JPH01224480A - Variable capacity compressor with swash plate - Google Patents

Abstract

PURPOSE:To vary the discharge capacity of a compressor continuously and accurately by furnishing an aux. loading means to give a load in the direction of suppressing the change of a spool to the max. capacity position of the compressor, and making the setting load non-linear. CONSTITUTION:A shaft 1 is supported rotatably by cylinder blocks 5, 6, and a swash plate 10 is coupled with this shaft 1 swingably. A piston 7 reciprocating with swing of this swash plate 10 is installed in a cylinder chamber 64, and working chambers 50, 60 are formed at the ends of this piston 7. Coaxially with the shaft 1, a spool 30 is installed which displaces the rotation supporting position of the swash plate 10 in the axial direction of the shaft 1. A control means 400 is installed which displaces the angle of inclination and the rotational center position of the swash plate 10 by displacement of the spool 30. An aux. loading means 900 is installed, with which the suppressing force increases non- linearly with displacement of the spool 30 toward the max. capacity position.

Description

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

[Background of the invention]

The present inventors have previously proposed a compressor in which the swash plate 10 is swingably engaged with the shaft 1, and the rotation center position of the swash plate is displaceable in the axial direction of the shaft 1 (second (Illustrated)

According to the swash plate type compressor shown in FIG. 2, the slope of the swash plate 10 decreases as the spool 30 is displaced in the axial direction of the shaft 1, thereby varying the reciprocating stroke of the piston 7.

Furthermore, when the stroke of this spool is varied, the swash plate 100
Since the rotation center position is also displaced, although there is a significant increase in dead volume in one working chamber 50 of the piston 7, the capacity in the other working chamber 60 is gradually reduced without a significant increase in dead volume. Can be done.

Therefore, in the compressor shown in FIG. 2, the compressor discharge capacity can be continuously controlled according to changes in the spool 30.

That is, according to the compressor shown in FIG. 2, the compressor discharge capacity can be reduced as shown by the solid line A in FIG. 3. In FIG. 3, a broken line C indicates a change in discharge capacity due to the working chamber 50, and a dashed line B indicates a change in discharge capacity due to the working chamber 60.

Further, the horizontal axis of the third tooth indicates the displacement of the spool 30.

As shown in FIG. 3, according to the compressor shown in FIG.
Compared to the continuously variable capacity (solid line ")", the spool 30
When the displacement is between e and l, the discharge capacity of the compressor decreases rapidly (solid line a+).

However, when the displacement of the spool is equal to or less than e, the rate of decrease in the compressor discharge capacity becomes gradual (solid line at).

However, according to the experimental studies conducted by the present inventors, it has been found that there are cases in which it is actually difficult to maintain the displacement of the spool 30 at an appropriate position.

If the back pressure applied to the spool 30 is gradually increased as shown in FIG. 4, the spool will be displaced as the back pressure increases, as shown by the solid line XY in the figure, until the spool back pressure reaches the predetermined pressure F2. Note that in FIG. 4, 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. 4, the back pressure of the spool 30 is set to a predetermined value F2.
It has been confirmed that when the stroke is increased above, the stroke does not continuously displace but immediately increases to the maximum stroke (solid line YZ). That is, when the spool 30 back pressure is higher than the predetermined value F2,
The sign will always be held at the position where the stroke of the spool 30 is maximum.

Conversely, when reducing the back pressure of the spool 30, the maximum back pressure load F3 is reduced to a predetermined load F2, and even if a smaller load F1 is further reduced, the spool 30 is maintained at the maximum displacement position. (dashed line Z
K). Then, when the back pressure of the spool 30 becomes lower than the predetermined value F on the low pressure side, the spool 30 is suddenly displaced by a certain amount of displacement (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. 3, 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 this origin and found that this is because the relationship between the axial force of the shirt 1 applied to the spool 30 at each stroke of the spool 30 tends to be as shown in FIG. It was done. 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 0 in FIG. 5). 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. 5). However, if the stroke of the spool 30 is to be increased further, the stroke of the spool 30 must be
On the contrary, it is recognized that the force required for a displacement of 0 becomes smaller (indicated by the solid line PQ in FIG. 5). 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, and is a state in which the discharge capacity of the compressor is slightly reduced from the increased discharge capacity.

