VARIABLE DISPLACEMENT SWASH PLATE TYPE COMPRESSOR

TECHNICAL FIELD

The present invention relates to a variable displacement swash plate type compressor.

BACKGROUND ART

Japanese Patent Laid-Open No. 4-94470 discloses a conventional variable displacement swash plate type compressor (hereinafter simply referred to as a compressor). As shown in FIGS. 4 and 8 of this publication, the compressor includes a drive shaft, a housing, a swash plate, a link mechanism, double-headed pistons, a conversion mechanism, an actuator, and a control mechanism.

The housing rotatably supports the drive shaft. First cylinder bores are formed at a rear side in the housing, i.e., at one end side of the drive shaft. Second cylinder bores are formed at a front side in the housing, i.e., an opposite end side of the drive shaft. The diameter of the second cylinder bores is smaller than the diameter of the first cylinder bores. A swash plate chamber is formed in the housing between the first cylinder bores and the second cylinder bores.

A first thrust bearing is provided between the one end side of the drive shaft and the housing. A second thrust bearing is provided between the opposite end side of the drive shaft and the housing. The swash plate is disposed in the swash plate chamber in a state in which the drive shaft is inserted therethrough, and is rotatable by rotation of the drive shaft. The link mechanism is provided between the drive shaft and the swash plate and permits change of an inclination angle of the swash plate. Here, the inclination angle refers to an angle of the swash plate with respect to a direction perpendicular to an axis of the drive shaft.

The double-headed pistons define first compression chambers in the first cylinder bores and define second compression chambers in the second cylinder bores. Heads of the pistons on the side of the second cylinder bores have a smaller diameter than heads of the pistons on the side of the first cylinder bores so as to correspond to the diameters of the first and second cylinder bores. The conversion mechanism reciprocates the pistons in the first and second cylinder bores at a stroke corresponding to the inclination angle along with the rotation of the swash plate. The actuator is capable of changing the inclination angle. The control mechanism controls the actuator.

In this compressor, by reciprocation of the pistons in the first and second cylinder bores, refrigerant is introduced into the first and second compression chambers, compressed therein, and discharged therefrom. At this time, the first thrust bearing receives a thrust force generated by a compression reaction force in the second compression chambers, which acts on the drive shaft in a direction toward the one end side, and the second thrust bearing receives a thrust force generated by a compression reaction force in the first compression chambers, which acts on the drive shaft in a direction toward the opposite end side. In addition, in this compressor, since the actuator is capable of changing the inclination angle of the swash plate, discharge capacity of the refrigerant can be changed.

However, in the above-described conventional compressor, no difference is present between the first thrust bearing and the second thrust bearing. On the other hand, in this compressor, since the diameter of the second cylinder bores and the diameter of the piston heads on the side of the second cylinder bores are smaller than the diameter of the first cylinder bores and the diameter of the piston heads on the side of the first cylinder bores, pressure receiving area of the piston heads is larger on the side of the first cylinder bores than on the side of the second cylinder bores. Thus, as the discharge capacity increases, the thrust force toward the opposite end side of the drive shaft becomes larger than the thrust force toward the one end side of the drive shaft. Then, if the second thrust bearing is excessively deformed in the axial direction of the drive shaft, durability may be decreased. Furthermore, there is a risk that the drive shaft is excessively displaced toward the opposite end and a gap is created between the drive shaft and the first thrust bearing or between the first thrust bearing and the housing. As a result, the drive shaft wobbles, and this leads to an increase of vibration and noise caused by the vibration at the time of operation.

One possible solution to this problem may be to configure both the first and second thrust bearings to have a large spring constant. In this case, although the decrease in durability and the wobbling of the drive shaft as described above may be inhibited, when the discharge capacity is small, drag resistances of the first and second thrust bearings acting on the drive shaft increases. As a result, substantial power loss occurs in this compressor.

The present invention has been made in view of the conventional circumstances described above, and an object of the invention is to provide a variable displacement swash plate type compressor capable of exhibiting excellent durability and inhibiting vibration and noise caused by the vibration at the time of operation with less power loss.

SUMMARY OF THE INVENTION

A variable displacement swash plate type compressor of the present invention comprises: a drive shaft; a housing that rotatably supports the drive shaft, the housing including a first cylinder bore formed at one end side of the drive shaft, a second cylinder bore formed at an opposite end side of the drive shaft, and a swash plate chamber formed between the first cylinder bore and the second cylinder bore; a swash plate that is rotatable in the swash plate chamber by rotation of the drive shaft; a link mechanism that is provided between the drive shaft and the swash plate and permits change of an inclination angle of the swash plate in a direction perpendicular to an axis of the drive shaft; a double-headed piston that defines a first compression chamber in the first cylinder bore and defines a second compression chamber in the second cylinder bore; a conversion mechanism that reciprocates the piston in the first and second cylinder bores at a stroke corresponding to the inclination angle along with the rotation of the swash plate; an actuator capable of changing the inclination angle; and a control mechanism that controls the actuator. The second cylinder bore has a smaller diameter than the first cylinder bore. A first thrust bearing is provided between the one end side of the drive shaft and the housing so as to receive a thrust force acting on the drive shaft in a direction toward the one end side. A second thrust bearing is provided between the opposite end side of the drive shaft and the housing so as to receive a thrust force acting on the drive shaft in a direction toward the opposite end side. A second spring constant of the second thrust bearing is greater than a first spring constant of the first thrust bearing.

Other aspects and advantages of the present invention will be apparent from the embodiments disclosed in the following description and the attached drawings, the illustrations exemplified in the drawings, and the concept of the invention disclosed in the entire description and drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a cross-sectional view of a compressor according to Embodiment 1 at the time of minimum displacement.

FIG. 2 is a cross-sectional view of the compressor according to Embodiment 1 at the time of maximum displacement.

FIG. 3 is a schematic diagram showing a control mechanism of the compressor according to Embodiment 1.

FIG. 4 is an enlarged cross-sectional view of an essential part of the compressor according to Embodiment 1, showing first cylinder bores, a first thrust bearing, and the like.

FIG. 5 is an enlarged cross-sectional view of an essential part of the compressor according to Embodiment 1, showing second cylinder bores, a second thrust bearing, and the like.

FIG. 6 is a schematic diagram of a compressor according to a comparative embodiment, showing states in which the drive shaft is supported by first and second thrust bearings; (A) shows the state when a thrust force acting on the drive shaft is small; and (B) shows the state when a thrust force acting on the drive shaft in a direction toward the opposite end of the drive shaft is large.

FIG. 7 is a schematic diagram of the compressor according to Embodiment 1, showing states in which the drive shaft is supported by the first and second thrust bearings; (A) shows the state when a thrust force acting on the drive shaft is small; and (B) shows the state when a thrust force acting on the drive shaft in a direction toward the opposite end of the drive shaft is large.

FIG. 8 is an enlarged cross sectional view of an essential part of a compressor according to Embodiment 2, showing second cylinder bores, a second thrust bearing, and the like.

FIG. 9 is an enlarged cross sectional view of an essential part of a compressor according to Embodiment 3, showing second cylinder bores, a second thrust bearing, and the like.

FIG. 10 is an enlarged cross sectional view of an essential part of a compressor according to Embodiment 4, showing second cylinder bores, a second thrust bearing, and the like.

DETAILED DESCRIPTION OF EXEMPLARY EMBODIMENTS

Embodiments 1 to 4, which embody the present invention, will be described below with reference to the drawings. These compressors are mounted on vehicles and constitute refrigeration circuits of vehicle air-conditioning apparatus.

