and the description is not repeated.
[0043] The circular groove K is provided from an intersection
portion of the flange 21a and cylindrical pivot portion 21b
constituting the main bearing 21 to the cylindrical pivot portion
21b. The circular groove K comprises an opening end Kd that is
opened to the cylinder chamber S, and the circular groove K is
formed deeper from the opening end Kd toward the side of the motor
portion 11 that is the opposite side of the cylinder chamber S.
[0044] The opening end Kd of the circular groove K is concentric
with the bearing hole N made in the main bearing 21, and the
opening end Kd is formed into a ring shape having a predetermined
width. In the circular groove K, a distance between an outer
circumferential surface Km and a circumferential surface of the
bearing hole N is kept constant from the opening end Kd in a depth
direction, while a distance between an inner circumferential
surface Kq and the circumferential surface of the bearing hole N
increases gradually.
[0045] In other words, while the diameter is evenly formed along
the axis direction in the outer circumferential surface Km of the
circular groove K, the inner circumferential surface Kq is tapered
such that the diameter increases gradually along the axis
direction. Therefore, the wall thickness from the circumferential
surface of the bearing hole N to the inner circumferential surface
Kq of the circular groove K becomes minimum (thinnest) at the
opening end Kd of the circular groove K and increases gradually
from the opening end Kd in the depth direction.
[0046] On the assumption that the inner circumferential surface Kq
of the circular groove K is tapered such that the diameter
increases gradually from the opening end Kd that is the side of the
cylinder chamber S toward the opposite side of the cylinder chamber
S, a depth L of the circular groove K is set to at least 40% of a
diameter D of the bearing hole N for the later-described reason,
where L is a depth of the circular groove K and D is a diameter
(that is also a shaft diameter of the rotating shaft 13) of the
bearing hole N.
[0047] In the circular groove K, because the inner circumferential
surface Kq is tapered, a wall thickness b that is a distance
between the inner circumferential surface Kq and the
circumferential surface of the bearing hole N becomes minimum at
the opening end Kd facing the cylinder chamber S. For the
later-described reason, the wall thickness b between the inner
circumferential surface Kq of the circular groove K and the
circumferential surface of the bearing hole N is set so as to
satisfy a relationship of an equation (1):
0.09.times.diameter D of bearing hole N.gtoreq.minimum wall
thickness b.gtoreq.0.04.times.diameter D of bearing hole N (1)
[0048] Assuming that e is an eccentric amount of the eccentric
portion 13a of the rotating shaft 13 and r is an outer
circumferential radius of the roller 25, the outer circumferential
radius g of the circular groove K is set so as to satisfy
relationships of equations (2) and (3) for the later-described
reason:
0.5 mm.ltoreq.[outer circumferential radius r (mm) of roller
25-eccentric amount e (mm) of eccentric portion 13a]-outer
circumferential radius g (mm) of circular groove K (2)
outer circumferential radius g (mm) of circular groove
K>diameter D (mm) of bearing hole N/2+minimum wall thickness b
(mm) (3)
[0049] The action of the compressor 1 and freezing action of the
refrigeration cycle apparatus R will be described below.
[0050] When a current is passed through the motor portion 11
constituting the compressor 1, the rotor 15 is rotated by a
rotating magnetic field generated by the stator 16, thereby
rotating the rotating shaft 13 integrated with the rotor 15. A
driving torque acts on the rotating shaft 13 from the motor portion
11, and the eccentric portion 13a provided in the rotating shaft 13
is eccentrically rotated while integrated with the roller 25 in the
cylinder chamber S.
[0051] Therefore, a negative pressure is partially generated in the
cylinder chamber S, and the refrigerant is introduced from the
accumulator 5 through the refrigerant pipe P. The refrigerant is
introduced into the space region partitioned by the circumferential
surface of the roller 25, the circumferential surface of the
cylinder chamber S, and the blade, and a volume of the space region
is reduced in association with the eccentric rotation of the roller
25, thereby compressing the refrigerant.
