JPH0444112B2 - - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION [Industrial Field of Application] The present invention relates to capacity control of a swash plate compressor, and is effective when used as a refrigerant compressor for an automobile air conditioner, for example.

[Conventional technology and problems]

As a conventional capacity control method for swash plate compressors, the angle of the swash plate is continuously changed to reduce the compressor capacity to 0%.
It is known that the value can be changed continuously up to 100%. (For example, Japanese Patent Application Laid-open No. 162780/1983). However, in this type of conventional swash plate compressor, the position of the center of inclination of the swash plate is fixed;
Since the swash plate is rotatably supported on the shaft, as the slope of the swash plate becomes smaller, the stroke of the piston decreases, but the dead volume in the cylinder increases uniformly on both sides of the piston. I had a problem with putting it away. If the dead volume increases in this way, the capacity will change greatly with a slight change in the angle of the swash plate due to re-expansion of the high pressure gas in the dead volume, resulting in poor controllability.

In order to provide a variable capacity swash plate compressor in which the dead volume does not increase, for example, the piston is not a double-headed type as described in JP-A-60-175783, and the swash plate compressor is This will change the angle and position of. However, in a compressor in which the piston is provided only on one side of the swash plate, there are problems in that the discharge pulsation and torque fluctuation during operation are large, and the capacity per unit of the compressor is small. .

Because of these problems, it is desirable that the swash plate type compressor be a compressor in which pistons are arranged on both sides of the swash plate. However, in a swash plate compressor that uses this double-ended piston, there are working chambers on both sides, so pressure adjustment at the back of the piston depends on the angle of the swash plate.
It is difficult to control the position.

[Problem that the invention seeks to solve]

The present invention was devised in view of the above drawbacks, and an object of the present invention is to perform continuous capacity control while preventing the dead volume from increasing uniformly on both sides of the piston in a swash plate compressor employing a double-ended piston. shall be.

Furthermore, in the swash plate compressor according to the second aspect of the present invention, it is an object of the present invention to enable variable control of the capacity of the compressor continuously and with good responsiveness.

Further, the swash plate compressor according to the third aspect of the present invention allows the capacity of the compressor to be continuously and responsively variably controlled, and moreover, the control range is such that the stroke of the piston begins to decrease from the maximum stroke. The purpose is to enable continuous capacity control even in the initial position. In other words, it is an object of the swash plate compressor according to the third aspect of the present invention to be able to continuously control the displacement even when the compressor displacement is slightly reduced from the maximum displacement.

In the swash plate compressor according to the fourth aspect of the present invention, the capacity of the compressor can be continuously and variably controlled with good responsiveness, and the capacity can be reliably controlled even when the stroke reduction amount of the piston is large. The purpose is to provide control. In other words, an object of the swash plate compressor according to the fourth aspect of the present invention is to ensure that the discharge capacity of the compressor can be reduced to the minimum capacity.

[Configuration and operation]

In order to achieve the above object, the swash plate compressor according to the present invention employs a configuration in which the swash plate, which is rotationally driven by the shaft, decreases its inclination as the spool moves in the axial direction, thereby varying the stroke of the piston. do. Furthermore, the center of the swash plate is a spherical bearing, and this bearing is also configured to be displaced in synchronization with the spool. Therefore, although there is a significant increase in the dead volume in one working chamber, the capacity in the other working chamber is gradually reduced without a significant increase in the dead volume. Therefore, the capacity of the compressor can be continuously controlled according to the displacement of the spool.

Further, in the swash plate compressor according to the second aspect of the present invention, the axial displacement of the spool can be variably controlled by adjusting the pressure in the control pressure chamber using the control valve. Therefore, the inclination angle of the swash plate can be continuously and responsively varied by switching control of the control valve.

Furthermore, in the swash plate compressor according to the third aspect of the present invention, the displacement of the spool is assisted when the stroke reduction amount of the piston is small, in other words, when the discharge capacity of the compressor is slightly reduced from the maximum discharge capacity. An auxiliary loading means shall be provided. When the discharge capacity of the compressor is somewhat reduced, the spool will not be displaced properly due to the dead space generated in one of the working chambers formed at both ends of the piston. However, in the swash plate compressor according to the third aspect of the present invention, by providing the auxiliary load means, it is possible to perform good control even in a state where the spool is displaced.

Further, in the swash plate compressor according to the fourth aspect of the present invention, a displacement load is applied in the direction where the inclination angle of the swash plate is minimized, in other words, the stroke of the piston is minimized and the discharge capacity of the compressor is minimized. An auxiliary displacement means is provided to give the That is, when the discharge capacity of the compressor is at its minimum, the inclination angle of the swash plate becomes difficult to decrease due to the influence of the fluid pressure difference within the working chambers formed at both ends of the piston. However, in the swash plate type compressor according to the fourth aspect of the present invention, the operation of the auxiliary displacement means reliably displaces the inclination angle of the swash plate, and the compressor can be continuously and satisfactorily moved to the position where the compressor has its minimum capacity. can be controlled.

[Example] An example of the present invention will be described below based on the drawings.
FIG. 1 is a longitudinal sectional view of a variable displacement swash plate compressor. The aluminum alloy front housing 4, front side plate 8, suction valve 9, front cylinder block 5, rear cylinder block 6, suction valve 12, rear side plate 11, and rear housing 13 are integrally fixed by through bolts (not shown). It forms the outer shell of the compressor. The cylinder blocks 5 and 6 each have five cylinders 64 (651 to 645) as shown in FIG.
However, the cylinders 64 are formed parallel to each other. A shaft 1 that rotates under the driving force of an automobile engine (not shown) is connected to a front housing 4 and a front cylinder block 5 via bearings 2 and 3, respectively.
It is rotatably supported on the shaft. Also, shaft 1
The thrust force (force acting in the left direction in the figure) applied to the shaft 1 is received by the front cylinder block 5 via the thrust bearing 15, and the movement of the shaft 1 in the right direction in the figure is restricted by the retaining ring 16. In addition, retaining ring 16
is locked by an annular groove formed in the shaft 1.

The rear shaft 40 is rotatably supported by the spool 30 via a bearing 14. The thrust force acting on the rear shaft 40 (the force acting in the right direction in the figure) is transferred to the spool 30 via the thrust bearing 116.
The retaining ring 17 prevents the rear shaft 40 from coming off the spool 30. This retaining ring is also locked in an annular groove formed in the rear shaft 40. The spool 30 is disposed within the cylindrical portion 65 of the rear cylinder block 6 and the cylindrical portion 135 of the rear housing 13 so as to be slidable in the axial direction.

A spherical part 107 is formed in the center of the swash plate 10 , and a spherical support part 405 fixed to the end of the rear shaft 40 is disposed on this spherical part 107 .
0 is supported by the spherical support portion 405 in a swingable state.

FIG. 3 is a perspective view of the shaft 1, the swash plate 10, and the rear shaft 40. As is clear from FIG. 3, a slit 105 is formed on the side surface of the shaft 1 of the swash plate 10.
A flat plate portion 165 is formed on the side end surface. By disposing the flat plate portion 165 in surface contact with the inner wall of the slit 105, the rotational driving force applied to the shaft 1 is transmitted to the swash plate 10.

Further, a shoe 18 and a shoe 19 are slidably disposed on both sides of the swash plate 10. On the other hand, a piston 7 is slidably disposed within the cylinder 64 of the front cylinder block 5 and the cylinder 64 of the rear cylinder block 6. As mentioned above, the shoes 18 and 19 are slidably attached to the swash plate 10. Further, the shoes 18 and 19 are rotatably engaged with the inner surface of the piston 7. Therefore, the rocking motion accompanied by the rotation of the swash plate 10 is as follows:
The reciprocating motion is transmitted to the piston via the shoes 18 and 19. The shoes 18 and 19 are formed so that their outer surfaces lie on the same spherical surface when assembled on the swash plate 10.

A long groove 166 is provided in the flat part 165 of the shaft 1.
The swash plate 10 also has pin holes 106 and 108 formed therein. Shaft 1
After the flat plate portion 165 is arranged in the slit 105 of the swash plate 10, it is locked in the pin through holes 106, 108 and the long groove 166 of the shaft 1 by the pin 80 and the retaining ring 81. The inclination of the swash plate changes depending on the position of the pin 80 in the long groove 166, and as the inclination changes, the position of the center of the swash plate (the spherical support portion 405 of the spherical portion 107) also changes. That is, in the second working chamber 60 on the right side in FIG. 1, even if the inclination of the swash plate 10 changes and the stroke of the piston 7 changes, the top dead center of the piston 7 on the working chamber 60 side hardly changes and remains a dead volume. The long groove 166 is provided so that there is substantially no increase in . On the other hand, in the working chamber 50 on the left side in the figure, the tilt of the swash plate changes and the top dead center of the piston 7 changes, so the dead volume also changes.

In this example, as described above, the long groove 166 is formed in such a shape that even if the inclination angle of the swash plate 10 changes, the top dead center position of the piston 7 on the working chamber 60 side does not change. Therefore, strictly speaking, the long groove 166 has a curved shape, but in actual formation, it can be approximated by a substantially straight long groove. Furthermore, in this example, the long grooves 166 are formed on the shaft 1 so that the shape of the flat plate portion 165 does not become excessively large due to the formation of the long grooves 166.
is located on the axis of In this way, the long groove 16
6 is formed on the axis of the shaft 1, and the flat plate portion 165
It is particularly effective to reduce the size of the swash plate type compressor in which the flat plate portion 165 is disposed inside the piston 7.

Reference numeral 21 in the figure is a shaft sealing device, which prevents refrigerant gas and lubricating oil from leaking to the outside through the shaft 1. Reference numeral 24 in the figure is a discharge port that opens into the working chambers 50, 60 and communicates with the discharge chambers 90, 93, and this discharge port 24 is opened and closed by a discharge valve 22. The discharge valve 22 and a valve holder 23 are fixed to the front side plate 8 and the rear side plate 11 with bolts (not shown). Reference numeral 25 in the figure indicates the working chambers 50, 60 and the suction chambers 72, 74.
A suction port that communicates with the suction valve 9 and the suction valve 12.
It is opened and closed by.

Reference numeral 400 in the figure is a solenoid valve for controlling the internal pressure of the control pressure space 200, and is controlled by a control circuit 500. FIG. 4 is a diagram showing the connection state of the solenoid valve 400. One end of the solenoid valve 400 is connected to the rear suction space 74 by a low pressure introduction passage 97. In addition, the other side is the throttle 99 and the high pressure introduction passage 9.
It is connected to the discharge space 93 via the control pressure passage 98 and to the control pressure chamber 200 via the control pressure passage 98 . When the solenoid valve 400 is not energized, it is marked 4 in the figure.
00a, and the discharge space 93 of the space 200 is connected via the aperture 99.