That is, as shown in Fig. 5, there is a maximum load F between the 300 strokes of the spool and the thrust force required for its movement.
z (point P) is recognized, and the stroke of the piston corresponding to this local maximum value F2 is P2. The stroke P2 of this spool 30 corresponds to point Y in FIG. As mentioned above, when the thrust force is made larger than the predetermined value F2, the spool 30 immediately advances to the maximum amount (point Q in Fig. 5 and point Z in Fig. 4), and in this state, the spool 30 is held at the maximum position. This will continue until the back pressure on the spool 30 decreases to a value below the thrust force F1 required for this. If the back pressure on the spool 30 becomes less than the thrust force F, the spool 30 will immediately be displaced from point Q to point R in FIG. Spool 30 at this R point
The displacement is Pl, and this position corresponds to the midpoint in FIG.

The characteristics shown in FIG. 5 are obtained because, 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. It is. This operation will be explained below using FIG. 6.

FIG. 6 shows the relationship between the stroke of the piston 7 and the internal pressure of the working chamber 50, in other words, the relationship between the internal volume of the working chamber 50 and the internal pressure of the working chamber 50. The state shown by the solid line A in FIG. 6 is the state in which the piston 7 moves forward to its maximum stroke, that is, the maximum displacement state of the compressor. In addition, in Figure 6 - dotted chain line B
2 shows 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 generated between the piston 7 and the side plate 8. Also, what is indicated by the broken line C in FIG. 6 is the swash plate 10.
The inclination angle of the swash plate 10 becomes smaller (and the dead volume becomes larger). In addition, the two-dot chain line in FIG. , indicates the state when the dead volume reaches 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 advances from the most retracted position (indicated by a in the figure), the volume of the working chamber 50 decreases and the pressure within the working chamber 50 increases (indicated by a-b-○ in the figure). ). When the pressure within the working chamber 50 reaches a predetermined pressure Pd, the discharge valve 24 opens, and the pressure within the working chamber 50 does not rise any further (indicated by c-de in the figure). After piston 7 has been displaced to its maximum stroke (
), when 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.
(indicated by f in the figure), the piston returns again to the rear end position (indicated by a in the figure). That is, when the piston is at its maximum displacement, d in the working chamber 50 is a, c, e, f.
, a, the pressure is changed in cycles.

If the angle of inclination of the swash plate 10 becomes somewhat small (and a dead volume is generated at the tip of the piston 7), as shown in FIG.
As shown by the dotted chain line B, a predetermined capacity is maintained within the working chamber 50. Therefore, even if the piston 7 retreats from this state, the refrigerant held in the working chamber 50 will expand again (indicated by the dotted chain line d-g in the figure), and during that time the suction pressure in the working chamber 50 will exceed Ps. pressure will be maintained.

As the inclination angle of the swash plate 10 becomes 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 not open even when the piston 7 moves forward. become. 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 in the dashed line in FIG. 6 is finally reached. In this case, suction and discharge are not performed within the working chamber 50,
The relationship between the volume and pressure of the actuator 50 is a-ba-a.

As explained above, the dead volume in the working chamber 50 causes the pressure in the working chamber 50 to fluctuate during the reciprocating cycle of the piston.

FIG. 7 is a graph showing the relationship between the pressure within the working chamber 50 and the reciprocating period of the piston 7. As shown in FIG. 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 line B in FIG. 7 corresponds to the dashed line B in FIG. recognized. 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. 7 corresponds to the broken line C in FIG. 6, and when the dead volume increases to this state, the pressure fluctuation in the working chamber 50 becomes almost sinusoidal, and the suction pressure falls below Ps. A decrease in the pressure inside the working chamber 50 causes vomiting.

In addition, the dashed dotted line in Figure 7 corresponds to the dashed double dotted line in Figure 6, but in this state, the pressure fluctuations are almost sinusoidal, similar to the case shown by dashed line C, and compression and suction are not performed. . Furthermore, in the state shown by the two-dot chain line, 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. 5 indicates the region in which the pressure-volume state within the cycle ranges from solid line A to broken line C in FIG.

That is, in this region, as is clear from FIG. 7, as the pressure remains in the working chamber 50, the force of the pressure in the first working chamber 50 that urges the piston 7 in the right direction in FIG. 1 increases. It turns out.

