US20030113211A1

US20030113211A1 - Piston type compressor

Info

Publication number: US20030113211A1
Application number: US10/292,043
Authority: US
Inventors: Takahiro Moroi; Tomoji Tarutani; Masahiro Kawaguchi; Noriyuki Shintoku; Masaki Ota; Hisato Kawamura; Shigeki Kawachi; Yoshinori Inoue; Akio Saiki
Original assignee: Toyota Industries Corp
Current assignee: Toyota Industries Corp
Priority date: 2001-11-12
Filing date: 2002-11-12
Publication date: 2003-06-19
Also published as: DE10252447B4; CN1419051A; JP3891099B2; DE10252447A1; JP2003239856A

Abstract

A rotary valve is used as a suction valve mechanism of a compressor. The rotary valve rotates integrally with a drive shaft to selectively open or close a suction passage of refrigerant gas from a suction chamber to a compression chamber in accordance with relative positions of the suction communicating passage and the suction guiding groove. An outer surface of the rotary valve directly slides on an inner surface of a valve accommodating chamber formed in a cylinder block, thereby constituting a slide bearing surface. The slide bearing surface rotatably supports the rear end of the drive shaft on housings. Accordingly, a piston type compressor which is comfortably used, inexpensively manufactured, and has high compression efficiency is provided.

Description

BACKGROUND OF THE INVENTION

The present invention relates to a piston type compressor for compressing gas by reciprocating movement of a piston.

A suction valve of a reed valve type is generally used for the piston type compressor. A suction stroke of the piston causes a reduction in a pressure of a compression chamber. The suction valve of the reed valve type is opened by a difference between the reduced pressure of the compression chamber and a pressure of a suction pressure region to permit suction of gas from the suction pressure region into the compression chamber.

However, the suction valve of the reed valve type generates abnormal noise caused by self-excited vibration, consequently inhibit comfortable use of the compressor. Therefore, Japanese Unexamined Patent Publication Nos. Hei 5-312146 and Hei 7-63165 disclose suction valves of rotary valve types which generate no self-excited vibration.

In Japanese Unexamined Patent Publication No. Hei 5-312146, as shown in FIG. 12, a

drive shaft

94 is rotatably supported on a housing 91 of a compressor 90 through rolling bearings 95 a and 95 b. In

cylinder blocks

92 and 93 constituting a part of the housing 91, a valve accommodating chamber 96 for rotatably accommodating a rotary valve 97 is formed in a position coaxial to the drive shaft 94. A suction communicating passage 98 communicated with a compression chamber 99 is opened on the inner surface of the valve accommodating chamber 96. The rotary valve 97 is fixed to the drive shaft 94 so as to be integrally rotated. A suction guiding groove 100 always communicated with a suction pressure region is formed on the outer surface of the rotary valve 97.

The

rotary valve

97 can connect or disconnect the suction guiding groove 100 to or from the suction communicating passage 98 in accordance with its rotational position. Especially, the rotary valve 97 connects the suction guiding groove 100 to the suction communicating passage 98 during a suction stroke of a piston to permit suction of gas from the suction pressure region into the compression chamber 99.

In the prior art described above, the

drive shaft

94 is supported on the housing 91 through the rolling bearings 95 a and 95 b. The rolling bearings are complex in structure, and high-precision finishing is required, consequently increasing manufacturing costs. Costs of the compressor 90 that uses the rolling bearings 95 a and 95 b are inevitably increased. In addition, core misalignment occurs because of an error in positional accuracy between the inner surface of the valve accommodating chamber 96 and an inner surface of each of accommodating chambers for accommodating the rolling bearing 95 a and 95 b in the housing 91. Consequently, there is a possibility that galling may occur between the outer surface of the rotary valve 97 and the inner surface of the valve accommodating chamber 96. Therefore, the accommodating chamber 96 and the accommodating chambers of the rolling bearings 95 a and 95 b must all be machined with high precision. Hence, in the compressor that employs the rolling bearings, not only costs of the rolling bearings but also manufacturing costs of the compressor for housing the rolling bearings are high. Such compressor as an end product is inevitably expensive.

If a clearance between the outer surface of the

rotary valve

97 and the inner surface of the valve accommodating chamber 96 is set large, the core misalignment can be absorbed by the clearance, i.e., allowance, thereby eliminating galling. However, if the clearance is set large between the rotary valve 97 and the valve accommodating chamber 96, leakage of gas through the clearance causes a reduction in compression efficiency of the compressor.

In Japanese Unexamined Patent Publication No. Hei 7-63165, as shown in FIG. 13, a

rotary valve

102 is integrally formed in the end of a drive shaft 101. In FIG. 13, slide bearings are arranged on the inner surface of a valve accommodating chamber 105. That is, supporting of the drive shaft 101 by a cylinder block 106 is carried out through the rotary valve 102 and the

slide bearings

103 and 104.

In the prior art described above, the slide bearing 104 is provided between the outer surface of the rotary valve 102 and the inner surface of the valve accommodating chamber 105, in other words, between a suction guiding groove 107 of the rotary valve 102 and a suction communicating passage 108 of the cylinder block 106. Accordingly, in order to increase compression efficiency of a compressor, leakage of gas must be prevented both between the outer surface of the rotary valve 102 and the slide bearing 104 and between the slide bearing 104 and the inner surface of the valve accommodating chamber 105. Therefore, it is necessary to machine the valve accommodating chamber 105, into which the slide bearing 104 is inserted, and the inner surface of the slide bearing 104, into which the rotary valve 102 is inserted, with high precision. This requirement increases costs of the compressor.