[0052] When the space region is minimized, the refrigerant is
raised to a high temperature while attaining a predetermined
high-pressure state. The discharge valve mechanism 27 is opened by
the compressed gas refrigerant, the refrigerant is introduced to
the closed vessel 10 through a valve cover 28, and the closed
vessel 10 is filled with the gas refrigerant. The high-temperature,
high-pressure gas refrigerant with which the closed vessel 10 is
filled is discharged from the discharge portion 1a to the
refrigerant pipe P.
[0053] The condenser 2 performs heat exchange of the gas
refrigerant for outside air or water, and the gas refrigerant is
condensed and liquefied into a liquid refrigerant. The liquid
refrigerant is introduced to the expansion valve 3 to perform
adiabatic expansion, the liquid refrigerant is introduced to the
evaporator 4 to perform the heat exchange for air around a region
where the evaporator 4 is disposed, and the liquid refrigerant is
evaporated.
[0054] Evaporative latent heat is seized from the surrounding
region in association with the evaporation of the refrigerant. That
is, the freezing action acts on the surrounding region. The
refrigerant evaporated in the evaporator 4 is introduced to the
accumulator 5 to perform gas-liquid separation. The refrigerant is
sucked into the cylinder chamber S of the compressor 1, the
refrigerant is compressed again to change into the
high-temperature, high-pressure gas refrigerant, and the
refrigeration cycle is repeated.
[0055] Thus, a suction stroke in which the refrigerant to which the
gas-liquid separation is performed is sucked from the accumulator
5, a compression stroke in which the sucked refrigerant is
compressed, and a discharge stroke in which the compressed
refrigerant is discharged are continuously performed in the
cylinder chamber S constituting the compression mechanism portion
12.
[0056] Particularly, in the compression stroke, the compressive
load is applied to the rotating shaft 13 by the compressed
high-pressure gas refrigerant, whereby the flexural deformation of
the rotating shaft 13 is generated, from a macroscopic point of
view. Specifically, the flexural deformation of the rotating shaft
13 is generated in an opposite direction to the compressive load
direction during the compression action.
[0057] However, because the main bearing 21 and the sub-bearing 22
comprise the circular grooves K set to the above-described
conditions, the uneven contact of the rotating shaft 13 with the
main bearing 21 and sub-bearing 22 is not generated, irrespective
of the flexural deformation of the rotating shaft 13, and the
smooth rotation is secured.
[0058] More specifically, the bearing hole N that is the inner
surface of the main bearing 21 is deformed so as to follow the
rotating shaft 13 in which the flexural deformation is generated by
receiving the load, and an area where the evenness of the gap
between the rotating shaft 13 and the main bearing 21 is retained
is expanded. Accordingly, the ability to form an oil film of
lubrication oil between the rotating shaft 13 and the main bearing
21 is improved, and the oil film is securely formed even if the
rotating shaft 13 is rotated at low speed.
[0059] There are circumstances in which the formation of the oil
film can hardly be maintained, such conditions being when the
number of rotations of the rotating shaft 13 is decreased,
viscosity of the lubrication oil is reduced, or the compressive
load is increased. That is, the contact between the rotating shaft
13 and the main bearing 21 makes a transition to a mixed
lubrication state in which not only the rotating shaft 13 and the
main bearing 21 come into contact with each other while the oil
film is interposed therebetween, but also metallic materials come
into solid-state contact with each other due to the surface
roughness of the rotating shaft 13 and main bearing 21 to support
the load.
[0060] Even if the solid-state contact cannot be avoided, the
surface of the bearing hole N of the main bearing 21 is
continuously deformed to prevent the generation of a locally high
contact force. The generation of seizing or local bearing abrasion
can be prevented to provide the high-reliability main bearing 21.
Because the sub-bearing 22 comprises the circular groove K having
completely the same structure, a similar effect can be obtained in
the sub-bearing 22.
[0061] The circular groove K of the embodiment will be described in
comparison with a flexible-structure groove described in Jpn. Pat.