Here, the force in which the pressure within the first working chamber 50 presses the piston 7 in the rightward direction acts in a direction that increases the inclination angle of the swash plate 5°. That is, the pressure remaining in the working chamber 50 increases the inclination angle of the swash plate 10, 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. 5, and in this region, as the dead volume increases, the pressure remaining in the working chamber 50 increases. Therefore, in this region, the thrust force required to press the spool 30 to the left in FIG. 2 increases as the dead volume increases.

As described 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 state P (FIG. 5) to the maximum position. As the amount of displacement of the spool 30 becomes maximum under 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 becomes discontinuous 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 constantly control the discharge capacity of the compressor, it is necessary to have the characteristics shown by the broken line PS in FIG. Therefore, the return spring 90 returns the spool 30 to the lower capacity side.
An auxiliary loading means such as 0 is used. That is, this return spring 900 improves the reverse gradient characteristic shown in FIG.

The region on which the return spring 900 acts is the region where the stroke of the spool 30 reaches its maximum from position P2 in FIG. 5 where the thrust force reaches its maximum value. Further, the spring constant of the return spring 900 is set to a value that is at least sufficient to compensate for the reverse slope of the thrust load.

For example, in a compressor as shown in FIG. 2, the stroke of the spool 30 at the position where the inclination angle of the swash plate 10 is the maximum is Omm, and the spool 30 is set at the position where the inclination angle of the swash plate is the minimum.
The maximum stroke of is 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 180
cc, the suction pressure Ps is 3 kg/c++1 a
b s, and the discharge pressure Pd is about 12 kg/cJ a b s to 18 kg/aft a b s, the above-mentioned reversal region shown in FIG. 8 occurs when the stroke of the spool 30 is 7
This will occur at the position where the value exceeds the value. Therefore, in the embodiment shown in FIG. 2, a return spring 900 is installed which generates a compressive load when the stroke of the spool 30 exceeds 7M. The spring constant of this return spring is, for example, 33 kg/body.

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 OF in FIG. 9, and as the back pressure of the control spool 30 is increased, the stroke of the spool 30 is The amount will be large. Further, in the region where the stroke of the spool 30 is 7 cylinders or more, the return spring 900 is activated.
Unless a thrust load greater than the set load of 0 is applied to the back surface of the spool 30, the spool 30 will not be displaced to the right in FIG. That is, this return spring 9°O eliminates the reversibility of the displacement load. Note that FIG. 2 shows a state in which the spool 30 has been displaced by 7 mm or more and the return spring 906 has started to be compressed.

In this way, in the compressor previously proposed by the present inventors, the reversibility of the thrust load and the spool 300 stroke displacement caused by the dead volume of the working chamber 50 can be achieved by using the compression force of the return spring. (Illustrated in Figure 8).

However, the above explanations of FIGS. 5 to 8 are given under the condition that the suction pressure Ps and the discharge pressure Pd are each constant, and when the compressor is used for compressing refrigerant in a refrigeration cycle, The suction pressure P depends on the required operating conditions of the refrigeration cycle, the ambient temperature of the compressor, etc.
s and the discharge pressure Pd (compression ratio) will change variously.

For example, in the low-load operating state of the refrigeration cycle, the suction pressure Ps
is 2-5 kg/cIlla b s and the discharge pressure Pd is about 16 kg/c+fl a b s. When the heat load of the refrigeration cycle is large, the suction pressure P
s increases to about 4 kg/c+fl a b s, and the discharge pressure Pd also increases to about 26 kg/cI a b s. In this way, the suction Ps and the discharge pressure Pd change,
The compression ratio ε of the compressor will also change accordingly.

FIG. 9 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. 9, the thrust load increases as the discharge pressure increases. In particular, as shown in FIG.
The fluctuation of the zero thrust load becomes large. 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. In other words, when the discharge pressure is high, the pressure within the working chamber 50 due to the dead volume increases (as a result, the thrust load required to displace the spool 30 in the axial direction also increases). .And as is clear from Figure 9,
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. 3 also changes depending on the displacement of the suction pressure Ps and discharge pressure Pd of the compressor, as shown in FIG. 9. In FIG. 10, a solid line indicates a steady compression state where the compression ratio is 5, a broken line indicates a low load state where the compression ratio is 4, and a dashed line indicates a high load state where the compression ratio is 6.