An object of the present invention is to provide a piston type compressor guaranteed for prevention of noise generated during use, and good use.

Another object of the present invention is to provide an inexpensive compressor.

Still another object of the present invention is to provide a compressor having excellent compression efficiency.

BRIEF SUMMARY OF THE INVENTION

A piston type compressor includes a drive shaft rotatably supported on a housing. The housing includes a cylinder block. The cylinder block includes a cylinder bore and a valve accommodating chamber. Reciprocation of a piston operably connected to the drive shaft in the corresponding cylinder bore changes the volume of a compression chamber in the cylinder bore to compress gas supplied from a first region, on which a suction pressure acts, to the compression chamber. The compressed gas is sent to a second region, on which a discharge pressure acts. The compressor includes a rotary valve accommodated in the valve accommodating chamber. The rotary valve rotates integrally with the drive shaft to selectively open or close a passage of gas from the first pressure region to the compression chamber. The rotary valve has a substantially cylindrical shape and an outer surface. The valve accommodating chamber has a circular cross-section and an inner surface. The outer surface and the inner surface constitute slide bearing surfaces for receiving a radial load applied to the drive shaft by sliding on each other.

Other aspects and advantages of the invention will become apparent from the following description, taken in conjunction with the accompanying drawings, illustrating by way of example the principles of the invention.

BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS

The invention, together with objects and advantages thereof, may best be understood by reference to the following description of the presently preferred embodiments together with the accompanying drawings in which: [0015]
FIG. 1 is a vertical sectional view of a piston type variable displacement compressor. [0016]
FIG. 2 is a partially enlarged side view of a rotary valve of the compressor of FIG. 1. [0017]
FIG. 3 is a transverse sectional view of a portion in the vicinity of the rotary valve taken along the line [0018] 1-1 of FIG. 1.
FIG. 4 is a transverse sectional view of a portion in the vicinity of a rotary valve according to a second embodiment. [0019]
FIG. 5 is a vertical sectional view of a portion in the vicinity of a rotary valve according to a third embodiment. [0020]
FIG. 6 is a vertical sectional view of a portion in the vicinity of a rotary valve according to a fourth embodiment. [0021]
FIG. 7 is a vertical sectional view of a piston type variable displacement compressor according to a fifth embodiment. [0022]
FIG. 8 is a vertical sectional view of a portion in the vicinity of a rotary valve according to another example of FIG. 6. [0023]
FIG. 9 is a vertical sectional view of a portion in the vicinity of a rotary valve according to yet another example of FIG. 6. [0024]
FIG. 10 is a vertical sectional view of a double-head piston type compressor. [0025]
FIG. 11 is a partial vertical sectional view of another embodiment of the compressor of FIG. 10. [0026]
FIG. 12 is a vertical sectional view of a compressor according to the prior art. [0027]
FIG. 13 is a vertical sectional view of a compressor according to the prior art.[0028]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Next, description will be made of first to fourth embodiments of piston type variable displacement compressors used for vehicle air conditioners. For the second to fifth embodiments, only differences from the first embodiment will be described. Identical or equivalent members will be denoted by similar reference numerals, and description thereof will be omitted. In FIG. [0029] 1, left and right sides respectively show front and rear portions of a compressor.
First Embodiment [0030]
As shown in FIG. 1, a housing of the piston type variable displacement compressor (simply referred to as a compressor, hereinafter) includes a [0031] cylinder block 11, a front housing 12 and a rear housing 14. The cylinder block 11 is made of an aluminum-series metallic material in order to reduce weight of the compressor. The front housing 12 is coupled to the front end of the cylinder block 11. The rear housing 14 is coupled through a valve plate 13 to the rear end of the cylinder block 11.
A [0032] crank chamber 15 is defined by the cylinder block 11 and the front housing 12. A drive shaft 16 is rotatably arranged in the crank chamber 15. The drive shaft 16 is made of an iron-series metallic material. The drive shaft 16 is operably connected to an engine (not shown) as a driving source for traveling a vehicle, and rotated by receiving power from the engine.
A lug plate [0033] 21 is fixed onto the drive shaft 16 in the crank chamber 15 so as to be integrally rotated. A swash plate 23 is housed in the crank chamber 15. The swash plate 23 is supported on the drive shaft 16 so as to slide and tilt. A hinge mechanism 24 is provided between the lug plate 21 and the swash plate 23. Accordingly, the swash plate 23 can be rotated integrally with the lug plate 21 and the drive shaft 16 by hinge connection with the lug plate 21 through the hinge mechanism 24, and support by the drive shaft 16. The swash plate 23 can also tilt with respect to the drive shaft 16, and slide in an axial direction of the drive shaft 16.
A plurality of cylinder bores [0034] 11 a (only one is shown in FIG. 1) are formed in the cylinder block 11. The cylinder bores 11 a are formed to surround a rear end of the drive shaft 16. A single-head piston 25 is housed to be reciprocated in each cylinder bore 11 a. Openings on front and rear sides of the cylinder bore 11 a are closed by the valve plate 13 and the piston 25. The cylinder bore 11 a defines a compression chamber 26 changed in volume in accordance with the reciprocation of the piston 25. Each piston 25 is retained on the outer peripheral portion of the swash plate 23 through shoes 27. Thus, rotation of the swash plate 23 accompanying rotation of the drive shaft 16 is converted through the shoes 27 into reciprocation of the piston 25.
A [0035] suction chamber 28 and a discharge chamber 29 are defined in the rear housing 14. The suction chamber 28 is formed in the center of the rear housing 14. The discharge chamber 29 is formed to surround the outer surface of the suction chamber 28. An discharge port 32 and a discharge valve 33 are formed in the valve plate 13. The discharge port 32 communicates the compression chamber 26 and the discharge chamber 29 with each other. The discharge valve 33 is a reed valve for selectively opening/closing the discharge port 32. A suction valve mechanism 35 equipped with a rotary valve 41 is provided in the cylinder block 11.
When the [0036] piston 25 is moved from a top dead center to a bottom dead center, refrigerant-gas of the suction chamber 28 is sucked through the suction valve mechanism 35 into the compression chamber 26 (suction stroke). When the piston 25 is moved from the bottom dead center to the top dead center, the refrigerant gas sucked into the compression chamber 26 is compressed to a predetermined pressure, and then discharged through the discharge port 32 and the discharge valve 33 to the discharge chamber 29 (discharge stroke).
A [0037] bleed passage 36 and a supply passage 37 is formed in the housing of the compressor. The bleed passage 36 connects the crank chamber 15 with the suction chamber 28. The supply passage 37 connects the discharge chamber 29 with the crank chamber 15. A control valve 38 is provided in the supply passage 37. The supply passage 37 is an electromagnetic valve.
Varying the opening size of the [0038] control valve 38 controls the balance of the amount of incoming high-pressure discharge gas, which flows from the discharge chamber 29 through the supply passage 37 to the crank chamber 15, and the amount of outgoing gas, which flows from the crank chamber 15 through the bleed passage 36 to the suction chamber 28. Thus, the internal pressure of the crank chamber 15 is determined. The change in the internal pressure of crank chamber 15 cause the difference between the internal pressure of the crank chamber 15 and the internal pressure of the compression chamber 26 via the piston 25 to change. Then, Thus, the inclination angle of the swash plate 23 is varied and piston stroke, or the compressor displacement, is adjusted.