Appln. KOKAI Publication No. 2004-124834. From the viewpoint of the
formation of the oil film, desirably a gap is evenly formed along
the axis line direction between the main bearing 21 that journals
the rotating shaft 13 and the rotating shaft 13 in which the
flexural deformation is generated by receiving the compressive load
in the cylinder chamber S.
[0062] The flexural deformation of the rotating shaft 13 is
maximized on the side of the cylinder chamber S in which the
compressive load is applied to the rotating shaft 13 and decreases
gradually with distance from the side of the cylinder chamber S. As
described above, when the circular groove K is formed in the main
bearing 21, rigidity of the internal diameter of the main bearing
21 is low on the side of the cylinder chamber S in which the
rotating shaft 13 has the large flexural deformation, and the
rigidity increases gradually with distance from the side of the
cylinder chamber S.
[0063] Therefore, the inner surface of the main bearing 21 is
deformed by following the deformation of the rotating shaft 13, and
the deformable circular groove K is formed deeper than the
flexible-structure groove, so that the circular groove K is greatly
deformed in a wide area to follow the rotating shaft 13.
Additionally, the rigidity of the internal diameter of the main
bearing 21 increases gradually with distance from the side of the
cylinder chamber. S, so that a fluctuation in load applied to the
main bearing 21 in the axis direction can be reduced.
[0064] On the other hand, in the flexible-structure groove, because
the wall thickness between the groove inner surface and the
circumferential surface of the bearing hole is kept constant over
the total length of the groove, the rigidity of the circumferential
surface of the bearing hole is kept constant. Therefore, the
rigidity is small in the groove portion, the rigidity increases
rapidly in the portion in which the groove is terminated, and the
fluctuation in load applied to the bearing also increases.
Accordingly, the oil film is easily broken in the portion in which
the groove is terminated. This cannot be solved even if the groove
depth is simply increased.
[0065] In the embodiment, the circular groove K is provided, and
the depth of the groove K and the wall thickness between the groove
K and the bearing hole N are increase to enhance the strength. The
rigidity of the internal diameter of the main bearing 21 increases
with distance from the side of the cylinder chamber S, the oil film
is evenly formed in the whole of the main bearing 21, and the fluid
lubrication state can be maintained in the wide operating area.
[0066] Even if the contact between the rotating shaft 13 and the
main bearing 21 makes the transition from the fluid lubrication
state to the mixed lubrication state in which the lubrication state
including the solid-state contact state is maintained, because the
circular groove K is deep and flexible, the solid-state contact is
generated in the depth range of the circular groove K in which the
elastic deformation can be generated, and the main bearing 21 is
elastically deformed to prevent the uneven contact with the
rotating shaft 13. Therefore, seizing and the like are not
generated.
[0067] As described above, there is the setting condition that the
inner circumferential surface Kq of the circular groove K is
tapered. The setting condition is fixed on the following basis.
First, the basis on which the depth L of the circular groove K is
set to at least 40% of the diameter D of the bearing hole N will be
described, on the assumption that the inner circumferential surface
Kq of the circular groove K is tapered such that the diameter
increases gradually from the surface facing the cylinder chamber S
toward the opposite side of the cylinder chamber S.
[0068] In the bearing hole N of the main bearing 21, the portion in
which the circumferential surface of the rotating shaft 13 is
particularly effectively journaled is a portion from an end portion
of the bearing hole N to a length corresponding to the diameter of
the bearing hole N. The depth L of the circular groove K is formed
equal to or more than 40% of the diameter D of the bearing hole
N.
[0069] Therefore, the inner surface (bearing hole N) of the main
bearing 21 is deformed so as to follow the deformation of the
rotating shaft 13, which desirably affects the formation of the oil
film between the rotating shaft 13 and the main bearing 21 and the
contact of the rotating shaft 13 with the main bearing 21 due to
the deformation of the rotating shaft 13.
[0070] This can be described with reference to FIG. 6. FIG. 6 is a
characteristic diagram illustrating a groove depth effect. In FIG.