A discharge space 90 on the front side in FIG. 1 is led to a discharge port 92 by a discharge passage 91 formed in the cylinder block 5, and a discharge space 90 on the rear side
3 is a discharge passage 9 formed in the cylinder block 6
4 to the discharge port 95. Since the discharge port 92 and the discharge port 95 are connected by external piping, the internal pressures of the discharge space 90 and the discharge space 93 are the same pressure. Further, the front side suction space 72 is led to the suction space 70 formed in the center of the housing by the suction passage 71, and similarly the rear side suction space 74 is guided to the suction space 70 by the suction passage 73.
guided by. In addition, the symbols 51, 52, 5 in the figure
3, 54, 55, and 56 are O-rings.

The operation of the compressor with the above configuration will be described.
An electromagnetic clutch (not shown) is connected, and shaft 1
The compressor starts when the driving force from the engine is transmitted to the compressor.

When the compressor drive signal is input to the controller 500, the controller 500 outputs an electric signal to the control valve to connect the low pressure introduction passage 97 and the signal pressure passage 98. Therefore, at startup, the third
The control pressure chamber 20 switches to the state 400b in the figure.
0, the pressure inside the suction chamber 74 is introduced. Therefore, in this state, spool 3
There is no longer a pressure difference before and after zero. That is, at the time of startup, no load is applied in the direction of tilting the swash plate 10 via the support portion 107.

When the shaft 1 starts rotating in this state, the rotation of the shaft 1 causes the piston 7 to reciprocate via the swash plate 10. This piston 7
As the refrigerant moves back and forth, the refrigerant is sucked, compressed, and discharged within the working chambers 50 and 60.

However, in this case, a force based on the pressure difference between the rear working chamber 60 and the front working chamber 50 is applied to the swash plate 10 via the piston 7 and the shoes 18 and 19. In particular, the swash plate 10 has a spherical support portion 40
5, and receives the rotational force of the shaft 1 by fitting the slit 105 and the flat plate part 165, so that the force applied to the piston 7 is applied to the inclination of the swash plate 10. It will act as a moment in the direction of decreasing the angle.

For example, when the pin 80 is positioned on the axis X in FIG. 2, the piston disposed in the first cylinder space 641 does not generate a moment that changes the inclination angle with respect to the swash plate 10. However, the second to fifth cylinder spaces 642, 643,
From the piston 7 arranged at 644, 645,
A rotational moment is generated in a direction that reduces the inclination angle of the swash plate 10. This rotational moment Fi×Ri is
It is received by the moment Fpm×R generated around the pin 80 (as shown in FIG. 5). Further, the rotational moment generated by the piston 7 applies a pressing force of Fbx to the spherical support portion 405.

That is, when the control valve introduces suction pressure into the control pressure chamber 200, the spherical support portion 405 and the spool 30 are displaced to the right in the figure, as shown in FIG. As a result, the swash plate 10 reduces its angle of inclination. However, the swash plate 10 has a long groove 166 in the shaft 1.
Since the swash plate 1 is regulated by the pin 80,
0 reduces the inclination and applies force to the spherical surface 405 at the center of the swash plate 10 in the right direction in the figure, causing the spherical portion 405 to move to the right. Spherical support part 405
The force acting on the rear shaft 40 in the right direction in the figure is transmitted to the spool 30 via the thrust bearing 16, and the spool 30 moves until it hits the bottom of the rear housing 13. In this state, the discharge capacity of the compressor is at its minimum in the state shown in FIG.

The refrigerant gas sucked in from a suction port (not shown) (connected to the evaporator of the refrigeration cycle) enters the suction space 70 in the center, then passes through suction passages 71 and 73, and then passes through the suction chambers 72 and 72 on the front and rear sides. Enter 74. Thereafter, during the suction stroke of the piston 7, the air is sucked into the working chambers 50, 60 from the suction port 25 via the suction valve 12. The sucked refrigerant gas is compressed in a compression stroke, and when it is compressed to a predetermined pressure, it is discharged from the discharge port 24 into the discharge chambers 90 and 93 by pushing open the discharge valve 22 . High-pressure refrigerant gas passes through discharge passages 91 and 94 and discharge ports 92 and 9.
5 and is discharged to a condenser (not shown) of the refrigeration cycle.

At this time, since the first working chamber 50 on the front side has a large dead volume, the compression ratio is smaller than that of the second working chamber 60 on the rear side, and the pressure of the refrigerant gas in the first working chamber 50 is equal to the pressure in the discharge space 90. (the discharge output of the second working chamber 60 on the rear side is led), and the action of sucking and discharging refrigerant gas in the first working chamber 50 on the front side is not performed.

When starting the compressor, the compressor discharge capacity is set to the minimum capacity as described above. However, when the compressor capacity required by the refrigeration cycle is high, the control valve 400 is in the state 400a in FIG. 4. That is, the control valve 400 shuts off the control pressure passage 98 and the low pressure introduction passage 97. Here, in this example, the control pressure chamber 200 communicates with a high pressure introduction passage 96 via a throttle 99. Therefore, in this state where the connection with the low pressure introduction passage 97 is cut off, the influence of the discharge pressure received from the high pressure introduction passage 96 on the control pressure chamber 200 increases. Therefore, the control pressure chamber 200
The pressure inside is rising.

Therefore, the force acting on the spool 30 in the left direction in FIG. 6 due to the pressure difference (due to the pressure difference between the control pressure chamber 200 and the suction space 74) gradually increases as the compressor rotates. When this force overcomes the force pushing the spherical support portion 405 to the right in the figure, the spool 30 gradually begins to move to the left in the figure. And the long groove 166 of the shaft 1 and the pin 80
Due to this action, the swash plate 10 moves its center of rotation (spherical support portion 405) to the left in the figure and increases its inclination. As the pressure inside the control pressure chamber 200 further increases, the spool 30 moves to the left in the figure until its shoulder 305 touches the rear side plate 11, achieving the maximum capacity state. This is the first
This is the state shown in the figure. In the state shown in FIG. 1, refrigerant gas sucked through a suction port (not shown) enters the central suction space 70 and flows into suction chambers 72 and 74 through suction passages 71 and 73, respectively. In the suction stroke, the suction valves 9 and 1 are connected to the suction port 25.
2, enter the working chambers 50 and 60, respectively, and are then compressed with the displacement of the piston 7, enter the discharge spaces 90 and 93, respectively, through the discharge port 24 and the discharge valve 22, and the discharge passages 91 and 94.
The gas is discharged from the discharge ports 92 and 95, and is merged with the external piping. In this state, both the working chamber 50 and the working chamber 60 are performing the action of sucking and discharging refrigerant gas.

The solid line a in FIG. 7 is a diagram showing the relationship between the piston stroke and compressor capacity of the variable displacement swash plate compressor according to the present invention. The capacity control method according to this example uses the piston 7 by changing the inclination of the swash plate 10.
Since the stroke of the piston is changed and the center position of the swash plate 10 is also changed, there is almost no increase in the dead volume in the second working chamber 60 on the rear side due to a reduction in the piston stroke. Therefore, as shown by the dashed line b, the discharge capacity gradually decreases according to the piston stroke. Conversely, in the front-side first working chamber 50, the dead volume increases as the piston stroke decreases, and as the dead volume increases, the compression ratio decreases, and the discharge capacity rapidly decreases as shown by the broken line c in FIG. . Then, when the maximum pressure (discharge output) in the front side working chamber 50 becomes lower than the discharge pressure in the working chamber 60 (7th
At point d in the figure), the suction and discharge operations of the front side working chamber 50 are no longer performed, and the suction, compression, and discharge operations of refrigerant gas are performed only in the rear side working chamber 60. The piston stroke d at which the first working chamber 50 stops sucking and discharging refrigerant gas is the maximum piston stroke L, suction pressure Ps (Kg/cm ² · abs),
If discharge pressure Pd (Kg/cm ²・abs), adiabatic index k of refrigerant gas, piston radius R, and pi ratio π, then Ps・(πR ² L) ^k = Pd・{πR ²・(L−d) } From ^k, it is expressed as d=L・{1−(Ps/Pd) ^1/k }, and the capacity b in this case is b=1/2・d/L・100(%).

Here, Ps=3Kg/cm ²・abs, Pb=16Kg/cm ²・abs,
If K=1.14, then d=0.77L b=38.5(%).

Note that this piston stroke is based on spool 30.
If the state in which the spool 30 has gone all the way to the right in the figure is 0, and the state in which the spool 30 has gone all the way to the left in the figure is 1, then the spool will move as shown in Figure 7. The relationship between the amount of movement and the compressor capacity can be seen (L∝l).

Now, the solid line a in FIG. 7 is the capacity change characteristic of the compressor according to the present invention. In the movement distance l to e section of the spool 30, the capacity changes as shown by the solid line a, and as shown by the thin line f in the figure. For the amount of spool travel,
Although the compressor capacity changes linearly, the slope is steep and controllability is poor, but the spool displacement e~0
In the section, the capacity changes as shown by the solid line _a2 in the figure,
The slope is gentler than that of the thin line f, and the controllability is particularly excellent at low capacity.

Next, when the cooling load decreases, the control circuit 500 in FIG.
By alternately switching between the states 0a and 400b (for example, duty ratio control), the pressure inside the space 200 is gradually lowered from the discharge pressure, and the first
The capacity can be reduced by moving the spool 30 to the right in the figure. In order to obtain the minimum capacity, the solenoid valve 400 in FIG. 4 is set to the state 400b, the pressure inside the space 200 is communicated with the suction space 74, and if the suction pressure is set, the minimum capacity is set as shown in FIG. be able to.

As shown in Fig. 7, the compressor capacity is variably controlled according to the amount of displacement of the spool, but according to the inventors' experimental study, it is actually
It was observed that there were cases in which it was difficult to maintain the zero displacement at an appropriate position.

As shown in FIG. 8, if the back pressure applied to the spool 30 is gradually increased, the spool back pressure reaches the predetermined pressure F ₂
Until then, the spool is displaced as the back pressure increases, as shown by the solid line XY in the figure. In FIG. 7, the vertical axis indicates the displacement of the spool 30, and the displacement value of the spool corresponds to the amount of tilt angle displacement of the swash plate 10, and further corresponds to the reciprocating stroke of the piston 7. .

As shown in Fig. 8, it has been confirmed that when the back pressure of the spool 30 is increased to a predetermined value F2 _or more, the stroke does not displace continuously but immediately increases to the maximum stroke. (solid line YZ). That is, when the back pressure of the spool 30 is higher than the predetermined value _F2 , the sign is always held at the position where the stroke of the spool 30 is the maximum.

Conversely, when reducing the back pressure of the spool 30, reduce the maximum back pressure load _F3 to the predetermined load _F2 ,
Even if the load is further reduced to a smaller load _F1 , the spool 30 remains at its maximum displacement position (broken line ZK). Then, when the back pressure of the spool 30 becomes lower than a predetermined value _F1 on the low pressure side, the spool 30 suddenly shifts to a certain position (indicated by the broken line KL).

That is, even if it is attempted to continuously control the back pressure of the spool 30 as shown in FIG. 7, it is difficult to accurately maintain and control the actual displacement of the spool 30 near the maximum displacement position of the spool 30. Ta.

