Control valve for variable displacement compressor

Description

CROSS-REFERENCES TO RELATED APPLICATIONS, IF ANY

This application claims priority of Japanese Application No. 2004-269661 filed on Sep. 16, 2004 and entitled “CONTROL VALVE FOR VARIABLE DISPLACEMENT COMPRESSOR” and No. 2005-161179 filed on Jun. 1, 2005, entitled “CONTROL VALVE FOR VARIABLE DISPLACEMENT COMPRESSOR”.

BACKGROUND OF THE INVENTION

(1) Field of the Invention

The present invention relates to a control valve for a variable displacement compressor, and more particularly to a control valve for a variable displacement compressor, which is mounted on a variable displacement compressor as a component of a refrigeration cycle of an automotive air conditioner, for controlling the flow of refrigerant discharged therefrom to a constant flow rate.

(2) Description of the Related Art

A compressor used in the refrigeration cycle of an automotive air conditioner, for compressing refrigerant, uses an engine as a drive source, and hence is incapable of performing rotational speed control. To eliminate the inconvenience, a variable displacement compressor capable of varying the compression capacity of refrigerant is employed so as to obtain an adequate cooling capacity without being constrained by the rotational speed of the engine.

In such a variable displacement compressor, a wobble plate fitted on a shaft driven by the engine for rotation has compression pistons connected thereto, and by varying the inclination angle of the wobble plate, the stroke of the pistons is varied to vary the discharge amount of refrigerant.

The inclination angle of the wobble plate is continuously changed by introducing part of compressed refrigerant into a hermetically closed crankcase, and causing a change in the pressure of the introduced refrigerant, thereby changing the balance of pressures acting on the opposite sides of each piston.

The pressure in the crankcase is changed by a control valve for a variable displacement compressor, which is disposed between the discharge chamber of the compressor and the crankcase or between the crankcase and the suction chamber. This control valve provides control e.g. such that communication therethrough is allowed or blocked so as to maintain the differential pressure across a valve that controls the flow rate of refrigerant flowing from the discharge chamber into the crankcase, at a predetermined value, and more particularly, the differential pressure can be set to the predetermined value by externally changing a value of control current supplied to the control valve. With this configuration, when the rotational speed of the engine rises to increase the discharge pressure, the pressure introduced into the crankcase increases to reduce the capacity of refrigerant that can be compressed, whereas when the rotational speed of the engine lowers, the pressure introduced into the crankcase decreases to increase the capacity of refrigerant that can be compressed, whereby the amount of compression capacity of the compressor is maintained constant irrespective of the rotational speed of the engine.

One known method of controlling the compression capacity of such a variable displacement compressor uses a control valve therefor, which provides control such that the flow rate of refrigerant discharged from the compressor becomes constant (see e.g. Japanese Unexamined Patent Publication (Kokai) No. 2001-107854 (Paragraph numbers [0034] to [0036], FIGS. 2 and 3)).

According to this control valve for a variable displacement compressor, two pressure sensors are provided at locations spaced from each other in the direction of flow through the refrigerant passage toward the suction chamber; the differential pressure between the pressure monitoring points of the two sensors is detected to thereby indirectly grasp the flow rate of refrigerant drawn in; the control valve controls the flow rate of refrigerant introduced from the discharge chamber into the crankcase such that the suction flow rate of refrigerant becomes constant to thereby control the flow rate of refrigerant discharged from the compressor such that it becomes constant.

However, the conventional control valve for a variable displacement compressor that controls the compressor such that the discharge flow rate of refrigerant becomes constant requires expensive pressure sensors and a control device for detecting the differential pressure across the refrigerant circulation passage and controlling the control valve for the compressor based thereon, which leads to an increase in the cost of the automotive air conditioner.

SUMMARY OF THE INVENTION

The present invention has been made in view of these problems, and an object thereof is to provide a control valve for use in a variable displacement compressor, for controlling the discharge flow rate of refrigerant to a constant flow rate, which can be constructed compact in size without necessitating pressure sensors for detecting a differential pressure.

To solve the above problem, the present invention provides a control valve for a variable displacement compressor, for controlling a flow of refrigerant discharged from the compressor to a constant flow rate, comprising a first valve that has a passage area thereof set according to a flow rate of refrigerant introduced from a discharge chamber of the compressor and flowing out to an outlet port of the compressor, a second valve that controls pressure in a crankcase of the compressor in a manner interlocked with the first valve such that a differential pressure across the first valve is maintained at a predetermined differential pressure, and a solenoid that sets a differential pressure across the passage having the passage area set by the first valve to the predetermined differential pressure dependent on a flow rate of refrigerant to which the flow of refrigerant is to be controlled.

The above and other objects, features and advantages of the present invention will become apparent from the following description when taken in conjunction with the accompanying drawings which illustrate preferred embodiments of the present invention by way of example.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a central longitudinal cross-sectional view showing details of the construction of a control valve for a variable displacement compressor, according to a first embodiment of the present invention.

FIG. 2 is an enlarged cross-sectional view of essential parts of the control valve according to the first embodiment in a state immediately after energization thereof.

FIG. 3 is an enlarged cross-sectional view of essential parts of the control valve according to the first embodiment in a transitional state after energization thereof.

FIG. 4 is an enlarged cross-sectional view of essential parts of the control valve according to the first embodiment in a balanced state.

FIG. 5 is an enlarged cross-sectional view of essential parts of the control valve according to the first embodiment in a transitional state at the time of a rapid increase in discharge pressure.

FIG. 8 is an enlarged cross-sectional view taken on line A-A of FIG. 7.

FIG. 9 is an enlarged cross-sectional view of essential parts of the control valve according to the third embodiment in a non-energized state.

FIG. 10 is an enlarged cross-sectional view of essential parts of the control valve according to the third embodiment in a state immediately after energization thereof.

FIG. 11 is an enlarged cross-sectional view of essential parts of the control valve according to the third embodiment in a state being controlled.

FIG. 16 is a view showing, by way of example, an application of the control valve according to the seventh embodiment to a variable displacement compressor of a carbon dioxide system.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

Hereafter, embodiments of the present invention will be described in detail with reference to the drawings showing control valves applied to a variable displacement compressor of a flow rate control type in which the flow rate of discharged refrigerant is controlled to be constant, by way of example.

