JPS6157955B2 - - Google Patents

Description

[Detailed description of the invention]

The present invention particularly relates to a vane compressor used in car air conditioners and the like. Before explaining the present invention, first, a sliding vane type rotary compressor for a car cooler will be explained. As shown in Fig. 1, a general sliding vane type compressor consists of a cylinder 1 having a cylindrical space inside, and a blade chamber 2, which is the internal space of the cylinder, fixed to both sides of the cylinder 1. A sealing side plate (not shown in FIG. 1), a rotor 3 eccentrically arranged in the cylinder 1, and a vane 5 slidably engaged in a groove 4 provided in the rotor 3. configured. 6 is a suction hole formed in the side plate, and 7 is a discharge hole formed in the cylinder 1.
The vanes 5 are projected outward by centrifugal force as the rotor 3 rotates, and their tip surfaces slide on the inner wall surface of the cylinder 1 to prevent gas leakage from the compressor. Compared to reciprocating compressors, which have a complex structure and many parts, this type of sliding vane rotary compressor has a smaller and simpler structure, and has recently been applied to compressors for car coolers. It became like that. However, this rotary type has the following problems compared to the reciprocating type. That is, in the case of a car cooler, the driving force of the engine is transmitted to the pulley of the clutch via a belt, which drives the rotating shaft of the compressor. Therefore, when a sliding vane compressor is used, its refrigerating capacity increases almost linearly in proportion to the rotational speed of the car engine. On the other hand, when using a conventionally used reciprocating type compressor, the followability of the suction valve becomes poor at high speed rotation, and the compressed gas cannot be sufficiently sucked into the cylinder, resulting in a decrease in refrigeration capacity. At high speeds, it becomes saturated. In other words, with a reciprocating type, the refrigerating capacity is automatically suppressed when running at high speeds, whereas with a rotary type, this effect does not occur, resulting in a reduction in efficiency due to increased compression work, or overcooling (cooling). (too much). As a method to solve the above-mentioned problems of the rotary compressor, a control valve is configured in the flow passage leading to the suction hole 6 of the rotary compressor, and the opening area of the flow passage changes, and the opening area is narrowed during high-speed rotation. Conventionally, methods have been proposed for performing capacity control using the suction loss. However, in this case, the above-mentioned control valve must be added separately, resulting in a complicated configuration and high cost. As another method for solving the problem of excessive capacity of a rotary compressor at high speeds, a structure has been proposed in which a fluid clutch, a planetary gear, etc. are used to prevent the rotational speed from increasing beyond a certain level. However, for example, the former has a large energy loss due to frictional heat generated by the relative moving surfaces, and the latter has a large size and shape due to the addition of a planetary gear mechanism with many parts, and with the trend toward energy saving, it has become increasingly simpler and more compact. It is difficult to put it into practical use in these days when there is a demand for technology. In order to solve the above-mentioned problems associated with the use of rotary refrigeration cycles for car coolers, the present inventors conducted a detailed study of the transient phenomenon of blade chamber pressure when using a rotary compressor. However, by appropriately selecting and combining parameters such as the suction hole area, discharge amount, and number of blades, it is possible to effectively self-suppress the refrigerating capacity during high-speed rotation, just as in the conventional reciprocating system. We have already found a special patent application filed in 1982.
No. 134048 (Japanese Unexamined Patent Publication No. 57-70986) is pending. In the invention of the above application, between the rotor and the cylinder,
In a vane compressor where the cylinder top is the closest part to the other, the angle from the cylinder top to the end of the vane on the cylinder side is θ radian, centered on the rotation center of the rotor. , the value of the angle θ radian at the end of the suction stroke is θ _s radian, and the value of the angle θ radian at the end of the suction stroke is θ _s radian.
