JPH0235160B2 - KAITENATSUSHUKUKI - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION The present invention relates to a compressor that is effective for use in a vehicle cooling system and the like, and particularly relates to the structure of a vane type rotary compressor.

Reciprocating type compressors have been the mainstream of this type of compressor, but recently vane type rotary compressors have come into use due to their small size and light weight. However, in conventional vane-type compressors, the sliding speed in the sliding parts between the rotor and liner, and the vanes and liner is high, resulting in large friction losses and a low cooling effect (coefficient of performance) for the input power. It has the following drawbacks. Furthermore, the sliding portion has a serious drawback in that it is subject to severe wear and is susceptible to seizure during high speed rotation.

The present invention was devised in view of the above-mentioned drawbacks of the vane type rotary compressor, and has a configuration that slows down the sliding speed in the sliding portion, thereby reducing wear loss and achieving a high coefficient of performance during high-speed rotation. The purpose of the present invention is to provide an excellent rotary compressor that does not cause seizure.

An embodiment of the present invention will be described below.

In FIG. 1, reference numeral 1 denotes a drive shaft, which is rotatably supported by a bearing 14 with respect to the housing 8, and supported by a bearing 10 with respect to a front plate 6 that is integrally fixed with the housing 8. There is. Therefore, the shaft 1 is assembled so as to be rotatable about the rotation axis A. The driven shaft 2 is supported by a bearing 11 on a rear plate 7 that is integrally fixed to the housing 8, and is assembled so as to be rotatable about the rotation axis A.

The rotor 3 is provided on the drive shaft 1 and the driven shaft 2, and is located at a second rotor that is eccentric by a certain amount with respect to the rotation axis A.
Both ends are pivotally supported by bearings 12 and 13 located on a rotation axis B, and the assembly is rotatably assembled around the rotation axis B. The liner 5 is a member that is integrally fixed in the housing 8 by being sandwiched between the front plate 6 and the rear plate, and has a circular inner circumferential surface centered on the rotation axis A as shown in FIG. The outer diameter portion of the rotor 3 forms an extremely small gap.

The vane 4 is slidably installed in a slit formed in the rotor 3. Chip vanes 4a are arranged at both ends of the vane 4,
It slides on the inner peripheral surface of the liner 5. As shown in FIG. 3, the front plate 6 has an internal gear 6a having 24 teeth centered on the rotation axis A, and has a 24-tooth internal gear 6a centered on the axis of the rotor 3 (coinciding with the second rotation axis B). It is assembled so as to mesh with an external gear 3a having 16 teeth provided as an external gear.

The housing 9 has a suction port 9a and a discharge port 9b, and is integrally fixed with the housing 8, the front plate 6, the liner 5, the rear plate 7, etc. with bolts 18. 1a and 2a are the driving shaft 1 and the driven shaft 2.
These are balance weights that are assembled and fixed with bolts 19 and 20, respectively. 15 and 16 are discharge valves and valve stoppers located above the three discharge ports 7a provided in the rear plate 7, and 17 is a shaft sealing device for preventing leakage of fluid from within the rotary compressor.

Next, the operation of the compressor having the above configuration will be explained.

When the drive shaft 1 rotates in the direction of arrow _N1 in FIG. 1 about the rotation axis A, the rotor 3, which is supported on a second rotation axis B eccentric to the rotation axis A, rotates as shown in FIG. It revolves on the orbit R shown by the dashed line in the direction of arrow _N1 . The driven shaft 2 rotates around the rotation axis A by supporting the rotor 3 on the rotation axis B. At this time, the rotor 3 revolves while maintaining a small gap between its outer diameter and the inner peripheral surface of the liner 5.

On the other hand, the internal gear 6a machined on the front plate 6, which is integrally fixed to the housing 8, meshes with the external gear 3a on the shaft of the rotor 3, so that the rotor 3 is N ₁ as shown in FIG. It will revolve on the orbit R in the direction. Therefore, the rotor 3 rotates about the rotation axis B in the _N2 direction. Furthermore, as shown in FIG. 2, a vane 4 that is slidable in the radial direction and a tip vane 4a that is slidable in the radial direction in slits formed at both ends of the vane are incorporated in the rotor 3. The vanes 4 and the tip vanes 4a slide against the inner circumferential surface of the liner 5 due to the centrifugal force caused by the revolution and rotation of the rotor 3. The working space of this rotary compressor has a cross-sectional shape surrounded by the outer circumference of the rotor 3, the inner circumference of the liner 5, the vanes 4, and the tip vanes 4a, and is constituted by the end faces of the front plate 6 and rear plate 7 in the axial direction. be done.