The control valve 10 for the variable displacement compressor according to the first embodiment comprises a first valve 11 that operates in dependence on the flow rate of refrigerant discharged from the compressor, a second valve 12 that controls pressure Pc in the crankcase of the compressor, a third valve 13 that controls the amount of leakage of refrigerant, and a solenoid 14 that is capable of externally setting the flow rate of refrigerant discharged from the compressor.

The first valve 11 is formed in a first body 15 disposed at an upper end location as viewed in FIG. 1. The first body 15 has a port 16 that is communicated with the discharge chamber of the compressor when the control valve 10 is mounted in the compressor, to thereby introduce refrigerant at discharge pressure Pdh therethrough, and a port 17 that is communicated with an outlet port of the compressor to thereby deliver refrigerant at discharge pressure Pdl, and is formed therethrough with a refrigerant passage 18 that connects these ports 16 and 17. A valve seat 19 is formed in the refrigerant passage 18 at an intermediate portion thereof, and a valve element 20 is disposed on a side toward the port 17 with respect to the valve seat 19 in a manner movable to and away from the valve seat 19. The vale element 20 is urged by a spring 21 having a weak spring force in the direction of closing the refrigerant passage 18. Now, the valve seat 19 and the valve element 20 form the first valve 11, and is constructed as a check valve that opens when the discharge pressure Pdh at the port 16 is higher than the discharge pressure Pdl at the port 17 by more than the urging force of the spring 21, and closes otherwise.

The second valve 12 is formed in a second body 22 that has the first body 15 secured to an upper part thereof by press-fitting. The second body 22 has a refrigerant introducing space defined between the same and the first body 15 by having the first body 15 fitted thereto, and into the refrigerant introducing space, refrigerant at the discharge pressure Pdh is introduced from the discharge chamber of the compressor via a refrigerant passage 23 formed through the first body 15. A strainer 24 is provided on an refrigerant inlet port side of the refrigerant passage 23 formed through the first body 15. The second body 22 has a port 25 communicated with the crankcase of the compressor when the control valve 10 is mounted in the compressor, to thereby discharge refrigerant at controlled pressure Pc to the crankcase. In the center of the upper part of the second body 22, there is formed a valve hole between the refrigerant introducing space and an internal space communicating with the port 25, and a valve element 27 is held by the second body 22 such that the valve element 27 is movable to and away from a vale seat 26 formed at a lower end face of the valve hole. The valve element 27 is urged by a spring 28 in a direction away from the valve seat 26. The valve seat 26 and the valve element 27 form the second valve 12 that controls the flow rate of refrigerant at the discharge pressure Pdh to thereby supply refrigerant at the pressure Pc to the crankcase, i.e. a Pd-Pc valve.

In the first body 15, further, a through hole is formed in an axial direction thereof, and a shaft 29 is disposed in a manner extending through the through hole and the valve hole of the second valve 12 and movable axially back and forth. An upper end of the shaft 29 as viewed in FIG. 1 is loosely fitted in the valve element 20 of the first valve 11, and a lower end of the same as viewed in FIG. 1 is loosely fitted in the valve element 27 of the second valve 12. Further, the shaft 29 has an upper part, as viewed in FIG. 1, which has an outer diameter larger than an inner diameter of the through hole of the first body 15. With this construction, when the shaft 29 moves downward as viewed in FIG. 1, a tapered stepped portion at a boundary between the upper part and a lower part is brought into abutment of the upper end face of the through hole to thereby close clearance between the shaft 29 and the through hole. This valve mechanism forms the third valve 13.

The second body 22 has a hole formed in the center of a lower part thereof, as viewed in FIG. 1. The rim of an opening of a bottomed sleeve 30 is tightly connected to the hole. The bottomed sleeve 30 has a core 31 and a plunger 32 of the solenoid 14 arranged therein. The core 31 is fixed to the hole in the center of the lower part of the first body 15 and the bottomed sleeve 30 by press-fitting. The plunger 32 is axially slidably disposed in the bottomed sleeve 30, and fixed to one end of a shaft 33 disposed in a manner axially extending through the core 31. Further, the plunger 32 is urged toward the core 31 by a spring 34 such that the other end of the shaft 33 is brought into contact with a lower end face of the valve element 27 of the second valve 12, as viewed in FIG. 1. Disposed around the outer periphery of the bottomed sleeve 30 is a coil 35, and a harness 36 for supplying electric current to the coil 35 is led to the outside of the solenoid 14. The inside of the bottomed sleeve 30 is communicated with internal space communicating with the port 25 via a pressure equalizing hole 37 formed through the second body 22.

In the control valve 10 constructed as above, as shown in FIG. 1, when the solenoid 14 is not energized, the second valve 12 is fully open since the valve element 27 is urged by the spring 28 of the second valve 12 against the urging force of the spring 34 of the solenoid 14 in the direction away from the valve seat 26, and the first valve 11 is fully closed, by having the valve element 20 seated on the valve seat 19 by the urging force of the spring 21.

In addition, FIG. 1 shows a state of the control valve 10 exhibited immediately after stoppage of the operation of the automotive air conditioner. This corresponds to the case where after the automotive air conditioner has been in operation, the solenoid 14 is deenergized. In this case, the solenoid 14 ceases to create the force attracting the plunger 32 toward the core 31, and hence the second valve 12 is fully opened since the spring 28 of the second valve 12 urges the valve element 27 against the urging force of the spring 34 of the solenoid 14 in the valve-opening direction. Therefore, the introduced refrigerant at the discharge pressure Pdh is supplied from the port 25 to the crankcase via the strainer 24, the refrigerant passage 23, and the second valve 12, whereby the compressor is shifted to the minimum capacity operation. This causes the compressor to do almost no work, so that the discharge pressure Pdh introduced into the compressor 10 becomes lower than the discharge pressure Pdl at the outlet port of the compressor, and the differential pressure between the discharge pressure Pdh and the discharge pressure Pdl causes the first valve element 11 to be fully closed.