6, a horizontal axis indicates the depth of the circular groove K,
and a vertical axis indicates the oil film thickness of the
lubricant oil formed between the rotating shaft 13 and the main
bearing 21 and the contact force between the rotating shaft 13 and
the main bearing 21. In FIG. 6, a solid-line indicates the contact
force and a broken-line indicates the oil film thickness. Where the
depth of the circular groove K is indicated by a ratio to the shaft
diameter (diameter) D of the rotating shaft 13 (bearing hole
N).
[0071] When the depth of the circular groove K in which the inner
circumferential surface Kg is tapered becomes zero, the contact
force between the rotating shaft 13 and the main bearing 21 becomes
maximum (100), and the oil film is hardly formed. The oil film is
formed in the thinnest state at a point where the contact force is
weakened to some extent. The contact force decreases rapidly with
increasing depth of the circular groove K, and the oil film
thickness is thickened in inverse proportion to the decreasing
contact force.
[0072] Particularly, when the depth of the circular groove K
exceeds 0.4 (40% of the shaft diameter ratio), a degree to which
the contact force decreases changes from the rapidly decreasing
state to the gradually decreasing state, the oil film thickness
exceeds a necessary oil film thickness (1), and the oil film
thickness is maintained at 1 or more.
[0073] In the fluid lubrication "state in which only the oil film
of the lubrication oil is interposed between the rotating shaft 13
and the main bearing 21, the oil film thickness is thickened by
increasing the groove depth, a tilt of the rotating shaft 13
increases to keep the oil film thickness substantially constant
when the depth of the circular groove K becomes at least 40% of the
shaft diameter ratio of the rotating shaft 13.
[0074] On the other hand, the contact load of the rotating shaft 13
and the main bearing 21 in the mixed lubrication state exhibits a
characteristic in which the contact load can be reduced with
increasing depth of the circular groove K. However, when the depth
of the circular groove K becomes at least 40% of the shaft diameter
ratio of the rotating shaft 13, the tilt of the rotating shaft 13
increases, and a decreasing ratio of the contact load becomes
small.
[0075] In the circular groove K whose inner circumferential surface
Kq is tapered, the wall thickness b that is the distance between
the inner circumferential surface Kq and the bearing hole N becomes
minimum (thinnest) at the opening end Kd facing the cylinder
chamber S.
[0076] The minimum wall thickness b between the inner
circumferential surface Kg of the circular groove K and the
circumferential surface of the bearing hole N is set so as to
satisfy the relationship of the equation (1):
0.09.times.diameter D of bearing hole N.gtoreq.minimum wall
thickness b.gtoreq.0.04.times.diameter D of bearing hole N (1)
[0077] This can be described with reference to FIG. 7. FIG. 7 is a
characteristic diagram illustrating a circular groove minimum wall
thickness effect. In FIG. 7, the horizontal axis indicates the
minimum wall thickness (shaft diameter ratio) b of the circular
groove K, and the vertical axis indicates the contact force. In
FIG. 7, the solid-line change indicates the contact force, and a
maximum allowable contact force is set to 0.5.
[0078] When the minimum wall thickness b of the circular groove K
decreases excessively, a lack of rigidity is generated in the main
bearing 21, and the deformation becomes large. At this point, even
if the oil film thickness can be secured in the fluid lubrication
state, the contact load increases in the mixed lubrication
state.
[0079] On the other hand, when the minimum wall thickness b of the
circular groove K increases excessively, the rigidity increases
excessively to hardly generate the deformation, and the contact
load also increases in the mixed lubrication state. Therefore, the
proper value of the minimum wall thickness to the contact load is
set as illustrated in FIG. 7 and the equation (1).
