Piston type compressor

Abstract

The piston type compressor includes a cylinder block, a rotary shaft, a plurality of pistons, a rotary valve and radial load transmission means. The radial load transmission means transmits a radial load caused by compression reaction force acting on the piston, which is on its compression stroke of the discharge stroke, to the rotary valve, thereby pressing the rotary valve against an inner peripheral surface of the valve chamber. The radial load transmission means has a direction turning portion for turning the radial load toward the inner peripheral surface of the valve chamber between the suction port communicating with the cylinder bore whose piston has completed the discharge stroke and the suction port communicating with the cylinder bore whose piston is on the compression stroke in a rotation direction of the rotary valve from the suction port communicating with the cylinder bore whose piston has completed the discharge stroke.

Description

BACKGROUND

The present invention relates to a piston type compressor and more particularly to an improvement of sealing performance of a rotary valve of the compressor.

There has been known a piston type compressor in which a plurality of cylinder bores is arranged around a valve chamber and the cylinder bores and the valve chamber are interconnected through suction ports, respectively, so that a rotary valve received in the valve chamber selectively opens and closes the suction ports. In such a compressor, though the suction port communicating with the cylinder bore whose piston is on its discharge stroke is closed by the outer periphery of the rotary valve, there is a fear that refrigerant gas in the cylinder bore may leak from tho suction port into the valve chamber flowing along the outer peripheral surface of the rotary valve because the pressure of the refrigerant gas in the cylinder bore whose piston is on the discharge stroke is increased high.

Unexamined Japanese Patent Publication No. 2003-222075 has proposed a compressor in which compression reaction force acting on a piston which is on its compression or discharge stroke is transmitted to the rotary valve for urging the rotary valve toward the suction port communicating with the cylinder bore whose piston is on the discharge stroke.

By urging the rotary valve toward the suction port then communicating with the cylinder bore whose piston is on the discharge stroke by the compression reaction force, leakage of the refrigerant gas from the suction port is prevented. In the case of a compressor having a larger number of cylinder bores arranged around a rotary shaft, the piston in the cylinder bore located adjacent to the cylinder bore whose piston is on the discharge stroke is on the discharge or compression stoke and, therefore, the pressure of refrigerant gas in the adjacent cylinder bore is relatively high. Therefore, even when a rotary valve 32 received in a valve chamber 31 is urged toward a suction port 34 communicating with the cylinder bore whose piston is on the discharge stroke by urging force 33, as shown in FIG. 9, the refrigerant gas may leak from the suction port communicating with the adjacent cylinder bore, as indicated by arrow in FIG. 9.

SUMMARY

The present invention is directed to a piston type compressor which effectively prevents leakage of refrigerant gas even by using a rotary valve.

The present invention has the following features. The piston type compression includes a cylinder block, a rotary shaft, a plurality of pistons, a rotary valve and radial load transmission means. The cylinder block has a valve chamber, a plurality of cylinder bores formed around the valve chamber and a plurality of suction ports. Each suction port connects the valve chamber and the respective cylinder bore. The rotary shaft is rotatably supported in the cylinder block. Each piston is received in the respective cylinder bore and is reciprocated therein in accordance with rotation of the rotary shaft. The rotary valve is received in the valve chamber and connected to the rotary shaft so as to be rotated in accordance with the rotation of the rotary shaft, thereby selectively closing the suction port communicating with the cylinder bore whose piston is on its discharge stroke. The radial load transmission means transmits a radial load caused by compression reaction force acting on the piston, which is on its compression stroke or the discharge stroke, to the rotary valve, thereby pressing the rotary valve against an inner peripheral surface of the valve chamber. The radial load transmission means has a direction turning portion for turning the radial load toward the inner peripheral surface of the valve chamber between the suction port communicating with the cylinder bore whose piston has completed the discharge stroke and the suction port communicating with the cylinder bore whose piston is on the compression stroke in a rotation direction of the rotary valve from the suction port communicating with the cylinder bore whose piston has completed the discharge stroke.

