JPH08319942A - Double end swash plate compressor - Google Patents

Abstract

PURPOSE: To make reduction of a consumption power compatible with suppression of vibration by a method wherein a plurality of disc springs are located in parallel combination with a thrust bearing between a pressure receiving seat formed at the boss part of a swash plate and a pressure receiving seat formed at the support part of a cylinder block. CONSTITUTION: When the two boss parts of a swash plate 5 are nipped by a cofastening member, containing two cylinder blocks 2 and 3, through thrust bearings 6A and 6B and disc springs 7 and 8, an axial fastening margin is absorbed by a buffering function based on resilient deformation adherent to the disc springs 7 and 8. During load speed operation, the disc springs 7 and 8 are easy to effect relative rotation as a complete different body and since support rigidity of the swash plate 5 based on a total of the repulsion forces of the disc springs 7 and 8 is low, a consumption force during low speed operation is reduced. Meanwhile, during high speed operation, the disc springs 7 and 8 are integrally formed and relative rotation is hard to make and since support rigidity of the swash plate 5 based on a repulsion force exceeding a total of the repulsion forces of the disc springs 7 and 8 is high, vibration and noise of the swash plate 5 are suppressed.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a double-headed swash plate compressor, and more particularly to improvement of a thrust bearing portion for receiving an axial load of the swash plate.

[0002]

2. Description of the Related Art Most of the double-headed swash plate type compressors used for vehicle air conditioning have a pair of cylinder blocks 1 as disclosed in JP-A-64-63669 (see FIG. 7).
Swash plate 1 on drive shaft 11 supported by 0A, 10B
The thrust bearings 13, 13 mounted on the front and rear of the swash plate 12 are attached to the cylinder blocks 10
A, 10B and both housings 1 for closing the outer ends thereof
The bolts 4 and 15 are clamped together by the through bolt 16 so as to be sandwiched. Then, in order to absorb the axial interference caused by the co-fastening by elastic deformation generated in the inner ring and the outer ring of the thrust bearing 13, the inner ring 13a has an annular pressure receiving member formed on both boss portions of the swash plate 12 in the vicinity of its outer diameter. Seat 12a
On the other hand, the outer ring 13b is configured to come into contact with the annular pressure receiving seat 10a formed in the support portion of the cylinder blocks 10A, 10B in the vicinity of the inner diameter thereof.

[0003]

However, in the structure in which the absorption of the interference in the axial direction due to the co-fastening is dependent only on the elastic deformability of the thrust bearing (race) as described above,
If the interference is set high to increase the support rigidity of the swash plate, the power consumption increases.On the other hand, if the interference is set low and the support rigidity of the swash plate is decreased, the vibration due to the displacement of the swash plate is different. This causes a problem of inducing sound (noise).
Needless to say, such a problem is not limited to the above-described conventional configuration, but similarly exists in a configuration in which another elastic member is required to absorb interference.

The present invention aims to solve the technical problems of reducing power consumption and suppressing vibrations.

[0005]

[Means for Solving the Problems]

(1) In the double-headed swash plate compressor according to claim 1, the outer ends of a pair of front and rear cylinder blocks are closed by front and rear housings, and a drive shaft is provided in a common central shaft hole of each cylinder block. Are supported through a pair of front and rear radial bearings, and thrust is formed between the pressure receiving seat formed on the boss portion of the swash plate that cooperates with the drive shaft and the pressure receiving seat formed on the bearing portion of each cylinder block. In a double-headed swash plate compressor having bearings sandwiched between the pressure receiving seat formed on the boss portion of the swash plate on at least one side and the pressure receiving seat formed on the support portion of the cylinder block, It is characterized in that at least two disc springs are interposed in parallel combination together with the thrust bearing.

(2) A double-headed swash plate compressor according to a second aspect of the present invention is the double-headed swash plate compressor according to the first aspect, wherein the disc springs are allowed to move in a radial direction and to rotate relative to each other. It is characterized by the presence of a detent to suppress. (3) The double-headed swash plate compressor according to claim 3 is the double-headed swash plate compressor according to claim 2, characterized in that the detent is constituted by a rough surface provided on the facing surface of each disc spring. To do.