[0080] Assuming that e is the eccentric amount of the eccentric
portion 13a that is provided integral with the rotating shaft 13
and r is the outer circumferential radius of the roller 25, the
outer circumferential radius g of the circular groove K is set so
as to satisfy the relationships of the equations (2) and (3):
0.5 mm.ltoreq.[outer circumferential radius r (mm) of roller
25-eccentric amount e (mm) of eccentric portion 13a]-outer
circumferential radius g (mm) of circular groove K (2)
outer circumferential radius g (mm) of circular groove
K>diameter D (mm) of bearing roller N/2+minimum wall thickness b
(mm) (3)
[0081] When the opening end Kd of the circular groove K is
communicated with the cylinder chamber S, the refrigerant
introduced to the cylinder chamber S remains partially in the
circular groove K, and the circular groove K becomes a dead volume.
Therefore, in order to prevent the dead volume of the circular
groove K, a minimum seal width is formed to exert a seal function
between an external diameter of the roller 25 and an external
diameter of circular groove K.
[0082] Particularly, the equation (2) can be described with
reference to FIG. 8. FIG. 8 illustrates a minimum seal width
effect. In FIG. 8, the horizontal axis indicates a minimum seal
width (mm), and the vertical axis indicates a performance
ratio.
[0083] The performance ratio is 0.2 when the minimum seal width
becomes 0, and the performance ratio does not change even if the
minimum seal width increases to about 0.3 mm. The performance ratio
increases when the minimum seal width exceeds about 0.3 mm, and the
performance ratio increases rapidly when the minimum seal width
exceeds 0.4 mm.
[0084] The performance ratio becomes a peak when the minimum seal
width is about 0.5 mm, and the performance ratio is substantially
kept constant even if the minimum seal width increases from about
0.5 mm.
[0085] In the equation (2), [outer circumferential radius r (mm) of
roller 25-eccentric amount e (mm) of eccentric portion 13a]-outer
circumferential radius g (mm) of circular groove K is the minimum
seal width. As can be seen from FIG. 8, the minimum seal width of
0.5 mm or more is required.
[0086] As described above, the inner circumferential surface Kq of
the circular groove K is tapered, and setting of a slope angle
.theta. becomes one of necessary conditions. That is, the contact
force between the rotating shaft 13 and the main bearing 21 varies
depending on the slope angle .theta.. The circular groove K is
formed such that the slope of the inner circumferential surface Kq
increases (the slope angle .theta. decreases) as much as possible,
thereby exerting the large contact load reducing effect.
[0087] FIG. 9 is a characteristic diagram illustrating a groove
slope effect. In FIG. 9, the horizontal axis indicates the slope of
the inner circumferential surface Kg of the circular groove K, and
the vertical axis indicates the contact force between the rotating
shaft 13 and the main bearing 21.
[0088] The contact force is maximized (1 or more) when the slope of
the circular groove K is close to zero (0). With increasing groove
slope, the contact force decreases, and therefore the oil film
thickness increases as described above.
[0089] Further, as illustrated in FIG. 2, there is another setting
condition that the main bearing 21 comprises the flange 21a whose
wall thickness H is set to the depth L of the circular groove K or
less.
[0090] Therefore, the rigidity of the coupling portion between the
cylindrical pivot portion 21b and the flange 21a that supports the
whole of the main bearing 21 is reduced to deform the whole of the
main bearing 21, whereby a property of following the rotating shaft
13 is enhanced to improve the effect of the circular groove K.
[0091] FIG. 3 is an enlarged longitudinal sectional view of a
compression mechanism portion 12 according to a second embodiment
of the invention. Because a basic configuration of a compression
mechanism portion 12 is identical to that of FIG. 2, the same
component is designated by the same numeral (only main part), and
the description of the same component is not repeated. In the
second embodiment, a diameter D1 of a portion (bearing hole Na)
that is journaled in a main bearing 21 of a rotating shaft 13
differs from a diameter D2 of a portion (bearing hole Nb) that is
journaled in a sub-bearing 22. Actually, the diameter D1 of the
portion journaled in the main bearing 21 of the rotating shaft 13
is formed larger than the diameter D2 of the portion journaled in
the sub-bearing 22 (D1>D2).
