JPS603494A - Vane compressor - Google Patents

Abstract

PURPOSE:To aim at reducing fluctuations in torque and pulsations in discharge, by providing such an arrangement that the contact positions between a rotor outer peripheral surface and a cylinder inner peripheral surface and as well discharge ports are set at unequal angular intervals with respect to the center of a rotor, and therefore, discharging timings of the discharge ports are shifted from each other. CONSTITUTION:The contact positions A, B between a rotor outer peripheral surface 42 and a cylinder inner surface are set at unequal angular intervals. Therefore, although vanes 40 are arranged at equal angular intervals, discharge timings at discharge ports 50, 54 are shifted from each other. That is, at the discharge ports 54, the discharge is carried out with a delay from that at the discharge port 50. Suction timings at suction ports 52, 58 are also shifted from each other. As a result, complication in the structure, an increase in power loss and a decrease in volmetric efficiency are avoided as possible as it can, and as well, fluctuations in torque necessary for rotating a rotor 36 and pulsations in discharge may be reduced.

Description

DETAILED DESCRIPTION OF THE INVENTION Technical Field The present invention relates to a vane compressor, and in particular to reducing discharge pulsation and load torque fluctuation thereof.

A type of prior art vane compressor is a type of vane compressor in which the rotor, which is an outer peripheral surface or a cylindrical surface, is brought into contact with the inner peripheral surface of the cylinder at multiple locations, thereby dividing the space within the cylinder into multiple sections, and the rotor having multiple spaces within the cylinder. There is a type in which the hones are provided at equal angular intervals.

The suction hole and the discharge hole are provided on both sides of the contact point between the rotor and the cylinder, and the number of vanes is often an integral multiple of the number of divisions of the cylinder interior space.

In this type of vane compressor, discharge is performed from multiple discharge holes at the same time. For example, in a vane compressor in which the inner peripheral surface of the cylinder is elliptical and contacts the outer peripheral surface of the rotor at two places, four vanes are normally provided, and in this case, there are two discharge holes. In this case, torque fluctuations and discharge pulsations corresponding to a four-cylinder reciprocating compressor occur.

It is desirable to reduce these torque fluctuations and discharge pulsations as much as possible from the viewpoints of downsizing the drive source, improving durability, and reducing operating noise. Therefore, various countermeasures have been taken in the past. For example, JP-A-58
．． -1318.5 Publication and JP-A-57'-210196
No. 3, No. 2003-119119 and the like describe a compressor in which a cylinder and a rotor are in contact with each other at three locations, discharge holes are provided at three locations, and the rotor is provided with five vanes. In this way, the timing of discharge from the three discharge holes is shifted, torque fluctuation and discharge pulsation are reduced, and the compressor becomes equivalent to a 15-cylinder reciprocating compressor. However, on the other hand, it is inevitable that the structure becomes complicated, the power increases, and the volumetric efficiency decreases.

Further, Japanese Utility Model Application Publication No. 57-144287 describes that a plurality of vanes are provided at unequal angular intervals to stagger the timing of discharge from a plurality of discharge holes. This is effective in terms of suppressing discharge pulsations and torque fluctuations, but on the other hand, it creates compression chambers with poor volumetric efficiency among the multiple compression chambers (space between two adjacent vanes). easy.

For example, before the volume of a compression chamber reaches its maximum, the trailing vane of the two vanes that partition the compression chamber may come off the suction hole, or even if the volume of a compression chamber begins to decrease, the trailing vane may come off the suction hole. There is a tendency for the vanes to not completely pass through the suction holes, resulting in gas backflow, and it is difficult to adjust the volumetric efficiency between the plurality of compression chambers. 1 Further, Japanese Utility Model Application Publication No. 5'l-32093 describes that the inner circumferential surface of the cylinder is expanded larger in a portion closer to the suction hole than in a portion closer to the discharge hole.

In this way, the suction stroke can be completed within a relatively small rotation angle range of the rotor, and the compression and discharge strokes can be performed within a large rotation angle range, thereby making it possible to reduce torque fluctuations. However, since ejection is still performed simultaneously from a plurality of ejection holes, no great effect can be expected in reducing ejection pulsation.

Purpose of the Invention The present invention has been made against the background of the above-mentioned circumstances, and its purpose is to reduce torque fluctuation while avoiding as much as possible the complication of the structure, increase in power loss, and decrease in volumetric efficiency. Another object of the present invention is to provide a vane compressor that can reduce discharge pulsation.