The inventors investigated the cause of this and found that it is because the relationship between the axial force of the shaft 1 applied to the spool 30 at each stroke of the spool 30 tends to be as shown in FIG. Ta. That is, the stroke of the spool 30 is increased from a state where the stroke of the spool 30 is the minimum, the inclination angle of the swash plate 10 is the minimum, and the amount of reciprocating movement of the piston 7 is the minimum (indicated by O in FIG. 9). As the stroke increases, the amount of reciprocating movement of the piston 7 increases, and the thrust force used to displace the spool 30 increases accordingly (as indicated by the solid line OP in FIG. 9). However, it is recognized that if the stroke of the spool 30 is to be increased further, the force required to displace the spool 30 will become smaller (No. 9
(Indicated by solid line PQ in the figure). The state indicated by the solid line PQ is a region in which the reciprocating stroke of the piston 7 is controlled to the maximum amount, or in other words, a state in which the discharge capacity of the compressor is slightly reduced from the maximum discharge capacity.

That is, as shown in FIG. 9, a maximum load F ₂ (point P) is observed between the stroke of the spool 30 and the thrust force required for its movement, and this maximum value
The stroke of the piston corresponding to F ₂ is P ₂ .
And the stroke P ₂ of this spool 30 is the 8th
This corresponds to point Y in the figure. As mentioned above, when the thrust force is made larger than the predetermined value _P2 , the spool 30 immediately moves forward to the maximum amount (point Q in Figure 9 and point Z in Figure 8), and this state holds the spool 30 at the maximum position. This will continue until the back pressure on the spool 30 decreases to a value less than the thrust force F ₁ required to achieve this. When the back pressure on the spool 30 becomes less than the thrust force _F1 , the spool 30
It will immediately be displaced from the point to the R point. The displacement of the spool 30 at this point R is P ₁ ,
This position corresponds to point L in FIG.

The reason for the characteristics shown in FIG. 9 is that in the swash plate compressor of this example, dead volume occurs only in the first working chamber 50 when the displacement of the spool 30 is small. be. This operation will be explained below using FIG. 10.

Figure 10 shows the stroke of the piston 7 and the working chamber 5.
The relationship with zero internal pressure, in other words, the relationship between the internal volume of the working chamber 50 and the internal pressure of the working chamber 50 is shown. The state shown by the solid line A in FIG. 10 is the state in which the piston 7 moves forward to its maximum stroke, that is, the maximum displacement state of the compressor. Furthermore, the dashed line B in FIG. 10 indicates a state in which the inclination angle of the swash plate 10 has decreased somewhat, and the amount by which the piston 7 can move forward has decreased accordingly. In the state shown by this dashed line B, a predetermined dead volume is therefore created between the piston 7 and the side plate 8. Furthermore, the broken line C in FIG. 10 shows a state in which the inclination angle of the swash plate 10 is further reduced and the dead volume is increased. Furthermore, the dashed double-dot line D in FIG.
The inclination angle of the piston 7 becomes the minimum, the reciprocating stroke amount of the piston 7 becomes the minimum, and the dead volume becomes the maximum.

First, a state in which the piston 7 is displaced to the maximum position (indicated by solid line A in the figure) will be described. As the piston 7 moves forward from the most retracted position (indicated by a in the figure), the capacity of the working chamber 50 decreases.
And the pressure inside the working chamber 50 increases (a-b in the figure).
-c). When the pressure inside the working chamber 50 reaches a predetermined pressure Pd, the discharge valve 24 opens and the working chamber 50
The pressure inside does not rise any further (c-d-e in the figure)
). After the piston 7 has been displaced to its maximum stroke (point e in the figure), the piston 7 begins to retreat, the suction port 25 opens, and the pressure inside the working chamber 50 immediately decreases to the suction pressure Ps (point f in the figure). ), then the piston returns again to the rear end position (indicated by a in the figure). That is, when the piston is at maximum displacement, d in the working chamber 50 is a, c, e, f,
Pressure fluctuation will be performed in cycle a.

When the angle of inclination of the swash plate 10 becomes somewhat smaller and a dead volume is created at the tip of the piston 7, a predetermined volume will be maintained in the working chamber 50, as shown by the dashed line B in FIG. . Therefore, even if the piston 7 retreats from this state, the refrigerant held in the working chamber 50 will expand again (indicated by the dashed line d-g in the figure), and during that time the working chamber 50 will have a suction pressure higher than Ps. pressure will be maintained.

As the inclination angle of the swash plate 10 becomes further smaller, the stroke amount of the piston 7 decreases, and a large dead volume is created in the working chamber 50, the discharge valve 24 will eventually not open when the piston 7 moves forward. . That is, the pressure within the working chamber 50 when the piston 7 moves forward does not exceed the discharge pressure Pd. This state is shown by the broken line C in FIG. 10, and in this case, the pressure and volume within the working chamber 50 repeat the operations a-b-c-ba in the figure. If the inclination angle of the swash plate 10 becomes even smaller and the movement stroke of the piston 7 becomes even smaller, the state as shown by the dashed line D in FIG. 10 is finally reached. In this case, suction and discharge are not performed within the working chamber 50, and the relationship between the volume and pressure of the working chamber 50 is ab-a.

As explained above, due to the dead volume in the working chamber 50, the pressure within the working chamber 50 changes during the reciprocating cycle of the piston.

FIG. 11 is a graph showing the relationship between the pressure within the working chamber 50 and the reciprocating period of the piston 7. Solid line A in the figure corresponds to the state of solid line A in FIG.
In this state, no dead volume is generated at the tip of the piston 7, and when the piston 7 begins to retreat, the pressure within the working chamber 50 immediately drops to the suction pressure Ps.
Furthermore, the dashed dot line B in FIG. 11 corresponds to the dashed dot line B in FIG. That is, even when the piston 7 performs a backward movement, the pressure within the working chamber 50 does not immediately decrease to the suction pressure, but gradually decreases from the discharge pressure Pd toward the suction pressure Ps. Furthermore, the broken line C in FIG. 11 corresponds to the broken line C in FIG. 10, but when the dead volume increases to this state, the pressure fluctuation in the working chamber 50 becomes almost sinusoidal, and the operation is performed below the suction pressure Ps. Room 50
The internal pressure will no longer decrease.

Furthermore, the dashed double-dot line D in FIG. 11 corresponds to the dashed double-dashed line D in FIG. do not have. Further, in the state shown by the two-dot chain line D, the pressure fluctuation within the working chamber 50 is reduced, and the maximum pressure within the working chamber 50 is reduced.

The region indicated by PQ in FIG. 9 indicates the region in which the pressure-volume state within the cycle ranges from solid line A to broken line C in FIG. 10. That is, in this region, as is clear from FIG.
By staying in the inner position, the force of the pressure in the first working chamber 50 that urges the piston 7 in the right direction in FIG. 1 increases.

Here, the force of the pressure in the first working chamber 50 pressing the piston 7 in the right direction is
The operation is in the direction of increasing the inclination angle. That is, due to the pressure remaining in the working chamber 50, the swash plate 1
The inclination angle at zero becomes larger, and the reciprocating stroke amount of the piston 7 increases. The behavior during this period is a region indicated by a solid line PQ in FIG. 9, and in this region, as the dead volume increases, the pressure remaining in the working chamber 50 increases. Therefore, in this region, as the dead volume increases, the thrust force required to press the spool 30 to the left in FIG. 1 increases.

Note that the explanation of FIGS. 9 to 11 above has been made with the suction pressure Ps and the discharge pressure Pd constant, respectively. However, when the compressor is used for compressing refrigerant in the refrigeration cycle, The suction pressure Ps and the discharge pressure Pd will vary depending on the required operating conditions. For example, in the low-load operating state of a refrigeration cycle, the suction pressure
Ps is 2.5Kg/cm ² abs and discharge pressure Pd is about 16Kg/cm ² abs. When the heat load of the refrigeration cycle is large, the suction pressure Ps increases to about 4 kg/cm ² abs, and the discharge pressure Pd also increases to about 26 kg/cm ² abs. In this way, suction Ps and discharge pressure Ps change,
Accordingly, the compression ratio ε of the compressor will also change. FIG. 12 shows the thrust load required to displace the spool 30 of the compressor in the axial direction when the discharge pressure Pd varies in this way. As shown in FIG. 12, the thrust load increases as the discharge pressure increases. In particular, as shown in FIG. 12, the thrust load of the spool 30 fluctuates greatly at the initial stage when a dead volume begins to form at the tip of the piston 7. This is because, as described above, the pressure caused by the dead volume acts in a direction to push back the spool 30 via the piston 7 and the swash plate 10. That is, when the discharge pressure is high, the pressure within the working chamber 50 due to the dead volume increases accordingly, and as a result, the spool 3
The thrust load required to displace 0 in the axial direction also increases. As is clear from FIG. 12, when the dead volume becomes larger than a predetermined value, the influence of the discharge pressure no longer remains in the working chamber 50. Therefore, the thrust load required to further displace the spool 30 in the axial direction when the spool 30 is displaced by a predetermined value or more is always maintained at a constant pressure regardless of fluctuations in the discharge pressure. Therefore, the relationship between the spool displacement amount and the compressor capacity as shown in Fig. 7 will also change depending on the displacement of the compressor suction pressure Ps and discharge pressure Pd as shown in Fig. 13. . In Fig. 13, the solid line indicates a steady compression state with a compression ratio of 5, and the broken line indicates a compression ratio of 4.
The dashed line indicates a low load state where the compression ratio is 6, and the dashed line indicates a high load state where the compression ratio is 6.

As mentioned above, the thrust load that displaces the spool 30 in the axial direction is caused by the pressure in the first working chamber 50 when the stroke of the spool 30 is displaced from the P ₂ state (FIG. 9) to the maximum position. As the amount of displacement of the spool 30 becomes maximum due to the influence,
That is, as the spool 30 is displaced to the left in FIG. 1, the load required for that displacement becomes smaller. Therefore, the relationship between the stroke of the spool 30 and the thrust force required to displace the spool 30 in the axial direction is non-linear as shown in FIG. In such a state, the capacity of the compressor cannot be accurately controlled only by controlling the pressure within the control pressure chamber 200. In order to control the discharge capacity of the compressor continuously, follow the dotted line in Figure 9.
It is necessary to have the characteristics shown in PS. Therefore,
Auxiliary loading means, such as a return spring 900 that returns the spool 30 to the lower capacity side, is used. That is, this return spring 900 improves the reverse slope characteristic shown in FIG.

The region where the return spring 900 acts is the region where the stroke of the spool 30 reaches its maximum from position _P2 in FIG. 9 where the thrust force reaches its maximum value. Return spring 900
The spring constant shall be at least sufficient to compensate for the reverse slope of the thrust load.

As a compressor, for example, in a compressor as shown in FIG. 14, the stroke of the spool 30 at the position where the angle of inclination of the swash plate 10 is maximum is 0 mm, and the maximum stroke of the spool 30 where the angle of inclination of the swash plate is the minimum is set. The length shall be 10mm. In this maximum stroke state, the reciprocating stroke amount of the piston 7 is 20
mm. And the maximum capacity of this compressor is 180c.c.
In this case, the suction pressure Ps is 3Kg/cm ² abs, and the discharge pressure is
If it is about 12Kg/cm ² abs to 18Kg/cm ² abs, the above-mentioned reversal region shown in FIG. 12 will occur at a position where the stroke of the spool 30 is 7 mm or more. Therefore, in the embodiment shown in FIG. 14, a return spring 900 is installed which generates a compressive load when the stroke of the spool 30 becomes 7 mm or more. This return spring has a spring constant of
For example, let's say 33Kg/mm.