Embodiment 1

As shown in FIGS. 1 and 2, the compressor of Embodiment 1 is a double-headed piston compressor and comprises a housing 1, a drive shaft 3, a swash plate 5, a link mechanism 7, a plurality of double-headed pistons 9, a plurality of pairs of shoes 11a, 11b, and an actuator 13. This compressor further comprises a control mechanism 15 as shown in FIG. 3.

As shown in FIGS. 1 and 2, the housing 1 has a first housing 17, a second housing 19, a first cylinder block 21, a second cylinder block 23, a first valve forming plate 39, and a second valve forming plate 41. In this embodiment, the front-rear direction of the compressor is defined on the assumption that the side on which the first housing 17 is disposed is the front side of the compressor, and the side on which the second housing 19 is disposed is the rear side of the compressor. The front side of the compressor corresponds to “one end side of the drive shaft” in the present invention, and the rear side of the compressor corresponds to “an opposite end side of the drive shaft” in the present invention.

The first housing 17 has a boss 17a projecting frontward. A shaft seal device 25 is provided in the boss 17a. The first housing 17 includes therein a first suction chamber 27a and a first discharge chamber 29a. The first suction chamber 27a is formed into an annular shape and located on an inner circumferential side in the first housing 17. The first discharge chamber 29a is also formed into an annular shape and located on an outer circumferential side of the first suction chamber 27a in the first housing 17.

The first housing 17 further includes a first front communication passage 18a. The front end of the first front communication passage 18a communicates with the first discharge chamber 29a, and the rear end thereof opens at the rear end face of the first housing 17.

The second housing 19 includes apart of the control mechanism 15 mentioned above. The second housing 19 also includes a second suction chamber 27b, a second discharge chamber 29b, and a pressure regulation chamber 31. The pressure regulation chamber 31 is located in the central part of the second housing 19. The second suction chamber 27b is formed into an annular shape and located on an outer circumferential side of the pressure regulation chamber 31 in the second housing 19. The second discharge chamber 29b is also formed into an annular shape and located on an outer circumferential side of the second suction chamber 27b in the second housing 19.

The second housing 19 further includes a first rear communication passage 20a. The rear end of the first rear communication passage 20a communicates with the second discharge chamber 29b, and the front end thereof opens at the front end face of the second housing 19.

The first cylinder block 21 is provided on the front side in the compressor and located between the first housing 17 and the second cylinder block 23. The first cylinder block 21 includes a plurality of first cylinder bores 21a extending in the direction of the axis O of the drive shaft 3 and arranged at equiangular intervals in a circumferential direction. As shown in FIG. 4, the first cylinder bores 21a have a bore diameter of the length L1.

The first cylinder block 21 has a first shaft hole 21b through which the drive shaft 3 is inserted. A first sliding bearing 22a is provided in the first shaft hole 21b. The first cylinder block 21 further includes a first recess 21c that communicates with the first shaft hole 21b from the rear side of the compressor. The first recess 21c is coaxial with the first shaft hole 21b and has an inner diameter of the length L2, which is larger than the inner diameter of the first shaft hole 21b. At the front wall of the first recess 21c, a first recessed surface 21d that is recessed toward the front side of the compressor is formed into an annular shape.

A first thrust bearing 35a is provided in the first recess 21c. The first thrust bearing 35a has an outer diameter of the length D1. The first thrust bearing 35a includes a first race 351, a second race 352, first rolling elements 353 held between the first and second races 351, 352, and a retainer (not shown) that retains the first rolling elements 353 between the first and second races 351, 352. The first race 351 corresponds to the one end-side first race in the present invention, and the second race 352 corresponds to the one end-side second race in the present invention. The first race 351 and the second race 352 of the first thrust bearing 35a have a thickness T1.

The first cylinder block 21 includes a first retainer groove 21e that regulates a maximum opening degree of first suction reed valves 391a, which will be described later. Furthermore, as shown in FIGS. 1 and 2, the first cylinder block 21 includes a first connecting path 37a and a second front communication passage 18b. Front ends of the first connecting path 37a and the second front communication passage 18b open at the front end face of the first cylinder block 21, and rear ends thereof open at the rear end face of the first cylinder block 21.

The second cylinder block 23 is provided on the rear side of the compressor and located between the first cylinder block 21 and the second housing 19. The second cylinder block 23 is joined to the first cylinder block 21, whereby a swash plate chamber 33 is formed therebetween. The swash plate chamber 33 communicates with the first recess 21c. Thus, the first recess 21c forms a part of the swash plate chamber 33.

The second cylinder block 23 includes a plurality of second cylinder bores 23a extending in the direction of the axis O of the drive shaft 3. Similarly to the first cylinder bores 21a, the second cylinder bores 23a are arranged at equiangular intervals in the circumferential direction. Each of the second cylinder bores 23a is coaxially aligned with and forms a pair with the corresponding one of the first cylinder bores 21a in the front-rear direction. As shown in FIG. 5, the second cylinder bores 23a have a bore diameter of the length L3. Here, the length L3 is shorter than the length L1 shown in FIG. 4, i.e., shorter than the bore diameter of the first cylinder bores 21a. That is, the second cylinder bores 23a have a smaller diameter than the first cylinder bores 21a. The numbers of the first cylinder bores 21a and the second cylinder bores 23a may be suitably selected as long as they form pairs. Furthermore, the respective pairs of the first cylinder bores 21a and the second cylinder bores 23a may not be aligned coaxially.

As shown in FIG. 5, the second cylinder block 23 has a second shaft hole 23b through which the drive shaft 3 is inserted. A second sliding bearing 22b is provided in the second shaft hole 23b. Alternatively, instead of the first sliding bearing 22a and the second sliding bearing 22b, roller bearings may be provided.

The second cylinder block 23 further includes a second recess 23c that communicates with the second shaft hole 23b from the front side of the compressor. The second recess 23c is coaxial with the second shaft hole 23b and has an inner diameter of the length L4, which is larger than the inner diameter of the second shaft hole 23b. Here, the length L4, i.e., the inner diameter of the second recess 23c, is longer than the length L2 shown in FIG. 4, i.e., the inner diameter of the first recess 21c. That is, the second recess 23c has a larger diameter than the first recess 21c. As shown in FIG. 5, at the rear wall of the second recess 23c, a second recessed surface 23d that is recessed toward the rear side of the compressor is formed into an annular shape.

A second thrust bearing 35b is provided in the second recess 23c. Similarly to the first thrust bearing 35a described above, the second thrust bearing 35b has an outer diameter of the length D1. The second thrust bearing 35b includes a first race 354, a second race 355, a plurality of second rolling elements 356 held between the first and second races 354, 355, and a retainer (not shown) that retains the second rolling elements 356 between the first and second races 354, 355. The first race 354 corresponds to the opposite end-side first race in the present invention, and the second race 355 corresponds to the opposite end-side second race in the present invention. The first race 354 and the second race 355 of the second thrust bearing 35b also have a thickness T1. That is, the first thrust bearing 35a shown in FIG. 4 and the second thrust bearing 35b shown in FIG. 5 are the same in shape. Alternatively, instead of the first and second thrust bearings 35a, 35b, other configurations such as sliding bearings may be employed.