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 DRAWINGS

The features of the present invention that are believed to be novel are set forth with particularity in the appended claims. 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 drawing, in which:

FIG. 1 is a sectional view showing a piston type compressor according to a first preferred embodiment of the present invention;

FIG. 2 is a schematic view showing a rotary valve and its vicinities according to the first preferred embodiment of the present invention;

FIG. 3 is a view showing force acting on the rotary valve according to the first preferred embodiment of the present invention;

FIG. 4 is a schematic view showing a rotary valve and its vicinities according to a second preferred embodiment of the present invention;

FIG. 5 is a schematic view showing a rotary valve and its vicinities according to a fourth preferred embodiment of the present invention;

FIG. 6 is a schematic view showing the rotary valve and its vicinities according to the fourth preferred embodiment of the present invention;

FIG. 7 is a schematic view showing a rotary valve and its vicinities according to a fifth preferred embodiment of the present invention;

FIG. 8 is a schematic view showing the rotary valve and its vicinities according to the fifth preferred embodiment of the present invention; and

FIG. 9 is a schematic view showing a rotary valve and its vicinities according to a prior art compressor.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

The following will describe preferred embodiments of the present invention with reference to the drawings.

FIG. 1 shows a piston type compressor according to a first preferred embodiment of the present invention. The left side of the drawing is a front side and the right side thereof is a rear side. The compressor has a cylinder block 1 which is connected at the front end thereof to a front housing 2 and at the rear end thereof to a rear housing 4 through a valve plate assembly 3. The cylinder block 1 and the front housing 2 cooperate to define a crank chamber 5. A rotary shaft 6 extending through the crank chamber 5 is rotatably supported by bearings 7, 28 which are provided in the cylinder block 1 and the front housing 2, respectively. The front end of the rotary shaft 6 protrudes from the front housing 2 and is connected to a drive source (not shown) such as a vehicle engine or a vehicle motor. A rotary support 8 is fixed on the rotary shaft 6 for rotation therewith and a swash plate 9 is engaged with the rotary support 8 in the front housing 2. The swash plate 9 has formed at the center thereof a through hole which the rotary shaft 6 extends and is rotated integrally with the rotary shaft 6 through a linkage 10. In addition, the swash plate 9 is supported in such a way that it is capable of sliding in the axial direction of the rotary shaft 6 and of inclining relative to the axial direction. Further, the rotary support 8 is supported for rotation by a thrust bearing 11 arranged in the recess formed in the front end of the front housing 2.

A plurality of cylinder bores 12 is formed in the cylinder block 1 around the rotary shaft 6. Each cylinder bore 12 has slidably received therein a piston 13. Each piston 13 engages with the outer periphery of the swash plate 9 through a pair of shoes 14. As the swash plate 9 rotates with the rotary shaft 6, the piston 13 reciprocates in the axial direction of the rotary shaft 6 in its associated cylinder bore 12 through the shoes 14.

The rear housing 4 has formed in the middle portion thereof a suction chamber 15 located in facing relation to the valve plate assembly 3. The rear housing 4 has also formed in the outer peripheral portion thereof a discharge chamber 16 surrounding the suction chamber 16.

The cylinder block 1 and the rear housing 4 have formed therein a supply passage 25 which connects the crank chamber 5 and the discharge chamber 16. On the supply passage 25 is located a displacement control valve 17 formed by an electromagnetic valve. In addition, the crank chamber 5 and the suction chamber 15 are in communication via a bleed passage 26.

The cylinder block 1 has formed therethrough at the middle portion thereof a valve chamber 18 extending in the axial direction of the rotary shaft 6. In the valve chamber 18 is received a rotary valve 19 which is mounted to the rear end of the rotary shaft 6 for rotation. The rotary valve 19 is loosely fitted on an eccentric pin 20 which is integrally Formed with the rotary shaft 6. Thus, the rotary valve 19 is rotated in the valve chamber 18 in accordance with the rotation of the rotary shaft 6.

A compression chamber 27 defined in each cylinder bore 12 by the valve plate assembly 3 and the respective piston 13 is in communication with the valve chamber 18 in the cylinder block 1 via a suction port 21. The suction ports 21 of the respective compression chambers 27 are selectively opened and closed by the outer peripheral surface of the rotary valve 19.

It is noted that the difference between the inside diameter of the valve chamber 18 and the outside diameter of the rotary valve 19 (hereinafter referred to “first clearance”) is smaller than the difference between the inside diameter of the bearing 7 and the outside diameter of the rotary shaft 6 (hereinafter referred to “second clearance”).

Tho magnitude relation between the first clearance and the second clearance provides radial load transmission means for transmitting a radial load caused by compression reaction force acting on the piston 13, which is then on are discharge stroke, to the rotary valve 19 through the shoes 14, the swash plate 9 and the rotary shaft 6.