(4) The double-headed swash plate compressor according to a fourth aspect is the double-headed swash plate compressor according to the second or third aspect, wherein the detent is formed by an uneven surface provided on the opposing surface of each disc spring. It is characterized by being (5) The double-headed swash plate type compressor according to claim 5 is characterized by
In the double-headed swash plate compressor described in 3 or 4, each disc spring has flat seats at inner and outer peripheral ends of conical surfaces facing each other.

(6) A double-headed swash plate compressor according to a sixth aspect is the double-headed swash plate compressor according to the first, second, third, fourth, fifth or sixth aspect, wherein each disc spring is formed on a boss portion of the swash plate. It is characterized in that it is located on the side of the pressure receiving seat. (7) The double-headed swash plate compressor according to claim 7 is the compressor according to claim 1,
In the double-headed swash plate compressor described in 3, 4, 5 or 6, the thrust bearing is a compound type slide bearing.

[0009]

[Action]

(1) In the compressor according to the first aspect, the interference at the time of assembly is set to be relatively small. Also, at low speed operation,
Since each disc spring located in the radial direction among the disc springs inserted in parallel combination acts only a centrifugal force which is not so large on the one located in the radial direction among them, each disc spring is a completely separate body. Is relatively easy to rotate,
The swash plate is supported based on each repulsive force. Therefore, during low speed operation, the supporting rigidity of the swash plate based on the total repulsive force of each disc spring is low. Therefore,
The power consumption during low speed operation is kept at a sufficiently low value.

During high-speed operation, among the disc springs that are interposed in parallel, the disc springs located in the radial direction exert a large centrifugal force on the disc springs located in the radial direction. The springs are in a state in which they cannot rotate relative to each other as an integral unit, and support the swash plate based on the repulsive force equal to or greater than the repulsive force of each spring. Therefore, during high-speed operation, the supporting rigidity of the swash plate is increased based on the repulsive force that is equal to or greater than the total repulsive force of each disc spring. Therefore, during high-speed operation, vibration of the swash plate and noise due to this vibration are suppressed well.

Further, the disc springs mounted in parallel combination move in the radial direction on the respective facing surfaces due to the vibration during low speed and high speed operation. At this time, a frictional force in a direction parallel to each of the opposed surfaces is generated between the opposed surfaces, and an axial component of the frictional force damps the vibration. Therefore, the vibration of the swash plate at the time of low speed and high speed operation and the noise due to this vibration are further suppressed.

Further, according to the disc spring, there is an advantage that the mounting space in the axial direction is small and the load management is easy. (2) During high-speed operation, it is considered that the disc springs are less likely to rotate relative to each other due to a large centrifugal force, and that a gap is unlikely to exist between the facing surfaces. For this reason, it is considered that a large frictional force is generated in the direction parallel to the opposed surfaces due to the vibration, and the vibration damping effect by the frictional force generated between the opposed surfaces of the disc springs is effective. However, when the disc springs are completely integrated with each other, the disc springs do not move in the radial direction, and a vibration damping action due to frictional force cannot be expected.

On the other hand, at low speed operation, the disc springs are likely to relatively rotate due to a small centrifugal force, and it is considered that there is a gap between the facing surfaces in the state of relative rotation. For this reason, the frictional force generated by the vibration is also small, and the damping action of the vibration cannot be considered to be very effective. In this respect, in the compressor according to the second aspect, the detents between the disc springs suppress the relative rotation of each other. Since this detent allows the radial movement of each disc spring, the radial movement of each facing surface due to vibration is possible, and the damping action of the vibration due to the frictional force is not impaired. Therefore, during low-speed operation, it is considered that the gap between the facing surfaces of the disc springs is unlikely to exist, so that a frictional force is reliably generated by the vibration, and the vibration damping action can be reliably exhibited. . However, depending on the anti-rotation means and the degree thereof, at the time of low speed operation, the disc springs become integrated due to the frictional force, and the supporting rigidity of the swash plate is increased despite the small centrifugal force. For this reason, it is necessary to set the rotation stopping means and degree in consideration of the range of the number of rotations that is normally used.

(3) In the compressor according to the third aspect, the unevenness of the rough surface provided on the facing surface of each disc spring has the function of the detent of the second aspect. Roughening is also relatively easy. (4) In the compressor according to the fourth aspect, the concavo-convex surface provided on the facing surface of each disc spring functions as the detent of the second aspect. (5) In the compressor according to the fifth aspect, since each disc spring has flat seats at the inner peripheral edge and the outer peripheral edge of the conical surfaces facing each other, no acute angle contact occurs. Therefore, even if the disc springs rotate relative to each other, the disc springs are less likely to deteriorate.
Further, even if each disc spring rotates relative to the thrust bearing or the pressure receiving seat of the swash plate or cylinder block, each disc spring, thrust bearing, swash plate or cylinder block is less likely to deteriorate.