If the return spring 900 is provided in this way, the stroke of the spool 30 from about 0 mm to about 7 mm is the region OP in FIG.
As the back pressure increases, the stroke amount of the spool 30 increases. Furthermore, in the range where the stroke of the spool 30 is 7 mm or more, the spool 30 will not be displaced to the right in FIG. 14 unless a thrust load greater than the set load of the return spring 900 is applied to the back surface of the spool 30 by the operation of the return spring 900. become. In other words, this return spring 9
00 eliminates the reversibility of the displacement load. Note that FIG. 15 shows a state in which the spool 30 has been displaced by 7 mm or more and the return spring 900 has started to be compressed.

That is, in the compressor of this example, the working chamber 50
Spool 3 generated by the dead volume of
The reversibility of zero stroke displacement and thrust load is eliminated by using the compression force of the return spring.

14 and 15, the return spring 900 is a coil spring and is disposed between the end of the shaft 1 and the locking plate 901 disposed at the end of the rear shaft 40. Here, displacement of the shaft 1 in its axial direction is restricted by a thrust bearing 15. Further, the rear shaft 40 is slidably disposed on the shaft 1 and is configured to be displaced integrally with the spool 30 in the axial direction via a thrust bearing 116. Therefore, when the spool 30 is displaced to the left in FIG. 14, the displacement of the spool 30 is transmitted to the spherical support portion 107 and the rear shaft 40 via the thrust bearing 116. Under this displacement,
If the locking plate 901 fixed to the rear shaft 40 is also displaced and the stroke of the locking plate becomes 7 mm or more,
The end of the return spring abuts the shaft 1. Therefore, when the stroke of the spool 30 becomes 7 mm or more, the pressing load of the return spring 900 is applied to the spool 3.
0 (as shown in FIG. 15).

In the embodiment shown in FIGS. 14 and 15, the shaft 1 passes through the swash plate 10 and connects to the rear housing 6.
The shaft 1 is supported from both sides by a rear bearing 14 and a front bearing 3. Therefore, this Figure 14 and Figure 1
In the embodiment shown in Figure 5, the shaft is supported on both sides so that it can be smoothly rotated.
Furthermore, in the embodiment shown in FIGS. 14 and 15, the pin 80 is supported in the pin holes 106 and 108 of the slit 105 via a bearing 909. Therefore, when the swash plate 10 changes its inclination angle in response to the displacement of the spool 30, the frictional force around the pin 80 is extremely small. As a result, the displacement of the spool 30 is smoothly transmitted to the change in the inclination angle of the swash plate 10, and the continuous change in the compressor discharge amount is better exhibited.

In the embodiment shown in FIGS. 14 and 15, the return spring 90 is disposed at the rear end of the shaft 1, but this return spring 900 is designed to absorb a compressive load when the spool 30 is displaced leftward by a predetermined value or more. It can be placed in any position as long as it can be used effectively.

FIG. 16 shows another example of the arrangement position of the return spring arranged as the auxiliary load means. As shown at 910 in the figure, the return spring is attached to the spool 30.
and the rear end plate 11. Further, as shown at 911 in the figure, a return spring may be interposed between the spool 30 and the rear housing 6. Furthermore, a return spring may be disposed between the support portion 107 and the flat plate portion 165 formed on the shaft 1, as indicated by 912 in the figure.

Further, in the above example, a return spring was provided as an auxiliary load means to improve the reverse slope characteristic of the spool, but it is of course possible to use other means other than the return spring. For example, instead of a spring, other means may be used. It is also possible to use pressure or the like. Furthermore, even when a spring is used as the auxiliary load means, a nonlinear spring may be used as the spring.

As explained above, since the swash plate compressor according to this example uses an auxiliary load means, even when the stroke amount of the spool 30 becomes large, the pressure in the control pressure chamber can be continuously maintained. By controlling this, the stroke amount of the spool 30 can be well controlled.

However, according to experimental studies conducted by the present inventors, it has been confirmed that in a region where the stroke of the spool 30 is small, there are situations in which the spool does not displace satisfactorily even if the pressure in the control pressure chamber 200 is lowered.

In this case, as shown in Fig. 12, if the stroke ratio of the spool 30 is 0.3, that is, the stroke amount of the spool 30 is 10 mm, the displacement will be 0 mm to 3 mm.
This is because the thrust load required to press the spool 30 to the left in FIG. 1 becomes negative in the region up to this point.

When the amount of displacement of this spool is near the minimum value,
As shown by the two-dot broken line D in Fig. 11, the first working chamber 5
The pressure remaining within 0 is decreasing. On the other hand, since no dead space is generated on the second working chamber 60 side, the pressure fluctuates between the suction pressure and the discharge pressure in accordance with the reciprocating movement of the piston 7. Therefore,
Even if this stroke is small, the pressure within the second working chamber 60 rises to the discharge pressure Pd.

Next, the thrust load when the stroke of the spool 30 is small will be explained in more detail. FIG. 17 shows the load applied to the spool 30 in this state. In the diagram
FPSi indicates the total force of pushing the swash plate 10 due to the compression of the piston 7. Further, FPN in the figure indicates the force applied to the pin 80. Furthermore, Ps indicates the pressure within the suction chamber 74, and Pc indicates the pressure within the control pressure chamber 200. Furthermore, the thrust loads generated by these pressures Ps and Pc are expressed as FPS and FPC. Furthermore, in the figure
FB, FS, and FO represent the friction force between the spherical support portion 107 and the swash plate 10, the friction force between the shaft 1 and the rear shaft 40, and the friction force between the spool 30 and the housing inner surface 135, respectively. This first
As is clear from Figure 7, in order to move the spool 30 from a state where it is operating at a certain capacity in a direction where the amount of displacement further decreases, in other words, to the right in the figure, FPSi + FPS > FPN + FPC + FB + FS + FO must be established. No.

However, when the discharge capacity of the compressor is small, the pressure on the first working chamber 50 side does not increase much, and on the other hand, the pressure on the second working chamber 60 side can rise to the discharge pressure, so FPSi has a small value. Therefore, the above-mentioned inequality does not hold, and the capacity cannot be lowered.

In order to make this inequality hold, it is possible to reduce the value of the first term (FPN) on the right side of the equation, or to increase the value of the first term (FPSi) on the left side of the equation.

However, FPN is caused by the pressure difference between the suction chamber 74 and the control pressure chamber 200,
It is structurally difficult to reduce this value. That is, it is difficult to reduce the pressure Pc in the control pressure chamber 200 to less than the suction pressure Ps. Furthermore, FB, FS,
Each FO is a frictional force, and there is a limit to reducing the friction coefficient of each part.

Therefore, the present inventors focused on FPSi, and by increasing this value, it was possible to reliably displace the spool 30 to the minimum position.
Here, in order to increase the value of FPSi, there are a method of increasing the pressure within the first working chamber 50 and a method of decreasing the pressure within the second working chamber 60.

Therefore, the inventors of the present invention first formed a boosting passage means that communicates the first working chamber 50 with the outside as an auxiliary displacement means for bringing the spool to the minimum position. FIG. 18 shows a boosting hole 9 as this boosting passage means.
An example in which 50 is formed on the discharge valve 22 is shown. In this example, the pressurizing hole 950 is a minute opening with a diameter of about 0.2 mm, so that the refrigerant at the discharge pressure is returned to the first working chamber 50 side by a minute flow rate Q ₀ . Due to the recirculation of the refrigerant with this discharge pressure increased, the first working chamber 50
The pressure inside increases.

FIG. 19 shows the average value of the internal pressure of the first working chamber 50 with respect to the opening area of the pressure increase hole 950. This first
As shown in FIG. 9, if the opening area of the pressure increase hole 950 is increased, the pressure within the first working chamber 50 will also increase. Therefore, in order to ensure the average pressure necessary to obtain the FPTi necessary to satisfy the above-mentioned inequality, the opening area should be set to A ₀ or more.

The measurement results of the compressor capacity C with respect to the differential pressure ΔP across the spool 30 in the case where a pressure increase hole having an opening area of A ₀ or more is provided as described above is shown by a solid line H in FIG. In addition, the differential pressure ΔP before and after the spool 30, etc. is determined by the suction pressure chamber 73,
Suction pressure Ps in 74 and control pressure in control pressure chamber 200
Indicates the differential pressure with PC. As shown in this Figure 20,
It was confirmed that when the pressure increase hole 950 is provided, by reducing the differential pressure across the spool 30, the spool can be reliably displaced to the minimum position and the discharge capacity of the compressor can be minimized. In addition,
What is shown by the broken line I in FIG. 20 is the relationship between the differential pressure across the spool 30 and the compressor capacity C in a compressor without the booster hole 950.

Note that the above-mentioned pressurizing hole 95 is used as the auxiliary displacement means.
In addition to 0, other pressure increasing passage means may be used. As shown in FIGS. 21 and 22, a pressure increasing groove 951 bypasses the end plate 82 and the discharge valve 22.
may be provided. When the pressure increase groove 951 is formed in this way, there is an advantage that foreign matter is less likely to clog the pressure increase hole 950.

FIG. 23 shows a communication passage 952 as an auxiliary displacement means.
This is an example in which a One end 953 of this communication passage 952 opens into the discharge chamber 90, and the other end 954 opens into the first working chamber 95 near the top dead center (the position indicated by the broken line in the figure) of the piston 7 in the minimum capacity state.
It opens at 0. By providing such a communication passage 952, it is also possible to introduce high pressure from the discharge chamber 90 in a state where the pressure in the first working chamber 50 needs to be increased. Moreover, with this communication passage 952, the opening end 954 is covered by the side surface of the piston 7 during normal compression operation, so that the passage can have a relatively large diameter.

In the above-described embodiment, a pressure increasing means for mainly introducing the high pressure of the discharge chamber 90 into the first working chamber 50 was provided as the auxiliary displacement means, but other methods may be used. Second
FIG. 4 shows another example of the auxiliary displacement means. In this illustration in FIG. 24, the first working chamber 60 of the piston 7
A dead space is also provided on the side. That is, as shown in FIGS. 38 and 39, the apparent hinge point 1001 of the swash plate 10
Shaft 1 from the center point position of shafts 18 and 19
When the swash plate 10 is displaced to the side and the angle of inclination of the swash plate 10 becomes small, the center position of the shoes 18 and 19 is at the 24th position.
It is designed to be displaced by a minute amount to the left in the figure.

As shown in FIG. 11 above, if such a small dead volume is created, the average pressure within the second working chamber 60 increases.