Similar to FIG. 9, FIG. 11 shows the relationship between the spool stroke ratio and the spool load. FIG. 11 shows how the spool load changes with respect to changes in the compression ratio. 11th
In the figure, a solid line indicates a state where the compression ratio is 2, a solid line indicates a state where the compression ratio is 4, and a solid line M indicates a state where the compression ratio is 7.

As described above, since the state of change in the spool load differs depending on the compression ratio, for example, if a return spring 900 having spring characteristics as shown in FIG. 8 is used, the displacement of the spool load will vary depending on the compression ratio. There is a situation where is not on a straight line. In the example shown in FIG. 8, the return spring 900 has a stroke ratio C
In order to exert the biasing force in the state of 1, the compression ratio is in state B (
When the maximum points Pk and PI are on the small side of the stroke ratio, such as when the compression ratio is 4 (0 in Fig. 11) or when the compression ratio is 4 (0 in Fig. 11), the load gradient is as shown in Fig. 12. A negative state occurs.

In FIG. 12, this situation occurs in the area shown by the solid line indicating the compression ratio. KI occurs between the stroke C2 corresponding to the local maximum point Pk when the compression ratio is 2 and the stroke ratio C1 at which the biasing force of the return spring 900 begins to occur.

In addition, when the maximum value Pm is larger than the stroke ratio C2 corresponding to the starting point of the biasing force of the return spring 900 (cm2 in FIG. 11), as in the case of the compression ratio 7 shown by the solid line M in FIG. In this case, a situation occurs in which the gradient of change in spool load becomes negative. This state corresponds to the state shown by M in FIG. That is, when the compression ratio is 7, the maximum point Pm occurs when the stroke ratio is greater than the 01 point. If the reverse gradient of the spool load becomes larger than the spring characteristic of the biasing spring shown in FIG. 8, as indicated by 0 in FIG. 11, the difference will appear in the reverse slope characteristic of FIG. 12.

As described above, according to the studies of the present inventors, it has been confirmed that simply providing the return spring 900 does not ensure a simple correlation between the stroke ratio and the spool load.

According to the studies conducted by the present inventors, it has been confirmed that what is important in the above-mentioned simple correlation between the stroke ratio and the spool load is the slope of the spool load change state on the higher stroke ratio side than the maximum value. Ta. Furthermore, it was confirmed that the slope of the spool load change on the higher stroke ratio side than this maximum value increases as the stroke ratio increases.

As mentioned above, the solid line 0 in Figure 11 is the characteristic of the spool load when the compression ratio is 7, and this characteristic has a steeper slope than the spool load gradient when the compression ratio is 6, as shown by the solid line P in Figure 11. ing.

Further, both the solid line O and the solid line P have steeper slopes than the spool load slope when the compression ratio is 5.

In FIG. 11, slopes are shown at three points when the compression ratio is 7, when the compression ratio is 4, and when the compression ratio is 2, but in reality there are an infinite number of compression ratios. Therefore, by measuring the point at which the maximum load gradient occurs according to each compression ratio and plotting it with the stroke ratio at which the maximum load gradient occurs on the horizontal axis, the tendency shown by the solid line F in Fig. 11 is obtained. becomes.

[Problem to be solved by the invention]

The present invention has been devised based on the above-described study results by the present inventors, and an object of the present invention is to be able to continuously and accurately control the compressor capacity regardless of the compression ratio of the compressor.

[Configuration and operation]

In order to achieve the above object, the present invention provides an auxiliary load means that applies a load to the spool in a direction that suppresses displacement toward the maximum capacity displacement side.

In the present invention, as this auxiliary load means, one whose load is displaced nonlinearly is used.

That is, in the compressor of the present invention, the load by the auxiliary load means is set to be larger at the maximum capacity position than at the small capacity position.

With such a configuration, in the compressor of the present invention, the relationship between the displacement of the spool and the load acting on the spool always changes continuously without being accompanied by a reverse gradient. Therefore, in the compressor of the present invention, by displacing the load acting on the spool, the discharge capacity of the compressor can be continuously controlled.

〔Example〕

An embodiment of the present invention will be described below based on the drawings.