The suction and discharge operations of this rotary compressor will be explained below with reference to FIG. The figure shows the angle at which the rotation axis B revolves around the rotation axis A in 60° steps. 3b is the suction port, which is indicated by 9a in FIG. It communicates with a refrigerant passage led into the rotor 3 through a space in the rear plate 7, and the vane 4
There are two locations near the. Reference numeral 7a denotes discharge ports, which are provided at three locations at positions where the inner circumference of the liner 5 is equally divided into three parts from the rotation axis A.

When the rotor 3 rotates 60 degrees from state 1 in Fig. 4, it becomes state 2 in Fig. 4, and the rotor 3 revolves 60 degrees in the _N1 direction and N2 _.
It rotates 30 degrees in the direction. This is due to the nature of the planetary gear mechanism, which has a tooth ratio of 3/2, with the internal gear 6a having 24 teeth and the external gear 3a having 16 teeth, as shown in FIG. Similarly, the cross section of the working space changes in 60 DEG steps as shown in FIGS. 4-6.

Focusing on the hatched working space a in Fig. 4 and looking at the changes in each 60° step, Fig. 4 1~
In FIG. 4, the refrigerant is in the process of expansion, and during this period, the refrigerant is sucked into the working space a from the suction port 3a. Between FIG. 4 3 and FIG. 4 , the working space a is reduced, but during this period, the suction port 3 b is in communication with the working space a, so the refrigerant gas flows back from the working space a to the suction port 3 b. The working space a shown by hatching in FIG. 4 is the suction volume of the rotary compressor.
Between FIGS. 4-6, the working space a performs a reduced compression and discharge action.

As mentioned above, the working space a of the rotary compressor of this example performs the suction-compression-discharge stroke in one rotation of the rotor 3, but by having three working spaces, the rotary compressor of this example performs the suction-compression-discharge stroke in one rotation of the rotor 3. Perform the suction-compression-exhalation stroke.

Comparing the sliding speeds of the sliding parts between the rotor and liner and between the vane tip and liner, which are the causes of friction loss and seizure phenomena in conventional vane type compressors, with this rotary compressor, we find the following: become that way.

First, the sliding speed V ₁ between the rotor 3 and liner 5 is determined by the rotor diameter Ro,
If the angular velocity is Wo, then V ₁ = RoWo...(1). On the other hand, the sliding speed V ₂ in the rotary compressor according to this example is as follows, assuming that the revolution radius of the rotor 3 is r:
When the angular velocity of revolution of the rotor 3 is Wo, the angular velocity of rotation is -Wo/2, so _V2 =(r-Ro/2)Wo...(2). As an example, liner radius 40mm, rotor radius
Comparing the case of 30 mm, the revolution radius ro of the rotary compressor according to this example is 10 mm, and V ₁ = 30Wo ……(1)′ V ₂ = (10−30/2)Wo=−5Wo ……( 2)', V ₁ :V ₂ =30:-5=6:-1, and the sliding speed _V2 of the rotary compressor according to this example is actually ₁ / of the sliding speed V1 of the conventional vane compressor. 6, resulting in an extremely slow sliding speed.

Furthermore, since the sliding parts of the rotor 3 and liner 5 according to this example change every moment as the rotor 3 revolves, the seizure phenomenon that occurs in conventional vane type compressors does not occur at all.

Next, compare the sliding speeds of the vane and liner. In a conventional vane type compressor, the vane makes one revolution along the liner for one rotation of the rotor, so the average speed V ₃ is as follows, where R _{1 is the liner radius: V 3 =} R ₁ Wo (3). In the rotary compressor according to this example, since the rotor 3 performs 1/2 rotation during one revolution, the average speed V ₄ is V ₄ = 1/2・R ₁ Wo, which is different from that of the conventional vane compressor. It becomes 1/2. Therefore, in the rotary compressor according to this example, the friction loss between the vane and the liner is reduced to about 1/2, and excellent effects such as improved cooling effect and less wear on input power can be obtained.

In addition, the balance weights indicated by 1a and 2a in Fig. 1 rotate integrally with the drive shaft 1 and the driven shaft 2, canceling out the rotational imbalance caused by the revolution of the rotor 3 and suppressing vibrations in the rotary compressor. This has the effect of reducing the

FIG. 5 shows a second embodiment.