As shown in FIGS. 1 and 2, a [0039] valve accommodating chamber 42 for accommodating the rotary valve 41 is formed in the housing of the compressor. The accommodating chamber 42 is extended through the center surrounded with the cylinder bore 11 a in the cylinder block 11 to the center of the rear housing 14. The valve accommodating chamber 42 has a circular section, and communicated with the suction chamber 28 at its rear end. The valve accommodating chamber 42 and each compression chamber 26 are communicated with each other through each of a plurality of suction communicating passages 43 (see FIG. 3) formed in the cylinder block 11.
The [0040] rotary valve 41 is rotatably accommodated in the valve accommodating chamber 42. The rotary valve 41 is made of an aluminum-series metallic material. The rotary valve 41 has a cylindrical shape. A hole 41 a is bored through the center of a bottom in the forward portion of the rotary valve 41. The rear end of the drive shaft 16 is placed in the valve accommodating chamber 42. The drive shaft 16 is fixed to the rotary valve 41 by fitting a small-diameter portion of its rear end into the hole 41 a. Accordingly, the rotary valve 41 and the drive shaft 16 are integrated, and aligned along one axis. The rotary valve 41 is rotated in synchronization with rotation of the drive shaft 16, i.e., reciprocation of the piston 25.
As shown in FIG. 3, an internal space of the [0041] rotary valve 41 forms an introduction chamber 44 communicated with the suction chamber 28. A suction guiding groove 45 is formed in a fixed section of a circumferential direction on an outer surface 41 b of the rotary valve 41. The suction guiding groove 45 and the suction communicating passage 43 constitute a refrigerant gas passage between the introduction chamber 44 as a suction pressure region and the compression chamber 26. By its rotation, the rotary valve 41 selectively opens/closes the refrigerant gas passage from the suction pressure region to the compression chamber 26 in accordance with relative positions of the suction communicating passage and the suction guiding groove.
When the [0042] piston 25 changes to a suction stroke, the rotary valve 41 is rotated in a direction where one surface 45 a of the suction guiding groove 45 precedes to open the suction communicating passage 43. Thus, the refrigerant gas of the suction chamber 28 is sucked through the introduction chamber 44 and the suction guiding groove 45 of the rotary valve 41 and the suction communicating passage 43 of the cylinder block 11 into the compression chamber 26.
When the [0043] piston 25 finishes its suction stroke, the rotary valve 41 is rotated in a direction where the other surface 45 b of the suction guiding groove 45 closes the suction communicating passage 43, and stops the suction of the refrigerant gas into the compression chamber 26. When the piston 25 changes to a discharge stroke, the suction communicating passage 43 is accommodated in a closed state by the outer surface 41 b of the rotary valve 41. Thus, compression of the refrigerant gas and discharge thereof to the discharge chamber 29 make progress.
As shown in FIG. 1, the front end of the [0044] drive shaft 16 is rotatably supported on the front housing 12 through a front bearing 47 composed of a rolling bearing. The rear end of the drive shaft 16 is rotatably supported on the housings 11, 12 and 14 by direct sliding of the outer surface 41 b of the rotary valve 41 and an inner surface 42 a of the valve accommodating chamber 42. That is, the rotary valve 41 constitutes a slide bearing surface for supporting the rear end side of the drive shaft 16 by its outer surface 41 b together with the inner surface 42 a of the valve accommodating chamber 42, and receiving a radial load.
A thrust load to the front side of an axis, acted on the [0045] drive shaft 16, is received by a thrust bearing 17 between the inner wall surface of the front housing 12 and the lug plate 21. The bearing 17 is composed of a rolling bearing. A thrust load to the rear side of the axis, acted on the drive shaft 16, is received by sliding of a rear end surface 41 f of the rotary valve 41 on an inner wall surface 14 a of the rear housing 14. The press-fit position of the rotary valve 41 relative to the drive shaft 16, or the press-fit distance between the rotary valve 41 and the drive shaft 16, is arranged such that the movable distance of the drive shaft 16 in the axial direction is within 100 μm.
Although not shown, a coating (e.g., later-described [0046] coating 48 of FIG. 2) is applied to the rear end surface 41 f of the rotary valve 41 so as to improve sliding characteristics between the rear housing 14 and the rotary valve 41. Other than the rear end surface 41 f, this coating may also be applied to the inner wall surface 14 a of the rear housing 14, or on both of the surfaces 41 f and 14 a. In place of the coating, a thrust bearing composed of a rolling bearing may be provided between the rear end surface 41 f of the rotary valve 41 and the inner wall surface 14 a of the rear housing 14.
The [0047] inner surface 42 a of the valve accommodating chamber 42 is set to a largest diameter on the assumption that strength of the cylinder block 11, i.e., predetermined minimum strength of a portion between the valve accommodating chamber 42 and the cylinder bore 11 a, can be maintained. Thus, the rotary valve 41 accommodated in the valve accommodating chamber 42 also has a large diameter. In the present embodiment, this diameter is larger than that of the drive shaft 16.
As shown in FIG. 2, the [0048] coating 48 is applied to the entire outer surface 41 b of the rotary valve 41 so as to improve sliding characteristics between the outer surface 41 b and the inner surface 42 a of the valve accommodating chamber 42. FIG. 2 shows only a part of the coating 48, which is indicated by a netted portion. The coating 48 is made of, for example a fluorocarbon resin. The fluorocarbon resin includes polytetrafluoroethylene (PTFE), tetrafluoroethylene and perfluoroalkylvinyl ether copolymer (PFA), tetrafluoroethylene and hexafluoropropylene copolymer (FEP), tetrafluoroethylene and ethylene copolymer (ETFE), polyvinylidene fluoride (PVDF), and polychlorotrifluoroethylene (PCTFE).
A [0049] helical pumping groove 49 is formed around an axis of the drive shaft 16 on the outer surface 41 b of the rotary valve 41. When the rotary valve 41 is rotated, the pumping groove 49 is operated as a pump in cooperation with the inner surface 42 a of the valve accommodating chamber 42. Therefore, between the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42, refrigerants present between the peripheral surfaces 41 b and 42 a and lubricating oil contained therein are caused to actively flow. A helical direction of the pumping groove 49, i.e., a flowing direction of the refrigerants and the lubricating oil, may be set in a direction of the crank chamber 15 (left side in the drawing) or in a direction of the suction chamber 28 (right side). In FIG. 2, the helical direction of the pumping groove 49 is set in the direction of the crank chamber 15.