The second cylinder block 23 includes a second retainer groove 23e that regulates a maximum opening degree of second suction reed valves 411a, which will be described later. Furthermore, as shown in FIGS. 1 and 2, the second cylinder block 23 includes an outlet port 230, a confluence discharge chamber 231, a third front communication passage 18c, a second rear communication passage 20b, an inlet port 330, and a second connecting path 37b. The outlet port 230 and the confluence discharge chamber 231 communicate with each other. Via the outlet port 230, the confluence discharge chamber 231 is connected to a condenser (not shown) which constitutes a refrigeration circuit. Via the inlet port 330, the swash plate chamber 33 is connected to an evaporator (not shown) which constitutes the refrigeration circuit.

The third front communication passage 18c communicates with the second front communication passage 18b and the confluence discharge chamber 231. The front end of the second rear communication passage 20b communicates with the confluence discharge chamber 231, and the rear end thereof opens at the rear end face of the second cylinder block 23. The front end of the second connecting path 37b opens to the swash plate chamber 33, and the rear end thereof opens at the rear end face of the second cylinder block 23.

As shown in FIG. 4, the first valve forming plate 39 is provided between the first housing 17 and the first cylinder block 21. The first housing 17 and the first cylinder block 21 are joined together with the first valve forming plate 39 interposed therebetween.

The first valve forming plate 39 includes a first valve plate 390, a first suction valve plate 391, a first discharge valve plate 392, and a first retainer plate 393. The first valve plate 390 and the first suction valve plate 391 extend to the outer circumferences of the first housing 17 and the first cylinder block 21. First suction ports 390a, the number of which is the same as that of the first cylinder bores 21a, are formed through the first valve plate 390, the first discharge valve plate 392, and the first retainer plate 393. First discharge ports 390b, the number of which is the same as that of the first cylinder bores 21a, are formed through the first valve plate 390 and the first suction valve plate 391. Furthermore, a first suction communication hole 390c is formed through the first valve plate 390, the first suction valve plate 391, the first discharge valve plate 392, and the first retainer plate 393. A first discharge communication hole 390d is formed through the first valve plate 390 and the first suction valve plate 391.

The first cylinder bores 21a communicate with the first suction chamber 27a through the respective first suction ports 390a. Also, the first cylinder bores 21a communicate with the first discharge chamber 29a through the respective first discharge ports 390b. The first suction chamber 27a communicates with the first connecting path 37a through the first suction communication hole 390c. The first front communication passage 18a communicates with the second front communication passage 18b through the first discharge communication hole 390d.

The first suction valve plate 391 is provided on the rear surface of the first valve plate 390. The first suction valve plate 391 has the plurality of first suction reed valves 391a which are elastically deformable to open and close the first suction ports 390a. The first discharge valve plate 392 is provided on the front surface of the first valve plate 390. The first discharge valve plate 392 has a plurality of first discharge reed valves 392a which are elastically deformable to open and close the first discharge ports 390b. The first retainer plate 393 is provided on the front surface of the first discharge valve plate 392. The first retainer plate 393 regulates a maximum opening degree of the first discharge reed valves 392a.

As shown in FIG. 5, the second valve forming plate 41 is provided between the second housing 19 and the second cylinder block 23. The second housing 19 and the second cylinder block 23 are joined together with the second valve forming plate 41 interposed therebetween.

The second valve forming plate 41 includes a second valve plate 410, a second suction valve plate 411, a second discharge valve plate 412, and a second retainer plate 413. The second valve plate 410 and the second suction valve plate 411 extend to the outer circumferences of the second housing 19 and the second cylinder block 23. Second suction ports 410a, the number of which is the same as that of the second cylinder bores 23a, are formed through the second valve plate 410, the second discharge valve plate 412, and the second retainer plate 413. Second discharge ports 410b, the number of which is the same as that of the second cylinder bores 23a, are formed through the second valve plate 410 and the second suction valve plate 411. Furthermore, a second suction communication hole 410c is formed through the second valve plate 410, the second suction valve plate 411, the second discharge valve plate 412, and the second retainer plate 413. A second discharge communication hole 410d is formed through the second valve plate 410 and the second suction valve plate 411.

The second cylinder bores 23a communicate with the second suction chamber 27b through the respective second suction ports 410a. Also, the second cylinder bores 23a communicate with the second discharge chamber 29b through the respective second discharge ports 410b. The second suction chamber 27b communicates with the second connecting path 37b through the second suction communication hole 410c. The first rear communication passage 20a communicates with the second rear communication passage 20b through the second discharge communication hole 410d.

The second suction valve plate 411 is provided on the front surface of the second valve plate 410. The second suction valve plate 411 has the plurality of second suction reed valves 411a which are elastically deformable to open and close the second suction ports 410a. The second discharge valve plate 412 is provided on the rear surface of the second valve plate 410. The second discharge valve plate 412 has a plurality of second discharge reed valves 412a which are elastically deformable to open and close the second discharge ports 410b. The second retainer plate 413 is provided on the rear surface of the second discharge valve plate 412. The second retainer plate 413 regulates a maximum opening degree of the second discharge reed valves 412a.

As shown in FIGS. 1 and 2, a first discharge communication passage 18 is formed by the first front communication passage 18a, the first discharge communication hole 390d, the second front communication passage 18b, and the third front communication passage 18c. Also, a second discharge communication passage 20 is formed by the first rear communication passage 20a, the second discharge communication hole 410d, and the second rear communication passage 20b.

The swash plate chamber 33 communicates with the first and second suction chambers 27a, 27b via the first and second connecting paths 37a, 37b and the first and second suction communication holes 390c, 410c. Thus, the pressure in the first and second suction chambers 27a, 27b and the swash plate chamber 33 are substantially equal. Because a low pressure refrigerant gas which has passed through the evaporator is introduced into the swash plate chamber 33 through the inlet port 330, the pressure in the swash plate chamber 33 and the first and second suction chambers 27a, 27b is lower than the pressure in the first and second discharge chambers 29a, 29b.

The drive shaft 3 includes a drive shaft body 30, a first support member 43a, and a second support member 43b. The first support member 43a corresponds to the first annular portion in the present invention. The second support member 43b corresponds to the second annular portion in the present invention.

The drive shaft body 30 extends rearward from the front side of the housing 1 in an axial direction, i.e., along the axis O of the drive shaft 3. A first small-diameter portion 30a is formed on the front end side of the drive shaft body 30. A second small-diameter portion 30b is formed on the rear end side of the drive shaft body 30. The drive shaft body 30 is inserted through the shaft seal device 25 and the first and second sliding bearings 22a, 22b in the housing 1. Thus, the drive shaft body 30 and therefore the drive shaft 3 are supported in the housing 1 so as to be rotatable about the axis O of the drive shaft 3. The front end of the drive shaft body 30 is inserted through the shaft seal device 25 in the boss 17a. The rear end of the drive shaft body 30 projects into the pressure regulation chamber 31.

The above-mentioned swash plate 5, the link mechanism 7, and the actuator 13 are provided on the drive shaft body 30. The swash plate 5, the link mechanism 7, and the actuator 13 are all disposed in the swash plate chamber 33.

A threaded portion 3a is formed on the front end of the drive shaft body 30. The drive shaft 3 is connected to a pulley or an electromagnetic clutch (not shown) via the threaded portion 3a.

As shown in FIG. 4, the first support member 43a is formed into an annular shape and its central axis coincides with the axis O of the drive shaft 3. The first support member 43a is press-fitted to the first small-diameter portion 30a of the drive shaft body 30 and supported by the first sliding bearing 22a in the first shaft hole 21b. A first flange 430 and a mounting portion (not shown) that is configured to allow a second pin 47b, which will be described later, to be inserted therethrough, are formed at the rear end side of the first support member 43a.