In this embodiment, a suction valve 22 is disposed between the front plate 6 and the liner 5 to block or allow communication between the suction port 6b of the front plate 6 and the working space.
The structure is such that refrigerant gas flowing in from a suction port 8a provided in the working space is guided into the working space. Further, the liner 5, the rotor 3, and the vanes 21 that constitute the working space are arranged as shown in FIG. 6, and in this embodiment, two vanes 21 are installed so as to be able to slide through the slits in the rotor.

The operation of the rotary compressor in this embodiment will be explained next.

As in the first embodiment, when the rotor 3 is driven to revolve in the direction shown in FIG. 6 _N1 , the rotor 3
N Rotates in _two directions, expanding the volume of the working space.
Reduce the size and perform the suction, compression, and exhalation strokes. Figure 7 shows how the cross section of the working space changes at every 60° rotation angle step. If we pay attention to the working space a shown by hatching in the same figure, we can see that
In this state, the working space a is in the process of expanding, and at this time a suction stroke occurs in which refrigerant gas flows in from the suction port 6b via a suction valve (not shown). The suction volume per working space of this rotary compressor is the volume of the working space a in FIG. 7, 3.

In the states shown in FIGS. 7-3 to 7-6, the working space a is in the process of shrinking, and at this time the suction port 6b is shut off by the suction valve, and the refrigerant gas in the working space is compressed, not shown. It passes through the discharge valve and is discharged from the discharge port 7a. In this embodiment as well, there are three working spaces, so the number of working spaces per rotation is 3.
The suction, compression, and discharge strokes are performed.

In addition, in this embodiment, the suction volume per rotation is three times the volume of the working space a shown by the hatching in FIG. This has the effect of increasing the volume of the working space a (three times the volume of the working space a).

Furthermore, the sliding speed between the rotor and liner and between the rotor and vane tips is extremely slow compared to conventional vane compressors, reducing friction loss, which significantly increases the cooling effect achieved with respect to input power. The effects and the effect of avoiding the abrasion and calcination phenomenon of the sliding portion are as shown in the above-mentioned first embodiment.

In the rotary compressor according to the present invention, in both the first and second embodiments, the number of teeth of the internal gear 6a provided on the front plate 6 and the external gear 3 provided on the rotor 3 are different.
Although the ratio of the number of teeth of a is set to 3:2, this is not necessarily limited to 3:2. That is, by rotating the rotor 3 by 1/2 during one revolution of the rotor 3, the rotation speed of the drive shaft 1 is 3.
In order to be able to perform suction, compression, and discharge twice, at least the discharge port 7a is arranged three times along the liner 5.
In the first embodiment, two suction ports were provided in the rotor, and in the second embodiment, three suction ports were provided along the liner. According to this configuration, the machining of these discharge ports and suction ports can be reduced, but the ratio of the number of teeth is not limited to 3:2 and can be set arbitrarily.

Furthermore, in the first and second embodiments of the present invention, the working space was configured to have three working spaces each consisting of a liner, a rotor, and a vane, but by reducing or increasing the number of vanes, the working space was It goes without saying that the effects of the present invention can be obtained even when the number is decreased or increased.

As explained above, in the compressor of the present invention, the rotor is configured to revolve around the inner circumferential surface of the liner as the drive shaft rotates, and also to rotate in the opposite direction to the revolution direction due to the meshing of the gears. The sliding speed between the vane tip and the liner becomes relatively slow,
Therefore, it has an excellent effect of effectively preventing problems such as image sticking.

[Brief explanation of drawings]

FIG. 1 is a cross-sectional view showing an embodiment of the compressor of the present invention, FIG. 2 is a cross-sectional view taken along the - arrow in FIG. 1, FIG. 3 is a cross-sectional view taken along the - arrow in FIG. 2,
3, 4, 5, and 6 are explanatory views for explaining the operation of the compressor shown in FIG. 1, FIG. 5 is a sectional view showing another example of the compressor of the present invention, and FIG. The sectional views and FIGS. 7, 1, 2, 3, 4, 5, and 6 are explanatory diagrams for explaining the operation of the compressor shown in FIG. 1... Drive shaft, 3... Rotor, 3a...
Internal gear, 4...vane, 5...liner, 6a...
...Internal gear.

Claims (1)

[Scope of Claims] 1. A cylindrical liner, a cylindrical rotor disposed on the inner surface of the liner, and a vane that is slidably incorporated into a slit formed in the rotor and abuts on the inner surface of the liner. A rotary compressor comprising a drive shaft having a crank mechanism for revolving the rotor, an external gear integrally fixed to the rotor, and a fixed internal gear meshing with the external gear. 2. In the rotary compressor according to claim 1, the ratio of the number of teeth of the external gear and the internal gear is 3:
2. A rotary compressor characterized by: