JPH04148083A - Variable displacement type swash type compressor - Google Patents

Abstract

PURPOSE:To remarkably reduce the thrust load so as to enhance the durability of a compressor by setting the rotational center position of a coupling means for coupling a shaft and a swash plate, in a direction away from a working chamber in a compressed condition, off from the cross point between the center axis of the shaft and the center axis of the swash plate. CONSTITUTION:A shaft 1 rotating by a drive force from an engine is rotatably journalled to a front cylinder block 5 and a spool 30, and a thrust force exerted to this shaft 1 is received by the front cylinder block 5 through the intermediary of a thrust bearing 15. The rear end of the shaft is slidably inserted in a support part 405, and a thrust force acting upon the support part 405 is received by the spool 30 through the intermediary of a thrust bearing 14. The support part 405 is formed in a substantially cylindrical shape, and a pin 407 serving as a coupling means is projected from the outer periphery thereof. The center axis of pin 407 is set to a position off from the center axis of the shaft. A swash plate 10 is coupled to this pin 407.

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.

[Prior art and its problems] As a variable displacement swash plate type compressor, a piston is arranged so as to straddle the swash plate, and a first piston is provided on one end surface of the piston.
In a so-called swash plate type compressor that forms a working chamber and a second working chamber on the other side of the piston, the inclination angle of the swash plate is varied and at the same time the center of rotation of the swash plate is shifted. On the chamber side, the top dead center position is controlled to be almost constant regardless of changes in the tilt angle of the swash plate, and on the second working chamber side, the dead volume increases according to the tilt angle of the swash plate. This method has already been proposed by the present inventors.

The present invention relates to improvements to this type of variable displacement swash plate compressor, and in particular to improving the durability of the thrust bearing that supports the load in the thrust direction of the swash plate.

That is, according to the studies of the present inventors, in the conventional variable displacement swash plate type compressor, a large thrust load is applied to the thrust bearing in combination with displacing the inclination angle and rotation center position of the swash plate. As a result, it was confirmed that the durability of the thrust bearing was an issue.

Here, if the durability of the thrust hair ring is impaired, the smooth rotation of the swash plate compressor is also impaired, which shortens the life of the compressor as a whole.

[Problems to be Solved by the Invention] The present invention was devised in view of the above points, and an object of the present invention is to extend the life of a thrust bearing of a variable capacity swash plate compressor.

〔composition〕

In order to achieve the above object, the present inventors conducted various studies on the fluctuation of the thrust load peculiar to the swash plate type variable displacement compressor. In other words, in this type of variable displacement compressor,
The swash plate is swingably engaged with the shaft via a connecting means such as a bottle, and by changing the position of the connecting means of the bracket in the axial direction of the shaft, the rotational center position and inclination angle of the swash plate are changed. are made to fluctuate at the same time. Therefore, during operation of a swash plate compressor, the distance from the momentary rotation center position of the swash plate during the compression process to the rotation center position of the connecting means is important in determining the moment of force inside the swash plate compressor. It becomes a factor.

A force that balances this moment is applied to the spool, and this acts as a thrust load. Therefore, in order to reduce the absolute value of the force without changing the moment, the distance from the rotation center of the coupling means to the instantaneous rotation center must be adjusted. All you have to do is make it longer.

Therefore, in the swash plate compressor of the present invention, the rotation center position of the coupling means is arranged at a position farther from the first working chamber in the compression process than the intersection of the central axis of the shaft and the central axis of the swash plate. This configuration is adopted.

[Operation]

When the compressor is in operation, compression reaction force by the piston is transmitted to the swash plate via the piston.

Further, in this state, a predetermined pressure is applied to the spool in order to maintain the inclination angle of the swash plate or to vary the inclination angle of the swash plate. In this case, the reaction force of the fluid applied to the piston is apparently concentrated at the instantaneous rotation center A located on a perpendicular line extending perpendicularly to the swash plate. On the other hand, the pressure applied to the spool is apparently concentrated on the center of rotation of the coupling means.

Therefore, in the compressor of the present invention, the rotational center position of the coupling means is arranged offset from the shaft center, so the apparent distance from the instantaneous rotational center point A above to the rotational center of the swash plate is long. It will become. As a result, the force generated inside the compressor as a thrust load can be reduced while maintaining the same moment.