The following will describe the operation of the piston type compressor according to the first embodiment. During suction stoke of the piston 13 moving in the cylinder bore 12 from the top dead center to the bottom dead center in accordance with rotation of the rotary shaft 6, the rotary valve 19 which is synchronously rotated with the rotary shaft 6 makes the suction port 21 for the cylinder bore 12 to communicate with the suction chamber 15, so that refrigerant gas in the suction chamber 15 flows into the cylinder bore 12 through the suction port 21.

Subsequently, during compression and discharge stroke of the piston 13 then moving in the cylinder bore 12 from the bottom dead center to the top dead center, the rotary valve 19 closes the suction port 21, so that the refrigerant gas in the cylinder bore 12 is compressed and then discharged into the discharge chamber 16 through a discharge port 24 and a discharge valve 29 of the valve plate assembly 3. In the present embodiment, the compression stroke is a state in which the discharge of the refrigerant gas from the compression chamber 27 to the discharge chamber 16 is not performed when the piston 13 moves in the cylinder bore 12 from the bottom dead center to the top dead center, and the discharge stroke is a state in which the discharge of the refrigerant gas from the compression chamber 27 to the discharge chamber 16 is performed when the piston 13 moves in the cylinder bore 12 from the bottom dead center to the top dead center. In addition, the piston 13 completes the discharge stroke when the piston 13 reaches the top dead center.

By setting the opening of the displacement control valve 17, the balance between the amount of refrigerant gas introduced from the discharge chamber 16 into the crank chamber 5 through the supply passage 25 and the amount of refrigerant gas flowing from the crank chamber 5 into the suction chamber 15 through the bleed passage 26 is controlled, and pressure Pc in the crank chamber 5 is determined, accordingly. If the pressure Pc in the crank chamber 5 is varied by changing the opening of the displacement control valve 17, the pressure differential between the crank chamber 5 and the cylinder bore 12 across the piston 13 is varied and the inclination of the swash plate 9 is varied, accordingly. Consequently, the stroke length of the piston 13 or the displacement of the compressor is adjusted.

As shown in FIG. 1, the compression reaction force which the piston 13 on completion of the discharge stroke receives acts on the radially outer peripheral portion of the swash plate 9 through the shoes 14 as a force F1, which urges the swash plate 9 upward as seen in FIG 1. As the swash plate 9 is urged upward, the rotary shaft 6 is also urged upward as seen in FIG. 1 through the through hole formed at the center of the swash plate 9. This urging force serves as a moment load around the position or engagement between the rotary shaft 6 and the bearing 28. At this time, a radial load acts on the rear end of the rotary shaft 6 toward the suction port 21 communicating with the cylinder bore 12 whose piston 13 has just completed the discharge stroke.

FIG. 2 is a schematic view showing the rotary valve 19 and its vicinities as seen from the rear end of FIG. 1. The rotary valve 19 is rotated in a counterclockwise direction as indicated by an arrow A of FIG. 2. For example, in the case where five suction-ports 21a, 21b, 21c, 21d, 21e are formed in the inner peripheral surface of the valve chamber 18 for fluid communication with the respective cylinder bores 12 as shown in FIG. 2, a radial load Fr acts on the rear end of the rotary shaft 6 toward the suction port 21a communicating with the cylinder bore 12 in which its piston 13 has just completed the discharge stroke and is located at the top dead center. Now, as a direction turning portion for turning direction of load, the eccentric pin 20 is formed on the rotary shaft 6 and the rotary valve 19 is loosely filled on the eccentric pin 20. The eccentric pin 20 is located between the suction port 21 communicating with the cylinder bore 12 whose piston 13 has just completed the discharge stroke and the suction port 21 communicating with the cylinder bore 12 whose piston 13 is on the compression stroke in the rotation direction of the rotary valve 19 from the suction port 21 communicating with the cylinder bore 12 whose piston 13 has just completed the discharge stroke. Specifically, when the position of the suction port 21a communicating with the cylinder bore 12 whose piston 13 has just completed the discharge stroke is defined as an angle of 0 degree, the eccentric pin 20 is located in a range of rotation angle between 0 degree and 90 degrees in the rotation direction of the rotary valve 19 from the auction port 21a. Therefore, when the radial load Fr acts on the rear end of the rotary shaft 6, firstly the rotary valve 19 is contracted with the inner peripheral surface of the valve chamber 18 in a range of rotation angle between 0 degree and 90 degrees in the rotation direction of the rotary valve 19 from the suction port 21a communicating with the cylinder bore 12 whose piston 13 has just completed the discharge stroke, by virtue of the magnitude relation between the second clearance (between the rotary shaft 6 and the bearing 7) and the first clearance (between the rotary valve 19 and the valve chamber 18), thereby receiving a drag Fb. Subsequently, the rotary shaft 6 is rotated with the eccentric pin 20 as an axis of rotation and is contacted with the inner peripheral surface of the bearing 7 in a range of rotation angle between 270 degrees and 360 degrees in the rotation direction of the rotary valve 19 from the suction port 21a, thereby receiving a drag Fa. Thus, the sum of the drags Fa and Fb balances with the radial load Fr.