The volume of the blade chamber when the angle is radian is V ₀ cc, and the effective area of the suction passageway from the evaporator to the blade chamber when the angle θ is radian is a(θ) cm ² .
When the weighted average is ＝∫〓 ₀ ^s θ ² a(θ)dθ／∫〓 ₀ ^{s θ 2} dθ, the vane shape is set so that the parameter θ _s /V ₀ is in the range of 0.025<θ _s /V ₀ <0.080. If the compressor is configured under the conditions found in the above invention, suction pressure loss can be minimized at low speeds, and effective pressure loss occurs only at high speeds. Therefore, effective capacity control can be achieved with a simple configuration that does not require any addition to a conventional rotary compressor. However, although the above invention can minimize the suction pressure loss at low speeds, in a small compressor with a relatively small refrigerating capacity,
This loss sometimes became a problem. The present invention is an improvement on the above-mentioned invention, and in a vane type compressor that utilizes suction loss, even if the same suction effective area is used, the speed is low even if the suction volume is the same by selecting the volume curve of the blade chamber. By focusing on the fact that the total weight of refrigerant sucked at different times is different, the cylinder shape is made non-circular, and the volume curve of the blade chamber during the suction stroke with respect to the rotation of the rotor has a roughly flat part. This ensures the amount of refrigerant supplied from the suction hole to the blade chamber during the suction stroke, and prevents suction pressure loss especially during low speed rotation. Hereinafter, as an example, a case will be described in which the present invention is applied to a four-vane type sliding vane compressor. FIG. 2 is a front sectional view of a compressor showing one embodiment of the present invention, in which 11 is a cylinder, 12 is a vane, 1
3 is a vane sliding groove, 14 is a rotor, 15 is a suction hole A, 16 is a suction groove, 17 is a suction hole B, and 22 is a discharge hole. Hereinafter, the suction stroke of this compressor will be explained using FIG. 3A to 3E. 18-1 is blade chamber A, 18-2 is blade chamber B, 1
9 is the top portion of the cylinder 11, 20-1 is the vane A, 20-2 is the vane B, and 21 is the end of the suction groove. The position where the tip of the vane A20-1 passes through the top part 19 of the cylinder 11 with the center of rotation of the rotor 14 as the center is θ=0, and the angle at any position of the tip of the vane with θ=0 as the origin is θ.
shall be. Focusing on the vane chamber A18-1, FIG. 3A shows a state in which the vane A20-1 passes through the top portion 19 and travels in the suction groove 16. FIG.
Refrigerant is supplied through the suction groove 16. In the embodiment, by forming the suction groove 16 sufficiently deep on the inner surface of the cylinder 11, the effective area of the suction groove ₁₆ is:
It was made so that a ₂ ≫ _{a 1} . Therefore, in the conditions shown in Figures A and B, the effective suction area of the flow passage connecting the blade chamber A18-1 and the refrigerant supply source is almost the same as the effective area of the suction hole A15:
a ₁ . Figure C shows the state immediately after the vane A20-1 passes over the suction hole B17 and the vane B20-2 passes through the suction groove end 21 at the same time. At this point, from the suction hole A15 to the blade chamber A18-
The supply of refrigerant to the refrigerant 1 is cut off by the vane B20-2, and supply from the suction hole B17 is started instead. In the example, when the effective area of the suction hole B17 is _a3 , the suction hole B17 is formed so that _a3 = _a1 . Therefore, in this compressor, the effective suction area of the suction flow path from the refrigerant supply source to the blade chamber A18-1 is always constant during the suction stroke. FIG. 2 shows a state in which the traveling angle θ of the vane A20-1 has reached 1/2 of the traveling angle of the entire stroke (suction/compression stroke). In a normal four-vane compressor configured with a perfectly circular cylinder, θ=θ _s1 ≒225°, and the blade chamber volume reaches its maximum at this point. However, in the embodiment of the present invention, the suction stroke is not yet completed at this point, and refrigerant is still supplied to the blade chamber A18-1 from the suction hole B17. Figure E shows the state immediately after the vane B20-2 passes the suction hole B17, and the supply of refrigerant from the suction hole B17 is blocked by the vane B20-2, so the suction stroke ends at this point. . Now, the compressor in one embodiment of the present invention is configured under the following conditions.