The first embodiment provides the following advantages. [0050]
(1) The [0051] rotary valve 41 constitutes the slide bearing surface by its outer surface 41 b together with the inner surface 42 a of the valve accommodating chamber 42. The drive shaft 16 is rotatably supported on the housings 11, 12 and 14 through the rotary valve 41. That is, the valve accommodating chamber 42 serves not only to accommodate the rotary valve 41 but also to bear and accommodate the drive shaft 16. Thus, by machining only the valve accommodating chamber 42 (i.e., inner surface 42 a) with high precision, galling caused by core misalignment between the valve accommodating chamber 42 and the bearing accommodating chamber of the drive shaft 16 in the prior art can be eliminated. Also, leakage of gas from the clearance between the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 can be prevented. Therefore, generation of noise can be prevented, and a compressor that is inexpensive and has high compression efficiency can be provided.
(2) The rotary valve [0052] 41 (slide bearing) has a larger diameter than the drive shaft 16. Thus, a surface pressure applied to the outer surface 41 b of the rotary valve 41 can be reduced, and a peripheral velocity of the rotary valve 41 can be increased. Thus, even in the case of oil film cutting which generally occurs easily when the slide bearing receives a high load and rotates at a low rotational speed, oil film cutting between the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 can be prevented. Therefore, it is possible to improve durability of the slide bearings 41 and 42.
When the diameter of the [0053] rotary valve 41 is large, the diameter of the inner surface 42 a of the valve accommodating chamber 42 becomes also large. In the cylinder block 11, a thickness of a portion between the cylinder bore 11 a and the valve accommodating chamber 42 can be reduced. Thus, the suction communicating passage 43 can be as short as possible, and volume efficiency of the compressor can be improved by reducing a dead volume of the compressor chamber 26.
(3) The [0054] drive shaft 16 and the rotary valve 41 are provided separately. Thus, for example, a shape dimension and a quality of material of the rotary valve 41 are independent of machining and functional restrictions of the drive shaft 16. The shape dimension and the quality of material of the rotary valve 41 can be selected by giving priority to its functions (including the function of the slide bearing).
That is, for the [0055] drive shaft 16, a straight shape (no outer diameter irregularities) is suitable for improving machining characteristics, and an iron-series metallic material is suitable when durability is taken into consideration.
On the other hand, for the [0056] rotary valve 41, a diameter as large as possible is suitable for improving its durability. A material similar to that of the cylinder block 11, for example, an aluminum-series metallic material, is suitable for preventing enlargement of a clearance between the cylinder block 11 and the rotary valve 41 caused by a difference of a coefficient of thermal expansion from that of the valve accommodating chamber 42 (cylinder block 11).
(4) The [0057] coating 48 applied to the outer surface 41 b of the rotary valve 41 improves sliding characteristics between the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42. Thus, even when the rotary valve 41 and the valve accommodating chamber 42 (cylinder block 11) are made of similar materials, it is possible to prevent adhesion caused by sliding between the cylinder block 11 and the rotary valve 41.
(5) The pumping [0058] groove 49 is formed on the outer surface 41 b of the rotary valve 41. Thus, for example, refrigerants and/or lubricating oil contained therein are caused to actively flow through the clearance between the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 by the pumping operation of the pumping groove 49. Therefore, it is possible to improve sliding characteristics between the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42. Moreover, the pumping groove 49 can be easily machined on the outer surface 41 b of the rotary valve 41.
Provision of the helical direction of the pumping [0059] groove 49 in the direction of the crank chamber 15 enables lubricating oil to be supplied to the crank chamber 15. Thus, a lubricating state between the swash plate 23 and the shoes 27 can be improved, thereby the durability of the compressor is improved. Conversely, provision of the helical direction of the pumping groove 49 in the direction to the suction chamber 28 prevents leakage of refrigerant gas from the suction pressure region, i.e., the suction chamber 28 or the introduction chamber 44, to the crank chamber 15. Therefore, the compression efficiency of the compressor can be much more increased.
As described above, a slide bearing surface can be easily formed between the [0060] outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42.
Second Embodiment [0061]
As shown in FIG. 4, according to a second embodiment, a [0062] supercharging blade 51 which serves as supercharging means is arranged between one surface 45 a and the other surface 45 b of a suction guiding groove 45 of a rotary valve 41. The supercharging blades 51 are provided in a plurality of places (three places in FIG. 4) to have angles substantially equal to those of the surfaces 45 a and 45 b.
Accordingly, when the [0063] rotary valve 41 is rotated in an arrow direction of FIG. 4 in accordance with the rotation of the drive shaft 16, refrigerant gas supplied from an introduction chamber 44 to the suction guiding groove 45 is transferred to a position communicated with a suction communicating passage 43 through a space formed by the end surfaces 45 a and 45 b of the suction guiding groove 45, the supercharging blades 51 adjacent thereto, and an inner surface 42 a of a valve accommodating chamber 42, or a space formed by the adjacent supercharging blades 51 and the inner surface 42 a of the valve accommodating chamber 42. The refrigerant gas that has been transferred to the position communicated with the suction communicating passage 43 is sent toward the suction communicating passage 43 with the aid of centrifugal force applied by rotation of the supercharging blades 51 or the surfaces 45 a and 45 b.
In the second embodiment, the supercharging [0064] blade 51 actively supplies the refrigerant gas to a compression chamber 26. Thus, a large amount of refrigerant gas can be sucked into the compressor 26, thereby increasing further compression efficiency of the compressor.
In addition, by supercharging of the refrigerant gas, refrigerants and/or lubricating oil contained therein are suitably supplied to transverse a clearance between an [0065] outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42. Thus, sliding characteristics between the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 can be improved. Therefore, the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 can easily constitute a slide bearing surface.