As the reaction of the drag Fb of the rotary valve 19 which acts on the eccentric pin 20, the rotary valve 10 receives a load Fc from the eccentric pin 20, thereby being pressed against the inner peripheral surface of the valve chamber 18 in a range of rotation angle between 0 degree and 90 degrees in the rotation direction of the rotary valve 19 from the suction port 21a. At this time, in the rotation direction of the rotary shaft 6 and the rotary valve 19 from the suction port 21a communicating with the cylinder bore 12 whose piston 13 has just completed the discharge stroke, the cylinder bore 12 for the suction port 21e has its piston 13 moving in the discharge or compression stroke, the cylinder bore 12 for the suction port 21d has its piston 13 moving in the compression stroke close to the suction stroke, and the cylinder bores 12 for the suction ports 21c, 21b have their pistons 13 moving in the suction stroke. Therefore, although the refrigerant gas in the cylinder bores 12 communicating with the suction ports 21a, 21c is relatively high in pressure, leakage of the refrigerant gas from the suction ports 21a,21e into the valve chamber 18 is effectively prevented because the rotary valve 19 is pressed in a range of rotation angle between 0 degree and 90 degrees by the load Fc.

When the rotary shaft 6 and the rotary valve 19 are rotated in the rotation direction from the above state, firstly the cylinder bore 12 for the suction port 21a has its piston 13 moving in the suction stroke, the cylinder bore 12 for the suction port 21e has its piston 13 having just completed the discharge stroke, the cylinder bores 12 for the suction ports 21d has its piston 13 moving in the discharge or compression stroke, the cylinder bore 12 for the suction port 21c has its piston 13 moving in the compression stroke, and the cylinder bore 12 for the suction port 21b has its piston 13 moving in the suction stroke. In this case, the radial load Fr acts on the rear end of the rotary shaft 6 toward the suction port 21e communicating with the cylinder bore 12 in which its piston 13 has just completed the discharge stroke and is located at the top dead center. When the position of the suction port 21e communicating with the cylinder bore 12 whose piston 13 has just completed the discharge stroke is defined as an angle of 0 degree, the rotary valve 19 is pressed against the inner peripheral surface of the valve chamber 18 in a range of rotation angle between 0 degree and 90 degrees in the rotation direction of the rotary valve 19 from the suction port 21c. Meanwhile, when the position of the suction port 21e is defined as an angle of 0 degree, the rotary shaft 6 is pressed against the inner peripheral surface of the bearing 7 in a range of rotation angle between 270 degrees and 360 degrees in the rotation direction of the rotary valve 19 from the suction port 21e, thereby receiving the drag Fa.

That is, during the operation of the compressor, the radial load Fr which is caused by the compression reaction force acting on the piston 13 and acts on the rear end of the rotary shaft 6 toward the suction port 21a is continuously supported at both of the position at which the rotary valve 19 is pressed against the inner peripheral surface of the valve chamber 18 by the load Fc and the position at which the rotary shaft 6 is pressed against the inner peripheral surface of the bearing 7 by the drag Fa. In this state, the rotary valve 19 and the rotary shaft 6 are rotated respectively in the valve chamber 18 and the bearing 7 while maintaining a predetermined positional relationship. Therefore, during the operation of the compressor, the rotary valve 19 prevents the leakage of the refrigerant gas from the suction port 21 communicating with the cylinder bore 12 whose piston 13 is on its discharge or compression stroke under relatively high pressure into the valve chamber 18.

Now, the relationship among the mounting angle of the eccentric pin 20 relative to the suction port 21a, the second clearance (between the rotary shaft 6 and the bearing 7) and the load which the rotary valve 19 receives from the eccentric pin 20 will be described with reference to FIG. 3.

Referring to FIG. 3, “00” is an angle made between the direction of the resultant force F0 of the load which the rotary valve 19 receives from the refrigerant gas in the cylinder bores 12 communicating with the suction ports 21a, 21b, 21c, 21d, 21e and an imaginary line, as indicated by dashed line, extending axially through the suction port 21a communicating with the cylinder bore 12 whose piston 13 has just completed the discharge stroke, and “α” is an angle made between the direction of the load Fc which the rotary valve 19 receives from the eccentric pin 20 and the above imaginary line.