Further, when the compressor is operating with a predetermined capacity, the valve element 20 of the first valve 11 has been moved away from the valve seat 19, and at this time, the shaft 29 is pushed downward, as viewed in FIG. 1, to place the third valve 13 in the closed state. In this state, when the solenoid 14 is deenergized to allow the valve element 20 of the first valve 11 to move upward as viewed in FIG. 1 and the valve element 27 of the second valve 12 to move downward, as viewed in FIG. 1, the shaft 29 receives the high discharge pressure Pdl on the top thereof and the discharge pressure Pdh lower than the discharge pressure Pdl on the bottom thereof, which holds the closed state of the third valve 13. This closes the clearance between the shaft 29 and the through hole, thereby preventing the discharge pressure Pdl held at high pressure from leaking to the upstream side of the second valve element 12 the pressure at which has become lower than the discharge pressure Pdl, via the clearance.

As a result, the pressure at the outlet port of the compressor can maintain the discharge pressure Pdl assumed before the stoppage of the operation of the automotive air conditioner. This provides an advantageous effect of improvement in the efficiency of the compressor since it is not necessary to compress refrigerant to the discharge pressure Pdl when the automotive air conditioner resumes its operation. Further, normally, a check valve is provided at the outlet port of the compressor for this purpose, but in the present embodiment, the first valve 11 of the control valve 10 serves as the check valve, which makes it possible to dispense with the check valve, and reduce the cost of the compressor.

Next, a detailed description will be given of the operation of the control valve 10, with reference to FIGS. 2 to 5.

FIG. 2 is an enlarged cross-sectional view of essential parts of the control valve according to the first embodiment in a state immediately after energization thereof; FIG. 3 is an enlarged cross-sectional view of essential parts of the control valve according to the first embodiment in a transitional state after energization thereof; FIG. 4 is an enlarged cross-sectional view of essential parts of the control valve according to the first embodiment in a balanced state; and FIG. 5 is an enlarged cross-sectional view of essential parts of the control valve according to the first embodiment in a transitional state at the time of a rapid increase in discharge pressure.

First, when the solenoid 14 shown in FIG. 1 is in the non-energized state, if a predetermined electric current is passed to the solenoid 14, then, immediately thereafter, the second valve 12 is instantaneously fully closed by the urging force of the solenoid 14. This causes the compressor to start the maximum capacity operation, but immediately after the energization, the discharge pressure Pdh is still lower than the discharge pressure Pdl at the outlet port of the compressor, and hence the first valve 11 is in the fully-closed state. At this time, the shaft 29 is pushed upward as viewed in FIG. 1 by the valve element 27 of the second valve 12, whereby the third valve 13 is opened, but the discharge pressure Pdl leaks to the discharge pressure Pdh via the clearance between the shaft 29 and the through hole. The leakage is slight, and the discharge pressure Pdh is also about to rise immediately. Therefore, no particular problem is brought about.

The compressor starts its operation with the maximum capacity, and when the discharge pressure Pdh becomes sufficiently higher than the discharge pressure Pdl, as shown in FIG. 3, the differential pressure therebetween causes the valve element 20 to move away from the valve seat 19. This causes the first valve 11 to open, and the refrigerant at the discharge pressure Pdh introduced into the port 16 is changed into refrigerant at the discharge pressure Pdl, which flows from the port 17 to the outlet part of the compressor. At this time, refrigerant flows through the first valve 11 at a flow rate corresponding to a value obtained by multiplying the passage area formed by opening of the first valve 11 by the differential pressure across the first valve 11.

The valve element 20 of the first valve 11 is moved in the valve-opening direction to be brought into abutment with the upper end of the shaft 29 which has been lifted upward, as viewed in FIG. 3, by the valve element 27 of the second valve 12. This causes the first valve 11 and the second valve 12 to operate in a manner interlocked with each other via the shaft 29, and the second valve 12 operates by detecting the differential pressure between the discharge pressure Pdh and the discharge pressure Pdl acting on the first valve 11, and the differential pressure between the discharge pressure Pdh and the pressure Pc.

Then, when the differential pressure between the discharge pressure Pdh and the discharge pressure Pdl acting on the first valve 11 becomes still larger, since the pressure-receiving area of the valve element 20 of the first valve 11 is larger than that of the valve element 27 of the second valve 12, the valve element 20 of the first valve 11 urges the valve element 27 of the second valve 12 by the differential pressure thereacross, and when the differential pressure between the discharge pressure Pdh and the discharge pressure Pdl reaches a predetermined value, as shown in FIG. 4, the second valve 12 slightly opens to a position where the differential pressures across the first valve 11 and the second valve 12, the loads of the springs 21 and 28, the urging force of the solenoid 14 dependent on the current value are balanced, whereby the controlled pressure Pc is supplied to the crankcase to place the compressor in the state in which the capacity or displacement thereof is controlled.

That is, in the control valve 10, the second valve 12 controls the pressure in the crankcase such that the differential pressure across the passage having the passage area produced by flow of refrigerant from the discharge chamber through the first valve 11 maintains a differential pressure set by the solenoid 14, to thereby control the discharge flow rate of the compressor to a constant flow rate. More specifically, e.g. when the rotational speed of the engine increases to increase the discharge pressure Pdh, the valve element 20 of the first valve 11 urges the valve element 27 of the second valve 12 in the direction of opening the same, by the increased amount of the differential pressure. This increases the pressure Pc in the crankcase, and hence the compressor operates in the direction of reducing the displacement thereof, whereby the discharge flow rate is controlled to a predetermined flow rate. Inversely, when the discharge pressure Pdh lowers to reduce the differential pressure across the first valve 11, the valve element 20 urges the valve element 27 of the second valve 12 in the direction of further closing the same. This reduces the pressure Pc in the crankcase and hence the compressor operates in the direction of increasing the displacement thereof, whereby the discharge flow rate is controlled to the predetermined flow rate.

By the way, the discharge pressure Pdh of the refrigerant discharged from the discharge chamber varies sensitively in response to a change in the rotational speed of the compressor. Therefore, when the rotational speed of the engine rapidly increases to rapidly increase the rotational speed of the compressor, the discharge pressure Pdh also rapidly increases. In such a case, in the control valve 10, since the pressure receiving area of the valve element 20 of the first valve 11 is set to be larger than that of the valve element 27 of the second valve 12, the force of the valve element 20 of the first valve 11 urging the valve element 27 of the second valve 12 in the direction of further opening the same by a change in the differential pressure across the valve element 20 is instantaneously increased, and as shown in FIG. 5, the second valve 12 operates instantaneously larger than during normal opening operation, whereby the compressor is promptly controlled in the direction of reducing the displacement. Inversely, when the rotational speed of the engine rapidly drops to rapidly decrease the discharge pressure Pdh, the valve element 20 of the first valve 11 also operates instantaneously largely in the valve-closing direction, and hence the second valve 12 also operates instantaneously largely in the valve-closing direction, whereby the compressor is promptly controlled in the direction of increasing the displacement thereof. Thus, after reacting to a rapid change in the discharge pressure Pdh with high responsiveness, the control valve 10 is capable of promptly restoring the compressor to a predetermined displacement thereof.