〔Effect of the invention〕

Therefore, according to the variable displacement swash plate compressor of the present invention, the thrust load generated inside the compressor can be significantly reduced while achieving variable displacement operation equivalent to that of the conventional compressor. As a result, the durability of the compressor can be improved and its life can be extended. Furthermore, as the thrust load generated inside the compressor is reduced, the thickness of the compressor housing can be reduced, and in this case, the overall weight of the compressor can also be reduced.

〔Example〕 An embodiment 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. Aluminum alloy front housing 4 front side plate 8. Suction valve9. Front cylinder block 5
．． Rear cylinder block 6° intake valve 12. The rear side plate 11 and the rear housing 13 are integrally fixed by through bolts 16 and form the outer shell of the compressor. Cylinder blocks 5 and 6 have cylinders 64 and 65
The cylinders 64 and 65 are formed parallel to each other at five locations. A shaft 1, which rotates under the driving force of an automobile engine (not shown), is rotatably supported by a front cylinder block 5 and a spool 30 via bearings 2 and 3, respectively. Further, the thrust force applied to the shaft 1 (force acting in the left direction in the figure) is received by the front cylinder block 5 via the thrust bearing 15.

The rear end of the shaft 1 is slidably inserted into a support portion 405, and the support portion 405 is rotatably supported by the spool 30 via a bearing 3.

Note that a spring 308 is disposed between the rear end of the shaft 1 and the support portion 405 to apply a preload to the spool 30 toward the right side in the figure. Further, the thrust force acting on the support portion 405 (the force acting in the right direction in the figure) is received by the spool 30 via the thrust bearing 14. The spool 30 is connected to the cylindrical portion 66 of the rear cylinder block 6 and the rear housing 13.
It is arranged so as to be able to slide in the axial direction within the cylindrical portion 135 of.

The support portion 405 has a substantially cylindrical shape as shown in FIG. 2, and the shaft 1 is inserted into the through hole 401 inside. Further, in this example, two pins 407 are formed to protrude perpendicularly from the outer periphery of the support portion, and these pins act as a connecting means, and the straight line of the central axis of this pin 407 is the central axis of the shaft, as will be described later. It's off than that. A cylindrical bush 409 is arranged on the bottle 407.

Furthermore, as shown in FIG.
07 is formed, and a support portion 40 is formed in this recess 107.
5 is inserted.

As shown in FIGS. 4 and 5, a holding groove 106 into which the bushing 406 is inserted is formed in the recess. The structure is such that the bushing 409 is held between the holding plate 108 while the bushing 409 is inserted into the holding groove 106. Note that the holding plate 108 is tightened and fixed to the swash plate lO.

Therefore, the swash plate IO is swingably connected to the support portion 405 via the pin 407.

Moreover, its swing position is displaced downward in FIG. 3 from the central axis of the shaft l.

A slit 105 is formed on the front side surface of the swash plate lO, and a flat plate portion 165 is formed on the shaft 1. 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 a cylinder 64 of the front cylinder block 5 and a cylinder 65 of the rear cylinder block 6. As mentioned above, the shoes 18 and 19 are slidably attached to the swash plate 10, and 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. Note that the shoes 18 and 19 are formed so that their outer surfaces lie on the same spherical surface when assembled on the swash plate lO.

A long groove 166 is provided in the flat plate portion 165 of the shaft 1, and a pin passage hole 109 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 long groove 166 of the shaft 1 by the pin 80 and the retaining ring. This long groove 166
The inclination of the swash plate changes depending on the position of the pin 8o inside, and as the inclination changes, the position of the center of the swash plate also changes. That is, in the first 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 working chamber 60 of the piston 7 hardly changes and remains dead. The long grooves 166 are provided so that substantially no increase in volume occurs. On the other hand, in the second 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.

Strictly speaking, the long groove 166 has a curved shape, but in actual formation, it can be formed proximally with a substantially straight long groove.

Further, in this example, the long grooves 166 are arranged on the axis of the shaft l so that the shape of the flat plate portion 165 does not become excessively large due to the formation of the long grooves 166.

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 discharge chamber 90.
The discharge port 24 is opened and closed by the discharge valve 23. The discharge valve is fixed to the front side plate 8 and the rear side plate 11 by bolts (not shown) together with a valve holder (not shown). In the figure, 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 suction bell 12.