Structure of the Invention The gist of the present invention is that a rotor whose outer circumferential surface is a cylindrical surface is brought into substantially contact with the inner circumferential surface of the cylinder at a plurality of locations, thereby dividing the space inside the cylinder into a plurality of spaces, and dividing the space into a plurality of spaces. In a vane compressor in which a suction hole and a discharge hole are provided on both sides of the contact point, and the rotor is provided with an integral number of vanes equal to the number of partitions of the space at equal angular intervals, the outer peripheral surface of the rotor and the inner periphery of the cylinder The point of contact with the surface and the formation position of the discharge holes are arranged at unequal angular intervals with respect to the center of the rotor.

Effects of the Invention With the above configuration, there will be a difference in the timing of discharge from multiple discharge holes. For example, if there are two contact points between the rotor and the cylinder and there are four vanes, torque fluctuations and The discharge pulsation is close to that of an 8-cylinder reciprocating compressor.

Moreover, the relationship between the number of contact points between the rotor and cylinder and the number of vanes is the same as that of a normal vane compressor, so there is no need for the structure to become double 5' (L) or for increased power loss. In addition, since the plurality of vanes are provided at equal angular intervals, the relative position of each vane and each suction hole and discharge hole is It is possible to optimally set the formation position and size of each suction hole and discharge hole without being different for each vane, and it is also possible to avoid a decrease in U volume efficiency.

Embodiments Hereinafter, some embodiments of the present invention will be described in detail with reference to the drawings.

2. FIGS. 1 and 2 show an example in which the present invention is applied to a vane compressor for compressing refrigerant gas in an air conditioner, and 10 in FIG. 1 is a cylinder. cylinder 1
Both end openings of 0 are closed by front side plates 12 and rear side plates 14, respectively.

A front housing 16 is disposed on the side of the freon 1 and the side plate 12 to form a suction chamber 18, and the refrigerant gas returned from the external circulation circuit is guided to the suction chamber 18 through the compressor inlet 20, and further It is designed to be sucked into the cylinder 10. Further, a rear housing 22 is provided to hold the cylinder 10 and both front and rear side plates 112 and 14 inside, and is connected to the front housing 16. As a result, a discharge chamber 24 is formed outside the cylinder 10, and an oil separation chamber 26 is formed behind the rear side plate 14. The refrigerant gas compressed in the cylinder 10 is discharged into the discharge chamber 24 , and is guided to the oil separation chamber 26 through a communication hole 28 formed in the tozide plate 14 , and when passing through the oil filter 30 . The oil present in the refrigerant gas as a mist is separated into the oil separation chamber 2.
It is stored at the bottom of 6. On the other hand, the refrigerant gas after oil separation is sent out from the compressor outlet 32 to the external circulation circuit.

As shown in FIG. 2, the cylinder 10 has an inner peripheral surface 34 with a cross-sectional shape close to an ellipse, and a cylindrical rotor 36 is disposed inside the inner peripheral surface 34. The rotor 36 has four vane grooves 38 provided at equal angular intervals of 90 degrees, and a vane 40 is slidably fitted into each vane groove 3'18. The aforementioned rear side plate 14 is formed with an oil passage (not shown) that allows oil to flow into and out of the bottom of the vane groove 38, and oil from the oil separation chamber 26 is supplied to this oil passage by the discharge pressure. As a result, the vane 40 can rotate together with the rotor 36 while slidingly contacting the cylinder inner circumferential surface 34.

The outer circumferential surface 42 of the rotor 36 is substantially in contact with the cylinder inner circumferential surface 34 at two locations. Here, the term "substantial contact" includes not only cases in which they are actually in contact, but also cases in which they are not in strict contact but are brought close enough to seal the refrigerant gas.

The contact positions between the rotor outer circumferential surface 42 and the cylinder inner circumferential surface 34 are indicated by A and B in FIG. 2, but for convenience, these will be abbreviated as rotor top positions. These rotor top positions Δ.

B is not located on one diameter of the rotor 36.

That is, the rotor top position B when measured clockwise from the rotor top position A to the rotor 360 rotation direction.
The angle α up to this point is less than 180 degrees, and the angle measured counterclockwise is greater than 180 degrees.