(6) In the swash plate and the cylinder block, in order that the swash plate can reliably receive the momentum when the discharge pressure rises, the swash plate is usually relatively relatively compared with the cylinder block. Made to be hard. In this regard, in the compressor according to claim 6, wherein each disc spring is positioned on the pressure receiving seat side formed on the boss portion of the swash plate, the repulsive force of each disc spring is reliably received by the relatively hard swash plate. However, the swash plate hardly deteriorates.

(7) In the compressor according to the seventh aspect, the thrust bearing is a slide bearing, and there is no noise generated from the bearing itself like a rolling bearing, which is more advantageous in that it can be manufactured at an extremely low cost. This tendency is even greater if all thrust bearings are slide bearings. If the plain bearing is a composite type, relative rotation easily occurs between the inner ring (swash plate side) and the outer ring (cylinder block side) that make up the composite type, and the relative rotation is likely to occur between the swash plate and the inner ring or between the outer ring and the cylinder. Relative rotation with the block can be reduced. Therefore, in particular, as described above, the relatively soft cylinder block is less likely to deteriorate.

[0017]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Embodiments 1 to 4 embodying the invention of claims 1 to 7 will be described below with reference to the drawings. Here, the overall structure of the compressor is the same as that of the conventional compressor, and detailed description thereof is omitted. (Embodiment 1) The compressor of Embodiment 1 embodies claims 1, 5 and 7. That is, in FIG. 1, reference numeral 1 denotes a drive shaft that is supported in a common central shaft hole of the front cylinder block 2 and the rear cylinder block 3 via a pair of front and rear radial bearings 4A and 4B. Thrust bearings 6A and 6B are provided in front of and behind the attached swash plate 5,
Both thrust bearings 6A and 6B are sandwiched between both boss portions of the swash plate 5 and bearing portions of both cylinder blocks 2 and 3 by joint tightening of through bolts as in the conventional case. In addition,
The swash plate 5 and the both cylinder blocks 2 and 3 are different from the two cylinder blocks 2 and 3 in order to ensure that the swash plate 5 receives the momentum when the discharge pressure rises. It is considered to be relatively hard.

The thrust bearings 6A and 6B, which are the characteristic features of this embodiment, have inner rings 61 and 63 on both boss sides of the swash plate 5, respectively.
And an outer ring 62, 64 on both cylinder blocks 2 and 3 side. Inner ring 61 and outer ring 6
No. 4 is a base material surface made of SPCC coated with a fluororesin. Also, the outer ring 62 and the inner ring 63
Is made up of SUJ2. In addition, as another embodiment, it is also possible to adopt a composite type sliding bearing having three or more sheets.

A swash plate 5 for holding the thrust bearing 6A at the front portion.
Of the front boss and the support portion of the front cylinder block 2 are formed with flat pressure receiving seats 5a, 2a facing each other, and the inner ring 61 and the outer ring 62 are respectively formed with the pressure receiving seats 5a, 5a.
The thrust bearing 6A is stably and rigidly sandwiched by being closely fitted to the entire surface 2a. In addition,
In another embodiment, the outer ring 62 is attached to the front cylinder block 2
It is also possible to stop the rotation at the bearing part.

On the other hand, flat pressure receiving seats 5b, 3b facing each other are also formed on the rear boss of the swash plate 5 for sandwiching the rear thrust bearing 6B and the support portion of the rear cylinder block 3. The outer ring 64 is closely fitted to the pressure receiving seat 3b almost entirely, and there is no space between the inner ring 63 and the pressure receiving seat 5b.
The disc springs 7 and 8 are interposed in parallel.
Thus, the disc springs 7 and 8 are located on the pressure receiving seat 5b side formed on the boss portion of the swash plate 5. The outer ring 64 can also be prevented from rotating on the support portion of the rear cylinder block 3.