FIG. 6 is a central longitudinal cross-sectional view showing details of the construction of a control valve for a variable displacement compressor, according to a second embodiment of the present invention. It should be noted that component elements in FIG. 6 identical or similar to those shown in FIG. 1 are designated by identical reference numerals, and detailed description thereof is omitted.

As distinct from the control valve 10 according to the first embodiment, in the control valve 10a according to the second embodiment, the pressure which the second valve 12 senses and operates in response thereto is changed. That is, while the second valve 12 of the control valve 10 according to the first embodiment senses the differential pressure between the discharge pressure Pdh on the discharge side and the pressure Pc in the crankcase, the second valve 12 of the control valve 10a according to the second embodiment senses the differential pressure between the discharge pressure Pdh on the discharge side and the suction pressure Ps in the suction chamber.

In the control valve 10a, the second body 22 is provided with a port 38 which is communicated with the suction chamber of the compressor to introduce the suction pressure Ps therein. The second body 22 also holds a shaft 39 formed integrally with the valve element 27 of the second valve 12, in a manner movable axially back and forth. The shaft 39 has an outer diameter which is approximately equal to an effective diameter of the second valve 12 which receives the discharge pressure Pdh. The second valve 12 is configured such that the valve element 27 receives the discharge pressure Pdh on the discharge side, and the lower end of the shaft 39, as viewed in FIG. 6, receives the suction pressure Ps in the suction chamber, to thereby sense the differential pressure between them. To accurately sense the differential pressure between the discharge pressure Pdh and the suction pressure Ps, the outer diameter of the shaft 39 is only required to be equal to the inner diameter of the valve hole of the second valve 12. In the second embodiment, to simplify the construction of a part holding the shaft 39, the outer diameter of the shaft 39 is made larger than the inner diameter of the valve hole of the second valve 12 to such an extent that there is no substantial influence on the operation of the second valve 12. At the lower end of the shaft 39, there is provided a spring 28 that urges the valve element 27 of the second valve 12 in the valve-opening direction. Further, in the control valve 10a, the bottomed sleeve 30 of the solenoid 14 is communicated with the port 38, whereby the suction pressure Ps is introduced therein.

The control valve 10a constructed as described above operates in the same manner as the control valve 10 according to the first embodiment except that the second valve 12 operates by sensing the differential pressure between the discharge pressure Pdh and the suction pressure Ps, and hence detailed description of the operation thereof is omitted.

FIG. 7 is a central longitudinal cross-sectional view showing details of the construction of a control valve for a variable displacement compressor, according to a third embodiment of the present invention; FIG. 8 is an enlarged cross-sectional view taken on line A-A of FIG. 7; FIG. 9 is an enlarged cross-sectional view of essential parts of the control valve according to the third embodiment in a non-energized state; FIG. 10 is an enlarged cross-sectional view of essential parts of the control valve according to the third embodiment in a state immediately after energization thereof; and FIG. 11 is an enlarged cross-sectional view of essential parts of the control valve according to the third embodiment in a state being controlled. It should be noted that component elements in FIGS. 7 to 11 identical or similar to those shown in FIG. 1 are designated by identical reference numerals, and detailed description thereof is omitted.

The control valve 10b according to the third embodiment is largely different from the control valves 10 and 10a according to the first and second embodiments in the construction of the first valve 11. More specifically, while in the control valves 10 and 10a according to the first and second embodiments, the first valve 11 opens to provide a passage area dependent on the flow rate of refrigerant, the first valve 11 of the control valve 10b according to the third embodiment is configured such that the passage area is not varied according to the flow rate of refrigerant, in a normal control region.

The first valve 11 has a valve element 20 disposed in a manner movable axially back and forth in the refrigerant passage 18 formed through the first body 15. The valve element 20 has, as shown in FIG. 8, a plurality of guides 40 integrally formed with the outer periphery thereof, which guide the valve element 20 in the axial direction along the refrigerant passage 18, and has refrigerant passages 41 formed between the valve element 20 and the inner wall of the refrigerant passage 18, which do not vary in the passage area even when the flow rate of refrigerant is varied and the valve lift is varied.

In the first valve 11, a hollow cylindrical valve seat-forming member 42 is disposed upstream of the valve element 20 in a manner opposed thereto. The valve seat-forming member 42 is press-fitted in the port 16 into which refrigerant at the discharge pressure Pdh is introduced, thereby forming the first valve 11 together with the valve element 20.

Further, in the valve element 20 of the first valve 11, the shaft 29 forming the valve element of the third valve 13 is disposed in a manner movable back and forth in the direction of axis of the valve element 20. Interposed between the valve element 20 and the shaft 29 is a spring 21 which urges the valve element 20 and the shaft 29 in respective directions of moving apart to thereby maintain the closed states of the first valve 11 and the third valve 13 when the solenoid 14 is deenergized.

In the second valve 12, the valve element 27 thereof is formed integrally with the shaft 33 of the solenoid 14, and the spring 28 urging the valve element 27 in the valve-opening direction is interposed between the core 31 of the solenoid 14 and the plunger 32. The shaft 33 of the solenoid 14 is terminated by the transmission shaft 43 extending through the valve hole of the second valve 12, and the transmission shaft 43 is disposed in a hole formed between the refrigerant passage 23 and the port 17 connected to the outlet port of the compressor in a manner movable back and forth in the direction of axis of the shaft 33.