Therefore, if a dead volume is generated on the second working chamber 60 side at a position where the stroke of the spool 30 is small as shown in the example shown in FIG. is pressed in a direction that reduces the angle of inclination. FIG. 25 shows this state. Fig. 25 solid line J indicates the apparent hinge point 1001 and the points 18, 19.
A state in which the distance Δ to the center is 0 mm is shown. That is, this solid line J indicates a state in which no dead space is generated on the second working chamber 60 side. In addition, the distance Δ of the dashed line K, one-dot chain line L, and two-dot chain line M is 2.5 mm, respectively.
5mm and 7.5mm. As is clear from the broken line K, one-dot chain line L, and two-dot chain line M, it is recognized that the load F increases as the dead space increases. If the load F increases in this way, the swash plate 10 due to the compression of the piston 7,
The total holding force FPSi increases, and as a result,
The above inequality holds and ensures that the compressor can be reduced to its minimum capacity.

26 and 27 show the capacity change state of the compressor in the compressor provided with the auxiliary displacement means related to the embodiment shown in FIG. 24. FIG. 26 shows a state in which the discharge capacity of the compressor is at its minimum. As is clear from this figure, in this state the inclination angle θ of the swash plate 10 is at its minimum, and the stroke S of the piston 7 is
is also the minimum. Therefore, in the state shown in FIG. 26, the pressure within the first working chamber 50 does not increase beyond the discharge pressure, and the discharge valve is always closed. On the other hand, FIG. 27 shows a state in which the discharge capacity of the compressor is at its maximum.

In the embodiment described above, by changing the apparent hinge point 1001, the second working chamber 6
Although a dead space is formed at the apparent hinge point 10 to obtain the minimum capacity, the apparent hinge point 10
01 may be changed to mechanically reduce the inclination angle of the swash plate 10.

This point will be explained below. As shown in FIG. 38, the center position of the piston 7 at the rear top dead center is an apparent hinge point 1001 of the swash plate 1.
When the zero inclination angle is changed, the high pressure generated in the second working chamber 60 hardly contributes to the moment in the direction of reducing the inclination angle of the swash plate 10.

Therefore, as shown in FIG. 39, if the apparent hinge point 1001 is shifted in the direction of the shaft 1,
Moment in the direction that reduces the inclination angle of the swash plate 10
Mr is generated, and this moment increases the force FPSi in the direction of minimum capacity. A detailed explanation of the load relationship in this case will be given based on FIG. 40.

Here, N is the number of pistons 7, FPS is the axial resultant force of the load and inertia force due to piston compression,
ARM (θ, i) is the arm length from the spherical support 405 to the center of the piston, and POPN is the spherical support 405.
Arm length from center to pin 80, ψ is long groove 166
The inclination angle of FPM, the reaction force generated on pin 80,
FSH is the axial component of pin 80 reaction force (FPN), β
(θ) is the angle MD made by the line connecting the center of the spherical support and the pin 80 with the center axis of the shaft 1, and the moment due to the centrifugal force of the swash plate 10, and the FPC is the spool 30.
If the pressing load is , the moment Mr around the center of the spherical support is _o 〓 〓 ⁱ⁼¹ {FPS/cosθARN(θ・i)}+FPN/cos{ψ−β(
θ)}POPN+Mr=0.

Also, the balance of the axial load is: _o 〓 ⁱ⁼¹ FPS+FSH+FPC=0 Here, FSH is FSH=FPN/cos {ψ−β(θ)}sinψ.

In other words, if the loads associated with the compression of the piston 7 are equal, the reaction forces FPN generated on the pin 80 are equal;
Therefore, the axial component FSH of the pin 80 reaction force increases as the inclination angle of the long groove 80 increases. Therefore, the spool load FPC also increases.

FIG. 41 schematically shows the load on each part of the long groove 80 and the pin 81 when the angle ψ of the long groove 80 is changed at the minimum capacity. Here, ψ ₁ , ψ ₂ (ψ ₂
is larger than ψ ₁ ), the axial component FSH of the reaction force of the pin 80 is FSH ₁ and FSH ₂ is larger than FSH ₁ , respectively. Therefore, it can be said that increasing the angle ψ of the long groove 80 is an effective means for ensuring the minimum capacity.

However, this effect is not limited to the minimum capacity.
It appears even at maximum capacity. That is,
As shown in Figure 42, even at maximum capacity,
The larger the angle ψ of the long groove 80, the more FSH ₂ '
This results in the effect that FSH becomes larger than ₁ '. This shows that if the long groove 80 is made almost direct and the offset amount of the apparent hinge point 1001 is constant, the characteristics will be as shown by the solid line K in FIG. 43. That is, as is clear from FIG. 43, the characteristics of the state in which the apparent hinge point 1001 is at the center position of the piston 7 (indicated by the broken line L) are simply raised as a whole.
In other words, the spool load required when the piston stroke is maximized becomes large. Here, since the pressure Pc supplied to the control chamber 200 is at most the same as the discharge pressure Pd,
Increasing the diameter of the spool 30 requires other measures.

Therefore, in this example, in order to solve this problem, in a region where the capacity of the compressor is relatively small, in other words, when the stroke of the spool 30 is relatively small,
The amount of offset of the apparent hinge point 30 is increased, and conversely, when the compressor discharge capacity is large, the offset is decreased. To achieve this, the long groove 16
6 is an arcuate shape as shown in FIG.
That is, the shape of the long groove 166 is determined so that the angle of inclination between the tangent to the long groove 166 and the axis of the shaft is ψ ₁ at maximum capacity, and ψ ₂ at minimum capacity.

FIG. 45 shows the effect when the long groove 166 is shaped like an arc in this way. Fig. 45 Solid line M
1 shows the long groove 166 having the above-mentioned arc shape. In addition, the dashed line K indicates that the long groove 166 is straight,
In addition, the apparent hinge point 1001 is shown offset by 5 mm. Furthermore, the broken line L indicates the long groove 16.
6 shows a straight groove and no offset. As is clear from this FIG. 45, the long groove 16
In the case where 6 is an arcuate groove, it was confirmed that the minimum capacity can be secured at a spool load of about 40 kg without changing the maximum spool load. Furthermore, as shown in FIG. 46, by making the shape of the long groove 166 into the above-mentioned arc shape, the minimum capacity of the compressor could be reliably achieved. In addition,
A solid line M in FIG. 46 indicates that the long groove is arcuate, and a broken line L indicates that the long groove 166 is linear and the apparent hinge point 1001 is located on the center of the piston 7.

In the embodiment described above, the load pressing the swash plate 10 is changed mechanically as an auxiliary displacement means, but the discharge of the compressor can also be changed by further lowering the pressure on the second working chamber 60 side. A state in which the capacity is minimized is obtained. Here, as a method for lowering the pressure on the second working chamber 60 side, it is possible to consider means for lowering the suction pressure and means for lowering the discharge pressure, but a method for lowering the suction pressure on the second working chamber 60 side is practical. It is proposed to throttle the intake refrigerant passage.

Figure 47 shows that the suction pressure on the second working chamber 60 side is 1
The effect on the spool pressing load when lowering Kg・^cm2 is shown. In FIG. 47, the solid line N indicates the second working chamber 60.
This shows the state where the suction is throttled on the side, and the broken line O is the second
This shows the state where no suction restriction is applied to the working chamber. Also,
The solid line P and the broken line Q indicate the load received from the first working chamber side. The solid line R and the broken line S indicate the total pressing load applied to the spool 30. A solid line R indicates a state where suction throttling is performed on the second working chamber 60 side, and a broken line S represents a state where suction throttling is not performed on the second working chamber 60 side. As is clear from FIG. 47, by using the suction throttle, the load received from the second working chamber 60 side is reduced;
Since the load received from the working chamber side does not change, the total load increases as a result, making it possible to secure the minimum capacity.

FIG. 48 shows an example of actually measuring the effect of reducing the suction pressure Ps on the compressor rotational speed when changing the aperture area at the minimum capacity. As is clear from FIG. 48, it can be confirmed that even if the capacity of the compressor is small and therefore the flow rate is small, sufficient effects can be obtained by selecting the aperture opening area.

FIG. 49 shows a configuration in which a throttle 2000 is disposed between the suction pressure chamber 70 and the rear suction chamber 74. Here, the refrigerant follows the path indicated by the arrow in the figure and is sucked into the second working chamber 60 from the suction chamber 74 via the suction passage 73. And at this time, aperture 200
0, the pressure inside the suction chamber 74 becomes lower than the suction pressure Ps without the throttle 2000, and the above-mentioned effect can be obtained.

However, with such a fixed throttle, the suction refrigerant is throttled even when the compressor capacity is large, which not only reduces compressor performance but also increases the spool pressing load, making it difficult to maintain maximum capacity. There is a problem in that the spool diameter must be increased in order to achieve this.

Therefore, as shown by the solid line V in Fig. 50, when the discharge capacity of the compressor is large, the suction passage is not throttled.
The above-mentioned problem is avoided by using a variable throttle that narrows the suction passage to a large extent only when the discharge capacity is small.
Note that the dashed line T in FIG. 50 indicates the second working chamber 60.
This shows a state in which a fixed throttle 2000 is arranged in the side suction passage, and a broken line U shows a state in which there is no throttle.

51 and 52 show a variable embodiment in which a variable throttle is disposed in the middle of the suction passage 73. FIG. FIG. 51 shows a state of maximum capacity, in which the spool 30 is maximally displaced to the left in the figure.
Here, in this example, a suction passage 200 is provided in the spool 30.
1 is formed, and this suction passage 2001
is formed in a tapered shape, so this 51st
In the illustrated state, there is a sufficiently large opening area between the suction passage formed in the spool 30 and the housing, and the suction passages 73 and 2001
The flow of refrigerant flowing through is not restricted.

On the other hand, FIG. 52 shows a state in which the compressor capacity has become smaller. As is clear from FIG. 52, the suction passage 2001 formed in the spool is
The opening area between the housing and the housing becomes smaller.
Therefore, the refrigerant passing through the suction passages 73 and 2001 is greatly throttled when the spool 30 passes through the suction passage 2001, and suction throttling is performed.

FIG. 53 is a graph showing the characteristics when suction throttling is performed based on the embodiment shown in FIGS. 51 and 52. FIG. Here, the broken line W in the figure shows a state without the suction throttle, and the solid line Z shows an example with the suction throttle according to this embodiment. As is clear from FIG. 53, in the example without a throttle, the control pressure chamber 20
The minimum capacity cannot be achieved unless the pressure Pc in the suction chamber 74 is lower than the pressure Ps in the suction chamber 74.
It has been confirmed that the compressor of this example achieves the minimum capacity when the internal pressure Pc of the control pressure chamber 200 is greater than the internal pressure Ps of the suction chamber 74.

In addition to the above-described embodiment shown in FIGS. 51 and 52, there are various other forms of the variable diaphragm. In FIG. 54, a suction passage 73 is formed on the inner surface of the cylindrical portion 65 of the housing, and a suction passage 2002 is formed on the outer surface of the spool 30. The end of the suction passage 2002 and the end of the suction passage 73 are formed at positions that narrow the passage when the spool 30 is displaced to the minimum capacity position.

In the embodiment described above, attention has been focused on the swash plate 10 and the holding force FPSi (shown in FIG.
(Illustrated in the figure) may also be used.