When the mounting angle “α” of the eccentric pin 20 (the angle of the load Fc) relative to the suction port 21a communicating with the cylinder bore 12 whose piston 13 has just completed the discharge stroke is variously changed with the first clearance (between the rotary valve 19 and the valve chamber 18) maintained at a predetermined value, the load Fc is decreased with an increase of the angle “α”. When the second clearance (between the rotary shaft 6 and the bearing 7) is decreased below a predetermined value, the load Fc is decreased rapidly, and the pressing force of the rotary valve 19 against the inner peripheral surface of the valve chamber 18 is reduced, accordingly.

This is because when the length of the second clearance (between the rotary shaft 6 and the bearing 7) approaches that of the first clearance (between the rotary valve 19 and the valve chamber 18) and the rotary shaft 6 then contacts the inner peripheral surface of the bearing 7 after the rotary valve 19 has contacted the inner peripheral surface of the valve chamber 18, contact position between the rotary shaft 6 and the bearing 7 approaches the position of rotation angle of 0 degree relative to the suction port 21a. Thus, the load Fc which the rotary valve 19 receives from the eccentric pin 20 is reduced.

Therefore, it is desirable that the second clearance (between the rotary shaft 6 and the bearing 7) should be larger than the first clearance (between the rotary valve 19 and the valve chamber 18), or more precisely, larger than the sum of the first clearance and a third clearance between the eccentric pin 20 and the bore of the rotary valve 19 which is loosely fitted on the eccentric pin 20.

In the present embodiment, when the position of the suction port 21 communicating with the cylinder bore 12 whose piston 13 has just completed the discharge stroke is defined as an angle of 0 degree, the rotary valve 19 is pressed against the inner peripheral surface of the valve chamber 18 in a range of rotation angle between 0 degree and 90 degrees in the rotation direction of the rotary valve 19 from the suction port 21. This is because when the position of the suction port 21 communicating with the cylinder bore 12 whose piston 13 has just completed the discharge stroke is defined as an angle of 0 degree, an angle of “80” made between the direction of the resultant force F0 of the load which the rotary valve 19 receives from the refrigerant gas in the cylinder bores 12 and the imaginary line is kept in a range of rotation angle between 180 degree and 270 degrees in the rotation direction of the rotary valve 19 from the suction port 21.

In addition, in the present embodiment, the rotary valve 19 is loosely fitted on the eccentric pin 20 formed on the rear end of the rotary shaft 6. In the structure, the first clearance and the second clearance are easily set in comparison with a case that the rotary valve 19 and the rotary shaft 6 are fixed to each other.

The first clearance between the rotary valve 19 and the valve chamber 18 is generally set to be extremely small in comparison with the second clearance between the rotary shaft 6 and the bearing 7 in order to prevent the leakage of the refrigerant gas from each suction port 21. Therefore, in the case where the rotary valve 19 is fixed to the rotary shaft 6 so as to be eccentric with respect to an axis of the rotary valve 19, unless the second clearance between the rotary shaft 6 and the bearing 7 is set with high accuracy, the radial loud Fr is not supported at both of the position at which the rotary valve 19 is pressed against the inner peripheral surface of the valve chamber 18 by the load Fc and the position at which the rotary shaft 6 is pressed against the inner peripheral surface of the bearing 7 by the drag Fa. At this time, the rotary valve 19 is not contacted with the inner peripheral surface of the valve chamber 18 in a suitable position.