[0092] Because the diameter D1 is formed larger than the diameter
D2, it is necessary to secure a seal width of a circular groove K
with respect to a cylinder chamber S in an end face of a roller 25.
Therefore, an inner circumferential surface Kq of the circular
groove K is hardly tapered, and a groove Ka having an even width in
the depth direction is provided.
[0093] That is, the tapered inner circumferential surface Kq of the
circular groove K is provided only in a rotating shaft portion that
is journaled in the sub-bearing 22 having a small diameter, and the
seal width of the end face of the roller 25 is secured with respect
to the cylinder chamber S.
[0094] Because the length in the axis direction of the cylindrical
pivot portion 22b is shorter than that of the main bearing 21, the
flexural deformation becomes large, and the load also becomes
large. Therefore, the circular groove K whose inner circumferential
surface Kq is tapered is extremely advantageously provided.
[0095] In the circular groove K whose inner circumferential surface
Kq is tapered, the dimensions and configuration are similar to
those of the first embodiment, and the effect similar to that of
the first embodiment is obtained. However, the overlapping
description is not repeated.
[0096] FIG. 4 is a longitudinal sectional view illustrating a
hermetic type compressor 1A according to a third embodiment of the
invention with part of the hermetic type compressor 1A omitted.
[0097] Basically, the configuration in which a motor portion 11 and
a compression mechanism portion 12A that is coupled to the motor
portion 11 with a rotating shaft 13 interposed therebetween are
accommodated in a closed vessel 10 is similar to that of the first
embodiment.
[0098] The compression mechanism portion 12A is a two-cylinder type
compressor 1A that comprises two cylinders 20A and 20B that are
provided above and below an intermediate partition plate 30. Each
of the cylinders 20A and 20B comprises an internal diameter hole
Sa. The internal diameter hole Sa of the cylinder 20A on the upper
side is closed by a main bearing 21 and the intermediate partition
plate 30 to form the first cylinder chamber Sa.
[0099] The internal diameter hole Sb of the cylinder 20B on the
lower side is closed by a sub-bearing 22 and the intermediate
partition plate 30. Eccentric portions 13a and 13b and a roller 25
are accommodated in the first cylinder chamber Sa and the second
cylinder chamber Sb, respectively. The eccentric portions 13a and
13b are provided while integrated with the rotating shaft 13, and
the eccentric portions 13a and 13b have a phase difference of
180.degree.. The roller 25 is fitted in the eccentric portions 13a
and 13b.
[0100] A diameter of a portion journaled in the main bearing 21 of
the rotating shaft 13 is equal to a diameter of a portion journaled
in the sub-bearing 22. In other words, diameters of bearing holes N
made in the main bearing 21 and sub-bearing 22 are equal to each
other.
[0101] Circular grooves K opened to the cylinder chambers Sa and Sb
are provided in the main bearing 21 and the sub-bearing 22. An
inner peripheral surface of the circular groove K is tapered such
that a diameter of the inner peripheral surface increases gradually
from the surface facing each of the cylinder chambers Sa and Sb
toward the opposite side of the cylinder chamber. The depth of the
circular groove K is set to at least 40% of the diameter of the
bearing hole.
[0102] Because all the above-described setting conditions are
included in the hermetic type compressor 1A of the third
embodiments, similar effects are obtained in both the main bearing
21 and the sub-bearing 22.
[0103] FIG. 5 is a longitudinal sectional view illustrating a
hermetic type compressor 1B according to a fourth embodiment of the
invention with part of the hermetic type compressor 1A omitted.
[0104] Basically, the hermetic type compressor 1B of the fourth
embodiment comprises a compression mechanism portion 12B having a
configuration similar to that of the two-cylinder type compression
mechanism portion 12A of the third embodiment (see FIG. 4).
[0105] In the fourth embodiment, a diameter D1 of a portion
journaled in a main bearing 21 of a rotating shaft 13 differs from
a diameter D2 of a portion journaled in a sub-bearing 22. The
diameter D1 of the portion journaled in the main bearing 21 of the
rotating shaft 13 is formed larger than the diameter D2 of the
portion journaled in the sub-bearing 22 (D1>D2).