【table】

[Table] In the examples of the present invention, a cylinder shape was used, which was formed from a combination of two perfect circles and had a center-to-center spacing of ε. As shown in Figure 4, O ₂ is the center of the left cylinder, O ₃ is the center of the right cylinder, and O ₂ is the center of the right cylinder.
At the same distance from O ₃ , the center of the rotor 14: O ₁
was placed. Vane traveling angle of the blade chamber formed by the cylinder 11, rotor 14, vane, and side plate: θ
The volume curve: Va (θ) for the distance is as shown in Fig. 5 (c) with the interval: ε as a parameter. By the way, curve A shows the volume curve of a conventional compressor in which the cylinder is formed of only one perfect circle, curve B shows the volume curve when ε=5 mm, and curve C shows the volume curve when ε=8 mm in the case of this example. shows the case where ε=10mm. Eccentricity: As ε increases, the change in the volume curve around θ = θ _s1 = 225° becomes smaller; for example, when ε
= 8 mm, it can be seen that it becomes almost flat in the range of 200° < θ < 250°. In the example, vane traveling angle: θ=θ _s2 =250
so that refrigerant is supplied to the blade chamber until
A suction hole B17 was arranged. Conventionally, in the case of 4 vanes, the angle at which the suction stroke ends is set around θ = θ _s1 = 225°, where the volume of the vane chamber: Va is maximum, but by using this cylinder shape, θ = θ _s2 = 250°. It was possible to extend the suction stroke end angle: θ _s2 to °. When a conventional cylinder shape is used, if the above-mentioned θ _s1 is extended, the volume of the blade chamber decreases, resulting in suction loss. When this cylinder shape is used, the flat part of the volume curve can be used, and the above-mentioned suction loss does not occur. Hereinafter, a characteristic analysis performed to understand in detail the transient phenomenon of refrigerant pressure, which is an important point of the present invention, will be described. One blade chamber (for example, blade chamber A)
Focusing on 18-1), and assuming that the pressure of the refrigerant supply source: Ps is always constant, the transient characteristics of the blade chamber pressure can be described by the following energy equation. C _P /AGT _A −PadVa/dt+dQ/dt=d/dt
(C _V /AγaVaTa) (1 equation) In the above 1 equation, G: weight flow rate of refrigerant;
Va: Volume of the blade chamber, A: Heat equivalent of work, C _P : Specific heat at constant pressure, T _A : Supply side refrigerant temperature, κ: Specific heat ratio, R:
Gas constant, C _V : Specific heat at constant volume, Pa: Impeller chamber pressure,
Q: amount of heat, γa: specific weight of the refrigerant in the blade chamber, Ta: temperature of the refrigerant in the blade chamber. In addition, in the following equations 2 to 4, a: effective suction area, g: gravitational acceleration,
γ _A : Specific weight of supply side refrigerant, Ps: Supply side refrigerant pressure. In Equation 1, the first term on the left side is the thermal energy of the refrigerant that passes through the suction hole and is brought into the blade chamber per unit time, the second term is the work done by the refrigerant pressure on the outside per unit time, and the third term is the external wall. The right side shows the increase in the internal energy of the system per unit time. Assuming that the refrigerant follows the ideal gas law and that the suction stroke of the compressor is rapid, we assume that the change is adiabatic.
Since γa=Pa/RTa and dQ/dt=0, the following equation is obtained. G=dVa/dt(A/C _P T _A +1/κRT _A ) Pa +Va/κRT _A dPa/dt (2 equations) Also, using the relational expression 1/R=A/C _P +1/κR, G= 1/RT _A・dVa/dt・Pa+Va/κRT _A d
Pa/dt (3 formulas) Nozzle theory can be applied to the weight flow rate of refrigerant passing through the suction hole. Therefore, by solving equations 3 and 4 simultaneously, the transient characteristics of the blade chamber pressure can be obtained. Figure 7 shows t=
The transient characteristics of the blade chamber pressure were determined using the rotation speed as a parameter under the initial conditions of 0 and P=Ps. In addition, since R12 is usually used as the refrigerant for car cooler refrigeration cycles, κ=1.13 and R=
668Kg・cm/〓Kg, γ _A = 16.8×10 ^-6 Kg/cm ³ , T _A =
The analysis was conducted as 283〓.