In the figure, 500 indicates the signal pressure introduced into the control pressure chamber 200.
This is a control valve that continuously controls the pressure within the discharge space 93 and the pressure within the suction space 74.

The operation of the compressor with the above configuration will be described.

When a t4Wa clutch (not shown) is connected and driving force from the engine is transmitted to the shaft 1, the compressor is started.

When the compressor is started after being stopped for a long time, there is no pressure difference inside the compressor.

Therefore, the pressure within the control pressure chamber 200 is not significantly different from the pressure within the suction space 74. In this way, no pressure difference occurs before and after the spool 30. That is,
At startup, no load is applied in the direction of tilting the swash plate 10 with respect to the support portion 405, and the spool 30 is displaced to the right in the figure due to the set load of the spring 308, and the swash plate 10 is tilted. The corners are kept to a minimum.

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

The refrigerant gas sucked through the suction port (connected to the evaporator of the refrigeration cycle) enters the central suction space 70, then passes through the suction passage 71.73 and enters the front and rear suction chambers 72.74. enter. Thereafter, during the suction stroke of the piston 7, the piston 7 is sucked into the working chamber 50°60 through the suction port 25 through the suction valve 12. The sucked refrigerant gas is compressed in a compression stroke, and when compressed to a predetermined pressure, the discharge valve 23 is pushed open through the discharge port 24 and discharged into the discharge chamber 90.93. 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 second working chamber 50 on the front side has a large dead volume, the compression ratio is smaller than that of the first working chamber 60 on the rear side, and the pressure of the refrigerant gas in the second working chamber 50 is the pressure in the discharge space ( The discharge pressure of the first working chamber 60 on the rear side is lower than that of the first working chamber 60 (the discharge pressure is guided). Therefore, the action of sucking and discharging refrigerant gas in the second 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, if the compressor capacity required by the refrigeration cycle is high, it is necessary to increase the discharge capacity of the compressor.

Here, it is known that the capacity required of the compressor, that is, the cooling load, has a correlation with the suction side pressure of the compressor. That is, when the cooling load is high and the compressor requires a large capacity, the suction side pressure increases due to superheating in the evaporator. Conversely, when the cooling load is small and the discharge capacity required of the compressor is small, there is no large superheat in the evaporator and the suction side pressure is low.

In the control valve 500 of this example, when the pressure on the suction side becomes low, the diaphragm 503 overcomes the biasing force of the spring 502 and is displaced, and the valve body 504 seats on the valve seat 506 and connects the signal pressure passage 402 to the low pressure The introduction passage 403 is cut off. Therefore, the pressure within the control pressure chamber 200 increases.

When the pressure in the discharge space 93 increases as the compressor is started, the pressure in the control pressure chamber 200 also increases in response to this pressure increase.

Therefore, the force acting on the spool 30 to the left in the figure 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 and the resultant force of the spring 308, the spool 30 gradually begins to move to the right in the figure. Then, due to the action of the long groove 166 of the shaft l and the pin 80, the swash plate 10 moves its center of rotation (pin 407 on the support part 405) to the left in the figure and increases its inclination. As the internal pressure increases, the spool 30 moves to the left in the figure until its shoulder 305 hits the rear side plate 11, achieving the maximum capacity state. This is the situation shown in FIG.

In the state shown in FIG. 1, refrigerant gas drawn from the suction boat enters the central suction space 70 and flows into the suction chambers 72 and 74, respectively, through the suction passage.

In the suction stroke, the suction valves 9 and 12 are connected to the suction port 25.
, enters the working chambers 50 and 60, respectively, and is then compressed with the displacement of the piston 7, enters the discharge space from the discharge port 24 via the discharge valve 23, passes through the discharge passage, is discharged from the discharge port, and is discharged from the outside. They are joined by piping. In this state, both the working chamber 50 and the working chamber 60 perform the action of sucking in and discharging refrigerant gas.

After the compressor starts operating, when the cooling load decreases and the pressure on the suction side decreases again, the control valve 5
00 controls the pressure output to the signal compression [1402. That is, the suction pressure introduced through the low pressure passage 403 is outputted as a pressure signal as appropriate.