As shown in FIGS. 2 and 3, each of the disc springs 7 and 8 has flat seats 7b and 8b at the inner peripheral ends of conical surfaces (opposing surfaces) 7a and 8a facing each other, and the conical surface 7
Flat seats 7c and 8c are also provided at the outer peripheral ends of a and 8a.
The conical surfaces 7a and 8a of the disc springs 7 and 8 are not roughened to prevent rotation. In the compressor of the first embodiment configured as described above, as shown in FIG. 1, both bosses of the swash plate 5 are thrust bearings 6A, 6B and disc springs 7,
When clamped by a clamping member including both cylinder blocks 2 and 3 via 8, the axial interference is skillfully absorbed by the cushioning function based on the elastic deformation inherent to the disc springs 7 and 8. The tightening and tightening force can be adjusted easily and stably.

Further, as shown in FIG. 2, between the disc springs 7 located in the radial direction and the disc springs 8 located in the radial direction, the angular velocities of ω and the disc springs 7 in contact with each other are set. If the distance from the axis to the central position between the conical surface 7a and the conical surface 8a of the disc spring 8 is r and the mass of the disc spring 7 is m, a centrifugal force of F = mrω ² acts. From the disc spring 7 to the disc spring 8, the conical surface 7a in the centrifugal force F and the component force Fn in the direction perpendicular to each other act as forces to press each other, and the component force Ft in the centrifugal force F in the direction parallel to the conical surface 7a is exerted. They act as forces that shift each other.

For this reason, since the angular velocity ω is still small during low speed operation, the disc spring 7 acts on the disc spring 8 only by the component force Fn which is not so large.
Is a completely separate body and relatively easy to rotate. Therefore, during low-speed operation, the disc springs 7 and 8 support the swash plate 5 based on their repulsive forces, and the supporting rigidity of the swash plate 5 based on the total repulsive force of the disc springs 7 and 8 is low. Has been done.

According to the test results of the inventors, when the elastic coefficient of the disc spring 7 is K ₁ and the elastic coefficient of the disc spring 8 is K ₂ , these and the elastic modulus K of the disc springs 7 and 8 as a whole are In the meantime, it became clear that the number 1 is almost established.

[0025]

[Equation 1]

Therefore, the power consumption during low speed operation is kept at a sufficiently low value. During high-speed operation, since the angular velocity ω is large, the disc spring 7 acts on the disc spring 8 with a large component force Fn, so that the disc springs 7 and 8 are in a state in which they cannot rotate relative to each other. There is. Therefore, at the time of high speed operation, the disc springs 7 and 8 support the swash plate 5 based on the repulsive force of each repulsive force or more, and based on the repulsive force of the total repulsive force of each disc spring 7 or 8 or more. The supporting rigidity of the swash plate 5 is increased.

Here again, according to the test results of the inventors,
It has been clarified that the equation 2 is almost established.

[0028]

[Equation 2]

Therefore, during high speed operation, vibration and noise of the swash plate are well suppressed. In addition, these numbers 1, 2
Is considered to be caused by the fact that the elastic coefficient of a general plate material is proportional to the cube of the plate thickness. For this reason, the disc springs 7 and 8 are in a state of being easily rotated relative to each other as a completely separate body at the time of low-speed movement, while the disc springs 7 and 8 are at the time of high-speed operation.
It is clear that the components are in a state in which relative rotation is difficult as a unit, and it is considered that the above action is caused by this.

Further, as shown in FIG. 3, each disc spring 7, 8 is
Due to vibration during low speed and high speed operation, each conical surface 7a,
8a moves in the radial direction. At this time, each conical surface 7a, 8
A frictional force R in a direction parallel to each conical surface 7a, 8a is generated between a. For example, the friction coefficient between the conical surfaces 7a and 8a is μ,
Assuming that the force due to vibration is V, a frictional force R = μN is generated between the conical surfaces 7a and 8a due to the vertical reaction force N of the force V due to vibration. Of this frictional force R, the axial component R
₁ damps vibration. Therefore, the vibration of the swash plate 5 at the time of low speed and high speed operation and the noise due to this vibration are further suppressed. However, at the time of low-speed operation, the disc springs 7 and 8 are likely to rotate relatively due to the small centrifugal force F, and it is considered that there is a gap between the conical surfaces 7a and 8a. Since it is small, the damping effect of the vibration is not considered to be very effective. The radial component R ₂ of the frictional force R is considered to be added to the centrifugal force F.