The basic structure of the variable displacement swash plate compressor of the present invention is the same as the example shown in FIG. 2 described above. Front housing made of aluminum alloy 4. Front side plate 8. Suction valve 9
．． Front cylinder block 5, rear cylinder block 6, intake valve 12. The rear side plate 11 and the rear housing 13 form an outer shell of the compressor that is integrally fixed by through bolts 16. cylinder block 5,
6 has cylinders 64 at five locations, and the cylinders 64 are formed in parallel to each other. A shaft 1 that rotates under the driving force of an automobile engine (not shown)
is rotatably supported by a front cylinder block 5 via a bearing 3. In addition, the thrust force applied to the shaft 1 (force acting in the left direction in the figure) is transferred to the thrust bearing 15.
The shaft 1 is received by the front cylinder block 5 via the retaining ring, and movement of the shaft 1 to the right in the figure is restricted by a retaining ring.

Note that the retaining ring is retained by an annular groove formed in the shaft 1.

The rear side of the shaft 1 is rotatably supported by a spool 30 via a slide portion 40 and a bearing 14. The thrust force acting on the slide portion 40 (the force acting in the right direction in the figure) is absorbed by the thrust bearing 11 and the retaining ring formed on the slide portion 40 . The spool 30 includes the cylindrical portion 65 of the rear cylinder cylinder and the cylindrical portion 13 of the rear housing 13.
5 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 formed at the end of the slide part 40 is disposed on this spherical part 107. It is supported by a support section 405.

A slit 105 is formed on the side surface of the shaft 1 of the swash plate 10, and a flat plate portion 16 is formed on the end surface of the shaft 1 on the swash plate 10 side.
5 is formed. 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.

Furthermore, shoes 18 and 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 described above, the shoes 18 and 19 are slidably attached to the swash plate 10. Furthermore, the shoes 18 and 19 are rotatably engaged with the inner surface of the piston 7. Therefore, the rocking motion accompanying the rotation of the swash plate 10 is transmitted to the piston via the shoes 18 and 19 as a reciprocating motion. 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 plate portion 165 of the shaft 1, and a pin hole is formed in the swash plate 10. After the flat plate portion 165 of the shaft 1 is arranged in the slit 105 of the swash plate 10, it is locked in the pin passage hole and the long groove 166 of the shaft 1 by the pin 80 and the retaining ring. The inclination of the swash plate changes depending on the position of the pin 80 in this long groove 166, and as the inclination changes, the center of the swash plate (the spherical part
The position of the spherical support portion 405) of No. 7 also changes. That is, in the second working chamber 60 on the right side in FIG. 2, 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 remains the same. The long grooves 166 are arranged so that there is no substantial increase in dead volume.
is provided. On the other hand, in the working chamber 50 on the left side in the figure, the top dead center of the piston 7 changes as the inclination of the swash plate 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. Further, in this example, the long grooves 166 are arranged on the axis of 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. In this way, the long groove 166
Forming the flat plate part 165 on the axis of the shaft 1 to reduce the size of the flat plate part 165 is particularly effective in a swash plate type compressor in which the flat plate part 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 along the shaft 1. Reference numeral 24 in the figure opens into the working chamber 50.60, and the discharge chamber 90.60.
The discharge port 24 is opened and closed by the discharge valve 22 . The discharge valve 22 has a valve holder 23
It is also fixed to the front side plate 8 and the rear side plate 11 by bolts (not shown). In the drawing, reference numeral 25 denotes a suction port that communicates the working chamber 50.60 and the suction chamber 72.74, and is opened and closed by the suction valve 9 and the suction valve 12.

Reference numeral 400 in the figure is a control valve for controlling the internal pressure of the control pressure space 200, one of which is connected to the rear suction space 74 by a low pressure introduction passage 97.

The other end is connected to the control pressure chamber 200 via a control pressure passage 98 . In addition, the control chamber 200 and the discharge space 93 are connected via a throttle 99.

A discharge space 90 on the front side in FIG. 1 is led to a discharge boat by a discharge passage formed in the cylinder block 5.
Further, the rear side discharge space 93 is led to a discharge port by a discharge passage formed in the cylinder block 6. Since the re-discharge port 95 is connected to 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.

A locking plate 901 is arranged on the rear end surface of the slide portion 40, and a spring means 900, which is an auxiliary displacement means, is arranged between the locking plate 901 and the rear end surface of the shaft 1. Therefore, the spring means 900 was set in view of the above-mentioned results of the study shown in FIGS. 5 to 13, and its spring constant k (shown in FIG. It becomes. FIG. 1 is an enlarged cross-sectional view of the spring means 900 portion.