In the control valve 10b constructed as described above, when the solenoid 14 is not energized, as shown in FIG. 9, the first valve 11 and the third valve 13 have the valve element 20 and the shaft 29 urged by the spring 21 in respective directions of moving away from each other, whereby the valve element 20 is seated on the end face of the valve seat-forming member 42 and the shaft 29 is seated on the opening end of the hole holding the transmission shaft 43, which makes both of them fully closed. Further, the spring 28 of the solenoid 14 urges the plunger 32 against the urging force of the spring 34 in a direction away from the core 31, whereby the valve element 27 of the second valve 12 integrally formed with the shaft 33 fixed to the plunger 32 is moved downward as viewed in FIG. 9, whereby the second valve 12 is in the fully open state.

Therefore, when the compressor is driven for rotation by the engine, the refrigerant at the pressure Pdh introduced from the discharge chamber of the compressor is completely supplied from the port 25 to the crankcase via the strainer 24, the refrigerant passage 23, and the second valve 12, so that the compressor is operated with the minimum capacity. If this deenergized state of the solenoid 14 corresponds to a state of the compressor which has stopped its operation after operating with a predetermined capacity, the discharge pressure Pdh at the port 16 connected to the discharge chamber becomes lower than the discharge pressure Pdl at the port 17 connected to the outlet port of the compressor, so that the differential pressure therebetween holds the first valve 11 having configuration of a check valve in the fully closed state.

Next, when the solenoid 14 is in the deenergized state as shown in FIG. 9, if a predetermined electric current is passed thereto, immediately thereafter, as shown in FIG. 10, the second valve 12 is instantaneously fully closed by the urging force of the solenoid 14. This causes the compressor to shift to the maximum capacity operation, but immediately after the energization, the discharge pressure Pdh is still lower than the discharge pressure Pdl at the outlet port of the compressor, so that the first valve 11 is in the fully-closed state. At this time, the shaft 29 of the third valve 13 is pushed upward as viewed in FIG. 9 by the transmission shaft 43 integrally formed with the valve element 27 of the second valve 12.

When the compressor continues to operate with the maximum capacity to make the discharge pressure Pdh higher than the discharge pressure Pdl by a predetermined value or more, as shown in FIG. 11, the differential pressure therebetween pushes open the valve element 20. This causes the refrigerant at the discharge pressure Pdh introduced into the port 16 passes through the refrigerant passages 41 formed between the valve element 20 and the inner wall of the refrigerant passage 18 to be changed into refrigerant at the discharge pressure Pdl, which flows from the port 17 to the outlet port of the compressor.

When the first valve 11 is opened, the valve element 20 thereof is brought into the upper end face of the shaft 29 as viewed in FIG. 9, and the lower end of the shaft 29 as viewed in FIG. 9 is abutment with the upper end face of the transmission shaft 43 as viewed in FIG. 9. As a consequence, after the first valve 11 is opened, the valve element 20 of the first valve 11, the valve element 27 of the second valve 12, and the shaft 29 of the third valve 13 come to operate in unison with each other. This causes the first valve 11 and the second valve 12 to operate in an interlocked manner via the shaft 29, whereby the second valve 12 operates by detecting the differential pressure between the discharge pressure Pdh and the discharge pressure Pdl acting on the first valve 11 and the differential pressure between the discharge pressure Pdh and the pressure Pc.

Here, let it be assumed that the control valve 10b is supplied with a predetermined energization current to control the compressor to a predetermined capacity, and is in a balanced state shown in FIG. 11. If the rotational speed of the engine increases to increase the discharge pressure Pdh, the valve element 20 of the first valve 11 is lifted by an amount corresponding to an increase in the differential pressure across the first valve 11, thereby urging the valve element 27 of the second valve 12 via the shaft 29 in a valve-opening direction. This increases the pressure Pc in the crankcase so that the compressor operates in the direction of reducing the capacity thereof, whereby it is controlled to a predetermined discharge flow rate. Inversely, when the discharge pressure Pdh lowers to reduce the differential pressure across the first valve 11, the valve element 20 moves the valve element 27 of the second valve 12 in the direction of further closing the same. This reduces the pressure Pc in the crankcase, so that compressor operates in the direction of increasing the capacity, whereby it is controlled to the predetermined discharge flow rate.

FIG. 12 is a central longitudinal cross-sectional view showing details of the construction of a control valve for a variable displacement compressor, according to a fourth embodiment of the present invention. It should be noted that component elements in FIG. 12 identical or similar to those shown in FIG. 7 are designated by identical reference numerals, and detailed description thereof is omitted.

The control valve 10c according to the fourth embodiment is distinguished from the control valve 10b according to the third embodiment, in that the port 44 for introducing refrigerant into the second valve 12 is provided independently of the port 16 for introducing refrigerant into the first valve 11. This port 44 is formed in a side of the second body 22, and O rings are provided on axially opposite sides of the valve element 27, with the port 44 located between the O rings.

It is possible to introduce part of refrigerant at the discharge pressure Pdh, which is introduced from the discharge chamber of the compressor into the port 16 of the first valve 11, into the port 44 of the second valve 12. However, preferably, the control valve 10c is applied to a variable displacement compressor equipped with an oil separator at a location downstream of the discharge chamber, whereby refrigerant at discharge pressure Pdh2 is supplied from an outlet port of the oil separator to the port 44.

FIG. 13 is a central longitudinal cross-sectional view showing details of the construction of a control valve for a variable displacement compressor, according to a fifth embodiment of the present invention. It should be noted that component elements in FIG. 13 identical or similar to those shown in FIG. 12 are designated by identical reference numerals, and detailed description thereof is omitted.

When compared with the control valve 10c according to the fourth embodiment, the control valve 10d according to the fifth embodiment has quite the same internal construction, but is distinguished therefrom in that the port 44 for introducing refrigerant into the second valve 12 and the port 25 for delivering refrigerant are reversed in location.

Due to this arrangement of the ports, refrigerant at the controlled pressure Pc is delivered between the valve element 27 of the second valve 12 formed integrally with the shaft 33 of the solenoid 14 and the transmission shaft 43, via the port 25. The pressure Pc is applied to the valve element 27 and the transmission shaft 43 from opposite directions to be cancelled out, which makes it possible to eliminate influence of the pressure Pc on the control operation of the control valve 10d. Therefore, the control valve 10d can cause refrigerant to flow at a flow rate determined by the passage area of the refrigerant passages 41 between the valve element 20 of the first valve 11 and the inner wall of the refrigerant passage 18 and the differential pressure between the discharge pressure Pdh and the discharge pressure Pdl on opposite sides of the refrigerant passage 41. The differential pressure between the discharge pressure Pdh and the discharge pressure Pdl is set by the solenoid 14, and the set differential pressure is held at a predetermined value by the first valve 11 and the second valve 12 operating in an interlocked manner to control the pressure Pc in the crankcase. As a result, the flow rate of refrigerant flowing through the first valve 11 to the outlet port of the compressor is held constant.