Further, in this compressor, since the disc springs 7 and 8 are adopted, the mounting space in the axial direction is small,
There is an advantage that load management becomes easy. Further, in this compressor, since the disc springs 7 and 8 are located on the pressure receiving seat 5b side formed on the boss portion of the swash plate 5, the repulsive force by the disc springs 7 and 8 is relatively hard. The plate 5 reliably receives the swash plate 5, and the swash plate 5 is less likely to deteriorate.

In addition, since the thrust bearings 6A and 6B are compound type slide bearings, relative rotation easily occurs between the inner ring 61 and the outer ring 62 and between the inner ring 63 and the outer ring 64, and the swash plate 5 and Between the inner ring 61, between the outer ring 62 and the front cylinder block 2, between the swash plate 5 and the disc spring 7, the disc spring 7
And the disc spring 8, the disc spring 8 and the inner ring 63, and the outer ring 64 and the rear cylinder block 3 can be reduced in relative rotation. Therefore, the cylinder blocks 2 and 3 which are generally relatively soft are less likely to deteriorate.

Here, relative rotation may occur between the swash plate 5 and the disc spring 7, between the disc spring 7 and the disc spring 8 and between the disc spring 8 and the inner ring 63. In this compressor, Disc springs 7, 8
Has flat seats 7a, 8a, 7c, 8c, so that no acute angle contact occurs. Therefore, the swash plate 5, each disc spring 7,
8 and the inner ring 63 are less likely to deteriorate. As another embodiment, it is also possible to make the circular and outer peripheral ends of the conical surfaces 7a and 8a rounded with a sufficiently large radius of curvature.

Therefore, this compressor can exhibit excellent durability. Further, in this compressor, since the thrust bearings 6A and 6B are slide bearings, there is no noise generated from the bearing itself like a rolling bearing, and it is further advantageous in that it can be manufactured at extremely low cost. The both disc springs 7 and 8 can be arranged on the front boss side of the swash plate 5 with exactly the same configuration. However, as in the above-described embodiment, the both disc springs 7 and 8 can be arranged on the swash plate 5. Of the swash plate which is arranged only on one side (rear side) of the boss portion and the thrust bearing 6A on the other side (front side) is rigidly sandwiched, the swash plate which is likely to be generated by mutual cushioning of the spring elements arranged on both front and rear sides. The unstable vibration of No. 5 can be suppressed together. (Embodiment 2) The compressor according to Embodiment 2 is claimed in claims 1 to 3, 5 to 7.
Is embodied in. Here, since the external configuration of the compressor is the same as that of the first embodiment, the different configuration, operation and effect will be described in detail with reference to FIGS.

That is, in this compressor, the conical surfaces 7a, 8a, etc. of the disc springs 7, 8 shown in FIGS. 1 to 3 are roughened by shot peening to prevent them from rotating. In this compressor, the rough surfaces between the conical surfaces 7a and 8a of the disc springs 7 and 8 suppress relative rotation between them. Since the rough surface allows radial movement of the disc springs 7 and 8, the radial movement of the conical surfaces 7a and 8a is possible due to vibration, and the vibration damping action by the frictional force R is impaired. Absent.

Therefore, in this compressor, while maintaining low supporting rigidity of the swash plate 5 during low speed operation, the gap between the conical surfaces 7a and 8a of the disc springs 7 and 8 is maintained even during low speed operation. Since it is hard to exist, the frictional force R is surely generated by the vibration, and the vibration damping action can be surely exhibited. Therefore, with this compressor, vibration and noise of the swash plate 5 during low-speed operation, which tends to be particularly insufficient, can be suppressed even better.

Further, in this compressor, since the rotation stoppers can be formed on the disc springs 7 and 8 by the shot peening process, which is relatively easy to perform, a low manufacturing cost can be realized. (Third Embodiment) A compressor according to a third embodiment embodies claims 1 to 3. This compressor is the same as that of the first embodiment except that the thrust bearing, the disc spring and the like are different. Therefore, the same configurations are denoted by the same reference numerals, and detailed description of the same operation and effect will be omitted.