[Table] In Figure 7, at ω = 1000 rpm, θ = 210
At °, ω = 1500 rpm, the blade chamber pressure reaches the supply pressure: Ps = 3.18 Kg/cm ² abs near θ = 240°.
Therefore, at the rotation speed within this range, sufficient refrigerant is supplied to the blade chamber at the end of the suction stroke, so
No inhalation losses occur. As the rotation speed increases, at the end of the suction stroke: θ=
It can be seen that since a large pressure drop occurs at 250°, the total weight of refrigerant in the blade chamber decreases, and capacity control works effectively. FIG. 8 shows the pressure characteristics of the blade chamber in the case where the compressor is configured with a conventional perfect circular cylinder (FIG. 5 A) and in the case of the embodiment of the present invention, with the same effective suction area:
The comparison was made using a ₁ = a ₂ = 0.2 cm ² . The solid line indicates a perfect circular cylinder, the chain line indicates the case of this embodiment, and a, b, c and A, B, C are for the case where ω=1000, 1500, and 2000 rpm, respectively. For example, when ω = 1000 rpm, despite the same effective suction area, in a perfectly circular cylinder (Figure a), θ
= θ _s1 = 225°, the blade chamber pressure: Pa has not yet reached the supply pressure: Ps, and there is a pressure loss of about ΔP=0.1 Kg/cm ² . However, in the case of this embodiment (Figure A), θ=
At 210°, the supply pressure has already reached Ps. In this way, the present invention is characterized by focusing on the fact that even if the same effective suction area is used, the total weight of refrigerant suction differs depending on the selection of the blade chamber volume curve or the selection of the cylinder shape. It is. The present invention can also be applied to a compressor in which the cylinder has conventionally been approximately elliptical in shape and the rotor is disposed at the center thereof. This type of compressor has a cylinder shape such as
Although it is often formed as a function of sin2θ, in order to apply the present invention, as in the case of the embodiment, the rate of change of the volume curve near the end of the suction stroke must be The cylinder shape may be selected so as to be smaller, and it is more preferable to have a substantially flat portion. An example is shown in FIG. 200 is a rotor circle centered on O ₃ with radius Rr, and 201, 202, 203, and 204 are cylinder circles with radius Rc centered on Q ₁ , O ₂ , O ₄ , and O _5, respectively. Distance between centers of O ₁ and O ₂ and O ₄ and O ₅ : ε is Rr,
It may be sufficiently small compared to various dimensions such as Rc, and other curves may be used for the point where the two circles intersect, which is sufficiently far from N, considering the running stability of the vane, etc. . As described above with reference to the embodiments, the vane compressor of the present invention improves the refrigerating capacity during high-speed operation by utilizing the suction loss in which the pressure in the vane chamber during the suction stroke drops below the refrigerant supply source pressure. A compressor that performs suppression includes a rotor with slidable vanes, a non-perfectly circular cylinder that accommodates the rotor and the vanes, and a cylinder that is fixed to both sides of the cylinder and that includes the vanes, the rotor, and the cylinder. The cylinder top is composed of a side plate that seals the space of the blade chamber formed by the cylinder on its side, and a suction hole and a discharge hole that are flow passages that communicate the blade chamber with the outside, and in which the rotor and the cylinder are close to each other. , with the rotation center of the rotor as the center, and an angle in the vicinity of an angle at which the volume of the blade chamber is at a maximum during the suction compression stroke with respect to the angle from the cylinder top part to the cylinder side end of the vane. , by forming the cylinder shape so that the volume curve of the blade chamber has a substantially flat part, and by forming the suction hole so that the flat part is included in the suction stroke, the blade chamber is in the suction stroke. This ensures the supply of refrigerant from the suction hole and prevents suction pressure loss, especially during low speed rotation. Note that although the embodiments of the present invention have been described with reference to a four-vane type, the present invention is not limited to the four-vane type.

[Brief explanation of the drawing]

FIG. 1 is a front sectional view of a conventional 4-vane type compressor, and FIG. 2 is a 4-vane type compressor showing an embodiment of the present invention.
A front cross-sectional view of the vane compressor, FIGS. 3A to 3E are explanatory views showing the suction stroke of the compressor of the same embodiment, FIG.
Figure 6 is a graph showing the volume curve of the compressor, Figure 6 is a graph showing the blade chamber pressure characteristics of the compressor in the embodiment of the present invention, and Figure 7 is the graph showing the blade chamber pressure characteristics of the conventional compressor and the compressor of this embodiment. Graph comparing machines,
FIG. 8 is a graph of pressure drop rate versus rotational speed, and FIG. 9 is an explanatory diagram showing another embodiment of the present invention. 11...Cylinder, 12...Vane, 14...
Rotor, 15, 17...Suction hole, 18-1...Blade chamber.

Claims (1)

[Claims] 1. In a compressor that suppresses the refrigerating capacity during high-speed operation by utilizing a suction loss in which the pressure in the blade chamber during the suction stroke is lower than the refrigerant supply source pressure,
a rotor on which vanes are slidably provided; a non-circular cylinder that accommodates the rotor and the vanes; and a blade chamber fixed to both sides of the cylinder and formed by the vanes, the rotor, and the cylinder. The rotor has a cylinder top portion in which the rotor and the cylinder are close to each other; With respect to the angle from the top of the cylinder to the end of the vane on the cylinder side, the cylinder shape is set near the angle at which the volume of the blade chamber is maximum during the suction compression stroke. A vane compressor, wherein the volume curve of the chamber is formed to have a substantially flat part, and the suction hole is formed so that the flat part is included in the suction stroke.