As described above, the cylinder inner circumferential surface 34 and the rotor outer circumferential surface 42 are substantially in contact with each other at two points, so that the space inside the cylinder 10 is reduced to the second chamber 44 and the second chamber 46.
It is separated into. A suction hole 48, a discharge hole 50 and a suction hole 5 are provided at both ends of each of these two chambers 44,46.
2 and a discharge hole 54 are provided. As mentioned above, the rotation angle α of the rotor from rotor top position A to rotor I/tube position B is smaller than the rotation angle β from rotor top position B to rotor top position A. The suction hole 48 and the discharge hole 50 are also opened in a narrower angular range than the suction hole 52 and the discharge hole 54, and the space between the compression chamber, the suspension, the rotor top position A or B, and the vane 40 is Or, while the volume of the space sandwiched between two adjacent vanes 40 increases, the suction hole 48 or 52 always communicates with these spaces, and in the process of decreasing the volume of these spaces, these The space of the vane 4 must be in communication with the suction holes 48 and 52.
It is cut off by 0.

Further, since the rotor top positions A and B are arranged at unequal angular intervals, the second chamber 44 is smaller than the second chamber 46, but the portion of the rotor inner circumferential surface 34 that defines the second chamber 44 is The portion of the hole 48 has a shape that bulges outward from the portion on the discharge hole 50 side, thereby reducing the difference in volume with the second chamber 46 as much as possible, and sucking in as much as possible of the rotor 36. The rotation angle range is as small as possible, and the discharge is performed in the widest possible rotation angle range. In this way, torque fluctuations are reduced and an increase in discharge pulsation is avoided.

If the discharge amount is the same, the rotation angle of the rotor during the discharge stroke is larger, and the discharge is performed more slowly. Therefore, the degree of overcompression, that is, the pressure inside the cylinder 10 is lower than the This makes it possible to reduce the degree to which the pressure rises above the internal pressure.

A mounted discharge valve 56 is fixed to each of the discharge holes 50 and 54 with a bolt 60 via a lift amount regulating member 58.

From both end surfaces of the rotor 36, the shaft 6
2 and 64 concentrically protrude, and the shaft 62 is rotatably supported by the front side plate 12 via a radial bearing 66 and is sealed by a shaft sealing device 68 while being attached to the front housing 16.
It is made to protrude to the outside. On the other hand, the shaft 64 is rotatably supported by the rear side plate 1, I4 via a radial bearing 6G, and the ends of the shaft 64 and the radial dovetail 6G are covered by a cap 70. In addition, both front and rear side plates 12 and 14 are provided on both end surfaces of the rotor 3G and the vane 40.
The refrigerant gas is essentially brought into contact with the refrigerant gas to prevent leakage of refrigerant gas.

In the vane compressor configured as above - C1 shaft 6
When rotor 36 and vane 40 are rotated by applying rotational power to cylinder 2, refrigerant gas is sucked into cylinder 10 through suction holes 48 and 52, and discharged into discharge chamber 24 through discharge holes 50 and 54. In this case, although the vanes 40 are arranged at equal angular intervals because the rotor top positions A and B are arranged at unequal angular intervals,
A difference occurs in the discharge timing between the discharge holes 50 and 54.

That is, as is clear from FIG. 2, the discharge is performed on the discharge hole 54 side later than on the discharge hole 50 side. Of course, there is also a difference in the timing of suction entering the suction hole.

Although suction starts earlier in the suction hole 52 than in the suction hole 48, the end of suction is delayed.

FIG. 3 conceptually shows this state.

In FIG. 3, the solid line indicates the pressure fluctuation in the first chamber 44, and the dashed line indicates the pressure variation in the second chamber 46. By shifting the timing of suction and discharge in this way, fluctuations in the torque required to rotate the rotor 36 are reduced. That is, the load torque of the rotor 36 reaches its maximum when the product of the pressure of the refrigerant scum and the area of the vane 40 that receives it reaches its maximum, but the time when this reaches its maximum occurs on the side of the chamber 44. Since the second chamber 46 side and the second chamber 46 side are shifted from each other, the maximum load torque of the rotor 36 becomes smaller. Furthermore, as is clear from FIG. 3, when the discharge is completed in the second chamber 46 side, compression has already progressed to some extent in the second chamber 44 (11J), so the minimum load torque of the rotor 36 is reversed. It is larger than in the conventional compressor,
In this way, if the maximum load torque becomes smaller and the minimum load torque becomes larger, the torque fluctuation range, which is the difference between the two, will of course become smaller.

As mentioned above, since the discharge timings of the discharge holes 5-0 and 54 are different, the discharge pulsation at the compressor outlet 32, where the discharged refrigerant gases from both the ridge holes 50 and 54 are combined, is the fourth. As shown by the solid line in the figure, two peaks appear for one rotation of the rotor 36, which is similar to an eight-cylinder reciprocating compressor. For comparison, the discharge pulsations in a conventional Hoehn compressor, which are at equal angular intervals from the rotor top position, are shown in broken lines in the same figure.
Compared to this pressure fluctuation width ΔP1, the pressure fluctuation width ΔP2 in the compressor of this embodiment is considerably smaller.