The first thrust bearing 35a is held between the first flange 430 and the front wall of the first recess 21c in the axial direction. The outer diameter of the first flange 430 is larger than the inner diameter of the first thrust bearing 35a and smaller than the outer diameter of the first thrust bearing 35a. Thus, the first thrust bearing 35a is in contact with the first flange 430 only in a region adjacent to the inner circumference of the second race 352. The inner diameter of the first recessed surface 21d on the front wall of the first recess 21c is larger than the inner diameter of the first thrust bearing 35a and smaller than the outer diameter of the first thrust bearing 35a. Thus, the first thrust bearing 35a is in contact with the front wall of the first recess 21c only in a region adjacent to the outer circumference of the first race 351.

More specifically, the first thrust bearing 35a is in contact with the first flange 430 in the annular area E1, which is adjacent to the inner circumference of the second race 352, and in contact with the front surface of the first recess 21c in the annular area E2, which is adjacent to the outer circumference of the first race 351. That is, the region where the first thrust bearing 35a is supported by the first support member 43a via the first flange 430 and the region where the first thrust bearing 35a is supported by the first cylinder block 21 via the front surface of the first recess 21c are radially shifted from each other. In this way, a predetermined preload is applied to the first thrust bearing 35a, and, in the preloaded state, the first thrust bearing 35a receives a thrust force which acts on the drive shaft 3 in a frontward direction at the time of operation of the compressor.

As shown in FIGS. 1 and 2, the front end of a first return spring 44a is inserted in the first support member 43a. This return spring 44a extends from a position near the first flange 430 toward the swash plate 5 in the direction of the axis O of the drive shaft 3.

As shown in FIG. 5, the second support member 43b is formed into an annular shape and its central axis coincides with the axis O of the drive shaft 3. The second support member 43b is press-fitted to the rear side of the second small-diameter portion 30b of the drive shaft body 30 and supported by the second sliding bearing 22b in the second shaft hole 23b. A second flange 431 is formed at the front end of the second support member 43b. The second flange 431 has a larger diameter than the first flange 430 shown in FIG. 4.

As shown in FIG. 5, O-rings 51b, 51c are provided on the second support member 43b at positions rearward of the second flange 431. By providing the O-rings 51b, 51c, the pressure regulation chamber 31 is hermetically sealed from the second recess 23c, and thus, hermetically sealed from the swash plate chamber 33.

The second thrust bearing 35b is held between the second flange 431 and the rear wall of the second recess 23c in the axial direction. The outer diameter of the second flange 431 is larger than the inner diameter of the second thrust bearing 35b and smaller than the outer diameter of the second thrust bearing 35b. Thus, the second thrust bearing 35b is in contact with the second flange 431 only in a region adjacent to the inner circumference of the second race 355. The inner diameter of the second recessed surface 23d on the rear wall of the second recess 23c is larger than the inner diameter of the second thrust bearing 35b and smaller than the outer diameter of the second thrust bearing 35b. Thus, the second thrust bearing 35b is in contact with the rear wall of the second recess 23c only in a region adjacent to the outer circumference of the first race 354.

More specifically, the second thrust bearing 35b is in contact with the second flange 431 in the annular area E3, which is adjacent to the inner circumference of the second race 355, and in contact with the rear surface of the second recess 23c in the annular area E4, which is adjacent to the outer circumference of the first race 354. That is, the region where the second thrust bearing 35b is supported by the second support member 43b via the second flange 431 and the region where the second thrust bearing 35b is supported by the second cylinder block 23 via the rear surface of the second recess 23c are radially shifted from each other. In this way, a predetermined preload is applied to the second thrust bearing 35b, and in the preloaded state, the second thrust bearing 35b receives a thrust force which acts on the drive shaft 3 in a rearward direction at the time of operation of the compressor.

As described above, since the first thrust bearing 35a is held between the first flange 430 and the front surface of the first recess 21c such that the region where the first thrust bearing 35a is supported by the first support member 43a and the region where the first thrust bearing 35a is supported by the first cylinder block 21 are radially shifted from each other, the first thrust bearing 35a is elastically deformable like a disk spring in the axial direction. Likewise, since the second thrust bearing 35b is held between the second flange 431 and the rear surface of the second recess 23c such that the region where the second thrust bearing 35b is supported by the second support member 43b and the region where the second thrust bearing 35b is supported by the second cylinder block 23 are radially shifted from each other, the second thrust bearing 35b is also elastically deformable like a disk spring in the axial direction.

In other words, in this compressor, the first spring constant K1 of the first thrust bearing 35a is set such that the first thrust bearing 35a is deformable in the direction of the axis O of the drive shaft 3. Also, the second spring constant K2 of the second thrust bearing 35b is set such that the second thrust bearing 35b is deformable in the direction of the axis O of the drive shaft 3. Here, as described above, the second flange 431 has a larger diameter than the first flange 430. Therefore, the area E3 where the second race 355 of the second thrust bearing 35b is in contact with the second flange 431 is larger than the area E1 where the second race 352 of the first thrust bearing 35a shown in FIG. 4 is in contact with the first flange 430. That is, the region where the second thrust bearing 35b shown in FIG. 5 is in contact with the second flange 431 is larger than the region where the first thrust bearing 35a shown in FIG. 4 is in contact with the first flange 430. As a result, the second thrust bearing 35b is less deformable in the direction of the axis O of the drive shaft 3 than the first thrust bearing 35a. That is, in this compressor, the second spring constant K2 of the second thrust bearing 35b is greater than the first spring constant K1 of the first thrust bearing 35a.

In Embodiment 1, the area E4 where the first race 354 of the second thrust bearing 35b shown in FIG. 5 is in contact with the rear surface of the second recess 23c is made equal to the area E2 where the first race 351 of the first thrust bearing 35a shown in FIG. 4 is in contact with the front surface of the first recess 21c. However, the second spring constant K2 of the second thrust bearing 35b can also be set greater than the first spring constant K1 of the first thrust bearing 35a by making the inner diameter of the area E4 smaller than the inner diameter of the area E2.

As shown in FIGS. 1 and 2, the swash plate 5 has a flat annular shape and includes a front surface 5a and a rear surface 5b. The front surface 5a faces frontward of the compressor in the swash plate chamber 33. The rear surface 5b faces rearward of the compressor in the swash plate chamber 33.

The swash plate 5 includes a ring plate 45. The ring plate 45 has a flat annular shape with an insertion hole 45a formed in its center. The drive shaft body 30 is inserted through the insertion hole 45a of the swash plate 5, whereby the swash plate 5 is attached to the drive shaft 3 in the swash plate chamber 33. The ring plate 45 has a coupling portion (not shown) to be connected to pull arms 132, which will be described below.

The link mechanism 7 has a lug arm 49. The lug arm 49 is disposed frontward of the swash plate 5 in the swash plate chamber 33 and located between the swash plate 5 and the first support member 43a. The lug arm 49 is formed substantially in an L shape from the front end toward the rear end. A weight portion 49a is formed at the rear end of the lug arm 49. The weight portion 49a extends in the circumferential direction of the actuator 13 over about a half of the circumference thereof. The shape of the weight portion 49a may be designed as appropriate.

The rear end side of the lug arm 49 is connected to the ring plate 45 with a first pin 47a. By doing so, when the axis of the first pin 47a is defined as a first pivot axis M1, the lug arm 49 is supported about the first pivot axis M1 so as to be pivotable with respect to the ring plate 45, i.e., the swash plate 5. The first pivot axis M1 extends in a direction perpendicular to the axis O of the drive shaft 3.