Note that FIG. 1 shows a state in which the spool 30 is displaced farthest to the left in FIG. 2, that is, a state in which the compressor has the minimum discharge capacity.

As shown in FIG. 1, when the spool 30 is displaced to the minimum capacity position of the compressor, one end 9 of the spring means 900
There is a predetermined gap l between 10 and the rear end portion 911 of the shaft.
is formed. Therefore, when the spool 30 is displaced toward the small capacity position, the biasing force of the spring means 900 will not be generated if the displacement is less than a predetermined value. The above-mentioned 13th
In the figure, the spring load F gradually increases from the position where the stroke ratio of the spool 30 is 50%.
This corresponds to the state in which the rear end portion 911 of the shaft begins to come into contact with the rear end portion 911 of the shaft. As is clear from FIG. 13, the spring load F of the spring means 900 in this example increases nonlinearly, and it approximately corresponds to the spring constant #gradient.

Next, the operation of the compressor with the above configuration will be described. When an electromagnetic clutch (not shown) is connected and driving force from the engine is transmitted to the shaft 1, the compressor is started.

Since no pressure difference occurs inside the compressor at startup, no pressure difference occurs before and after the spool 30. That is,
At the time of startup, no load is applied to the direction in which the swash plate 10 is tilted 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. As the piston 7 moves back and forth, the refrigerant is sucked, compressed, and discharged within the working chamber 50, 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, 19. In particular, since the swash plate 10 is slidably supported by the spherical support part 405 and receives the rotational force of the shaft l by fitting the slit 105 and the flat plate part 165, the force applied to the piston 7 acts as a moment in the direction of decreasing the inclination angle of the swash plate 10.

Therefore, in this state, the spherical support portion 405 and the spool 30 are displaced to the right in the figure. As a result, the swash plate 10 reduces its angle of inclination. However, the swash plate 10 is the shaft 1
Because it is regulated by the long groove 166 of the bottle 80,
The swash plate 10 reduces its inclination and applies a force to the right in the figure to the spherical support 405 at the center of the swash plate 10, moving the spherical support 405 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.

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 the suction passage 73 and enters the suction chamber 74 on the rear side. Thereafter, during the suction stroke of the piston 7, the working chamber 6o is
inhaled into the body. The sucked refrigerant gas is compressed in the compression stroke, and when it is compressed to a predetermined pressure, it is discharged from the discharge port 24 through the discharge valve 2.
2 is pushed open and discharged into the discharge chamber 93. The high-pressure refrigerant gas passes through the discharge passage and is discharged from the discharge port 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 punt volume, the pressure of the refrigerant gas in the first working chamber 50 is the pressure inside the discharge space 90 (the discharge pressure in the second working chamber 60 on the rear side is derived). ), and the suction and discharge of refrigerant gas in the front-side first working chamber 50 is not performed.

When the compressor is started, the compressor discharge capacity becomes the minimum capacity as described above. However, when the compressor capacity required by the refrigeration cycle is high, 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 the state where the connection with the low pressure introduction passage 97 is cut off, the control pressure chamber 2
00, the influence of the discharge pressure received from the high pressure introduction passage 96 becomes large. Therefore, the pressure within the control pressure chamber 200 increases.

Therefore, the force acting on the spool 30 in the left direction in FIG. 2 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 operates. 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. Then, due to the action of the long groove 166 of the shaft 1 and the pin 80, the swash plate 10 increases its inclination while moving its center of rotation (spherical support portion 405) to the left in the figure. 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 situation shown in FIG. In the state shown in FIG. 2, refrigerant gas sucked through a suction port (not shown) enters the central suction space 70, passes through suction passages 71 and 73, and flows into suction chambers 72 and 74, respectively. and,
In the suction stroke, it enters the working chambers 50 and 60 from the suction port 25 through the suction valves 9 and 12, respectively, and is then compressed with the displacement of the piston 7, and flows from the discharge port 24 through the discharge valve 22 into the discharge spaces 90 and 60, respectively. 93, passes through a discharge passage, is discharged from a discharge port, and joins at an external pipe.