Third Embodiment [0066]
As shown in FIG. 5, according to a third embodiment, a [0067] rotary valve 41 is received by a static pressure in a valve accommodating chamber 42. That is, a hole 55 is bored through as a pressure supply passage on a wall of the rotary valve 41. The hole 55 is arranged in a region opposite a suction guiding groove 45 with respect to an axis of a drive shaft 16.
The [0068] hole 55 communicates an introduction chamber 44 with a clearance between the rotary valve 41 and the valve accommodating chamber 42. Accordingly, by application of centrifugal force based on rotation of the rotary valve 41, refrigerants in the introduction chamber 44 and/or lubricating oil contained therein are supplied through the hole 55 to the clearance between the rotary valve 41 and the valve accommodating chamber 42. Thus, the rotary valve 41 is received by a static pressure in the valve accommodating chamber 42.
According to the third embodiment, improved sliding characteristics can be provided between an [0069] outer surface 41 b of the rotary valve 41 received by the static pressure in the valve accommodating chamber 42, and an inner surface 42 a of the valve accommodating chamber 42. Therefore, the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 can easily constitute a slide bearing surface.
In a region where the clearance between the [0070] outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 is adjacent to the suction guiding groove 45, proper leakage of refrigerants and/or lubricating oil contained therein through the suction guiding groove 45, therefore good sliding characteristics, can be expected. However, these advantages cannot be expected in a region where a clearance is not adjacent to the suction guiding groove 45.
However, according to the third embodiment, the [0071] hole 55 is arranged in the region opposite the suction guiding groove 45 with respect to the axis of the drive shaft 16. Thus, the refrigerants and/or lubricating oil contained therein are even supplied through the hole 55 to the region where the clearance is not adjacent to the suction guiding groove 45, thereby providing good sliding characteristics. This arrangement also facilitates constitution of a slide bearing surface by the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42.
Fourth Embodiment [0072]
As shown in FIG. 6, according to a fourth embodiment, an [0073] outer surface 41 b of a rotary valve 41 and an inner surface 42 a of a valve accommodating chamber 42 are tilted in such a direction as to approach an axis of a drive shaft 16 toward the rear side of a compressor, and formed in a taper state. Sliding to the rear side along the axis of the drive shaft 16 is regulated by abutment between the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42. That is, a slide bearing surface constituted of the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 receives not only the radial load acted on the drive shaft 16 described above with reference to the first to third embodiments, but also a thrust load acted rearward on the drive shaft 16.
In order to suitably receive both of the radial load and the thrust load acted on the [0074] drive shaft 16, tilt angles of slide bearing surfaces 41 b and 42 a with respect to the axis of the drive shaft 16 are set in a range of greater than 0° to 10° or less, more preferably in a range of 0.5° to 1°. In FIG. 6, for easier understanding, tilt angles of the slide bearing surfaces 41 b and 42 a are shown with exaggeration.
In the fourth embodiment, the slide bearing surfaces [0075] 41 b and 42 a can also receive the thrust load acted on the drive shaft 16. Thus, it is not necessary to provide means for receiving the thrust load between a rear end surface 41 f of the rotary valve 41 and an inner wall surface 14 a of a rear housing 14, thereby simplifying a structure of the compressor.
Fifth Embodiment [0076]
As shown in FIG. 7, according to a fifth embodiment, the [0077] bleed passage 36 is provided in the drive shaft 16 and in the front end of the rotary valve 41 in their axial direction. The bleed passage 36 includes a restriction 36 a at its downstream side where the bleed passage 36 connects with the introduction chamber 44. Thus, a refrigerant gas in the crank chamber 15 is introduced into the introduction chamber 44 through the bleed passage 36 and the restriction 36. Due to the drastically reduced cross-section in the restriction 36 and centrifugal force applied by rotation of the drive shaft 16, a lubricating oil is separated from refrigerant gas at the upstream side of the restriction 36 a.
In this embodiment, the front end of the [0078] rotary valve 41 has a smaller portion 41 g the diameter of which is smaller than that of the drive shaft 16. The rotary valve 41 is press-fit through the smaller portion 41 g into the attachment hole 16 b in the rear end of the drive shaft 16. A oil returning hole 57 is formed through the drive shaft 16 at the overlapping portion of the smaller portion 41 g of the rotary valve 41 and the rear end of the drive shaft 16. The hole 57 communicates the bleed passage 36 at the upstream side of the restriction 36 a with the crank chamber 15. Thus, the separated oil, which is separated from refrigerant gas at the upstream of the restriction 36 a, is returned through the hole 57 to the crank chamber 15.
The fifth embodiment has same advantages as the first embodiment. In addition, a lubricating oil, which is discharged from crank [0079] chamber 15 together with refrigerant gas, may be separated from refrigerant gas in the bleed passage 36 and returned to the crank chamber 15 rapidly. Thus, adequate amount of the lubricating oil can be maintained in the crank chamber 15, causing excellent contact and slidability between the parts in the crank chamber 15 (e.g. between the swash plate 23 and shoes 27 and between the shoes 27 and the piston 25).
It should be apparent to those skilled in the art that the present invention may be embodied in many other specific forms without departing from the spirit or scope of the invention. Particularly, it should be understood that the invention may be embodied in the following forms. [0080]
As shown in FIG. 8, the fourth embodiment (c.f. FIG. 6) may be changed to tapered [0081] portions 41 c and 42 b of the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42. In such a configuration, tapering of the rotary valve 41 and the valve accommodating chamber 42 can be facilitated.