That is, in FIG. 4, the thrust bearing 6
C and 6D are inner rings 71 and 73 on both boss sides of the swash plate 5,
Outer rings 72, 74 on both cylinder blocks 2, 3 side, and rollers 7 held between each inner ring 71, 73 and outer rings 72, 74.
The roller bearing is composed of 5,76. Flat pressure receiving seats 5a, 2a facing each other are formed on the front boss of the swash plate 5 that holds the front thrust bearing 6C and the support portion of the front cylinder block 2. The inner ring 71 is the pressure receiving seat 5
a is almost entirely closely fitted, and as a result, the thrust bearing 6
C is stable and clamped rigidly. The three disc springs 21 to 23 are provided between the outer ring 72 and the pressure receiving seat 2a.
Are installed in parallel combination.

On the other hand, flat pressure receiving seats 5b, 3b facing each other are also formed on the rear boss of the swash plate 5 which holds the rear thrust bearing 6D and the support portion of the rear cylinder block 3. The inner ring 73 is also closely fitted to the pressure receiving seat 5b almost entirely, so that the thrust bearing 6D is also stably and rigidly held. Then, the outer ring 74 and the pressure receiving seat 3
Three disc springs 24 to 26 are also interposed in parallel with b.

As shown in FIGS. 5 and 6, each of the disc springs 24 to 26 has a conical surface (opposing surface) 24a, 2 facing each other.
5a, 25b, 26a, each conical surface 24
Surfaces a and the like are roughened by shot peening as a rotation stop. The same applies to the disc springs 21 to 23. In this compressor, as shown in FIG. 4, the thrust bearing 6C,
Since 6D is a roller bearing that is rigidly sandwiched, the thrust bearings 6C and 6D almost absorb the relative rotation between the swash plate 5 and the cylinder blocks 2 and 3. For this reason,
The swash plate 5 and the cylinder blocks 2 and 3 are less likely to deteriorate.

However, the rotation of the disc springs 21 to 26 may occur to some extent. Therefore, as shown in FIG.
A centrifugal force F acts between the disc spring 26 located radially inward of the disc spring 25 and the disc spring 25 located radially outward thereof, as in the first and second embodiments. Belleville spring 26 to Belleville spring 25
The conical surface 26a of the centrifugal force F and the component force Fn in the right angle direction act on each other as a force pressing each other.
The component force Ft in the parallel direction with the conical surface 26a at the point acts as a force that shifts from each other. The disc spring 25 located radially inward with respect to the disc spring 24 and the disc spring 24 located radially outward with respect to the disc spring 24.
The same applies between and. In addition, each disc spring 21-2
The same is true for 3.

Further, as shown in FIG. 6, each disc spring 24,
25 is the conical surface 2 due to vibration during low speed and high speed operation.
4a and 25a move in the radial direction. At this time, a frictional force R in a direction parallel to the conical surfaces 24a and 25a is generated between the conical surfaces 24a and 25a, as in the first and second embodiments. An axial component (not shown) of the frictional force R damps the vibration. The same is true between the disc springs 25 and 26. The same applies to the disc springs 21 to 23.

Therefore, also in this compressor, the same effects as those of the first and second embodiments can be obtained. (Fourth Embodiment) The compressor of the fourth embodiment embodies claims 1, 2, and 4. Here, there is no particular difference in the external configuration of the compressor from that of the third embodiment, so that FIG.
6, different configurations, operations and effects will be described in detail.

That is, in this compressor, the disc springs 21 to 26 shown in FIGS. As the disc springs 21 to 26, the disc springs with a wave described in JP-A-2-134430 can be used. Even in this compressor,
The same operation and effect as those of the third embodiment can be achieved.

[0045]

As described above in detail, since the inventions of the respective claims adopt the configurations described in the respective claims, the following excellent effects can be exhibited. In the compressor according to claim 1, the fluctuation ratio of the swash plate support rigidity to the load is changed in two ways depending on the operating speed by the centrifugal force, and the vibration damping effect is exerted by the friction force. It is possible to skillfully achieve both suppression of vibration.

Therefore, in the compressor according to the first aspect, it is possible to reduce the power consumption, that is, to improve the fuel consumption, and to suppress the vibration of the swash plate and the derived noise satisfactorily. Further, in the compressor according to the first aspect of the present invention, by adopting the disc spring, there is an advantage that the mounting space in the axial direction is small and the load management becomes easy.

In the compressors according to claims 2 to 4, the detent between the disc springs ensures the damping action of the vibration during the low speed operation. Therefore, the vibration and noise of the swash plate during the low speed operation are further improved. Can be suppressed. In particular, claim 3
In the described compressor, since the detent can be configured by a relatively easy work, a low manufacturing cost can be realized.