For this purpose, a pressure higher than the discharge pressure Pd may be introduced as the signal pressure supplied to the control chamber 200 internal pressure Pc, or a pressure lower than the suction pressure Pc may be introduced to the control chamber pressure Pc. Therefore,
As a means for creating these pressures, the present inventors focused on pressure fluctuations within the working chamber 60 as shown in FIG. 55. The state shown in FIG. 55 shows actual pressure fluctuations within the working chamber. In other words, while the state shown in FIG. 11 represents an idealized relationship between pressure and volume within the working chamber, the state shown in FIG. 55 does not represent pressure fluctuations that actually occur within the working chamber. be.

That is, in an actual working chamber, due to the influence of the inertia of the discharge valve 22 and the suction valve 9, an overshoot of ΔPd occurs during compression, while an undershoot of ΔPs exists during suction. Therefore, by extracting these pressures ΔPd and ΔPs and using them as the control chamber pressure Pc, the range of control pressure can be expanded.

FIGS. 56 and 57 show an embodiment using overshoots ΔPd and ΔPs as this auxiliary displacement means. That is, as shown in FIGS. 56 and 57, a high pressure chamber 3000 and a low pressure chamber 3001, which are closed spaces, are formed in the rear housing 13 independently of the discharge chamber 93 and the suction chamber 74.
The high pressure chamber, the low pressure chamber, and the second working chamber 6
0, the high pressure take-off valve 3002 and the low-pressure take-off valve 30
03 to communicate with each other. Here, the high pressure takeoff valve 3002 and the low pressure takeoff valve 300
3 are made of materials with lower rigidity than the discharge valve 22 and the suction valve 9, respectively. In the suction stroke, when the internal pressure P of the second working chamber 60 drops to around the suction pressure Ps, the low pressure take-off valve first Only valve 3003 opens. Based on the difference in rigidity, the suction valve 9 will then open. Therefore, when only the low pressure take-off valve 3003 is open and the suction valve 9 is not yet opened, the pressure within the low pressure chamber 3001 is lower than the suction pressure Ps by the amount of undershoot ΔPs. Then, when the suction valve 9 opens, the pressure inside the second working chamber 60 changes to the suction pressure.
In order to return to Ps, the low pressure take-out valve 3003 closes as the pressure within the second working chamber 60 increases. Therefore, the low pressure chamber 3001 is always maintained at a pressure lower than the suction pressure Ps by the amount of undershoot ΔPs.

Next, when the piston 7 enters the compression stroke, the high pressure take-off valve 300 is lower than the discharge valve 22 due to the difference in rigidity.
2 opens first. Therefore, when the pressure in the second working chamber 60 rises further by the amount of overshoot ΔPd than the discharge pressure Pd, that high pressure can be introduced into the high pressure chamber 3000. When the discharge valve 22 opens after the high pressure take-off valve 3002 opens, the second working chamber 60
The pressure inside returns to the discharge pressure Pd. That is, the overshoot ΔPd is eliminated, and the pressure within the second working chamber 60 is relatively reduced. In this case, the high pressure take-off valve 3002 is immediately closed, so that a pressure higher than the discharge pressure Pd by the amount of overshoot ΔPd is always maintained in the high pressure chamber 3000.

Therefore, the high pressure chamber 3000 is communicated with the control valve 400 via the high pressure introduction passage 96, and the low pressure chamber 3001
By communicating with the control valve 400 via the low pressure introduction passage 97, a larger force can be exerted to drive the spool 30. In particular, low pressure chamber 3
In the state where the low pressure in 01 is introduced into the control pressure chamber 200, the function as an auxiliary displacement means is well exhibited, and the compressor can be reliably changed to the minimum capacity.

FIG. 58 shows an example of actual measurement of overshoot ΔPd and undershoot ΔPs obtained when the rotational speed of the compressor was changed. As is clear from FIG. 58, if overshoot ΔPd and undershoot ΔPs are used, the spool 30
The control pressure Pc _nax required to displace the spool 30 to the maximum position and the control chamber pressure Pc _nio necessary to displace the spool 30 to the minimum position can be satisfactorily achieved.

Figure 28 shows the effect of this example compressor.
This is a comparison with a compressor such as that described in Japanese Patent No. 175783 in which a piston is disposed only on one side of a swash plate. In the figure, solid lines A and B indicate the drive torque of the compressor of this example, and broken lines C and D indicate the drive torque of a type of compressor in which the piston is disposed only on one side of the swash plate. Further, solid line A and broken line C show torque fluctuations when the compressor has the maximum capacity, and solid line B and broken line D show torque fluctuations when the compressor has the minimum capacity. Note that the horizontal axis in FIG. 28 is the shaft rotation angle.

As shown in FIG. 28, the drive torque fluctuations in the compressor of this example and in the conventional compressor are almost the same when the compressor is at its minimum capacity. This is because, even in the compressor of this example, only the second working chamber 60 on one side of the piston operates, and the first working chamber 50 on the other side does not operate when the compressor has the minimum capacity.

However, when the compressor is operating at maximum capacity, its driving torque is significantly different.
That is, in the compressor related to this example, the working chamber 50,
60 are formed on both sides of the piston 7, the driving torque fluctuation per shaft rotation becomes extremely small. On the other hand, in conventional compressors, since the piston is actuated by only one side of the swash plate, the drive torque fluctuation per shaft rotation angle is extremely large.

Further, in the above-described embodiment, the flat plate portion 1 of the shaft 1
65 in the slit 105 of the swash plate 10 to transmit the driving force to the swash plate 10. Therefore, members (pin 80, long groove 1
66, spherical support portion 405, etc.) have the advantage that no rotational driving force is directly applied to them, and that they have excellent durability.

In other words, conventional compressors (for example, JP-A-58-
162780), it was difficult to transmit large rotational driving force because the driving force of the swash plate and the thrust force applied to the swash plate during compression were applied to the pin that was the center of inclination of the swash plate. In the compressor, the rotational driving force is efficiently transmitted as described above.

Furthermore, in this embodiment, by using the spool 30 that can slide based on the pressure difference to control the swash plate 10, a compact variable displacement swash plate type compressor can be realized without using a complicated and large-sized motor. Can be configured.

In the above embodiment, a solenoid valve was used as the control valve 400 to control the pressure inside the control pressure space 200 and change the inclination and position of the swash plate to perform capacity control. The pressure inside the space 200 may be controlled using the control valve 150. Generally, when the cooling load is large, the suction pressure is high, and as the cooling load becomes small, the suction pressure becomes low. In the embodiment shown in FIG. 29, a pressure control valve is used to detect this suction pressure and adjust the internal pressure of the space 200. The structure and operation of the compressor are the same as those in the embodiment shown in FIG. 1, so explanations thereof will be omitted.

A cylinder 1501 having two stages of diameter is formed in the main body of the pressure control valve 150.
Inside 01, there is a spool 150 with a diameter of two stages.
3 is slidably inserted. cylinder 150
A small diameter portion 1504 of a spool 1503 is inserted into a small diameter portion 1502 of the spool 1503, and a ball 150 is inserted into the tip of the spool 1503.
5 is fixed. A spring 1507 is inserted between the spool 1503 and the stopper 1506, and urges the spool 1503 toward the distal end.
A suction space 74 is connected to the cylinder 1501 via a low pressure introduction passage 97 on the tip side of the large diameter part of the spool 1503.
The space 200 is connected to the space 200 via the diaphragm 1508. A discharge space 93 is connected to the small diameter portion 1502 of the cylinder 1501 via a high pressure introduction passage 96 . A port 1503 that can be closed by a ball 1505 is connected to the small diameter portion 1502 of the cylinder 1501.
is connected to the space 200 via the communication hole 98.

Next, the operation of the embodiment shown in FIG. 29 will be explained. Suction pressure according to the cooling load that is the target of control
Spring 15 to obtain Ps (Kg/cm ²・abs)
07 Set pressure is P _ST . When the compressor starts operating, the cooling load is large, so the pressure in the suction space 74 decreases.
Ps is larger than the target set pressure _PST . This suction pressure Ps acts on the spool 1503, overcomes the urging force of the spring 1507, and moves the spool 1503 to the right in the figure. Then, the ball 1505 separates from the valve seat at the opening of the port 1509, the control pressure chamber 200 and the discharge space 93 communicate with each other, and the internal pressure of the control pressure chamber 200 becomes the discharge pressure Pd.
As a result, the compressor moves to its maximum capacity state, as described in the rising example. As the compressor continues to operate further, the cooling load decreases, and the pressure Ps in the suction space 74 reaches the target set pressure _PST .
Furthermore, the target set pressure P _ST becomes lower. Then, the urging force by the spring 1507 becomes larger than the force due to the suction pressure acting on the spool 1503, and the spool 1503 moves to the left in the figure. Due to this movement, the ball 1505 seats on the valve seat of the opening of the port 1509, and the control pressure chamber 200 and the discharge space 9
Cut off communication with 3. As a result, the control pressure chamber 200
Since it is connected to the suction space 74 via the throttle 1508, the suction pressure gradually becomes the suction pressure, and the compressor realizes the minimum capacity state as described in the previous embodiment. By repeating this operation, the suction space 7
4 will be maintained near the target set value _PST , and the capacity of the compressor can be adjusted according to the cooling load.

Further, in the embodiment shown in FIG. 4, when a solenoid valve is used as the control valve 400, only the conduction is interrupted between the control pressure passage 98 and the low pressure introduction passage 97, and the communication between the control pressure passage 98 and the high pressure introduction passage 96 is interrupted. Although the passages 96 and 97 are always communicated through the throttle 99, the control valve 400 may be used to switch between the passages 96 and 97.

FIG. 30 is a circuit diagram showing the configuration of the control valve 400 in this case, and in the state 400C in the figure, the high pressure introduction passage 96 and the control pressure passage 97 are communicated. Further, in the state 400D in the figure, the control pressure passage 96 is blocked from the other passages 97 and 98. Also, 400 in Figure 30
In state E, the control pressure passage 98 and the low pressure introduction passage 97
Conducts with. Therefore, in the embodiment illustrated in FIG. 30, when it is necessary to increase the pressure inside the control pressure chamber 200, the state is set to 400C in the figure, and conversely, when it is necessary to decrease the pressure inside the control pressure chamber 200, the state is set to 400C in the figure. The state is set to 400E. Further, when the pressure inside the control pressure chamber 200 is maintained at that state, the state is set as 400D in the figure. In the embodiment shown in FIG. 30, the discharge pressure chamber 93 is connected to the constant control pressure chamber 200.
The discharge pressure chamber 9 is no longer in communication with the discharge pressure chamber 9.
Leakage of the refrigerant inside 3 is prevented. Therefore, the discharge efficiency of the compressor as a whole is further improved.

Further, as shown in FIG. 31, a three-way valve that switches between a control pressure passage 98, a high pressure introduction passage 96, and a low pressure introduction passage 97 may be adopted as the control valve 400. In the embodiment illustrated in FIG.
In the state of 00F, the control pressure passage 98 and the high pressure introduction passage 96 communicate with each other via the throttle 99. Also, 4 in the diagram
In the 00G state, the control pressure passage 98 communicates with the low pressure introduction passage 97.