FIG. 14 is a central longitudinal cross-sectional view showing details of the construction of a control valve for a variable displacement compressor, according to a sixth embodiment of the present invention. It should be noted that component elements in FIG. 14 identical or similar to those shown in FIG. 12 are designated by identical reference numerals, and detailed description thereof is omitted.

The control valve 10e according to the sixth embodiment is distinguished from the control valve 10d according to the fifth embodiment in the improvement of influence of the internal leakage of refrigerant on variable displacement control.

More specifically, as distinct from the port arrangement of the control valve 10d, in the control valve 10e, a port 45 communicating with the suction chamber is formed between the port 17 at the discharge pressure Pdl and the port 25 of the crankcase at the pressure Pc. This lengthens the distance between the port 17 at the discharge pressure Pdl and the port 25 at the pressure Pc, and the transmission shaft 43 of the shaft 33 as well is lengthened, so that a shaft 46 as an additional component is interposed between the transmission shaft 43 and the third valve 13, and one end of the movable part of the solenoid 14 is supported by the transmission shaft 43 alone.

With this construction, when the control valve 10e is performing variable displacement control of the compressor, the third valve 13 is open, and hence even if the refrigerant at the discharge pressure Pdl leaks via the clearance between the shaft 46 and the first body 15 supporting the shaft 46, the leaked refrigerant flows via the port 45 into the suction chamber, but does not flow to the port 25 connected to the crankcase. Therefore, refrigerant leakage into the crankcase which directly determines the displacement of the compressor is prevented from occurring, and hence the pressure Pc in the crankcase does not vary, thereby making it possible to accurately carry out the displacement control of the compressor. It should be noted that also between the port 25 for delivering the controlled pressure Pc and the port 45 at the suction pressure Ps, refrigerant leakage occurs through a clearance between the transmission shaft 43 and the first body 15 holding the same. However, this clearance is smaller in passage area than an orifice provided within the compressor at a location between the crankcase and the suction pressure, for allowing refrigerant to flow from the crankcase into the suction pressure, and hence it does not adversely affect the displacement control of the compressor. If the passage area of the orifice is set by taking the clearance into account, the influence of the refrigerant leakage through the clearance can be substantially eliminated.

In the embodiments described above, it is assumed that the present invention is applied to a system using Hydrochlorofluorocarbon “HFC-134a” as a refrigerant of the refrigeration cycle. On the other hand, when the present invention is applied to a system using refrigerant very high in operating pressure, such as carbon dioxide, it is required to control high pressure, and hence, particularly, in the second valve 12, the valve diameter and the like thereof are required to be designed small so as to reduce the pressure-receiving area, and the method of sealing the control valve and the compressor in which the control valve is mounted, from outside, needs to be changed. Now, a description will be given of an embodiment which considers its application to a system using carbon dioxide as refrigerant.

FIG. 15 is a central longitudinal cross-sectional view showing details of the construction of a control valve for a variable displacement compressor, according to a seventh embodiment of the present invention. It should be noted that component elements in FIG. 15 identical or similar to those shown in FIG. 13 are designated by identical reference numerals, and detailed description thereof is omitted.

The control valve 10f according to the seventh embodiment is distinguished from the control valve 10d according to the fifth embodiment in that the valve element 27 of the second valve 12 and the transmission shaft 43, which are formed integrally with the shaft 33 of the solenoid 14 in the latter embodiment, are formed separately from the shaft 33, and are urged by the spring 28 in the direction of opening the second valve 12. This makes it possible to form the valve element 27 of the second valve 12 using a material thin but robust, and enhance the freedom of design. Further, in the control valve 10f, the second body 22 and the core 31 of the solenoid 14 are formed integrally with each other, and the core 31 is press-fitted into the bottomed sleeve 30 having an open end formed with a flange. The outer periphery of the flange has a packing 47 disposed thereon. The packing 47 is formed of a material which is free from penetration of refrigerant and leakage of the same to the outside. Further, a screw thread 49 for mounting the control valve 10f in the compressor is formed on an outer peripheral portion, close to the flange, of a casing 48 serving as a yoke of the solenoid.

FIG. 16 is a view showing, by way of example, an application of the control valve according to the seventh embodiment to a variable displacement compressor for a carbon dioxide system.

The variable displacement compressor includes a hermetically formed crankcase 51, which contains a rotating shaft 52 rotatably supported therein. One end of the rotating shaft 52 extends via a sealed bearing device to the outside of the crankcase 51, and a pulley 53 transmitted a drive force from an engine for an automotive vehicle is fixed to the one end of the rotating shaft 52. The rotating shaft 52 has a wobble plate 54 fitted thereon such that the inclination angle of the wobble plate 54 can be varied. Around the axis of the rotating shaft 52, there are arranged a plurality of cylinders 55 (one of which is shown in FIG. 16). Each cylinder 55 has a piston 56 disposed therein, for converting the rotating motion of the wobble plate 54 into reciprocating motion. The cylinder 55 is connected to a suction chamber 59 and a discharge chamber 60 via a suction relief valve 57 and a discharge relief valve 58, respectively. The control valve 10f is disposed between the discharge chamber 60 and an outlet port 61 formed to communicate therewith and between the discharge chamber 60 and the crankcase 51, and an orifice 62 is provided between the crankcase 51 and the suction chamber 59. The compressor further includes a passage, indicated by a broken line in FIG. 16, which extends from the discharge chamber 60 to the control valve 10f. The control valve 10f is inserted into a mounting hole of the compressor, and is mounted therein by screwing.

In the variable displacement compressor, the outlet port 61 is connected via a gas cooler 63 and an internal heat exchanger 64, to an expansion valve 65 by a high-pressure refrigerant conduit line, and the expansion valve 65 is connected via an evaporator 66, an accumulator 67, and again the internal heat exchanger 64, to an inlet port which is formed to communicate with a suction chamber 59, by a low-pressure refrigerant conduit line, whereby a refrigeration cycle as a closed circuit is formed.