FIG. 6 is a schematic diagram showing the pressure state during operation of the compressor. A force equal to the projected area of the spool multiplied by the pressure difference between the internal pressure of the signal pressure chamber 200 and the internal pressure of the suction chamber 74 is applied to the spool 30 in the axial direction. On the other hand, a reaction force due to the compression of the piston 7 is applied to the swash plate 10. Here, since the angle of the swash plate 10 is maintained by the engagement between the long groove 166 and the pin 80, the instantaneous center A of the pressure generated on the swash plate 10 is a line extending perpendicularly from the swash plate 10. It is determined as the intersection of the pin 80 and a line extending perpendicularly from the engagement surface of the long groove 166.

The moment M generated around the center A at this instant, which tends to tilt the swash plate, is balanced by the axial force generated on the spool 30. Further, the axial force generated on the spool 30 is in turn the axial force Fb applied to the pin 407,
matches. Therefore, if the distance from the instantaneous center A to the pin 407 is , then the above moment MA and the axial load Fb.

The relationship with MA/L, = Sc (Pc-Ps) = Fb.

It will be calculated as

As is clear from this equation, in order to reduce the load Fb in the thrust direction applied to the pin, it is sufficient to increase the distance La to the instantaneous center A. In other words, the position where the pin 407 is placed can be The swash plate 10 may be displaced in a direction away from the first working chamber 60, which is in the middle of compression (downward in FIG. 1), from the intersection of the central axis of the swash plate 10 and the central axis of the swash plate 10.

FIG. 7 is a schematic diagram showing the load applied to the guide bin 80 and the load applied to the front thrust bearing 15. As mentioned above, in the compressor according to the present invention, pin 4
07 is shifted downward in the figure from the central axis of the shaft 1, the axial load Fb generated on the pin 407
, can be made smaller. And along with this, guide pin 8
The load Fp applied to the front thrust bearing 15 and the load Fb applied to the front thrust bearing 15 can also be reduced. This is because, if we consider that the position of the pin 407 is B and the moment around this point B is M, then the load FP applied to the guide pin 80 is
and Fb generated in the front thrust bearing 15
, is Fbr=Fp°si°=,2o. 3. . -1),
Note that Lp represents the distance between the guide pin 80 and the pin 407, where Mm is considered to be approximately the coupled moment, so even if the pin 407 is displaced from the central axis of the shaft, the amount of displacement δ, When it is small, it can be considered that there is almost no difference. On the other hand, this displacement δ
7, the distance between the pin 407 and the guide pin 80 changes. That is, in FIG. 7, Lp is Lp.

becomes larger than Moreover, the elevation angle of the line connecting pin 407 and guide pin 80 is β. becomes larger than Here, since the elevation angle α of the long groove 166 is clearly greater than or equal to 0 degrees and less than or equal to 90 degrees, cos (α−β) > cos
The relationship is (α-βO). From this relationship, guide pin 8
Fp added to 0 is the load FP when the amount of shift δ is 0.

Obviously smaller, as well as thrust bearing 1
The load Fbf generated at 5 is also clearly smaller than the load Fbto when the shift position δ7 is O.

In this way, by shifting the position of the pin 407 downward in the figure, not only the load generated on the pin 407 but also the load generated on the guide pin 80 and the load generated on the front side thrust bearing 15 can be reduced.

Note that when the compressor is in the state shown in FIG. 1, that is, in a state in which it operates at 100% capacity, the load generated on the front thrust bearing 15 and the load generated on the rear thrust bearing 14 are equal.

FIG. 8 shows the position of the pin 407 necessary to exhibit the above-mentioned thrust load reduction effect.

When controlling the discharge capacity of the compressor by changing the inclination angle of the swash plate 10, if the behavior of the swash plate 10 varies depending on the position of the pin 407, good control cannot be achieved. must always be on the AOW,
Here, Ao is the smoothness amount δ.

shows the instantaneous center of the piston 7 when is O. W indicates the intersection of perpendicular lines drawn from the instantaneous center A0 to the swash plate 10.

Further, when the position of the pin 407 is shifted, the distance MLm from the instantaneous center A to the center of the pin 407 is the distance before changing the position of the pin 407. If the length is not longer, the effect of the present invention cannot be achieved, and the position of the pin 407 for that purpose is the eighth position.
It must be below the dashed line in the figure. In addition, the broken line in FIG. 8 shows the position of the pin 407 when the displacement amount δ is 0.