In this state, both the working chamber 50 and the working chamber 60 perform the action of sucking in and discharging refrigerant gas. The capacity control method according to this example changes the stroke of the piston 7 and also changes the center position of the swash plate 10 by changing the inclination of the swash plate lO. There is almost no increase in volume. Therefore, as shown by the dashed line in FIG. 3, the discharge capacity gradually decreases in accordance with the piston stroke.

On the contrary, in the front side first working chamber 50, the dead volume increases as the piston stroke decreases, and the compression ratio decreases due to the increase in dead volume, and the discharge capacity rapidly increases as shown by the broken line C in FIG. Decrease. And the maximum pressure (discharge pressure) in the front side working chamber 50
is lower than the discharge pressure in the working chamber 60 (3rd point)
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.

As shown in FIG. 13 above, in this example, the spring load k of the spring means 900 is set to correspond to the load gradient F, so the spool 3 is always applied regardless of the compression ratio.
0 will be displaced in response to an increase in the internal pressure of the control pressure chamber 200. In other words, in the compressor of this example, the relationship between the spool load and the spool stroke ratio has an inverse slope (M in FIG. 12), such as when the spring means 900 is a linear spring.
1°Kl) does not occur.

FIG. 14 shows the relationship between the spool stroke ratio and the spool load in a compressor using the nonlinear spring of this example as the spring means 900. As is clear from FIG. 14, in this example, the spool load simply increases upward to the right in the figure, regardless of the compression ratio. Therefore, according to the compressor of this example, if the pressure in the control pressure chamber 200 is increased, the spool 30 is reliably displaced to the right in FIG. 2 in response to the increase in pressure. As a result, in the compressor of this example, the discharge capacity of the compressor can be accurately and uniquely determined by controlling the pressure within the control pressure chamber 2° to increase or decrease using the control valve.

It should be noted that, although the above-mentioned embodiments are desirable measures of the present invention, the present invention is not limited to the above-described embodiments, and there are various other embodiments.

That is, in the above example, a nonlinear spring was used as the spring 900, which is the auxiliary load means, but two springs 905 and 906
By arranging them in series, the characteristics of the spring means as a whole may be made non-linear. FIG. 15 shows an example in which spring means are arranged in series, one end of the first spring 905 is locked to the rear end portion 911 of the shaft, and the other end is locked to the slide portion 40.
It is locked by a spring receiver 914 located within. Further, one end of the second spring 906 is locked by the spring receiver 914, and the other end is held by the locking plate 901.

Figure 16 shows the two springs 905°906 shown in Figure 15.
The spring constant of the spring means using . The illustrated shape in FIG. 1 indicates the spring constant of the first spring 905, and the solid line I indicates the spring constant of the second spring 905.
06 spring constant is shown. And these both springs 905,
The overall spring constant in combination with 906 has the shape shown in FIG. Further, the shape shown in FIG. 1 indicates the spring constant of the nonlinear spring shown in FIG. 1 described above. As is clear from FIG. 16, by arranging the two springs 905 and 906 in series, nonlinear characteristics similar to those of the above embodiment can be obtained, and the spool can be controlled reliably.

Furthermore, if two springs are arranged in series as shown in FIG.
5,906 can have its ends engaged with the shaft 1, the spring receiver 914 and the locking plate 901. As a result, in the embodiment shown in FIG. 15, the springs 905, 906 are prevented from freely rotating within the slide portion 40. In other words, with the spring 900 having the shape shown in FIG. 1, the gap 2 was created between the spring 900 and the rear end portion 911 of the shaft when the spool 30 was displaced to the small capacity position. In the example shown in FIG. 15, such a gap can be prevented from occurring. Therefore, the spring 900
When the spring 900 rotates inside the spring 900, the contact position of the spring 900 changes slightly.1. It is also possible that the spring properties change somewhat as a result of that change. However, if the springs 905 and 906 always engage their ends as in the example shown in FIG.
Rotation inside is well prevented. Therefore, in the example shown in FIG. 15, the spring loads produced by the springs 905, 906 can be uniformly determined.

FIG. 17 is a further modification of the example shown in FIG. 15. In the example shown in FIG. 17, the end of the first spring 905 comes into contact with the planar rear end 911 of the shaft without providing a special holding groove in the rear end 911 of the shaft.