In the compressor of the fifth aspect, since each disc spring has a flat seat to prevent an acute angle contact, deterioration of the disc spring and the like can be prevented and excellent durability can be exhibited. In the compressor of claim 6, since each disc spring is positioned on the pressure receiving seat side formed on the boss portion of the swash plate, the repulsive force of each disc spring is reliably received by the relatively hard swash plate. The swash plate hardly deteriorates.

In the compressor of claim 7, since the thrust bearing is a slide bearing, it is more advantageous from the viewpoint of noise reduction and economical efficiency due to the simplification of the structure. Further, in the compressor of the seventh aspect, by adopting the composite type slide bearing, it is possible to prevent deterioration of the cylinder block and exhibit excellent durability.

[Brief description of drawings]

FIG. 1 is a cross-sectional view showing a main part of a double-headed swash plate compressor according to first and second embodiments.

FIG. 2 is an enlarged cross-sectional view of a main part of the double-headed swash plate compressor according to the first and second embodiments.

FIG. 3 is an enlarged cross-sectional view of a main part of the double-headed swash plate compressor according to the first and second embodiments.

FIG. 4 is a cross-sectional view showing a main part of a double-headed swash plate compressor according to Examples 3 and 4.

FIG. 5 is an enlarged cross-sectional view of a main part of the double-headed swash plate compressor according to the third and fourth embodiments.

FIG. 6 is an enlarged cross-sectional view of a main part of a double-headed swash plate compressor according to Examples 3 and 4.

FIG. 7 is a cross-sectional view showing the entire contents of a conventional double-headed swash plate compressor.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 ... Drive shaft 2 ... Front cylinder block 3 ... Rear cylinder block 5 ... Swash plate 5a, 5b, 2a, 3b ... Pressure receiving seat 6A, 6B, 6
C, 6D ... Thrust bearing 61, 63, 71, 73 ... Inner ring 62, 64, 7
5, 76 ... Outer ring 75, 76 ... Rollers 7, 8, 21-2
6 ... Disc springs 7a, 8a, 24a, 25a, 25b, 26a ... Cone surface 7b, 8b, 7c, 8c ... Plane seat

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Satoshi Konomura 2-chome Toyota-cho, Kariya city, Aichi stock company Toyota Industries Corporation (72) Inventor Hayato Ikeda 2-chome Toyota-cho, Kariya city, Aichi stock Company Toyota Industries Corporation (72) Inventor Naoto Kawamura 2-chome Toyota Town, Kariya City, Aichi Stock Company, Toyota Industries Corporation (72) Inventor Masanori Yokoi 2-chome Toyota Town, Kariya City, Aichi (72) Inventor Hiromi, 2-chome, Toyota-cho, Kariya city, Aichi Prefecture

Claims (7)

[Claims] 1. A front and rear housings are respectively closed at both outer ends of a pair of front and rear cylinder blocks.
A drive shaft is supported in a common central shaft hole of each cylinder block through a pair of front and rear radial bearings, and a pressure receiving seat formed on a boss portion of a swash plate that cooperates with the drive shaft and a support of each cylinder block. In a double-headed swash plate type compressor in which thrust bearings are sandwiched between the pressure bearing seat formed on the swash plate and the pressure receiving seat formed on the boss of the swash plate on at least one side, and the bearing portion of the cylinder block. A double-headed swash plate compressor in which at least two disc springs are interposed in parallel together with the thrust bearing between the pressure receiving seat formed in the above. 2. The double-headed swash plate compressor according to claim 1, wherein detents are provided between the disc springs so as to allow the radial movements of the disc springs and suppress the relative rotation of the disc springs. 3. The double-headed swash plate compressor according to claim 2, wherein the detent is constituted by a rough surface provided on the opposing surface of each disc spring. 4. The double-headed swash plate compressor according to claim 2, wherein the detent is constituted by an uneven surface provided on the opposing surface of each disc spring. 5. The double-headed swash plate type compressor according to claim 1, wherein each disc spring has flat seats at an inner peripheral edge and an outer peripheral edge of conical surfaces facing each other. 6. The disc springs are located on the pressure receiving seat side formed on the boss portion of the swash plate.
The double-headed swash plate compressor according to 3, 4, 5 or 6. 7. The double-headed swash plate compressor according to claim 1, wherein the thrust bearing is a composite type slide bearing.