FIG. 32 shows a cross-sectional configuration of the control valve 400 shown in FIG. 31. In the figure, 4001 is a valve housing made of a non-magnetic material, and a coil housing 4002 made of a magnetic material is attached to the magnetic plate 4 at the end of this valve housing 4001.
Connect via 003. coil housing 40
A coil 4004 is wound inside the coil 4004, and the center portion of the coil housing 4002 is fitted into the inner surface of the coil 4004 as a stator core. And this coil housing 4002
A moving core 4005 made of a magnetic material is disposed at a position facing the stator core portion. Note that the moving core 4005 is urged in a direction to be separated from the coil housing 4002 by a spring 4006.

A first passage member 4010 and a second passage member 4011 are inserted into the valve housing 4001,
It is fixed by bolts 4012. A high pressure passage 4013 communicating with the high pressure introduction passage 96 is formed in the first passage member 4010, and a high pressure passage 4013 communicating with the high pressure introduction passage 96 is formed in the first passage member 4010.
11 is a low pressure passage 4 that communicates with the low pressure introduction passage 97.
014, and a control passage 4015 communicating with the control pressure passage 98. Note that the control passage 4015,
The high pressure passage 4013 and the low pressure passage 4014 communicate with the passages 98, 96, and 97 through passages formed in the valve housing 4001, respectively.

High pressure passage 401 of the first passage member 4010
3. A first valve seat 4018 is formed at the open end. On the other hand, the second passage member 4011 has a second valve seat 4019 at a position facing the first valve seat 4018.
is formed. And both valve seats 4018, 4019
A valve body 4020 made of a steel ball is disposed in between so as to be able to come into contact with both valve seats. Further, the tip of the moving core 4005 passes through the second valve seat 4019 and comes into contact with the valve body 4020 .

When the coil 4004 is not excited, the moving core 4005 is displaced upward in the figure by the biasing force of the spring 4006, and presses the valve body 4020 against the first valve seat 4018. As a result, the control passage 4015 is communicated with the low pressure passage 4014 via the second valve seat 4019. Therefore, in this state, the low pressure introduction passage 97 and the control pressure passage 98 are electrically connected.
Suction pressure is introduced into the control pressure chamber 200 (state 400E in FIG. 30).

Next, when the coil 404 receives an electric signal from the controller 500 and is excited, a magnetomotive force is generated between the coil housing 4002 and the moving core 4005, and this magnetic force causes the moving core 400 to
5 is sucked downward in the figure. As a result, the valve body 4
020 is no longer pressed down by the moving core 4005. Therefore, the pressure inside the high pressure passage 4013 is the same high pressure as the discharge pressure chamber 93, while the pressure on the low pressure passage 4014 side is the same as that in the suction pressure chamber 7.
It has the same low pressure as 4. Therefore, the valve body 4020 is pressed against the second valve seat 4019 based on this pressure difference. Therefore, the high pressure passage 4013 communicates with the control passage 4015 via the first valve seat 4018. Therefore, in this state, the high pressure introduction passage 96
is in communication with the control pressure passage 98, and high pressure is introduced into the control pressure chamber 200 (state 400F in FIG. 31).

Furthermore, in the above example, as shown in FIG.
66 is formed in the flat plate portion 165 of the shaft 1, while a pin through hole 10 is formed in the slit 105 of the swash plate 10.
However, as shown in FIG. 33, the pin through hole 106 and the long groove 166 may be reversed. In this way, the long groove 166 and the swash plate 10
Even if it is formed on the slit 105 side, the same operation as in the above embodiment can be obtained, as shown in FIGS. 34 and 35.

Further, in the above example, the swash plate 10 has a slit 10.
5 is formed on the shaft 1, and the flat plate portion 165 is formed on the shaft 1, but the formation position may be reversed. That is, a slit 105 is formed on the shaft 1 side,
The flat plate portion 165 may be formed on the swash plate 10 side.

Further, in the above-described embodiment, only the second working chamber 60 on the rear housing 13 side of the piston 7 is operated when the discharge capacity of the compressor is reduced.
This operable working chamber may also be reversed. That is, as shown in FIGS. 36 and 37, when the compressor is at its minimum discharge capacity, the first working chamber 50 on the front housing 4 side may operate as a suction compressor. In this case, the spool 30 is connected to the swash plate 1.
At the same time as reducing the inclination angle of 0, the spherical support part 4
05, the rotational center position of the swash plate 10 is displaced toward the front housing 4 side. That is, in the embodiment shown in FIGS. 36 and 37, when the discharge capacity of the compressor is made small, the control pressure chamber 200
The discharge pressure of the discharge chamber 93 is introduced into the discharge chamber 93.
When the discharge capacity of the compressor is increased, the pressure in the control pressure chamber 202 and the suction pressure chamber 74 is guided.

Furthermore, in the embodiment described above, the spool 30 in FIG.
As a means of moving the spool, a pressure difference is used, so the minimum capacity of the compressor is set to about 10% capacity. However, if a means such as external pressure or a motor is used to move the spool, the minimum capacity of the compressor can be reduced to 0%. It is also possible to do this.

[Brief explanation of drawings]

FIG. 1 is a cross-sectional view showing an embodiment of the compressor of the present invention, FIG. 2 is a cross-sectional view of the compressor shown in FIG.
FIG. 3 is a perspective view showing the main parts of the compressor shown in FIG. 1;
Fig. 4 is a circuit diagram of the solenoid valve shown in Fig. 1, Fig. 5 is an explanatory diagram showing the pressure state applied to the piston, swash plate, and spool of the compressor shown in Fig. 1, and Fig. 6 is a circuit diagram of the compressor shown in Fig. 1. 7 is a characteristic diagram showing the capacity variable characteristics of the compressor shown in FIG. 1,
Figure 8 is a graph showing the relationship between the load due to spool back pressure and the amount of spool movement, Figure 9 is a graph showing the relationship between the amount of spool movement and the thrust load acting on the spool, and Figure 10 is the compressor rotation speed. Fig. 11 is a graph showing the relationship between the reciprocating motion of the piston and the working chamber pressure, and Fig. 12 is a graph showing the relationship between the spool stroke ratio and the spool thrust load. Graph, 1st
FIG. 3 is a graph showing the relationship between the stroke range of the spool and the discharge capacity ratio of the compressor, FIG. 16 is a sectional view showing another embodiment of the compressor according to the present invention, FIG. 17 is an explanatory diagram showing the thrust load applied to the spool, and FIG. 18 is a sectional view showing another embodiment of the compressor according to the present invention. 19 is a sectional view showing the main parts of the compressor involved.
Figure 8 is a graph showing the relationship between the opening area of the relief hole and the pressure average value of the pressure chamber, Figure 20 is a graph showing the relationship between the differential pressure across the spool and the compressor discharge capacity,
FIG. 21 is a front view showing other main parts of the compressor according to the present invention, FIG. 22 is a sectional view of the embodiment shown in FIG. 21, and FIG. 24 is a cross-sectional view showing the main parts of still another example of the compressor according to the present invention,
Fig. 25 is a graph showing the relationship between the capacity ratio of the compressor and the thrust load of the spool, Fig. 26 is a sectional view showing an embodiment of the compressor related to the present invention, and Fig. 27 is a graph showing the relationship between the capacity ratio of the compressor and the thrust load of the spool. 28 is an explanatory diagram showing a torque fluctuation state of the compressor shown in FIG. 1; FIG. 29 is a circuit diagram showing a control valve according to another example of the compressor of the present invention; FIG. FIG. 31 is a block diagram showing another example of the control valve of the compressor of the present invention, FIG. 31 is a block diagram showing still another example of the control valve of the compressor of the present invention,
32 is a sectional view showing the control valve structure shown in FIG. 31, FIG. 33 is a perspective view showing a shaft and swash plate portion according to another example of the compressor of the present invention, and FIGS. 36 and 37 are sectional views showing still another example of the compressor of the present invention, and FIGS. 38 and 39 are sectional views showing an apparent hinge. Figure 40 is an explanatory diagram showing the state of the load applied to the swash plate, Figures 41 and 42 are explanatory diagrams showing the state of the load applied to the pin, and Figure 43 is an explanatory diagram showing the state of the load applied to the swash plate. Graph showing the relationship with spool load,
Fig. 44 is an explanatory diagram showing the long groove shape, Fig. 45 is an explanatory diagram showing the relationship between piston stroke ratio and spool load, Fig. 46 is an explanatory diagram showing the relationship between spool load and compressor capacity, and Fig. 47 is an explanatory diagram showing the relationship between piston stroke ratio and spool load. An explanatory diagram showing the relationship between the piston stroke and the spool pressing load, Fig. 48 is an explanatory diagram showing the relationship between the compressor rotation speed and the suction throttle amount, and Fig. 49 shows the main part of an embodiment using the suction throttle. 50 is an explanatory diagram showing the relationship between the piston stroke and the spool pressing load, FIGS. 51 and 52 are sectional views showing the main parts of an embodiment having a suction throttle, and FIG. 53 is a diagram showing the suction throttle. FIG. 54 is a sectional view showing another aspect of the embodiment having a suction throttle, FIG. 55 is an explanatory view showing fluctuations in working chamber pressure, FIG. 56 and FIG. 57 is a sectional view showing a main part of an embodiment having a high pressure chamber and a low pressure chamber, respectively. 1...Shaft, 5, 6...Housing, 7...
... Piston, 10 ... Swash plate, 30 ... Spool,
40...Rear shaft, 400...Solenoid valve.

Claims (1)