The front end of the lug arm 49 is connected to the first support member 43a with a second pin 47b. By doing so, when the axis of the second pin 47b is defined as a second pivot axis M2, the lug arm 49 is supported about the second pivot axis M2 so as to be pivotable with respect to the first support member 43a, i.e., the drive shaft 3. The second pivot axis M2 extends parallel to the first pivot axis M1. The lug arm 49 and the first and second pins 47a, 47b constitute the link mechanism 7 in the present invention.

The weight portion 49a is provided so as to extend at the rear end of the lug arm 49, i.e., on the opposite side of the second pivot axis M2 with reference to the first pivot axis M1. As the lug arm 49 is supported by the ring plate 45 with the first pin 47a, the weight portion 49a passes through a groove portion 45b of the ring plate 45 and reaches the rear surface side of the ring plate 45, i.e., the side of the rear surface 5b of the swash plate 5. Therefore, the centrifugal force generated by rotation of the swash plate 5 about the axis O of the drive shaft 3 acts on the weight portion 49a at the side of the rear surface 5b of the swash plate 5.

In this compressor, the swash plate 5 is able to rotate together with the drive shaft 3 because the swash plate 5 is connected to the drive shaft 3 by the link mechanism 7. Furthermore, the swash plate 5 is able to change its inclination angle from the minimum angle shown in FIG. 1 to the maximum angle shown in FIG. 2 because both ends of the lug arm 49 pivot about the first pivot axis M1 and the second pivot axis M2, respectively.

As shown in FIGS. 1 and 2, the pistons 9 are double-headed pistons, each having a first piston head 9a at its front end and a second piston head 9b at its rear end. Since the second cylinder bores 23a have a smaller diameter than the first cylinder bores 21a, the second piston heads 9b have a smaller diameter than the first piston heads 9a.

The first piston heads 9a are reciprocally accommodated in the respective first cylinder bores 21a. The first piston heads 9a and the first valve forming plate 39 define respective first compression chambers 53a in the respective first cylinder bores 21a. The second piston heads 9b are reciprocally accommodated in the respective second cylinder bores 23a. The second piston heads 9b and the second valve forming plate 41 define respective second compression chambers 53b in the respective second cylinder bores 23a.

In this compressor, the top dead center positions of the first piston heads 9a and the second piston heads 9b moves when the strokes of the pistons 9 changes according to the change in the inclination angle of the swash plate 5. Specifically, as the inclination angle of the swash plate 5 decreases, the top dead center positions of the first and second piston heads 9a, 9b move such that the volume of the second compression chamber 53d become larger than the volume of the first compression chamber 51d.

The actuator 13 is disposed in the swash plate chamber 33. More specifically, the actuator 13 is disposed rearward of the swash plate 5 in the swash plate chamber 33, i.e., on the side of the second cylinder block 23 with respect to the swash plate 5 in the swash plate chamber 33, in which the second cylinder bores 23a are formed. Accordingly, the actuator 13 is capable of advancing into the second recess 23c.

The actuator 13 includes a movable body 13a, a partition body 13b, and a control pressure chamber 13c. The control pressure chamber 13c is formed between the movable body 13a and the partition body 13b. The movable body 13a includes a rear wall 130, a circumferential wall 131, and a pair of pull arms 132. The pull arms 132 correspond to the connecting portion in the present invention. In FIGS. 1 and 2, only one of the pull arms 132 is shown.

The rear wall 130 is positioned rearward in the movable body 13a and extends radially away from the axis O of the drive shaft 3. The rear wall 130 has an insertion hole 130a through which the second small-diameter portion 30b of the drive shaft body 30 is inserted. The O-ring 51d is provided in the insertion hole 130a. The circumferential wall 131 is continuous with the outer edge of the rear wall 130 and extends toward the front in the movable body 13a. The pull arms 132 are disposed at the front end of the circumferential wall 131 such that the drive shaft axis O is interposed therebetween, and project frontward in the movable body 13a. The movable body 13a has a bottomed cylindrical shape which is formed by the rear wall 130, the circumferential wall 131, and the pull arms 132.

The partition body 13b is formed into a disc shape having a diameter that is substantially equal to the inner diameter of the movable body 13a. The partition body 13b has an insertion hole 133 extending through its center. The O-ring 51e is provided on the outer circumference of the partition body 13b.

An inclination angle reducing spring 44b is provided between the partition body 13b and the ring plate 45. Specifically, the rear end of the inclination angle reducing spring 44b is disposed in contact with the partition body 13b, and the front end of the inclination angle reducing spring 44b is disposed in contact with the ring plate 45. The inclination angle reducing spring 44b urges both the partition body 13b and the ring plate 45 so that they are spaced apart from each other.

The drive shaft body 30 is inserted through the insertion hole 130a of the movable body 13a. Thus, the movable body 13a is movable with respect to the drive shaft body 30 along the axis O of the drive shaft 3. On the other hand, the drive shaft body 30 is press-fitted to the insertion hole 133 of the partition body 13b. Thus, the partition body 13b is fixed to the drive shaft body 30 so that the partition body 13b is rotatable together with the drive shaft body 30. Alternatively, instead of press-fitting, the drive shaft body 30 may be inserted through the insertion hole 133 of the partition body 13b such that the partition body 13b is movable along the axis O of the drive shaft 3.

The partition body 13b is disposed within the movable body 13a at a position rearward of the swash plate 5, and its outer circumference is surrounded by the circumferential wall 131. Consequently, when the movable body 13a moves in the direction of the axis O of the drive shaft 3, the inner circumference of the circumferential wall 131 of the movable body 13a slides against the outer circumference of the partition body 13b.

The control pressure chamber 13c is formed between the movable body 13a and the partition body 13b by surrounding the partition body 13b with the circumferential wall 131. The control pressure chamber 13c is partitioned from the swash plate chamber 33 by the rear wall 130, the circumferential wall 131, and the partition body 13b.

The pull arms 132 are connected to the ring plate 45 with a third pin 47c. When the axis of the third pin 47c is defined as an action axis M3, the swash plate 5 is supported by the movable body 13a so as to be pivotable about the action axis M3. The action axis M3 extends parallel to the first and second pivot axes M1, M2. In this way, the movable body 13a is connected to the swash plate 5, and due to this connection, the partition body 13b and the swash plate 5 face each other.

The second small-diameter portion 30b has an axial path 3b extending frontward from the rear end in the direction of the axis O of the drive shaft 3, and a radial path 3c extending radially from the front end of the axial path 3b and opening at the outer circumferential surface of the drive shaft body 30. The rear end of the axial path 3b communicates with the pressure regulation chamber 31. The radial path 3c communicates with the control pressure chamber 13c. Thus, the control pressure chamber 13c communicates with the pressure regulation chamber 31 through the radial path 3c and the axial path 3b.

As shown in FIG. 3, the control mechanism 15 includes a bleed passage 15a, a supply passage 15b, a control valve 15c, an orifice 15d, the axial path 3b, and the radial path 3c.

The bleed passage 15a is connected to the pressure regulation chamber 31 and the second suction chamber 27b. The control pressure chamber 13c, the pressure regulation chamber 31, and the second suction chamber 27b communicate with one another through the bleed passage 15a, the axial path 3b, and the radial path 3c. The supply passage 15b is connected to the pressure regulation chamber 31 and the second discharge chamber 29b. The control pressure chamber 13c, the pressure regulation chamber 31, and the second discharge chamber 29b communicate with one another through the supply passage 15b, the axial path 3b, and the radial path 3c. The supply passage 15b has the orifice 15d.