In the variable displacement compressor constructed as above, as the rotating shaft 52 to which the drive force is transmitted from the engine is rotated, the wobble plate 54 fitted on the rotating shaft 52 wobbles while rotating. Then, each piston 56 connected to the outer peripheral part of the wobble plate 54 performs reciprocating motion in a direction parallel to the axis of the rotating shaft 52, whereby refrigerant at suction pressure Ps in the suction chamber 59 is drawn into the associated cylinder 55 and compressed therein, and the compressed refrigerant at discharge pressure Pdh is discharged into the discharge chamber 60. At this time, high-pressure refrigerant in the discharge chamber 60 is decompressed to discharge pressure Pdl when passing through the control valve 10f, and delivered from the outlet port 61 to the gas cooler 64. Part of the high-pressure refrigerant at the discharge pressure Pdh2 is introduced into the crankcase 51 via the control valve 10f. This causes the pressure Pc in the crankcase 51 to rise, whereby the inclination angle of the wobble plate 54 is set such that the bottom dead center of the piston 56 is brought to a position where the pressure in the cylinder 55 and the pressure Pc in the crankcase 51 are balanced. Thereafter, the refrigerant introduced into the crankcase 51 is returned to the suction chamber 59 via the orifice 62.

In the control valve 10f, the first valve 11 detects a flow rate of refrigerant sent from the discharge chamber 60 to the gas cooler 63, and the second valve 12 introduces the refrigerant into the crankcase 51 at a flow rate dependent on the detected flow rate, thereby providing control such that the flow rate of the refrigerant sent from the discharge chamber 60 to the gas cooler 63 becomes constant. For example, when the rotational speed of the engine increases, the discharge pressure Pdh rises. This increases the flow rate of refrigerant sent from the discharge chamber 60 to the gas cooler 63 via the control valve 10f, to increase the differential pressure across the first valve 11. According to an increase in the differential pressure, the second valve 12 opens, and the flow rate of refrigerant at the discharge pressure Pdh2 introduced into the crankcase 51 also increases, whereby the pressure Pc in the crankcase 51 increases. Accordingly, in the variable displacement compressor, the wobble plate 54 is inclined in such a direction as will cause the wobble plate 54 to become at right angles to the rotating shaft 52 to decrease the stroke of the pistons 56, which acts on the compression capacity of the cylinders 55 in the direction of reducing the same to reduce the discharge flow rate of refrigerant. Thus, even when the flow rate of discharged refrigerant is about to increase due to an increase in the rotational speed of the engine, the control valve 10f increases the flow rate of refrigerant introduced into the crankcase 51 according to the increase in the flow rate of refrigerant, whereby the pressure Pc in the crankcase 51 is increased to reduce the displacement of the compressor. Therefore, the flow rate of refrigerant discharged from the compressor is controlled to be constant. Similarly, when the rotational speed of the engine lowers, the flow rate of refrigerant at the discharge pressure Pdl sent from the discharge chamber 60 to the gas cooler 63 via the control valve 10f is decreased, whereby the flow rate of refrigerant at the discharge pressure Pd2 introduced into the crankcase 51 is also decreased to lower the pressure Pc in the crankcase 51. As a result, the compressor has the discharge flow rate of refrigerant controlled such that it is increased, whereby the flow rate of discharged refrigerant is controlled to be constant.

The control valve for a variable displacement compressor, according to the present invention, is configured such that the first valve indirectly measures the discharge flow rate of refrigerant, and the second valve is controlled based on a value of the discharge flow rate to thereby control the pressure in the crankcase. This is advantageous in that it is possible to dispense with expensive pressure sensors for detecting the discharge flow rate, and reduce the cost of the automotive air conditioner.

Further, the first valve has a structure that opens depending on the flow rate of refrigerant flowing from the discharge chamber of the compressor toward the outlet port, and hence closes when the compressor shifts to the minimum capacity operation to make the flow rate of refrigerant the minimum, and further, holds the closed state when the relationship in pressure between the discharge chamber and the outlet port of the compressor, as in the case of immediately after transition to the minimum capacity operation, to thereby hold the pressure at the outlet port at the pressure assumed before transition to the minimum capacity operation. This makes it possible to abolish a check valve provided at the outlet port of the compressor for this purpose, and thereby reduce the cost of the compressor.

Further, the first valve is configured to have a larger pressure-receiving area for receiving discharge pressure on the discharge chamber side than the second valve, and hence it is possible to construct a highly responsive variable displacement compressor that is operable when the rotational speed of the compressor has rapidly changed, to promptly react in the direction of suppressing a change in displacement caused thereby.

The foregoing is considered as illustrative only of the principles of the present invention. Further, since numerous modifications and changes will readily occur to those skilled in the art, it is not desired to limit the invention to the exact construction and applications shown and described, and accordingly, all suitable modifications and equivalents may be regarded as falling within the scope of the invention in the appended claims and their equivalents.