FIG. 18 shows yet another example of the present invention, and in the embodiment shown in FIG. 18, a first spring 905 is installed so as to fit inside a second spring 906. In the embodiment shown in FIG. 18, the nonlinear characteristic of the spring constant is achieved by two springs 905 and 906, similar to the above-described example shown in FIG. 15.

Furthermore, the springs 905 and 906 do not necessarily have to be disposed within the slide portion 40.

For example, as shown in FIG. 19, the first spring 905 may be disposed within the suction pressure chamber 74.

〔Effect of the invention〕

As explained above, in the compressor of the present invention, an auxiliary load means is provided to apply a load in a direction that suppresses the spool from changing to the maximum capacity position of the compressor, and the set load of the auxiliary load means is made non-linear. Regardless of an increase or decrease in the compression ratio in the working chamber of the compressor, the spool can always be reliably displaced as the pressure in the control pressure chamber increases.

Therefore, according to the compressor of the present invention, the discharge capacity of the compressor can be continuously and accurately variably controlled.

[Brief explanation of the drawing]

FIG. 1 is a sectional view showing the spring means used in the compressor of the present invention;
Fig. 2 is a sectional view showing a compressor according to the present invention, Fig. 3 is an explanatory diagram showing the relationship between spool displacement and compressor capacity of the compressor shown in Fig. 2, and Fig. 4 is an illustration without using auxiliary loading means. FIG. 5 is an explanatory diagram showing the relationship between spool back pressure and spool stroke in a compressor. FIG. 5 is an explanatory diagram showing the relationship between spool stroke and thrust force acting on the spool in a compressor that does not use auxiliary loading means. FIG. An explanatory diagram showing changes in the stroke of the piston in the first working chamber, Fig. 7 is an explanatory diagram showing changes in the pressure in the first working chamber, and Fig. 8 shows compression using an auxiliary load means having linear spring characteristics. An explanatory diagram showing the relationship between the spool stroke ratio and the spool load in the machine, Fig. 9 is an explanatory diagram showing the load change due to discharge pressure fluctuation, and Fig. 10 is an explanatory diagram showing the relationship between the compressor capacity ratio due to compressor fluctuation. Figure 11 is an explanatory diagram showing the relationship between compression ratio fluctuation and spool load, and Figure 12 is an explanatory diagram showing the relationship between spool stroke ratio and spool load when a spring with linear characteristics is used as an auxiliary load means. An explanatory diagram showing the relationship; FIG. 13 is an explanatory diagram showing the spring constant of the spring means shown in FIG. 1;
FIG. 14 is an explanatory diagram showing the relationship between the spool stroke ratio and the spool load when a spring means having nonlinear characteristics according to the present invention is used as an auxiliary load means, and FIG. FIG. 16 is a characteristic diagram showing the spring characteristics of the auxiliary load means shown in FIG. 15, and FIGS. 17 to 19 are cross sections showing other examples of the auxiliary load means according to the present invention. It is a diagram. DESCRIPTION OF SYMBOLS 1... Shaft, 7... Piston, 10... Swash plate, 30... Spool, 40... Slide part, 50.
60... Working chamber, 900... Spring means forming auxiliary load means.

Claims (2)

[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, and within the cylinder chamber. An actuation formed between a piston that is slidably disposed and reciprocates within the cylinder chamber in response to the rocking motion of the swash plate, and an inner surface of the cylinder chamber at each of both ends of the piston. a spool disposed coaxially with the shaft and displacing the rotational support position of the swash plate in the axial direction of the shaft; and a maximum capacity position where the spool is positioned at a maximum inclination angle of the swash plate; displacing the swash plate in the axial direction of the shaft between a minimum capacity position that minimizes the inclination angle of the spool, and displacing the rotational center position of the swash plate in the axial direction of the shaft by the displacement of the spool; a control means for displacing the inclination angle of the swash plate; and an auxiliary load means for suppressing displacement of the spool toward the maximum capacity position; 1. A variable capacity swash plate type compressor, characterized in that the variable capacity swash plate compressor is configured to increase non-linearly as the capacity is displaced toward a maximum capacity position. (2) The variable capacity type according to claim 1, wherein the auxiliary load means is comprised of a nonlinear spring means that is locked at both ends, and a locking force is transmitted to the spool from one end of the spring means. Swash plate compressor. (3) The auxiliary loading means is
2. The variable displacement swash plate compressor according to claim 1, further comprising a plurality of spring means arranged in series.