[0106] Accordingly, in the compression mechanism portion 12B,
similarly to the compression mechanism portion 12 of the second
embodiment (see FIG. 3), because the diameter D1 is formed larger
than the diameter D2, it is necessary to secure a seal width of a
circular groove K with respect to a cylinder chamber S in an end
face of a roller 25. Therefore, an inner circumferential surface Kg
of the circular groove K is hardly tapered, and a groove Ka having
an even width in the depth direction is provided.
[0107] The tapered inner circumferential surface Kq of the circular
groove K is provided only in a portion of a rotating shaft 13 that
is journaled in the sub-bearing 22 having a small diameter, and the
seal width of the end face of the roller 25 is secured with respect
to the cylinder chamber S.
[0108] Because the length in the axis direction of the cylindrical
pivot portion 22b is shorter than that of the main bearing 21, the
flexural deformation becomes large, and the load also becomes
large. Therefore, the circular groove K whose inner circumferential
surface Kq is tapered is extremely advantageously provided.
[0109] FIG. 10 is a longitudinal sectional view of the hermetic
type compressor 1A according to a modification of the third
embodiment of the invention, and the refrigeration cycle is omitted
in FIG. 10.
[0110] Basically, the hermetic type compressor 1A of the
modification of the third embodiment comprises the two-cylinder
type compression mechanism portion 12A of the third embodiment (see
FIG. 4), diameters of bearing holes N made in a main bearing 21 and
a sub-bearing 22 are equal to each other, and the main bearing 21
and the sub-bearing 22 comprise circular grooves K.
[0111] In the modification of the third embodiment, a discharge
valve mechanism 27 for a first cylinder chamber Sa is provided in
the main bearing 21, a discharge valve mechanism 27 for a second
cylinder chamber Sb is provided in the sub-bearing 22, and a
discharge valve mechanism 27A for the first cylinder chamber Sa and
a discharge valve mechanism 27A for the second cylinder chamber Sb
are provided in an intermediate partition plate 30A that is
interposed between two cylinders 20A and 20B.
[0112] Because the intermediate partition plate 30A comprises the
two discharge valve mechanisms 27A, the intermediate partition
plate 30A is divided into two in a thickness direction. As
described later, the two discharge valve mechanisms 27A of the
intermediate partition plate 30A are mounted while overlapping each
other when viewed from above.
[0113] FIG. 11 is a plan view of the intermediate partition plate
30A when viewed from a side of a surface in which the discharge
valve mechanisms 27A overlap each other.
[0114] As illustrated by a solid-line arrow of FIG. 11, a gas
refrigerant that is discharged from a discharge holes 26 made in
each of the divided intermediate partition plates 30A is guided to
the outside from a communication hole 32 through a groove 31
provided in each of the intermediate partition plates 30A.
[0115] FIG. 12 is a longitudinal sectional view of a region where
the discharge valve mechanisms 27A are provided in the intermediate
partition plates 30A divided into two.
[0116] The discharge valve mechanism 27A comprises a discharge
valve 33 and a discharge valve guard 34a. One end of the discharge
valve 33 is supported while separated from a discharge hole 26. The
discharge valve 33 is formed by a thin spring plate, and the other
end of the discharge valve 33 is in close contact with the
discharge hole 26 so as to close the discharge hole 26. The
discharge valve guard 34a is formed by a thick plate piece having
rigidity, and the discharge valve guard 34a is gently bent from a
support portion at one end toward the discharge hole 26 at the
other end.
[0117] A pressure at each of cylinder chambers Sa and Sb increases
by the compression action of the refrigerant, the discharge valve
33 is pressed when the pressure reaches a predetermined value, and
the discharge valve 33 is elastically deformed to open the
discharge hole 26. Accordingly, the high-pressure gas refrigerant
compressed by each of the cylinder chambers Sa and Sb is discharged
from the discharge hole 26. The discharge valve guard 34a receives
the elastically-deformed discharge valve 33 to regulate further
deformation, thereby preventing metal fatigue of the discharge
valve 33 as much as possible.