Claims

1. A control valve for a variable displacement compressor, for controlling a flow of refrigerant discharged from the compressor to a constant flow rate, comprising: a first valve that has a passage area thereof set according to a flow rate of refrigerant introduced from a discharge chamber of the compressor and flowing out to an outlet port of the compressor; a second valve that controls pressure in a crankcase of the compressor in a manner interlocked with the first valve such that a differential pressure across the first valve is maintained at a predetermined differential pressure; and a solenoid that sets a differential pressure across the passage having the passage area set by the first valve to the predetermined differential pressure dependent on a flow rate of refrigerant to which the flow of refrigerant is to be controlled.
2. The control valve according to claim 1, wherein the first valve comprises a first valve seat formed in a first refrigerant passage through which refrigerant is introduced from the discharge chamber and is allowed to flow out to the outlet port, and a first valve element disposed in a manner opposed to the first valve seat, in a state urged from a downstream side in a valve-closing direction, and wherein the second valve comprises a second valve seat formed in a second refrigerant passage through which refrigerant is introduced from the discharge chamber and is allowed to flow out into the crankcase, a second valve element disposed downstream of the second valve seat in a manner opposed thereto, the second valve element having a smaller pressure-receiving area than the first valve element, and a spring urging the second valve element against an urging force of the solenoid, in a valve-opening direction, the control valve further comprising a shaft that is disposed between the first valve element and the second valve element, for transmitting the urging force of the solenoid to the first valve element and transmitting a change in the differential pressure across the passage received by the first valve element to the second valve element.
3. The control valve according to claim 2, wherein the shaft is held in a through hole formed through a body of the first valve in a manner movable in a valve-opening direction or a valve-closing direction of the first valve element and the second valve element, the shaft having a larger outer diameter on a side toward the first valve element than an inner diameter of the through hole, thereby forming a third valve that closes a clearance between the shaft and the through hole when the second valve is fully opened by the urging force of the spring during deenergization of the solenoid.
4. The control valve according to claim 2, wherein the second valve is configured to operate by sensing a differential pressure between discharge pressure of refrigerant introduced from the discharge chamber and pressure of refrigerant allowed to flow out into the crankcase.
5. The control valve according to claim 2, wherein the second valve is configured to operate by sensing a differential pressure between discharge pressure of refrigerant introduced from the discharge chamber and suction pressure in a suction chamber of the compressor.
6. The control valve according to claim 5, wherein the second valve has the second valve element and another shaft integrally formed with each other, the another shaft having an outer diameter approximately equal to an effective diameter of the second valve element on which discharge pressure of refrigerant introduced from the discharge chamber is received, and receiving the suction pressure on an end face thereof.
7. A control valve for a variable displacement compressor, for controlling a flow of refrigerant discharged from the compressor to a constant flow rate, comprising: a first valve that is disposed in a first refrigerant passage through which refrigerant is introduced from a discharge chamber of the compressor and flows out to an outlet port of the compressor, the first valve having a configuration of a check valve; a second valve that is arranged in a second refrigerant passage through which refrigerant is introduced from the discharge chamber and flows out into a crankcase of the compressor; a shaft that is disposed between the first valve and the second valve, for transmitting a change in a differential pressure across the first valve to the second valve, to thereby cause the first valve and the second valve to operate in the same valve-opening or valve-closing direction in an interlocked manner; and a solenoid that gives an urging force in a valve-closing direction to the second valve according an electric current value, and sets via the shaft a differential pressure across the first valve to a predetermined differential pressure dependent on a flow rate of refrigerant to which the flow of refrigerant is to be controlled.
8. A control valve for a variable displacement compressor, for controlling a flow of refrigerant discharged from the compressor to a constant flow rate, comprising: a first valve that is lifted according to a flow rate of refrigerant introduced from a discharge chamber of the compressor and flowing out to an outlet port of the compressor, and is not varied in a passage area of a refrigerant passage irrespective of an amount of lift, after the first valve is opened; a second valve that controls pressure in a crankcase of the compressor in a manner interlocked with the first valve such that a differential pressure across the first valve is maintained at a predetermined differential pressure; and a solenoid that sets a differential pressure across the refrigerant passage assumed when the first valve is opened, to the predetermined differential pressure dependent on a flow rate of refrigerant to which the flow of refrigerant is to be controlled.
9. The control valve according to claim 8, wherein the first valve comprises a first valve seat formed in a first refrigerant passage through which the refrigerant is introduced from the discharge chamber and is allowed to flow out to the outlet port, and a first valve element disposed on a downstream side of the first valve seat in a manner opposed thereto and movable axially back and forth in a state in which an outer periphery of the first valve element is spaced from an inner wall of the first refrigerant passage by a predetermined distance, and a spring urging the first valve element in a valve-closing direction, and wherein the second valve comprises a second valve seat formed in a second refrigerant passage through which refrigerant is introduced from the discharge chamber and is allowed to flow out into the crankcase, a second valve element disposed on a side toward the solenoid in a manner opposed to the second valve seat, in a state movable axially back and forth and urged in a valve-opening direction the second valve element having a smaller pressure-receiving area than the first valve element, and a transmission shaft disposed in a through hole that communicates with the first refrigerant passage via a valve hole, for operating in unison with the second valve element, the control valve further comprising a shaft that transmits the urging force of the solenoid given to the second valve element to the first valve element via the transmission shaft, and transmits a change in the differential pressure across the refrigerant passage received by the first valve element to the second valve element.
10. The control valve according to claim 9, wherein the shaft is disposed in a manner movable axially back and forth in the first valve element and urged by the spring in a direction toward the second valve with respect to the first valve element, and forms a third valve that is abutted by the transmission shaft urged to move forward from the through hole into the first refrigerant passage, when the solenoid is energized, to operate in unison with the first valve element and the second valve element, and closes the through hole into which the transmission shaft is retracted due to the second valve element being urged in a valve-opening direction, when the solenoid is deenergized.
11. The control valve according to claim 9, wherein the second valve element and the transmission shaft of the second valve are formed integrally with a shaft of the solenoid.
12. The control valve according to claim 9, wherein a refrigerant inlet port of the first refrigerant passage through which refrigerant is introduced from the discharge chamber and a refrigerant inlet port of the second refrigerant passage are communicated with a first port and a second port formed independently of each other, respectively.
13. The control valve according to claim 12, wherein the second port as the refrigerant inlet port of the second refrigerant passage is formed toward the first valve with respect to the second valve seat, and on a side toward the solenoid with respect to the second valve seat, a third port is disposed which communicates with a refrigerant outlet port of the second refrigerant passage.
14. The control valve according to claim 12, wherein the second port as the refrigerant inlet port of the second refrigerant passage is disposed toward the solenoid with respect to the second valve seat, and on a side toward the first valve with respect to the second valve seat, a third port is disposed which communicated with a refrigerant outlet port of the second refrigerant passage.
15. The control valve according to claim 14, wherein a fifth port which communicates with a suction chamber of the compressor is disposed between the third port and a fourth port communicating with a refrigerant outlet of the first refrigerant passage, whereby refrigerant leaking through a clearance between the through hole and the transmission shaft is caused to flow into the fifth port.

Priority Claims (2)

Number	Date	Country	Kind
2004-269661	Sep 2004	JP	national
2005-161179	Jun 2005	JP	national

Control valve for variable displacement compressor

Information

Publication Number

Date Filed

Date Published

Inventors

Original Assignees

CPC

US Classifications

International Classifications

Abstract

Description

Claims

Priority Claims (2)