That is, in the case where the rotary valve 19 is fixed to the rotary shaft 6 so as to be eccentric with the axis of the rotary shift 6, even if the rotary valve 19 is contacted with the inner peripheral surface of the valve chamber 18, if the second clearance between the rotary shaft 6 and the bearing 7 is larger than the first clearance between the rotary valve 19 and the valve chamber 18 in an area other than the position at which the rotary valve 19 contacts the inner peripheral surface of the valve chamber 18, the rotary shaft 19 is not contacted with the bearing 7. At this time, the rotary valve 19 is not pressed against the inner peripheral surface of the valve chamber 18 in the range of rotation angle between 0 degree and 90 degrees. Therefore, in the case where the rotary valve 19 is fixed to the rotary shaft 6 so as to be eccentric with the axis of the rotary shaft 6, the second clearance between the rotary shaft 6 and the bearing 7 needs to be set to be smaller than the first clearance between the rotary valve 19 and the valve chamber 18 in an area other than the position at which the rotary valve 19 contacts the inner peripheral surface of the valve chamber 18. For the above reason, it is extremely difficult to set the first clearance and the second clearance. In the present embodiment, however, since the rotary shaft 6 is capable of rotating with the eccentric pin 20 as an axis in the bearing 7 relative to the rotary valve 19 by loosely fitting the rotary valve 19 on the eccentric pin 20 of the rotary shaft 6, even if the first clearance between the rotary valve 19 and the valve chamber 18 in an area other than the position at which the rotary valve 19 contacts the inner peripheral surface of the valve chamber 18 is smaller than the second clearance between the rotary shaft 6 and the bearing 7, the radial load Fr is capable of being supported at the above two positions by contacting the rotary shaft 6 with the inner peripheral surface of the bearing 7 in a range of rotation angle between 270 degrees and 360 degrees with the eccentric pin 20, which is loosely fitted on the rotary valve 19, as an axis.

FIG. 4 shows a rotary valve and its vicinities of a piston type compressor according to a second preferred embodiment of the present invention. In the first embodiment, the rotary valve 19 is loosely fitted on the eccentric pin 20 formed on the rear end of the rotary shaft 19. In the second embodiment, however, as the direction turning portion, a flat pin 22 is formed on the rear end of the rotary shaft 6, having a flat cross section and an elongated shape extending in a predetermined radial direction and a rotary valve 23 is loosely fitted on the flat pin 22 so as to be capable of sliding along the above predetermined radial direction. The structure of the second embodiment other than the flat pin 22 is substantially the same as that of the first embodiment.

When the position of the suction port 21a communicating with the cylinder bore 12 whose piston 13 has lust completed the discharge stroke is defined as an angle of 0 degree, the flat pin 22 is radially elongated in a range of rotation angle between 270 degrees and 360 degrees in the rotation direction of the rotary valve 23 from the suction port 21a. Therefore, the rotary valve 23 is capable of sliding radially in a range of rotation angle between 270 degrees and 360 degrees, but is incapable of sliding radially in a range of rotation angle between 0 degree and 90 degrees. When the eccentric pin 22 and the rotary valve 23 are used, if the radial load Fr acts on the rear end of the rotary shaft 6 toward the suction port 21a communicating with the cylinder bore 12 whose piston 13 has just completed the discharge stroke, the rear end of the rotary shaft 6 is contacted with the inner peripheral surface of the bearing 7 in a range of rotation angle between 270 degrees and 360 degrees in the rotation direction of the rotary valve 19 from the suction port 21a, thereby receiving the drag Fa. In addition, the rear end of the rotary shaft 6 receives the drag Fb from the rotary valve 23 through the flat pin 22. Thus, the sum of the drags Fa and Fb balances with the radial load Fr.

Further, as the reaction of the drag Fb of the rotary valve 23 which acts on the flat pin 22, the rotary valve 23 receives load Fc from the flat pin 22, thereby being pressed against the inner peripheral surface of the valve chamber 18 in a range of rotation angle between 0 degree and 90 degrees in the rotation direction of the rotary valve 23 from the suction port 21a.

Consequently, as is the case with the first embodiment, leakage of the refrigerant gas from the suction port 21a communicating with the cylinder bores 12 whose piston 13 has just completed the discharge stroke and the suction port 21e communicating with the cylinder bores 12 whose piston 13 is on its discharge or compression stroke under relatively high pressure into the valve chamber 18 is effectively prevented.

In the first and second embodiments, as the direction turning portion, the rotary valve 19 or 23 is loosely fitted on the eccentric pin 20 or the flat pin 22. In a third preferred embodiment, however, referring to FIG. 2, as the direction turning portion, the rotary shaft 6 and the rotary valve 19 are fixed to each other in such a way that the rotary valve 19 is eccentric with respect to the axis of the rotary shaft 6.

Thus, as is the case with the first and second embodiments, during the operation of the compressor, the rotary valve 19 prevents the leakage of the refrigerant gas from the suction port 21 communicating with the cylinder bore 12 whose piston 13 is on its discharge or compression stroke under relatively high pressure into the valve chamber 18.

FIG. 5 shows a rotary valve and its vicinities of a piston type compressor according to a fourth preferred embodiment of the present invention. In the first preferred embodiment, the rear end of the rotary shaft 6 is rotatably supported by the bearing 7 in the cylinder block 1, and the rotary valve 19 is loosely fitted on the eccentric pin 20 formed on the rear end of the rotary Shaft 6. In the fourth preferred embodiment, however, the compressor is not provided with the bearing 7 which rotatably supports the rear end of the rotary shaft 6, but is provided with a rotary valve 41 which is fixed to the rear end of the rotary shaft 6 and is received in the valve chamber 18 of the cylinder block 1. The other structure of the fourth embodiment is substantially the same as that of the first embodiment.