The control valve 15c is provided in the bleed passage 15a. The control valve 15c is capable of regulating the opening degree of the bleed passage 15a based on the pressure in the second suction chamber 27b.

In this compressor, the inlet port 330 shown in FIGS. 1 and 2 is connected to a pipe leading to the evaporator, and the outlet port 230 is connected to a pipe leading to the condenser. The condenser is connected to the evaporator through a pipe and an expansion valve. The compressor, the evaporator, the expansion valve, the condenser and the like constitute the refrigeration circuit of vehicle air-conditioning apparatus. Illustration of the evaporator, the expansion valve, the condenser and the pipes is omitted in the drawings.

In the compressor configured as described above, the rotation of the drive shaft 3 causes rotation of the swash plate 5, whereby the pistons 9 reciprocate in the first cylinder bores 21a and the second cylinder bores 23a. Therefore, the volumes of the first and second compression chambers 53a, 53b change in response to the piston strokes. Thus, in this compressor, the following phases are repetitively carried out: a suction phase in which the refrigerant gas is drawn into the first and second compression chambers 53a, 53b; a compression phase in which the refrigerant gas is compressed in the first and second compression chambers 53a, 53b; and a discharge phase in which the compressed refrigerant gas is discharged into the first and second discharge chambers 29a, 29b.

The refrigerant gas discharged into the first discharge chamber 29a flows through the first discharge communication passage 18 to the confluence discharge chamber 231. Likewise, the refrigerant gas discharged into the second discharge chamber 29b flows through the second discharge communication passage 20 to the confluence discharge chamber 231. The refrigerant gas that has reached the confluence discharge chamber 231 is discharged to the condenser from the outlet port 230 through the pipe.

During these suction phase and the like, piston compression forces to reduce the inclination angle of the swash plate 5 acts on a rotational body, which consists of the swash plate 5, the ring plate 45, the lug arm 49, and the first pin 47a. When the inclination angle of the swash plate 5 is changed, the strokes of the pistons 9 increase or decrease, and thereby it is possible to control the displacement.

Specifically, in the control mechanism 15 shown in FIG. 3, when the control valve 15c increases the opening degree of the bleed passage 15a, the pressure in the pressure regulation chamber 31 and hence the pressure in the control pressure chamber 13c are made to be substantially equal to the pressure in the second suction chamber 27b. As a result, a variable differential pressure, which is the difference in pressure between the control pressure chamber 13c and the swash plate chamber 33, decreases. Therefore, due to the piston compression forces acting on the swash plate 5, the movable body 13a of the actuator 13 moves frontward in the swash plate chamber 33 as shown in FIG. 1.

As a result, in the compressor, the swash plate 5 is urged toward a direction to reduce the inclination angle due to the compression reaction force, i.e., a resultant force of the piston compression forces, which acts on the swash plate 5 via the pistons 9, and thus the movable body 13a is pulled frontward in the swash plate chamber 33 via the pull arms 132 at the action axis M3 so that the ring plate 45 is brought into contact with the rear end of the return spring 44a. As the movable body 13a is pulled frontward in the swash plate chamber 33, the swash plate 5 pivots clockwise about the action axis M3 against the biasing force of the return spring 44a. Furthermore, the rear end of the lug arm 49 pivots counterclockwise about the first pivot axis M1 and the front end of the lug arm 49 pivots counterclockwise about the second pivot axis M2. Thus, the front end portion of the lug arm 49 approaches the first flange 430 of the first support member 43a. With such a configuration, the swash plate 5 pivots using the action axis M3 as a point of action and using the first pivot axis M1 as a point of pivot. This reduces the inclination angle of the swash plate 5 with respect to the direction perpendicular the axis O of the drive shaft 3 and thus decreases the strokes of the pistons 9. Consequently, the discharge capacity of the compressor per rotation of the drive shaft 3 decreases.

In this compressor, the centrifugal force acting on the weight portion 49a is also imparted to the swash plate 5. As a result, the swash plate 5 can be easily displaced in a direction to reduce its inclination angle.

In this compressor, as the inclination angle of the swash plate 5 becomes smaller and the strokes of the pistons 9 decrease, the top dead center position of the second piston heads 9b moves away from the second valve forming plate 41. As a result, when the inclination angle of the swash plate 5 becomes nearly zero degrees, no compression work is carried out in the second compression chambers 53b while slight compression work is carried out in the first compression chambers 53a.

On the other hand, in the control mechanism 15 shown in FIG. 3, when the control valve 15c reduces the opening degree of the bleed passage 15a, the pressure in the pressure regulation chamber 31 increases due to the pressure of the refrigerant gas in the second discharge chamber 29b, and thereby the pressure in the control pressure chamber 13c increases. As a result, the variable differential pressure, i.e., the difference in pressure between the control pressure chamber 13c and the swash plate chamber 33, increases. Therefore, against the piston compression forces acting on the swash plate 5, the movable body 13a of the actuator 13 moves rearward in the swash plate chamber 33 from the position shown in FIG. 1 so as to advance into the second recess 23c as shown in FIG. 2.

Consequently, in the compressor, the movable body 13a pulls the swash plate 5 rearward in the swash plate chamber 33 via the pull arms 132 at the action axis M3 against the biasing force of the inclination angle reducing spring 44b. Thus, the swash plate 5 pivots counterclockwise about the action axis M3. Furthermore, the rear end of the lug arm 49 pivots clockwise about the first pivot axis M1 and the front end of the lug arm 49 pivots clockwise about the second pivot axis M2. Therefore, the front end portion of the lug arm 49 moves rearward away from the first flange 430 of the first support member 43a. The swash plate 5 thus pivots in a direction opposite to the above-described direction to reduce the inclination angle, using the action axis M3 as a point of action and the first pivot axis M1 as a point of pivot. This increases the inclination angle of the swash plate 5 with respect to the direction perpendicular to the axis O of the drive shaft 3 and increases the strokes of the pistons 9. Consequently, the discharge capacity of the compressor per rotation of the drive shaft 3 increases.

Since this compressor has the first and second cylinder bores 21a, 23a so that the refrigerant gas is compressed in both the first and second compression chambers 53a, 53b, it is possible to provide a large discharge capacity per rotation of the drive shaft 3 as compared with the case in which the refrigerant gas is compressed only in the first compression chambers 53a, for example. Moreover, in this compressor, even at the time of small discharge capacity, the compression phase for compressing the refrigerant gas is carried out in the first compression chambers 53a, i.e., in the first cylinder bores 21a having a larger diameter than the second cylinder bores 23a, and therefore, it is possible to provide a required discharge capacity.

Here, in this compressor, since the second cylinder bores 23a have a smaller diameter than the first cylinder bores 21a and the second piston heads 9b have a smaller diameter than the first piston heads 9a, pressure receiving area of the first piston heads 9a is larger than that of the second piston heads 9b. As a result, as the discharge capacity increases, the thrust force toward the rear end side of the drive shaft 3 becomes larger than the thrust force toward the front end side of the drive shaft 3.