In the above embodiment, the cross-sectional shape of the cylinder inner circumferential surface 34 is asymmetrical, but this is not necessarily essential. For example, as shown in FIG. 5, the cross-sectional shape of the cylinder inner circumferential surface 80 remains elliptical as in the conventional compressor, and the diameter of the rotor 82 is slightly smaller than the diameter of a normal rotor, which is indicated by the two-dot chain line. By moving the center along the long axis of the ellipse and bringing the outer circumferential surface 84 into contact with the cylinder inner circumferential surface 80, it is also possible to make the rotor top positions A and B at unequal angular intervals. It is. However, in this case, the volumes of the first chamber 86 and the second chamber 88 are different, and the amount of protrusion of the vane 90 is particularly large on the second chamber 88 side, so there is a difference in the angular interval between the rotor top positions A and B. cannot be made that large. In the above embodiment, in order to avoid this problem, the cylinder inner peripheral surface 3
4 has an asymmetrical cross-sectional shape.

In addition, as shown in FIG. 6, without moving the center of the rotor 92, the cylinder inner circumferential surface 94 is slightly moved by placing an oval ring 96 shown by a two-dot chain line at a position slightly to the right of the rotor center. It is also possible to make the rotor top positions A and B at unequal angular intervals by making the rotor top positions A and B into a crushed shape.

It should be noted that the present invention is most effective when applied to a type of vane compressor in which the inner circumferential surface of the cylinder and the outer circumferential surface of the rotor are in contact at two places, but the present invention is most effective when applied to a type of vane compressor in which the inner circumferential surface of the cylinder and the outer circumferential surface of the rotor contact at two or more places. Of course, it is also possible to apply it to a machine. In this case, the number of vanes is usually at least twice the number of divisions in the cylinder internal space divided by the rotor top position, but in a simple compressor, the number of vanes is the same as the number of divisions in the cylinder internal space. For example, it is also possible to set both to 3 ([/, J or 4).

In addition, it goes without saying that the present invention can be implemented in various modifications and improvements based on the knowledge of those skilled in the art without illustrating each one or departing from the spirit of the present invention.

[Brief explanation of drawings]

FIG. 1 is a front sectional view of a vane compressor for compressing refrigerant gas, which is an embodiment of the present invention, and FIG.
−■ It is a sectional view. FIGS. 3 and 4 are graphs for explaining the advantages of the compressor shown in FIGS. 1 and 2. 5 and 6 are diagrams conceptually showing other embodiments of the present invention, respectively. IO=Cylinder ■8: Suction chamber 20: Compressor population 24: Discharge chamber 32: Compressor outlet 34.94 Two inner peripheral surfaces 36.82,9
2: Rotor 40.90: Vane 42, 84: Outer peripheral surface 44.86:
- Chamber 46.88: Second chamber 4B, 52: Suction hole 50
, 54: Discharge hole 80 Niji cylinder inner peripheral surface r

Claims (3)

[Claims] (1) By substantially contacting the inner circumferential surface of a rotor or cylinder whose outer circumferential surface is a cylindrical surface at multiple points, the space inside the cylinder is divided into multiple sections, and there are suction holes and discharge holes on both sides of the contact points. A vane compressor in which holes are provided in the rotor at equal angular intervals equal to an integral multiple of the number of partitions of the space, the contact point between the rotor outer circumferential surface and the cylinder inner circumferential surface and the discharge outlet. A vane compressor characterized in that the holes are formed at unequal angular intervals with respect to the center of the rotor. (2) The cross-sectional shape of the inner peripheral surface of the cylinder is divided into the plurality of sections even though the contact points between the inner peripheral surface of the cylinder and the outer peripheral surface of the rotor are arranged at unequal angular intervals due to the axially symmetrical shape. The vane compressor according to claim 1, wherein the vane compressor is deformed to the side where the difference in volume between the spaces is as small as possible. (3) The cross-sectional shape of the inner circumferential surface of the cylinder is such that, between at least one set of adjacent contact points between the outer circumferential surface of the rotor and the inner circumferential surface of the cylinder, the portion closer to the suction hole has a shape closer to the discharge hole. 3. The vane compressor according to claim 1, wherein the vane compressor has a shape that bulges more outward in the radial direction than a portion closer to the vane compressor.