JPH11324930A - Variable capacity type compressor - Google Patents

Abstract

PROBLEM TO BE SOLVED: To improve the fuel consumption of a vehicle (engine) without generating a rise of the manufacturing cost of a compressor (freezing cycle). SOLUTION: This compressor is mechanically interlocked with a differential pressure Δp in a first orifice 113 provided in a coolant passage 112 at a discharge side and controls a variable capacity mechanism VVc. With this structure, since the discharge capacity can be maintained nearly constant without using an electric means such as a solenoid valve, fuel consumption of the engine can be improved without generating a rise of the manufacturing cost of the compressor.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a variable displacement compressor for a refrigerating cycle driven by a vehicle engine.

[0002]

2. Description of the Related Art As a variable displacement compressor for a refrigeration cycle, for example, in the invention described in Japanese Patent Publication No. 6-15872, a variable displacement mechanism is operated by utilizing a differential pressure between a suction pressure and a discharge pressure. I have.

[0003]

In recent years, there has been a growing demand for improved fuel efficiency. In response to this requirement, in the variable displacement compressor (hereinafter abbreviated as compressor) described in the above publication, the heat load of the refrigeration cycle, that is, the suction pressure (pressure inside the evaporator) becomes equal to or less than a predetermined value. For example, even if the engine speed is increased in a state where the suction pressure (heat load) is high, the discharge capacity does not change until the suction pressure becomes equal to or less than a predetermined value.

[0004] For this reason, the mechanical work required to operate the compressor increases in proportion to the increase in the engine speed, which causes a problem that fuel efficiency deteriorates. In order to solve this problem, a means is known in which the discharge capacity of the compressor is controlled in accordance with the engine speed by adjusting the pressure on the discharge side with an electromagnetic valve (Japanese Patent Publication No. 2-55).
636).

However, in this means, in addition to the solenoid valve,
Since an electric component such as a control device for controlling the solenoid valve is required, a new problem occurs in that the manufacturing cost of the compressor (refrigeration cycle) increases. In view of the above, it is an object of the present invention to provide a variable displacement compressor suitable for improving fuel efficiency without increasing the manufacturing cost of the compressor (refrigeration cycle).

[0006]

The present invention uses the following technical means to achieve the above object. Claim 1
In the invention described in Item 3, the orifice (11) provided in the refrigerant passage (112) for the refrigerant discharged from the compression mechanism (Cp).
3) Differential pressure (ΔP) between the upstream and downstream of the refrigerant flow across
Variable mechanical mechanism (VVc) that operates mechanically in conjunction with
And a control mechanism (Cv) for mechanically controlling the operation of.

First, since the operation of the variable displacement mechanism (VVc) is mechanically linked to the differential pressure ΔP, the discharge displacement is controlled by using an electromagnetic valve as disclosed in Japanese Patent Publication No. 2-55636. , The manufacturing cost of the variable displacement compressor can be reduced. Second, differential pressure (ΔP)
Changes in proportion to the square of the discharge flow rate, the discharge capacity of the variable displacement compressor is determined by the set value of the orifice (113) and the control mechanism (Cv), as described later. Is mechanically controlled so as to achieve the discharge capacity. Therefore, even if the engine speed increases, the discharge capacity decreases and changes to maintain the discharge capacity substantially constant, so that it is possible to prevent an increase in mechanical work required for operating the variable displacement compressor. .

As described above, in the variable displacement compressor according to the present invention, it is possible to improve the fuel efficiency of the vehicle (engine) without increasing the manufacturing cost of the variable displacement compressor. As described in the second aspect, the variable displacement mechanism (VVc) has control pressure chambers (120, 133) communicating with the suction side and the discharge side of the compression mechanism (Cp). Changing the discharge capacity by changing the pressure in the control pressure chambers (120, 133);
Further, the control mechanism (Cv) mechanically interlocks with the differential pressure (ΔP) to open and close a passage for communicating the control pressure chamber (120, 133) with the suction side or the discharge side (125).
It is desirable to have a configuration.

The valve means (125) is provided with a differential pressure (ΔP)
It is desirable to be able to be opened and closed by a pressure responsive member (128) that is moved mechanically in conjunction with the first member. Incidentally, the reference numerals in parentheses of the above means are examples showing the correspondence with specific means described in the embodiments described later.

[0010]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS (First Embodiment) FIG. 1 is a schematic diagram of a vehicle refrigeration cycle using a scroll-type variable displacement compressor (hereinafter abbreviated as a compressor) 100 according to the present embodiment. And 200, a condenser (radiator) for cooling the refrigerant discharged from the compressor 100. 300 is the condenser 2
An expansion valve (decompressor) whose pressure is reduced so that the degree of heating at the outlet side of the evaporator 400, which will be described later, becomes a predetermined value. An evaporator for evaporating the decompressed liquid-phase refrigerant.

The compressor 100 is driven by a vehicle traveling engine (hereinafter abbreviated as engine) 500 via a V-belt and an electromagnetic clutch (not shown).
Next, the structure of the compressor 100 will be described. FIG. 2 shows a cross section of the compressor 100, and 101 is a shaft that is driven to rotate via an electromagnetic clutch. Reference numeral 102 denotes a front housing that holds a rolling bearing 103 that rotatably supports the shaft 101. A fixed scroll (fixed portion) 104 having a spiral tooth portion 104a is fixed to the front housing 102. .

The space formed by the fixed scroll 104 and the front housing 102 has a tooth 1
A movable scroll (movable portion) 105 provided with a spiral tooth portion 105a that meshes with the 04a is provided. The movable scroll 105 is rotatably mounted via a bearing 101b to a crank portion (eccentric portion) 101a formed at a position eccentric from the rotation center of the shaft 101 by a predetermined amount.

The orbiting scroll 105 orbits around the shaft 101 with the rotation of the shaft 101, so that the volume of the working chamber Vc formed by the scrolls 104 and 105 is enlarged or reduced to suck and compress the refrigerant. Hereinafter, a mechanism that sucks and compresses the refrigerant of the scrolls 104 and 105 and the like is referred to as a compression mechanism Cp. Also,
106 is a suction chamber communicating with a suction port (not shown) connected to the outlet side of the evaporator 400, and 107 is a condenser 20
The discharge chamber communicates with the discharge port 108 connected to the inlet side of the “0”. The discharge chamber 107 is provided with the fixed scroll 10.
4 communicates with the working chamber Vc through a discharge port 109 formed in the end plate portion 104b, and the refrigerant flows back from the discharge chamber 107 to the working chamber Vc to the discharge chamber 107 side of the discharge port 109. Valve-shaped discharge valve 1 for preventing air leakage
10 are provided.

Incidentally, the discharge valve 110 is fixedly fastened to the end plate 104b together with a valve stop plate (valve retainer) 111 for regulating the maximum opening of the discharge valve 110. A first orifice (first fixed throttle) 113 which decompresses the refrigerant and has a fixed opening is provided in the middle of the refrigerant passage 112 from the discharge port 109 (discharge chamber 107) to the discharge port 108.
The pressure on the upstream side of the refrigerant flow of the first orifice 113 is guided to a fourth control chamber 130 described later, and the pressure on the downstream side is guided to a third control chamber 129 described later.

A bypass port 114 communicating with the working chamber Vc during the compression stroke is formed in the end plate portion 104b.
5 and the suction chamber (suction side) 1 via the bypass passage 116
06. A reed valve-shaped bypass valve (bypass valve element) 117 for opening and closing the bypass port 114 is provided on the intermediate chamber 115 side of the bypass port 114. The bypass valve 119 is connected to the bypass port 114 when the pressure in the intermediate chamber 115 is higher than the pressure in the working chamber Vc to which the bypass port 114 communicates (hereinafter, this working chamber Vc is referred to as an intermediate pressure working chamber).
Is closed, and when the pressure in the intermediate chamber 115 is lower than the pressure in the intermediate pressure working chamber, the bypass port 114 is opened.