In the embodiment of FIG. 8, the tapered [0082] portion 41 c of the outer surface 41 b of the rotary valve 41 is provided at the rear of the opening of a suction guiding groove 45. Also, the tapered portion 42 b of the inner surface 42 a of the valve accommodating chamber 42 is provided at the rear of the opening of a suction communicating passage 43. That is, the tapered portions 41 c and 42 b are provided away from the opening/closing position of a refrigerant gas passage by the rotary valve 41.
Accordingly, even when the [0083] drive shaft 16 slides back and forth along the axis, a clearance between the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 is prevented from being changed in the vicinity of a connection region between the suction guiding groove 45 and the suction communicating passage 43. Thus, leakage of gas caused by enlargement of the clearance can be prevented, thereby maintaining compression efficiency of the compressor.
As shown in FIG. 9, the fourth embodiment (c.f. FIG. 6) may be changed to be [0084] curved portions 41 d and 42 c of the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 in such a direction as to approach the axis of the drive shaft 16 toward the rear side of the compressor.
In the embodiment of FIG. 9, the projected [0085] curved portion 41 d of the outer surface 41 b of the rotary valve 41 is provided at the rear of the opening of the suction guiding groove 45. Also, the recessed curved portion 42 c of the inner surface 42 a of the valve accommodating chamber 42 is provided at the rear of the opening of the suction communicating passage 43. That is, the curved portions 41 d and 42 c are provided away from the opening/closing position of the passage of refrigerant gas by the rotary valve 41.
Accordingly, even when the [0086] drive shaft 16 slides back and forth along the axis, the clearance between the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 is prevented from being changed in the vicinity of the connection region between the suction guiding groove 45 and the suction communicating passage 43. Thus, leakage of gas caused by enlargement of the clearance can be prevented, thereby maintaining compression efficiency of the compressor.
In the fourth embodiment of FIG. 6, the tapered portions of the [0087] outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 may be tilted in such a direction as to approach the axis of the drive shaft 16 toward the front side of the compressor. In this case, since the slide bearing surfaces 41 b and 42 a serve as thrust bearings 17, the thrust bearings 17 (see FIG. 1) can be omitted. In this embodiment, means (e.g., coating) is necessary for receiving a thrust load acted rearward on the drive shaft 16 between the rear end surface 41 f of the rotary valve 41 and the inner wall surface 14 a of the rear housing 14.
The pumping [0088] groove 49 may be formed not on the outer surface 41 b of the rotary valve 41 but on the inner surface 42 a of the valve accommodating chamber 42. Even in such a configuration, an advantage similar to the advantage (5) of the first embodiment can be provided.
The pumping [0089] groove 49 is not limited to the helical shape. For example, it may be an oblique groove tilted with respect to the axis of the drive shaft 16.
The [0090] coating 48 may be formed only on the inner surface 42 a of the valve accommodating chamber 42. Alternatively, the coating 48 may be formed on both of the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42.
The material for the [0091] rotary valve 41 is different from that (aluminum-series metallic material) of the cylinder block 11. Other metallic materials such as a synthetic resin having a coefficient of thermal expansion close to that of the aluminum-series metallic material, and brass having a coefficient of thermal expansion close to that of the aluminum-series metallic material but not adhered to the same may be used. In such a manner, sliding between the outer surface 41 b of the rotary valve 41 and the inner surface 42 a of the valve accommodating chamber 42 becomes sliding between different materials, thereby eliminating the necessity of the coating 48. When the rotary valve 41 is made of a synthetic resin, a glass fiber can be suitably used as a reinforcing material.
The [0092] rotary valve 41 and the cylinder block 11 may be formed of an iron-series metallic material having excellent durability.
The [0093] rotary valve 41 and the drive shaft 16 may be formed integrally. In this case, by setting the portion of the rotary valve 41 which has a larger diameter than other portions, an advantage similar to the advantage (2) of the first embodiment can be provided. The expression “rotary valve has a larger than drive shaft” includes an arrangement that the rotary valve 41 is formed integrally with the drive shaft 16, and the portion of the rotary valve 41 has a larger diameter than other portions.
The compressor is not limited to the single-head piston type compressor. As shown in FIG. 10, a fixed displacement compressor including a double-head piston may be employed. In the double-head piston type compressor, groups of cylinder bores [0094] 11 a are arranged not only on the rear side but also on the front side of the drive shaft 16. Thus, the rotary valve 41 is applied to the front suction valve mechanism 35.
In the compressor of FIG. 10, the front bearing (rolling bearing) [0095] 47 is removed, and the rotary valve 41 can be used as a rolling bearing for supporting the front end of the drive shaft 16. Therefore, it is not necessary to use expensive rolling bearings for all the radial bearings of the drive shaft 16, making it possible to reduce costs of the compressor further. In FIG. 10, identical or equivalent members is denoted by similar reference numerals, and description thereof will be omitted.
The [0096] drive shaft 16 and the rotary valve 41 may not necessarily be separate. As shown in FIG. 10, the drive shaft 16 and the rotary valve 41 may be formed integrally. This reduces the number of the components of the compressor and simplifies the manufacture of the compressor. To form the drive shaft 16 and the rotary valve 41 integrally, machining, casting, forging (e.g. cold forging) may be utilized.
FIG. 11 shows an another example of the embodiment of FIG. 10. In FIG. 10, a [0097] drive shaft 16 and a hollow (pipe-like) rotary valve 41 are separate members. These members may be integrate by press-fitting as in the first to fifth embodiments, by welding, or by pressure welding. Pressure welding refers to a method in which, for example in FIG. 11, the smaller portion 16 a of the drive shaft 16 is inserted in the hole 41 a of the rotary valve 41 without play and the drive shaft 16 and the rotary valve 41 are relatively rotated to weld the outer surface of the smaller portion 16 a in the hole 41 a with friction heat.
In place of the [0098] swash plate 23, a wave cam may be used.
Therefore, the present examples and embodiments are to be considered as illustrative and not restrictive and the invention is not to be limited to the details given herein, but may be modified within the scope and equivalence of the appended claims. [0099]