[0118] The discharge valve guard 34a has a specific thickness
because the discharge valve guard 34a has the necessary rigidity.
One end of the discharge valve guard 34a mounted on the
intermediate partition plate 30A is formed into a flat shape, and
the discharge valve guard 34a is bent into a predetermined curved
shape from the flat-shape leading end to the other end facing the
discharge hole 26. Therefore, the leading end of the discharge
valve guard 34a is formed at a certain level from a flat surface
formed in the mounting portion.
[0119] When the intermediate partition plate 30A directly comprises
the discharge valve mechanism 27A, the wall thickness of the
intermediate partition plate 30A increases considerably, and the
compression mechanism portion 12A is lengthened in the axis
direction, which leads to enlargement of the compressor 1A.
[0120] When the intermediate partition plate 30A is thickened, an
interval between the first cylinder chamber Sa and the second
cylinder chamber Sb is lengthened, and the distance between the
eccentric portions 13a of the rotating shafts 13 that are
accommodated in the first cylinder chamber Sa and the second
cylinder chamber Sb. This leads to the degradation of the rigidity
of the rotating shaft 13 to cause the increase of flexural
deformation, amplification of wobbling, and the degradation of the
reliability.
[0121] Therefore, as illustrated in a first example of FIG. 12, in
the discharge valve guard 34a, a flat portion mounted on the
intermediate partition plate 30A has the same wall thickness, and a
bent portion U facing the discharge hole 26 is tapered such that a
wall thickness decreases gradually toward the leading end and such
that the wall thickness in section becomes the thinnest in the
leading end portion.
[0122] Because the discharge valve guard 34a receives the force of
the discharge valve 33, strength is required for the discharge
valve guard 34a, and the discharge valve guard 34a is formed with a
predetermined thickness. However, a stress is not applied to the
leading end of the bent portion U too much, and no problem occurs
even if a section of the leading end of the bent portion U is
thinned into the tapered shape.
[0123] Therefore, a height of the discharge valve guard 34a can be
reduced to decrease the wall thickness of the intermediate
partition plate 30A. As the height of the compression mechanism
portion 12A is reduced, the distance between the eccentric portions
13a of the rotating shafts 13 can be shortened to reduce the
flexural deformation or wobbling of the rotating shaft 13, thereby
improving the reliability.
[0124] Meanwhile, the discharge valve mechanisms 27 of the main
bearing 21 and sub-bearing 22 are removed, the discharge valve
mechanism 27A for the first cylinder chamber Sa and the discharge
valve mechanism 27A for the second cylinder chamber Sb may be
provided only in the intermediate partition plate 30A.
[0125] Alternatively, as illustrated in a second example of FIG.
13, only a leading end Z of the bent portion is processed, although
the plate thickness is evenly formed from the mounting portion to
the bent portion without changing a configuration of the discharge
valve guard 34a.
[0126] That is, at the leading ends Z of the discharge valve guards
34a, surfaces facing each other that are the surfaces that do not
collide with the discharge valves 33 are cut into flat shapes so as
to be parallel. Therefore, the distance between the mounting
portions of the two discharge valve guards 34a can further be
shortened to minimize the thickness of the intermediate partition
plate 30A, so that the above-described effect is obtained.
[0127] The invention is not limited to the embodiments, but various
modifications can be made at an implementation stage without
departing from the scope of the invention. Various inventions can
be made by appropriately combining a plurality of constituents
disclosed in the embodiments.
[0128] According to the invention, according to the flexural
deformation of the rotating shaft due to the compressive load in
the cylinder, the uneven contact with the rotating shaft is
prevented in at least one of the main bearing and sub-bearing,
thereby achieving the enhancement of the reliability and the longer
operation life. Additionally, the hermetic type compressor is
provided to form the refrigeration cycle, thereby improving
refrigeration efficiency.
* * * * *