Incidentally, reference numeral 118 denotes a valve stop plate (valve retainer) for regulating the maximum opening of the bypass valve 119. The valve stop plate 118 is fixed together with the bypass valve 119 to the end plate portion 104b. In the bypass passage 116,
A spool-type bypass valve 119 that opens and closes the bypass passage 116 (intermediate chamber port 115 a) is provided.
The bypass valve 119 is slidably disposed in the inside.
A first control chamber (control pressure chamber) 120 is formed by the and the fixed scroll 104. And the first control room 1
Reference numeral 20 controls the opening / closing operation of the bypass valve 119 and communicates with both the discharge chamber 107 (discharge side) and the suction chamber 106 (suction side).

The first control chamber 120 and the discharge chamber 107 always communicate with each other via a second orifice (second fixed throttle) 121 which generates a relatively large pressure loss. 120 and the suction chamber 106 are connected to the control passage 1
22 (122a-122c). The pressure in the suction chamber 106 is guided to the side opposite to the first control chamber 120 with the bypass valve 119 interposed therebetween.
A second control chamber 124 in which a first coil spring (first elastic body) 123 for applying an elastic force to the bypass valve 119 to reduce the volume of the first control chamber 120 is formed.

For this reason, the pressure in the first control chamber
When the pressure is higher than the pressure in the control chamber 124, the bypass passage 1
16 (intermediate chamber port 115a) is closed while the first
When the pressure in the control chamber 120 is lower than or equal to the pressure in the second control chamber 124, the bypass passage 116 (the intermediate chamber port 115a) is opened. The control passage 122 is provided with a control valve (valve element) 125 for opening and closing the control passage 122 (122a). And this control valve 12
5 is provided on one side with a second coil spring (second elastic body) 126 for applying an elastic force to the control valve 125 in a direction to close the control passage 122 (122a), and on the other side, a control passage Retainer (push rod) 12 for applying a force in the direction of opening 122 (122a) to control valve 125
7 are provided.

Incidentally, 128 partitions the third control room 129 and the fourth control room 130 and both control rooms 12
A movable diaphragm (pressure-responsive member) interlocked with the differential pressure between the diaphragms 9 and 130, and a retainer 127 is connected (fixed) to the diaphragm 128. Next, the compressor 10
The characteristic operation of 0 will be described.

1. At the time of the maximum capacity operation (see FIG. 1) When the shaft 101 rotates and the compressor 100 operates, the compressed refrigerant is discharged from the discharge port 106 to the discharge chamber 107, and condensed from the discharge port 108 through the refrigerant passage 112. It is discharged toward the container 200. At this time, the pressure loss when the refrigerant passes through the first orifice 113 generates a pressure difference ΔP such that the pressure in the third control chamber 129 becomes lower than the pressure in the fourth control chamber 130.
28 and the retainer 127 apply a force for opening the control passage 122 (hereinafter, this force is referred to as a valve opening force) to the control valve 125.
To act on.

The elastic force of the second coil spring 126 (hereinafter, this force is referred to as a valve closing force). Is smaller than the valve opening force, the control passage 122 is closed, so that the pressure in the first control chamber 120 becomes equal to that of the discharge chamber 107, and the bypass passage 116 (the intermediate chamber port 115a) is closed. Therefore, the pressure in the intermediate chamber 115 becomes higher than the pressure in the intermediate pressure working chamber, and the bypass port 114 is closed (the closed state is maintained), so that the compressed refrigerant flows from the bypass port 114 to the suction chamber 106 ( Suction side)
Is discharged from the discharge port 109 without flowing out. That is, the compressor 100 operates with a discharge capacity (100% capacity) close to the theoretical discharge amount of the compressor 100.