Claims

1. A piston type compressor comprising a drive shaft rotatably supported on a housing, the housing including a cylinder block, the cylinder block including a cylinder bore and a valve accommodating chamber, wherein reciprocation of a piston operably connected to the drive shaft in the corresponding cylinder bore changes the volume of a compression chamber in the cylinder bore to compress gas supplied from a first region, on which a suction pressure acts, to the compression chamber, wherein the compressed gas is sent to a second region, on which a discharge pressure acts, the compressor comprising:

a rotary valve accommodated in the valve accommodating chamber, wherein the rotary valve rotates integrally with the drive shaft to selectively open or close a passage of gas from the first pressure region to the compression chamber, wherein the rotary valve has a substantially cylindrical shape and an outer surface, wherein the valve accommodating chamber has a circular cross-section and an inner surface, wherein the outer surface and the inner surface constitute slide bearing surfaces for receiving a radial load applied to the drive shaft by sliding on each other.

2. The piston type compressor according to claim 1, wherein the outer surface of the rotary valve and the inner surface of the valve accommodating chamber are tilted with respect to the axis of the drive shaft, and the slide bearing surfaces also receive a thrust load acting on the drive shaft.

3. The piston type compressor according to claim 2, wherein the tilted portions are provided away from a part of the refrigerant gas passage that is opened and closed by the rotary valve.

4. The piston type compressor according to claim 2, wherein the outer surface of the rotary valve and the inner surface of the valve accommodating chamber are tilted in such a direction as to approach the axis of the drive shaft toward the rear side of the compressor, and are tapered.

5. The piston type compressor according to claim 2, wherein at least a part of the outer surface of the rotary valve and a part of the inner surface of the valve accommodating chamber are tapered.

6. The piston type compressor according to claim 5, wherein the tapered portions are provided away from a part of the refrigerant gas passage that is opened and closed by the rotary valve.

7. The piston type compressor according to claim 6, further comprising a plurality of suction communicating passages, wherein the compression chamber is one of a plurality of compression chambers, wherein each suction communicating passage corresponds to one of the compression chambers, wherein the valve accommodating chamber and each compression chamber are communicated with each other through the corresponding suction communicating passage, wherein the rotary valve has a suction guiding groove which is communicated with the first region, wherein the tapered portion of the outer surface of the rotary valve is provided at the rear of an opening of the suction guiding groove, and the tapered portion of the inner surface of the valve accommodating chamber is provided at the rear of openings of the suction communicating passages.

8. The piston type compressor according to claim 2, wherein at least a part of the outer surface of the rotary valve and a part of the inner surface of the valve accommodating chamber are curved.

9. The piston type compressor according to claim 8, wherein the curved portions are provided away from a part of the refrigerant gas passage that is opened and closed by the rotary valve.

10. The piston type compressor according to claim 9, further comprising a plurality of suction communicating passages, wherein the compression chamber is one of a plurality of compression chambers, wherein each suction communicating passage corresponds to one of the compression chambers, wherein the valve accommodating chamber and each compression chamber are communicated with each other through the corresponding suction communicating passage, wherein the rotary valve has a suction guiding groove which is communicated with the first region, wherein the curved portion of the outer surface of the rotary valve is provided at the rear of an opening of the suction guiding groove, and the curved portion of the inner surface of the valve accommodating chamber is provided at the rear of openings of the suction communicating passages.

11. The piston type compressor according to claim 1, wherein the rotary valve has a larger diameter than the drive shaft.

12. The piston type compressor according to claim 1, wherein the drive shaft and the rotary valve are separately provided.

13. The piston type compressor according to claim 1, wherein the drive shaft and the rotary valve are integrally provided.

14. The piston type compressor according to claim 1, wherein, on at least one of the outer surface of the rotary valve and the inner surface of the valve accommodating chamber, a coating is provided for improving sliding characteristics between the surfaces.

15. The piston type compressor according to claim 1, wherein a pumping groove is provided on the outer surface of the rotary valve or the inner surface of the valve accommodating chamber to pump at least one of refrigerant gas and lubricating oil contained in the refrigerant gas by rotation of the rotary valve.

16. The piston type compressor according to claim 1, wherein the rotary valve includes supercharging means for actively supplying the refrigerant gas to the compression chamber by using rotational force of the valve.

17. The piston type compressor according to claim 1, wherein the rotary valve further includes a hole for communicating a discharge pressure region with a clearance between the rotary valve and the valve accommodating chamber.

18. The piston type compressor according to claim 17, wherein the hole is arranged in a region opposite the suction guiding groove with respect to the axis of the drive shaft.

19. The piston type compressor according to claim 1, wherein the cylinder block and the rotary valve are made of materials having coefficients of thermal expansion equal to or close to each other.

20. A piston type compressor comprising a drive shaft rotatably supported on a housing, the housing including a cylinder block, the cylinder block including a cylinder bore and a valve accommodating chamber communicating with a suction communicating passage, wherein reciprocation of a piston operably connected to the drive shaft in the corresponding cylinder bore changes the volume of a compression chamber in the cylinder bore to compress gas supplied from a first region through the suction communicating passage, on which a suction pressure acts, to the compression chamber, wherein the compressed gas is sent to a second region, on which a discharge pressure acts, the compressor comprising:

a rotary valve accommodated in the valve accommodating chamber, wherein the rotary valve has a suction guiding groove which is communicated with the first region, wherein the rotary valve rotates integrally with the drive shaft to selectively open or close a passage of gas from the first pressure region to the compression chamber in accordance with relative positions of the suction communicating passage and the suction guiding groove, wherein the rotary valve has a substantially cylindrical shape and an outer surface, wherein the valve accommodating chamber has a circular cross-section and an inner surface, wherein the outer surface and the inner surface constitute slide bearing surfaces for receiving a radial load applied to the drive shaft by sliding on each other.