In this regard, in this compressor, the second flange 431 has a larger diameter than the first flange 430, and the area E3 where the second race 355 of the second thrust bearing 35b shown in FIG. 5 is in contact with the second flange 431 is greater than the area E1 where the second race 352 of the first thrust bearing 35a shown in FIG. 4 is in contact with the first flange 430. Due to this configuration, the second spring constant K2 of the second thrust bearing 35b is greater than the first spring constant K1 of the first thrust bearing 35a. As a result, in this compressor, even when the thrust force toward the rear end side of the drive shaft 3 becomes larger than the thrust force toward the front end side of the drive shaft 3 as the discharge capacity increases, it is possible to sufficiently receive the large thrust force.

The above-described advantages of this compressor will be described more specifically below by comparing the compressor of Embodiment 1, the schematic diagrams of which are shown in FIGS. 7A and 7B, with a compressor of a comparative embodiment, the schematic diagrams of which are shown in FIGS. 6A and 6B. The compressor of the comparative embodiment is configured such that, unlike the compressor of Embodiment 1, the second spring constant K22 of the second thrust bearing 35b is equal to the first spring constant K1 of the first thrust bearing 35a. Except for this, the features of the compressor of the comparative embodiment are the same as those of the compressor of Embodiment 1.

As shown in FIG. 6A, in the compressor of the comparative embodiment, when the thrust force which acts on the drive shaft 3 in the frontward direction and the thrust force which acts on the drive shaft 3 in the rearward direction are small at the time of small discharge capacity, the first and second thrust bearings 35a, 35b are able to support the drive shaft 3 suitably. However, as shown in FIG. 6B, as the discharge capacity increases, the thrust force F acting on the drive shaft 3 in the rearward direction increases, and the second thrust bearing 35b is largely deformed in the direction of the axis O of the drive shaft 3. If the second thrust bearing 35b is excessively deformed in the direction of the axis O of the drive shaft 3, the durability of the second thrust bearing 35b is more likely to decrease.

In addition, in the case where the thrust force F acting on the drive shaft 3 in the rearward direction causes the large deformation of the second thrust bearing 35b in the direction of the axis O of the drive shaft 3 and consequently the drive shaft 3 is displaced toward the rear end of the compressor by the distance M1 from the position at the time of small discharge capacity, there is a risk that the first thrust bearing 35a is expanded to its limit and separated from the drive shaft 3. That is, the first thrust bearing 35a may become unable to sufficiently support the drive shaft 3 by elastic deformation and a gap may be created between the drive shaft 3 and the first thrust bearing 35a. As a result, in the compressor of the comparative embodiment, the drive shaft 3 wobbles, and this leads to an increase of vibration and noise caused by the vibration at the time of operation.

To solve these problems, the compressor of Embodiment 1 is configured such that the second spring constant K2 of the second thrust bearing 35b is greater than the first spring constant K1 of the first thrust bearing 35a. As a result, the first and second thrust bearings 35a, 35b are able to support the drive shaft 3 suitably not only at the time of small discharge capacity as shown in FIG. 7A but also at the time of large discharge capacity as shown in FIG. 7B when the thrust force F acting on the drive shaft 3 in the rearward direction is large. That is, in the compressor of Embodiment 1, even when the thrust force F acting on the drive shaft 3 in the rearward direction increases, the second thrust bearing 35b is less likely to be deformed in the direction of the axis O of the drive shaft 3. Thus, with the compressor of Embodiment 1, it is possible to inhibit excessive deformation of the second thrust bearing 35b in the direction of the axis O of the drive shaft 3 as compared with the compressor of the comparative embodiment, and therefore, it is possible to increase the durability of the second thrust bearing 35b.

Furthermore, since the second thrust bearing 35b is less likely to be deformed in the direction of the axis O of the drive shaft 3, although the thrust force F acts on the drive shaft 3 in the rearward direction, the drive shaft 3 is displaced toward the rear end of the compressor only by the distance M2, which is shorter than the distance M1, from the position at the time of small discharge capacity. As a result, in the compressor of Embodiment 1, the first thrust bearing 35a is able to sufficiently support the drive shaft 3 without creating a gap between the drive shaft 3 and the first thrust bearing 35a. Therefore, in the compressor of Embodiment 1, the drive shaft 3 is less likely to wobble even at the time of large discharge capacity.

Moreover, in the compressor of Embodiment 1, the first spring constant K1 of the first thrust bearing 35a does not need to be significantly large to support a large thrust force. Therefore, drag resistances of the first and second thrust bearings 35a, 35b acting on the drive shaft 3 do not increase excessively at the time of small discharge capacity shown in FIG. 7A.

Therefore, the compressor of Embodiment 1 exhibits excellent durability and inhibits vibrations and noise caused by the vibration at the time of operation with less power loss.

In particular, in this compressor, the first thrust bearing 35a is configured such that the region where the second race 352 is supported by the first flange 430 (i.e., the area E1) and the region where the first race 351 is supported by the front surface of the first recess 21c (i.e., the area E2) are radially shifted from each other. Likewise, the second thrust bearing 35b is configured such that the region where the second race 355 is supported by the second flange 431 (i.e., the area E3) and the region where the first race 354 is supported by the rear surface of the second recess 23c (i.e., the area E4) are radially shifted from each other.

As a result, in this compressor, both the first and second thrust bearings 35a, 35b are elastically deformable in the direction of the axis O of the drive shaft 3 like a disc spring. Therefore even when the drive shaft 3 is displaced in the direction of the axis O of the drive shaft 3 due to the thrust forces described above, the first and second thrust bearings 35a, 35b are able to follow the displacement suitably. Therefore, the first and second thrust bearings 35a, 35b are able to support the drive shaft 3 suitably. Furthermore, in this compressor, the first and second thrust bearings 35a, 35b may be provided with an interference fit, whereby product-to-product variations of the first and second spring constants K1 and K2 can be prevented at the time of assembly.

Moreover, in this compressor, the area E3 where the second race 355 of the second thrust bearing 35b is in contact with the second flange 431 is larger than the area E1 where the second race 352 of the first thrust bearing 35a is in contact with the first flange 430. Therefore, this compressor can be easily configured such that the second spring constant K2 of the second thrust bearing 35b is greater than the first spring constant K1 of the first thrust bearing 35a while allowing both the first and second thrust bearings 35a, 35b to be deformed in the direction of the axis O of the drive shaft 3 as described above.

Furthermore, in this compressor, since the second cylinder bores 23a have a smaller diameter than the first cylinder bores 21a, it is possible to configure the second recess 23c so as to have a larger diameter than the first recess 21c without increasing the size of the second cylinder block 23. As a result, by disposing the actuator 13 on the side of the second cylinder block 23 with respect to the swash plate 5 in the swash plate chamber 5, i.e., on the side where the second cylinder bores 23a are formed, the size of the actuator 13 can be increased without the need to increase the size of the compressor. By doing this, the pressure receiving area of the movable body 13a can be increased and the movable body 13a is movable with a large thrust force. Therefore, this compressor is capable of exhibiting high controllability.

Embodiment 2

The compressor of Embodiment 2 is provided with a second thrust bearing 35c shown in FIG. 8 in place of the second thrust bearing 35b in the compressor of Embodiment 1. In Embodiment 2, the second flange 431 has the same diameter as the first flange 430 (see FIG. 4).