2. At the time of variable displacement operation (see FIG. 2) When the rotation speed of the engine increases and the flow rate of the refrigerant flowing through the refrigerant passage 112 (refrigerant flow rate) increases, the differential pressure ΔP increases in conjunction with this, and the valve opening force decreases It becomes larger than the valve closing force, and the control passage 122 opens. For this reason, the first control room 12
Since the pressure in the intermediate chamber 115 decreases because the pressure in the intermediate chamber 115 decreases, the pressure in the intermediate chamber 115 decreases, so that the bypass port 114 opens and the discharge capacity decreases.

Although the discharge chamber 107 and the first control chamber 120 are always in communication, the opening area (throttle diameter) of the second orifice 121 is selected to be sufficiently smaller than that of the control passage 122. By opening and closing the control passage 122, the pressure in the first control chamber 120 can be changed (controlled). On the other hand, when the bypass port 114 is opened, the refrigerant flows out from the intermediate pressure working chamber to the suction chamber 106 (suction side), so that the flow rate of the refrigerant flowing through the refrigerant passage 112 decreases, and the differential pressure ΔP decreases. Therefore, the control valve 125
Moves in the direction to close the control passage 122 (to the right in the drawing), the bypass port 114 is closed, and the state shifts to the maximum capacity operation state.

Therefore, the compressor 10 according to the present embodiment
0, the amount of refrigerant flowing through the refrigerant passage 112 (the compressor 10
0 discharge capacity) and the valve opening force exceeds the valve closing force,
The discharge capacity changes so as to decrease. On the other hand, when the discharge capacity decreases and the valve opening force decreases, the discharge capacity changes so as to increase. That is, in the compressor 100 according to the present embodiment, the compressor 100 is mechanically controlled so that the predetermined displacement is determined by the valve opening force determined by the differential pressure ΔP and the valve closing force by the second coil spring 126. Will be controlled in a controlled manner.

As is apparent from the above description of operation, in this embodiment, the variable displacement mechanism VVc (FIG. 1) that changes the displacement of the refrigerant discharged from the compression mechanism Cp by the bypass port 114 and the bypass valve 119 and the like. The control passage 122, the control valve 125, the diaphragm 128, and the like constitute a control mechanism Cv (see FIG. 1) for mechanically controlling the operation of the variable displacement mechanism VVc.

Next, the features of this embodiment will be described. According to the present embodiment, the control passage 1 is mechanically linked to the differential pressure ΔP.
Since the discharge capacity is controlled by opening and closing the compressor 22, the compressor 100 is controlled in comparison with the control of the discharge capacity using an electromagnetic valve as disclosed in Japanese Patent Publication No. 2-55636.
Manufacturing cost can be reduced.

Further, even if the engine speed increases, the discharge capacity decreases and changes to maintain the discharge capacity substantially constant, thereby preventing an increase in the mechanical work required to operate the compressor 100. it can. As described above, in the compressor 100 according to the present embodiment, the fuel efficiency of the vehicle (engine) can be improved without increasing the manufacturing cost of the compressor 100.

(Second Embodiment) In the above-described embodiment, the present invention is applied to a scroll type compressor, but this embodiment is applied to a swash plate compressor as shown in FIG. That is, the variable displacement mechanism VVc of the compressor 100 according to the present embodiment changes the reciprocating stroke (stroke) of the piston 132 by changing the inclination angle of the swash plate 131, as is well known. Also, the control mechanism C
v is the same as in the first embodiment.

Then, the swash plate chamber 1 in which the swash plate 131 is stored.
33 is always in communication with the suction chamber 106 through a communication passage 134 and is in communication with the discharge port 109 through a control passage 122 opened and closed by a control valve 125. Note that, as will be apparent from the operation description to be described later, the swash plate chamber 133 corresponds to the first control chamber 120 in the first embodiment.

Next, the general operation of the present embodiment will be described. 1. At the time of the maximum capacity operation At the time of the maximum capacity operation, as in the first embodiment, since the control passage 122 is closed, the pressure in the swash plate chamber 133 becomes substantially equal to the pressure in the suction chamber 106 (suction side), and the swash plate 131 The compressor 100 operates in a state where is inclined most. For this reason,
Since the reciprocating stroke of the piston 132 is maximized, a maximum capacity operation (100% operation) state is set.