Furthermore, in order to be able to change the angle of inclination of the swash plate 10, the position of the pin 407 must be closer to the spool 30 than the position of the guide pin 80, that is, on the left side in FIG. In order to displace the inclination angle of the pin 407, the position of the pin 407 must be on the right side of the point W. Therefore, the position where the pin 407 can be placed is the eighth point.
This is within the range delimited by diagonal lines in the figure. According to the studies conducted by the present inventors, it has been confirmed that when the pin 407 is placed within the range delimited by diagonal lines in Fig. 8, the moment MA around the instantaneous center point A hardly changes within this range. has been done.

By the way, the load Fp generated on the guide pin 80 is
If the position of 7 is shifted to the right in Fig. 8, the guide pin load will obviously decrease, but if the pin 407 is shifted to the left in Fig. 8, the guide pin load may increase. .

Therefore, in order to reduce the thrust load of the compressor as a whole, the range δ in FIG. 8 is particularly desirable even if it is within the range delimited by the diagonal lines in FIG.

In addition, when shifting the position of the pin 407 in the axial direction,
It is necessary to vary the inclination angle of the inclined groove 166 so that the instantaneous center A is always on the line A, -W as described above.

Figure 9 shows the shift position δ7 of the pin 407 and the thrust load Fb.
The zero condition showing the relationship is that the discharge pressure of the compressor is 25 kg/a
iG, the suction pressure Ps is 3 kg/dG, the compressor rotation speed is 700 rpm, and the compressor discharge capacity is 100%. In this state, the thrust load Fbf generated on the front thrust bearing 15 and the rear Since the thrust load Fbr generated on the side thrust bearing 14 is equal, the thrust load can be reduced by about 20% when the slenderness Σ is taken as lO■.

FIG. 10 shows the load Fp generated on the guide pin 80 under the same conditions as FIG. 9. In this case as well, a load reduction of 20% is observed when the shift amount δV is set to 10-. Further, the load applied to the guide pin 80 can be reduced by about 10% when the axial clearance amount Σ is approximately lO-.

In this way, in the compressor of this example, the position of the pin 407 is shifted from the central axis of the shaft 1, so that the thrust load can be significantly reduced.

Note that FIG. 11 shows another embodiment of the present invention.

In this example, a spherical support part 405 is used instead of a pin as a connecting means.
is adopted. That is, the shaft 10 has a spherical support portion 40
5 so as to be swingably supported. The center position of this spherical support portion 405 is offset from the center axis of the shaft 1, similar to the pin shown in FIG. 1 described above.

[Brief explanation of the drawing]

Fig. 1 is a cross-sectional view showing the first embodiment of the compressor of the present invention;
The figure is a perspective view of the supporting part shown in Fig. 1, Fig. 3 is a sectional view showing a state in which the supporting part shown in Fig. 1 and the shaft are engaged, and Fig. 4 is a front view of Fig. 3 seen from the Figure 5 shows V-■ in Figure 3.
6 is a schematic diagram showing the pressure state of the compressor shown in FIG. 1, and FIG. 7 is a guide pin 80 of the compressor shown in FIG. 1.
A schematic diagram showing the surrounding pressure state, FIG. 8 is a schematic diagram showing possible positions of pins in the compressor shown in FIG. 1, and FIGS. 9 and 10 are graphs showing examples of effects of the compressor shown in FIG. 1. , FIG. 11 is a sectional view showing another embodiment of the compressor of the present invention. 1...Shaft, 5.6...Cylinder block, 7
···piston. O... Swash plate. 30...Spool. 5...Support part. 407...Pin.

Claims (1)

[Claims] A cylinder block having a cylinder chamber therein, a shaft rotatably disposed within the cylinder block, and a shaft connected to the shaft via a connecting means so as to be swingable around the connecting means. 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 formed between the inner surface of the cylinder chamber and for sucking, compressing, and discharging fluid; and a working chamber that engages with the connecting means to displace the position of the connecting means in the axial direction of the shaft, and the swash plate. In the first working chamber formed on one side of the piston, the top dead center position is substantially constant regardless of the inclination angle of the swash plate. A second working chamber formed on the other side includes a spool for controlling a dead volume to vary according to the inclination angle of the swash plate, and the rotation center position of the connecting means is aligned with the center line of the shaft. A variable capacity swash plate type compressor, characterized in that the variable capacity swash plate compressor is provided offset from the intersection of center lines of the swash plate in a direction away from a first working chamber in a compressed state.