The rotary valve 41 includes a cylindrical valve portion 41a which selectively opens and closes each suction port 21 which connects the compression chamber 27, which is defined in each cylinder bore 12 by the valve plate assembly 3 and the respective piston 13, to the valve chamber 18 in the cylinder block 1. The rotary valve 41 also includes a columnar support portion 41b which is integrally fixed to the end in an axial direction of the valve portion 41a. The support portion 41b is fixed to the rear end of the rotary shaft 6.

The support portion 11b has the same outside diameter as the rotary shaft 6, and is coaxially arranged on the rotary shaft 6. The valve portion 41a has larger outside diameter than the support portion 41b, and is arranged so as to be eccentric with respect to an axis of the support portion 41b. The support portion 41b and the valve portion 41a have substantially the same positional relation as the rotary shaft 6 and the rotary valve 19 of FIG. 2 according to the first preferred embodiment. That is, in the present embodiment, as the direction turning portion, the valve portion 41a is fixed to the support portion 41b of the rotary valve 41 so as to be eccentric with the axis of the support portion 41b.

As shown in FIG. 6, the valve chamber 18 of the cylinder block 1, which receives the rotary valve 41, includes a portion 18a, in which the valve portion 41a of the rotary valve 41 is received, and a portion 18b, in which the support portion 41b is received. The portion 18a and the portion 18b have a step therebetween. The inside diameter of the portion 18b of the valve chamber 18 is the same as that of the bearing 7 according to the first embodiment.

In the above structure, even if the compressor is not provided with the bearing 7 which rotatably supports the rear end of the rotary shaft 6, as is the case with the first through third embodiments, during the operation of the compressor, the rotary valve 41 prevents the leakage of the refrigerant gas from the suction port 21 communicating with the cylinder bore 12 whose piston 13 is on its discharge or compression stroke under relatively high pressure into the valve chamber 18.

It is noted that as shown in FIG. 6, coating 42 is desirably provided in the cylinder block 1 so as to be applied to the inner peripheral surface of the portion 18b of the valve chamber 18 which the outer peripheral surface of the support portion 41b of the rotary valve 41 contacts for reducing friction.

FIG. 7 shows a rotary valve and its vicinities of a piston type compressor according to a fifth preferred embodiment of the present invention. In the fifth preferred embodiment, as is the case with the fourth preferred embodiment, the compressor is not provided with the bearing 7 which rotatably supports the rear end of the rotary shaft 6, but is provided with a rotary valve 51 which is fixed to the rear end of the rotary shaft 6 and is received in the valve chamber 18 of the cylinder block 1. The other structure of the fifth embodiment is substantially the same as that of the first embodiment.

The rotary valve 51 includes a cylindrical valve portion 51a which selectively opens and closes each suction port 21 which connects the compression chamber 27, which is defined in each cylinder bore 12 by the valve plate assembly 3 and the respective piston 13, to the valve chamber 18 in the cylinder block 1. The rotary valve 51 also includes a columnar support portion 51b which is integrally fixed to the end in an axial direction of the valve portion 51a. The support portion 51b is fixed to the rear end of the rotary shaft 6.

The support portion 51b has larger outside diameter than the rotary shaft 6, and is arranged on the rotary shaft 6 so as to be eccentric with an axis of the rotary shaft 6. The valve portion 51a has slightly larger outside diameter than the support portion 51b, and is arranged on the support portion 51b so as to be eccentric with respect to an axis of the support portion 51b. The support portion 51b and the valve portion 51a have substantially the same eccentric relation as the rotary shaft 6 and the rotary valve 19 of FIG. 2 according to the first preferred embodiment, and the support portion 51b and the valve portion 51a are fixed to each other. That is, in the present embodiment, as the direction turning portion, the support portion 51b of the rotary valve 51 is fixed to the rotary shaft 6 so as to be eccentric with the axis of the rotary shaft 6, and the valve portion 51a is fixed to the support portion 51b so as to be eccentric with the axis of the support portion 51b.

As shown in FIG. 8, a portion of the valve chamber 18 in which the valve portion 51a of the rotary valve 51 is received has the same inside diameter as that of the valve chamber 18 in which the support portion 51b is received.