2. At the time of the variable capacity operation At the time of the maximum capacity operation, as in the first embodiment, since the control passage 122 is opened, the pressure in the swash plate chamber 133 increases. Therefore, the angle between the swash plate 131 and the shaft 101 (hereinafter, this angle is referred to as the swash plate angle θ) gradually becomes 90 ° by the balance between the pressure in the working chamber Vc and the pressure in the swash plate chamber 133. As approaching, the discharge capacity decreases.

When the discharge capacity decreases, the control passage 122 closes again, so that the swash plate angle θ decreases, the reciprocating stroke of the piston 132 increases, and the discharge capacity increases. That is, the compressor 100 according to the present embodiment
Similarly to the compressor 100 according to the first embodiment, the differential pressure ΔP
The compressor 100 is mechanically controlled so as to have a predetermined discharge capacity determined by the valve opening force determined by the above and the valve closing force by the second coil spring 126.

In the first embodiment, the discharge chamber 10
7 and the first control room 120 are always in communication, and the first control room 1
The control mechanism Cv for mechanically controlling the operation of the variable displacement mechanism VVc is configured by controlling the opening and closing of the control passage 122 that communicates the suction chamber 20 with the suction chamber 106, but the first embodiment is not limited to this. Not the first control room 12
0 and the suction chamber 106 side at all times, and by controlling the communication state between the discharge chamber 107 and the first control chamber 120,
The control mechanism Cv may be configured.

Further, the compression mechanism Cp of the compressor 100 according to the present invention is not limited to the scroll type or the swash plate type, but may employ another type of compression mechanism. Further, in the above-described embodiment, the thin film-shaped diaphragm 1 is used as the pressure responsive member that is movable mechanically in association with the differential pressure ΔP.
Although 28 is used, the present invention is not limited to this, and other members such as bellows bellows may be used.

[Brief description of the drawings]

FIG. 1 is a schematic diagram of a refrigeration cycle.

FIG. 2 is a cross-sectional view of the compressor according to the first embodiment (at the time of maximum capacity operation).

FIG. 3 is a cross-sectional view of the compressor during a variable displacement operation.

FIG. 4 is a sectional view of a compressor according to a second embodiment (at the time of maximum capacity operation).

[Explanation of symbols]

Vc: working chamber, Cp: compression mechanism, VVc: variable capacity mechanism, 109: discharge port, 113: first orifice, 1
14: bypass port, 122: control passage, 125: control valve (valve means), 128: diaphragm (pressure responsive member).

Claims (3)

[Claims] 1. A vehicle driven by a vehicle driving engine.
A variable displacement compressor for a refrigeration cycle, comprising: a compression mechanism (Cp) having a working chamber (Vc) for sucking and compressing refrigerant; and a variable displacement for changing the discharge capacity of refrigerant discharged from the compression mechanism (Cp). Mechanism (VVc) and a refrigerant passage (1) for refrigerant discharged from the compression mechanism (Cp).
And an orifice (11) provided in
3) mechanically controlling the operation of the variable displacement mechanism (VVc) by interlocking with the differential pressure (ΔP) between the upstream and downstream of the refrigerant flow across the orifice (113). A variable capacity compressor comprising: 2. The variable capacity mechanism (VVc) has control pressure chambers (120, 133) communicating with the suction side and the discharge side of the compression mechanism (Cp), and the control pressure chamber (120). , 133) to change the discharge capacity by changing the pressure in the control pressure chambers (120, 133) mechanically interlocking with the differential pressure (ΔP). )
2. The variable displacement compressor according to claim 1, further comprising a valve means (125) for opening and closing a passage communicating the suction side and the suction side or the discharge side. 3. 3. The valve means (125) is provided with the differential pressure (Δ
P) and a pressure responsive member (12)
The variable displacement compressor according to claim 2, wherein the compressor is opened and closed according to (8).