[Scope of Claims] 1. A cylinder block having a cylinder chamber inside; a shaft rotatably supported within the cylinder block; a swash plate pivotally connected to the shaft and rotating integrally with the shaft; a piston that is slidably disposed within the cylinder chamber and reciprocates within the cylinder chamber in response to the rocking motion of the swash plate; a working chamber for suctioning, compressing, and discharging fluid; a supporting portion disposed coaxially with the shaft and swingably holding the central point position of the swash plate; a spool that is displaced in the axial direction, the displacement of the spool displaces the center point position of the swash plate in the shaft axial direction, and also displaces the inclination angle of the swash plate, reciprocating movement of the piston within the cylinder chamber. A position where the piston can move forward in a working chamber formed on one end surface side of the piston, and a working chamber formed on the other surface side of the piston among the working chambers. A variable capacity swash plate type compressor, characterized in that the forward movement positions of the pistons are different from each other. 2. The support part is comprised of a spherical support part, and is rotatably and slidably held on a spherical part formed at the center point of the swash plate. variable capacity swash plate compressor. 3. The variable capacity swash plate according to claim 1, wherein the shaft is formed to pass through the swash plate, and the spool is slidably disposed on the shaft. mold compressor. 4. Claim 1, wherein the shaft is disposed only on one side of the swash plate.
Variable capacity swash plate compressor as described in . 5. A flat plate portion is formed on the shaft,
A patent characterized in that the swash plate is formed with a slit into which the flat part fits, and when the flat part engages with the slit, the swash plate is swingably connected to the shaft. A variable capacity swash plate compressor according to claim 1. 6 A long groove is formed in the flat plate portion, and a pin through hole is formed in the slit portion, and the swash plate is swung by the shaft by the pin through hole and the pin inserted into the long groove. 6. The variable displacement swash plate compressor according to claim 5, wherein the variable displacement swash plate compressor is configured to be movably connected. 7. A pin through hole is formed in the flat plate part, and a long groove is formed in the slit, and the swash plate can swing relative to the shaft by the long groove and the pin fitted into the pin through hole. A variable capacity swash plate compressor according to claim 5, which is configured to be connected to a compressor. 8. The variable displacement swash plate compressor according to claim 6 or 7, wherein the pin is rotatably held in the pin through hole via a bearing. . 9 When the reciprocating stroke of the piston in the cylinder chamber is varied, in the working chamber formed on one end surface side of the piston, the piston is moved regardless of the inclination angle displacement of the swash plate. 2. The variable displacement swash plate compressor according to claim 1, wherein the top dead center position is always a position where no dead volume is generated in the working chamber. 10 a cylinder block having a cylinder chamber therein; a shaft rotatably supported within the cylinder block; a swash plate pivotally connected to the shaft and rotating integrally with the shaft; and a swash plate that slides within the cylinder chamber. a piston that is freely disposed and reciprocates within the cylinder chamber in response to the rocking motion of the swash plate; , a working chamber that performs compression and discharge; a support section that is arranged coaxially with the shaft and that swingably supports the center point position of the swash plate; and a support section that displaces the support section in the axial direction of the shaft. a spool; a control pressure chamber formed in a portion of the spool on the opposite side of the support portion and displacing the spool in the axial direction of the shaft according to internal pressure; and a signal supplied to the control pressure chamber. and a control valve for controlling pressure, wherein the spool displaces the center point position of the swash plate in the axial direction of the shaft according to the signal pressure supplied to the control pressure chamber, and changes the inclination angle of the swash plate. A variable capacity swash plate type compressor, characterized in that it is configured to be driven to cause displacement. 11 The control valve switches and controls the signal pressure supplied to the control pressure chamber between the suction side pressure and the discharge side pressure of the compressor, and the control valve controls the discharge side pressure of the compressor to the control pressure chamber. When the spool is introduced into the control pressure chamber, the spool displaces the support portion in a direction in which the inclination angle of the swash plate increases, and the control valve introduces the suction side pressure of the compressor into the control pressure chamber. When this occurs, the compression reaction force of the piston causes the support portion and the spool to be displaced in a direction that becomes smaller than the inclination angle of the swash plate, and in the working chamber formed on one end surface side of the piston, , the piston is capable of advancing to a predetermined position for sucking, compressing, and discharging fluid regardless of the tilt angle displacement of the swash plate, and in the working chamber formed on the other side of the piston, 11. The variable displacement swash plate compressor according to claim 10, wherein a dead space is created in the working chamber depending on the inclination angle of the swash plate. 12. The variable displacement swash plate compressor according to claim 10, wherein a plurality of said cylinder chambers are formed in said cylinder block so as to be parallel to each other. 13 The control valve has a control pressure passage communicating with the control pressure chamber, a high pressure introduction passage introducing discharge side pressure of the compressor, a low pressure introduction passage introducing suction pressure side pressure of the compressor, and the control pressure passage. 11. The variable displacement swash plate compressor according to claim 10, further comprising a switching valve that selectively switches to either the high pressure introduction passage or the low pressure introduction passage. 14 The control pressure chamber communicates with the discharge pressure side of the compressor via a predetermined flow resistance, and the control valve introduces the suction pressure side pressure of the compressor into a control pressure passage communicating with the control pressure chamber. Claim 1, characterized in that it has a low pressure introduction passage and a valve body that switches and controls conduction/blocking between the low pressure introduction passage and the control pressure passage.
The variable capacity swash plate compressor according to item 0. 15 A cylinder block having a cylinder chamber therein, a shaft rotatably supported within the cylinder block, a swash plate pivotally connected to the shaft and rotating integrally with the shaft, and a swash plate that slides within the cylinder chamber. a piston that is freely disposed and reciprocates within the cylinder chamber in response to the rocking motion of the swash plate; and an operating chamber that is formed between the inner surface of the cylinder chamber and the inner surface of the cylinder chamber at each of both ends of the piston. a support part disposed coaxially with the shaft and for swingably holding the center point position of the swash plate; a spool for displacing the support part in the axial direction of the shaft; control means for displacing the shaft in the axial direction between a maximum capacity position where the angle of inclination of the swash plate is maximum and a minimum capacity position where the angle of inclination of the swash plate is the minimum; an auxiliary load means for applying a load to the spool in a direction that suppresses the spool from being displaced by the predetermined value or more toward the maximum capacity position when the spool is displaced by a predetermined value or more; A variable capacity swash plate type compressor, characterized in that the center point position of the swash plate is displaced in the axial direction of the shaft, and the inclination angle of the swash plate is also displaced. 16. The variable displacement swash plate compressor according to claim 15, wherein the auxiliary load means comprises a spring means. 17. A claim characterized in that the spring means is arranged between the spool and the shaft end, and starts compressing when the spool is displaced by more than the predetermined value. The variable displacement swash plate compressor according to item 16. 18. Claims characterized in that the spring means is arranged between the spool and the cylinder block, and is configured such that a spring load is generated when the spool is displaced by more than the predetermined value. 17. The variable displacement swash plate compressor according to item 16. 19 An end plate that covers the working chamber is arranged at the end of the cylinder block, and the spring means is arranged between the spool and the end plate, and when the spool is displaced by more than the predetermined value, the spring means is arranged between the spool and the end plate. 17. The variable displacement swash plate compressor according to claim 16, characterized in that the compressor is configured to exert an elastic load. 20 The shaft is formed with a locking portion facing the support portion, the spring means is disposed between the locking portion and the support portion, and the spool is displaced by the predetermined value or more; 17. The variable displacement swash plate compressor according to claim 16, wherein an elastic load is generated when the support section is displaced by more than the predetermined value due to this displacement. 21 A cylinder block having a cylinder chamber therein, a shaft rotatably supported within the cylinder block, a swash plate pivotally connected to the shaft and rotating integrally with the shaft, and a swash plate sliding in the cylinder chamber. a piston that is freely disposed and reciprocates within the cylinder chamber in response to the rocking motion of the swash plate; and an operating chamber that is formed between the inner surface of the cylinder chamber and the inner surface of the cylinder chamber at each of both ends of the piston. , a support portion disposed coaxially with the shaft and swingably supporting the center point position of the swash plate; a spool displacing the support portion in the axial direction of the shaft; axially displacing the swash plate, displacing the center point position of the swash plate in the axial direction of the shaft, and displacing the inclination angle of the swash plate to vary the reciprocating stroke of the piston within the cylinder chamber, A forwardly movable position of the piston in a first working chamber formed on one end surface side of the piston and a forwardly movable position of the piston in a second working chamber formed on the other end surface side of the piston among the working chambers. A variable capacity swash plate type compressor, comprising: control means for controlling the swash plate differently from each other; and auxiliary displacement means for applying a load in a direction to reduce the inclination angle of the swash plate. 22. The variable displacement swash plate type according to claim 21, wherein the auxiliary displacement means comprises a spring that applies a load to the spool in a direction that reduces displacement of the spool toward the support portion. compressor. 23. Claim 22, wherein the spring is arranged such that one end abuts the spool and the other end abuts the shaft.
Variable capacity swash plate compressor as described in . 24 In the first working chamber, a dead space is created at the tip of the piston when the tilt angle of the swash plate is displaced as small as possible, and the auxiliary displacement means is a pressure booster that guides the discharge side pressure into the first working chamber. 22. The variable displacement swash plate compressor according to claim 21, characterized by comprising passage means. 25. Claim 1, characterized in that the pressure boosting passage means is a communication means for communicating between the first working chamber and a discharge chamber that communicates with the first working chamber via a discharge port and a discharge valve. 24. The variable capacity swash plate compressor according to 24. 26 The end face of the cylinder block has the first
Claim 25, characterized in that a side plate having the discharge port communicating between the working chamber and the discharge chamber is disposed, and the relief passage is formed in this side plate. Variable capacity swash plate compressor. 27. The variable displacement swash plate compressor according to claim 25, wherein the relief passage means is comprised of a relief hole formed in the discharge valve. 28 In the second working chamber, the piston can move forward to a predetermined position for sucking, compressing, and discharging fluid regardless of changes in the inclination angle of the swash plate; A dead space is generated at the tip of the piston depending on the angle, and the auxiliary displacement means displaces the end of the piston that can be advanced toward the second working chamber to create a dead volume on the side of the second working chamber of the piston. 22. The variable displacement swash plate compressor according to claim 21, further comprising dead volume generating means. 29 A flat plate portion is formed in the shaft or the swash plate, and a slit into which the flat plate portion fits is formed in the swash plate or the shaft, and a long groove is formed in the flat plate portion. and a pin through hole is formed in the slit portion,
The swash plate is swingably connected to the shaft by the pin disposed in the pin through hole and the long groove, and the long groove of the flat plate portion adjusts the inclination angle of the swash plate as the spool is displaced. 29. The variable displacement swash plate compressor according to claim 28, wherein a dead volume is generated on the second working chamber side when the compressor becomes smaller. 30 A flat plate portion is formed in the shaft or the swash plate, and a slit into which the flat plate portion fits is formed in the swash plate or the shaft, and a long groove is formed in the flat plate portion. and a pin through hole is formed in the slit portion,
The swash plate is swingably connected to the shaft by the pin disposed in the pin through hole and the long groove, and furthermore, the long groove is connected to the swash plate when the inclination angle of the swash plate is at a maximum. When the apparent hinge point between the plate and the piston is located at the center of the piston and the angle of inclination of the swash plate is minimum, the apparent hinge point between the swash plate and the piston is located at the center of the piston. 22. The variable displacement swash plate compressor according to claim 21, wherein the variable displacement swash plate compressor is formed in an arc shape so as to be located closer to the support portion than the center position of the piston. 31 In the first working chamber, a dead space is created at the tip of the piston depending on the inclination angle of the swash plate, and in the second working chamber, fluid is sucked, compressed, and discharged regardless of the inclination angle displacement of the swash plate. 22. The variable valve according to claim 21, which is capable of moving forward to a predetermined position, and wherein the auxiliary displacement means comprises a restricting means formed in the middle of a suction passage for sucking fluid into the second working chamber. Capacitive swash plate compressor. 32 the throttle means is formed on the spool;
32. The variable displacement swash plate compressor according to claim 31, wherein the amount of restriction of the suction passage is made variable in accordance with the displacement of the spool. 33 The throttle means reduces the amount of restriction of the suction passage when the angle of inclination of the swash plate becomes large, and increases the amount of restriction of the suction passage when the angle of inclination of the swash plate becomes small. 32. The variable displacement swash plate compressor according to claim 31. 34 The control means drives the spool based on the differential pressure between the high pressure on the discharge side and the low pressure on the suction side of the compressor, and the auxiliary displacement means causes the control means to control the pressure below the low pressure on the suction side. 22. The variable displacement swash plate type compressor according to claim 21, wherein the variable displacement swash plate compressor supplies a low pressure of 1 or a high pressure higher than the discharge side high pressure.