It is noted that a clearance is formed between the outer peripheral surface on the rear end of the rotary shaft 6 and the inner surface of a through hole 1a of the cylinder block 1 through which the rear end of the rotary shaft 6 extends such that the outer peripheral surface on the rear end of the rotary shaft 6 contacts the inner surface of the through hole 1a of the cylinder block 1 when the rotary shaft 6 is rotated.

In the above structure, even if the compressor is not provided with the bearing 7 which rotatably supports the rear end of the rotary shaft 6, as is the case with the first through fourth embodiments, during the operation of the compressor, the rotary valve 51 prevents the leakage of the refrigerant gas from the suction port 21 communicating with the cylinder bore 12 whose piston 13 is on its discharge or compression stroke under relatively high pressure into the valve chamber 18.

According to the embodiments of the present invention, leakage of the refrigerant gas from the suction port is effectively prevented. Especially in the compressor which uses carbon dioxide as the refrigerant gas, the pressure of carbon dioxide in the cylinder bore whose piston is on its discharge or compression stroke is increased to an extremely high level and, therefore, the amount of carbon dioxide leaking from the suction port communicating with cylinder bore whose piston is on the discharge or compression stroke, increases unless an appropriate measure is taken effectively. Therefore, the present invention is advantageously applicable to the compressor using carbon dioxide as the refrigerant gas.

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.

Claims

1. A piston type compressor comprising: a cylinder block having a valve chamber, a plurality of cylinder bores formed around the valve chamber and a plurality of suction ports, wherein each suction port connects the valve chamber and the respective cylinder bore; a rotary shaft rotatably supported in the cylinder block, a plurality of pistons, each of which is received in the respective cylinder bore and is reciprocated therein in accordance with rotation of the rotary shaft; a rotary valve received in the valve chamber and connected to the rotary shaft so as to be rotated in accordance with the rotation of the rotary shaft, thereby selectively closing the suction port communicating with the cylinder bore whose piston is on its discharge stroke; and radial load transmission means for transmitting a radial load caused by compression reaction force acting on the piston, which is on its compression stroke or the discharge stroke, to the rotary valve, thereby pressing the rotary valve against an inner peripheral surface of the valve chamber, wherein the radial load transmission means has a direction turning portion for turning the radial load toward the inner peripheral surface of the valve chamber between the suction port communicating with the cylinder bore whose piston has completed the discharge stroke and the suction port communicating with the cylinder bore whose piston is on the compression stroke in a rotation direction of the rotary valve from the suction port communicating with the cylinder bore whose piston has completed the discharge stroke.

2. The piston type compressor according to claim 1, wherein the direction turning portion is formed by integrally fixing the rotary valve to the rotary shaft so that the rotary valve is eccentric with respect to an axis of the rotary shaft.

3. The piston type compressor according to Claim 1, wherein an eccentric pin is formed on an end of the rotary shaft on a side of the rotary valve between the suction port communicating with the cylinder bore whose piston has completed the discharge stroke and the suction port communicating with the cylinder bore whose piston is on the compression stroke in the rotation direction of the rotary valve from the suction port communicating with the cylinder bore whose piston has completed the discharge stroke, and the direction turning portion is formed by loosely fitting the rotary valve an the eccentric pin.

4. The piston type compressor according to claim 1, wherein a flat pin, having a flat cross section and an elongated shape extending in a radial direction perpendicular to a radial direction of the inner peripheral surface of the valve chamber between the suction port communicating with the cylinder bore whose piston has completed the discharge stroke and the suction port communicating with the cylinder bore whose piston is on the compression stroke in the rotation direction of the rotary valve from the suction port communicating with the cylinder bore whose piston has completed the discharge stroke, is formed on an end of the rotary shaft on a side of the rotary valve, and the direction turning portion is formed by loosely fitting the rotary valve on the flat pin so as to be capable of sliding only along the radial direction in which the elongated shape of the flat pin extends.

5. The piston type compressor according to claim 1, further comprising a bearing near the valve chamber for rotatably supporting the rotary shaft, wherein the difference between an inside diameter of the valve chamber and an outside diameter of the rotary valve is smaller than the difference between an inside diameter of the bearing and an outside diameter of the rotary shaft.

6. The piston type compressor according to claim 1, further comprising a coating near the valve chamber for rotatably supporting the rotary shaft wherein the difference between an inside diameter of the valve chamber and an outside diameter of the rotary valve is smaller than the difference between an inside diameter of the coating and an outside diameter of the rotary shaft.