As shown in FIG. 8, the second thrust bearing 35c has an outer diameter of the length D1, which is the same as those of the first and second thrust bearings 35a, 35b described above. The second thrust bearing 35c has a first race 357, a second race 358, a plurality of second rolling elements 359 held between the first and second races 357, 358, and a retainer (not shown) that retains the second rolling elements 359 between the first and second races 357, 358. The first race 357 corresponds to the opposite end-side first race in the present invention, and the second race 358 corresponds to the opposite end-side second race in the present invention. The second thrust bearing 35c is configured such that the first race 357 and the second race 358 both have a thickness T2. The thickness T2 is thicker than the thickness T1 of the first and second races 351, 352 of the first thrust bearing 35a shown in FIG. 4. That is, the first and second races 357, 358 of the second thrust bearing 35c have a larger thickness than the first and second races 351, 352 of the first thrust bearing 35a. Alternatively, for example, only the first race 357 may have the thickness T2 while the second race 358 may have the thickness T1 as with the compressor of Embodiment 1.

In this compressor, similarly to the compressor of Embodiment 1, the second thrust bearing 35c is held between the second flange 431 and the rear wall of the second recess 23c in the axial direction. The second thrust bearing 35c is in contact with the second flange 431 at an annular area E5, which is adjacent to the inner circumference of the second race 358, and in contact with the rear surface of the second recess 23c at an annular area E4, which is adjacent to the outer circumference of the first race 357. That is, also in this compressor, the region where the second thrust bearing 35c is supported by the second support member 43b via the second flange 431 and the region where the second thrust bearing 35c is supported by the second cylinder block 23 via the rear surface of the second recess 23c are radially shifted from each other. Here, as described above, the second flange 431 has the same diameter as the first flange 430. Thus, in this compressor, the area E5 where the second race 358 of the second thrust bearing 35b is in contact with the second flange 431 is equal to the area E1 where the second race 352 of the first thrust bearing 35a shown in FIG. 4 is in contact with the first flange 430. The other features of this compressor are the same as those of the compressor of Embodiment 1. The same reference numerals are given to the same components and detailed description thereof is omitted.

In this compressor, the first and second races 357, 358 of the second thrust bearing 35c have a thickness that is larger than the thickness of the first and second races 351, 352 of the first thrust bearing 35a. As a result, as compared with the first thrust bearing 35a, the second thrust bearing 35c is less likely to be elastically deformed in the direction of the axis O of the drive shaft 3 like a disc spring. That is, also in this compressor, the second spring constant K2 of the second thrust bearing 35c is greater than the first spring constant K1 of the first thrust bearing 35a. Consequently, this compressor is capable of providing the same advantages as those of the compressor of Embodiment 1.

Embodiment 3

The compressor of Embodiment 3 is provided with a second thrust bearing 35d shown in FIG. 9 in place of the second thrust bearing 35b in the compressor of Embodiment 1. In this compressor, the second flange 431 has the same diameter as the first flange 430 (see FIG. 4) as with the compressor of Embodiment 2.

As shown in FIG. 9, the second thrust bearing 35d has a first race 361, a second race 362, a plurality of second rolling elements 363 held between the first and second races 361, 362, and a retainer (not shown) that retains the second rolling elements 363 between the first and second races 361, 362. The first race 361 corresponds to the opposite end-side first race in the present invention, and the second race 362 corresponds to the opposite end-side second race in the present invention. The second thrust bearing 35d is configured such that the first race 361 and the second race 362 both have a thickness T1 and thus have a thickness equal to the thickness of the first and second races 351, 352 of the first thrust bearing 35a shown in FIG. 4. On the other hand, as shown in FIG. 9, the outer diameter D2 of the second thrust bearing 35d is smaller than the outer diameter D1 of the first thrust bearing 35a. That is, the second thrust bearing 35d has a smaller diameter than the first thrust bearing 35a.

In this compressor, similarly to the compressor of Embodiment 1, the second thrust bearing 35d is held between the second flange 431 and the rear wall of the second recess 23c in the axial direction. The second thrust bearing 35d is in contact with the second flange 431 at an annular area E5, which is adjacent to the inner circumference of the second race 362, and in contact with the rear surface of the second recess 23c at an annular area E6, which is adjacent to the outer circumference of the first race 361. That is, also in this compressor, the region where the second thrust bearing 35d is supported by the second support member 43b via the second flange 431 and the region where the second thrust bearing 35d is supported by the second cylinder block 23 via the rear surface of the second recess 23c are radially shifted from each other. Here, as described above, the outer diameter D2 of the second thrust bearing 35d is smaller than the outer diameter D1 of the first thrust bearing 35a. Thus, in this compressor, the area E6 where the first race 361 of the second thrust bearing 35d is in contact with the rear surface of the second recess 23c is smaller than the area E2 where the first race 351 of the first thrust bearing 35a shown in FIG. 4 is in contact with the front surface of the first recess 21c. The other features of this compressor are the same as those of the compressor of Embodiment 1.

In this compressor, since the outer diameter D2 of the second thrust bearing 35d is smaller than the outer diameter D1 of the first thrust bearing 35a, the second thrust bearing 35d is less likely to be elastically deformed in the direction of the axis O of the drive shaft 3 like a disc spring, as compared with the first thrust bearing 35a. That is, also in this compressor, the second spring constant K2 of the second thrust bearing 35d is greater than the first spring constant K1 of the first thrust bearing 35a. Consequently, this compressor is capable of providing the same advantages as those of the compressor of Embodiment 1.

Embodiment 4

As shown in FIG. 10, the compressor of Embodiment 4 does not have the second recessed surface 23d in the rear wall of the second recess 23c unlike the compressor of Embodiment 1. Furthermore, in Embodiment 4, the outer diameter of the second flange 431 is substantially equal to the outer diameter of the second thrust bearing 35b. Thus, in this compressor, the second thrust bearing 35b is held between the second flange 431 and the rear wall of the second recess 23c in the axial direction so that the second thrust bearing 35b is in contact with the second flange 431 over the entire area of the second race 355. Also, the second thrust bearing 35b is in contact with the rear wall of the second recess 23c over the entire area of the first race 354. That is, in this compressor, the second thrust bearing 35b is provided rigidly between the second flange 431 and the rear wall of the second recess 23c. The other features of this compressor are the same as those of the compressor of Embodiment 1.

In this compressor, since the second thrust bearing 35b is provided rigidly, the second thrust bearing 35b is less likely to be elastically deformed in the direction of the axis O of the drive shaft 3 as compared with the first thrust bearing 35a. Therefore, the second spring constant K2 of the second thrust bearing 35b is greater than the first spring constant K1 of the first thrust bearing 35a. The value of the first spring constant K1 of the first thrust bearing 35a needs to be smaller than the value of the second spring constant K2 of the second thrust bearing 35b. Consequently, this compressor is capable of providing the same advantages as those of the compressor of Embodiment 1.

Although the present invention has been described above in line with Embodiments 1 to 4, it is needless to say that the invention is not limited to the above-described Embodiments 1 to 4, but may be appropriately modified and applied without departing from the gist of the invention.

For example, a compressor may be configured by combining the features of the compressors of Embodiments 1 to 4 as appropriate.

Furthermore, the first thrust bearing 35a and the second thrust bearings 35b to 35d may be made of different materials so that the second spring constant K2 of the second thrust bearings 35b to 35d is set greater than the first spring constant K1 of the first thrust bearing 35a.

The control mechanism 15 may be configured such that the control valve 15c is provided in the supply passage 15b and the orifice 15d is provided in the bleed passage 15a. In this case, the opening degree of the supply passage 15b can be adjusted by the control valve 15c. As a result, the pressure of the control pressure chamber 13c can be increased rapidly due to the pressure of the refrigerant gas in the second discharge chamber 29b, and therefore, the discharge capacity can be increased rapidly.

VARIABLE DISPLACEMENT SWASH PLATE TYPE COMPRESSOR

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CPC

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Priority Claims (1)