JPS6235090A - Rotary compressor - Google Patents

Abstract

PURPOSE:To reduce the pulsation width of a pressure in a suction chamber, by a method wherein a crank angle, at which a pressure wave, produced during a suction stroke and having an incoming velocity to the suction chamber, is reflected by an open end on the accumulator side and returned, is controlled in a specified range. CONSTITUTION:In the case of thetaSR=thetaS+12LsN/aS, wherein thetas(deg) is a suction starting crank angle, Ls(m) is the length of a suction line 13, N(rpm) is the number of revolutions of a shaft, aS(m/s) is a sound velocity in the suction line 13, and thetas(deg) is the return angle of a pressure wave produced on the cylinder 10 side of the suction line 13 during the starting of suction, a compressor is so constituted that the return crank angle is in a range of 296(deg)<=thetaSR<=406(deg). With this constitution, a crank angle theta(deg), at which a pressure wave produced at the pipe end on the cylinder 10 side of the suction line 13 during the suction stroke of a rotary compressor and an incoming velocity to the suction chamber of the cylinder 10 is reflected by the open end of the accumulator 9 side, is controlled in a specified range in the vicinity of a subsequent suction stroke starting crank angle, and this enables reduction of a pressure pulsation width of the suction chamber.

Description

DETAILED DESCRIPTION OF THE INVENTION Field of Industrial Application The present invention relates to a rotary compressor used in an air conditioner.

2. Description of the Related Art A conventional rotary compressor of this type is disclosed in, for example, Japanese Patent Application Laid-Open No.
As shown in Japanese Patent No. 122192, it had a structure as shown in FIG.

In the figure, reference numeral 1 denotes a closed container in which a compressor element 2 and an electric motor element 3 for driving the compressor element 2 via a crankshaft 4 are built-in. Further, the suction muffler 5 is disposed outside the closed container 1, and a suction connecting pipe 6 attached to the suction muffler 5 so as to protrude into the inside from the lower part thereof is communicated with the cylinder 7 of the compressor element 2. Also, the length of the suction connection pipe 6 (im)
is determined by the following formula.

Ru.

The operation will be explained below.

The gas refrigerant enters the volume space within the suction muffler 5 and is sucked into the cylinder 7 via the suction connection pipe 6. When the suction gas passes through the suction connection pipe 6, pressure pulsations occur due to various factors. This pressure pulsation is influenced by the length of the suction connection pipe 6, the suction stroke period, and the sonic velocity of the suction gas. Figure 5 shows the relationship between the suction connecting pipe 6 and volumetric efficiency (refrigerant: R22).
However, if the suction connection pipe 6 is determined by equation (1), a supercharging effect can be obtained and the volumetric efficiency can be improved.

Problems to be Solved by the Invention However, in the above configuration, although the volumetric efficiency is certainly improved due to the supercharging effect, the EER is deteriorated. (1)
In the length of the suction connecting pipe 6 according to the formula, the pressure pulsation in the cylinder 7 during the suction stroke is at its maximum. When examined in terms of phase, the pressure within the cylinder 7 is at its lowest at a crank angle of about 1800, that is, about half of the suction stroke. That is, maximum overinflation occurs. In a rotary compressor, due to its structure, the rate of change in volume reaches its maximum near a crank angle of 180 degrees, so the loss of work due to overexpansion increases more than necessary. As a result, the input will not increase more than the increase in the amount of refrigerant circulated due to the improvement in volumetric efficiency. There are problems such as worsening noise and vibration from the intake muffler.

In view of the above problems, the present invention has high EER and low noise.
This provides a rotary compressor with low vibration.

Means for Solving the Problems In order to solve the above-mentioned problems, the present invention provides a suction pipe that connects an accumulator to a cylinder, and a suction port formed in the cylinder. is θs (deg), the length of the suction pipe is Ls (m), the number of shaft rotations is N (rpm), the sound velocity in the suction pipe is a5 (rn/s), and the suction pipe length is Ls (m). When the return crank angle of the pressure wave generated at the cylinder side pipe end of the pipeline is θSR (deq), that is, 296 (decr)≦θSR≦406 (deq)
-...G3).

According to the above-described structure, the present invention has the advantage that, at the start of the suction stroke of the rotary compressor, a pressure wave generated at the cylinder-side pipe end of the suction pipe and having a velocity flowing into the cylinder suction chamber is transmitted to the accumulator-side open end. The crank angle that is reflected and returned at the pump is controlled within a certain range around the crank angle at which the next suction stroke starts, depending on the length of the suction pipe and the compressor rotation speed, thereby reducing pressure pulsations in the cylinder suction chamber during the suction stroke. It is something.

EXAMPLE Hereinafter, a rotary compressor according to an example of the present invention will be described with reference to the drawings.

FIG. 1 shows a sectional view of a rotary compressor in an embodiment of the present invention.

In the figure, 8 is a rotary compressor, 9 is an accumulator, and 10 is a cylinder. 11 is a suction pipe, 12 is a suction port bored in the cylinder 10, and the suction pipe 11 and the suction port 12 constitute a suction pipe line 13. 14 is a shaft, 15 is a piston, 16 is a partition sheave 1, and the space inside the cylinder 10 is divided into a cylinder suction chamber 17 and a cylinder compression chamber 18 by the piston 15 and the partition sheave 16. θB is the suction start crank angle (dog), and Lj is the length (m) of the suction pipe 1a. Suction start crank angle Os (dog), length L of suction pipe 13
s, the compressor rotation speed, that is, the rotation speed N of the shaft 14;

The sound velocity in the suction pipe is set to satisfy equation (3).

The operation will be explained below.

The piston 15 rotates as the shaft 14 rotates, and when the crank angle passes θs (dog), the cylinder suction chamber 17
The volume of increases and the inhalation stroke begins.

That is, the gas refrigerant that has flowed into the accumulator 9 flows through the suction pipe line 1 which is constituted by the suction pipe 11 and the suction port 12.
3 and is sucked into the cylinder suction chamber 17. When the suction stroke starts, the suction pipe line 13 and the cylinder suction chamber 17
Pressure pulsations occur at

Assuming that the pressure pulsation (transient characteristics) in the cylinder suction chamber 17 is an adiabatic change since the suction stroke is rapid, it can be described as follows.

In the above equation (4), Pcs is the pressure of the cylinder suction chamber 17, t is the time, qcs is the weight of the refrigerant in the cylinder suction chamber 17, and Vcs is the volume of the cylinder suction chamber 17.

The pressure pulsations in the suction pipe 13 can be described as follows based on the equation of continuity, the law of conservation of momentum, and the first method of thermodynamics JIIJ.

10・
...V) Here, in the above equations (5), (6), and (7), ρ is the density, V
is the flow rate, P is the pressure, D is the pipe inner diameter, f is the pipe friction coefficient, A is the heat equivalent of work, q is the gravitational acceleration, Cv is the isovolumic specific heat, as
is the speed of sound.

Further, the amount of refrigerant flowing into the cylinder suction chamber mouth end of the suction pipe 13 can be described by the change in the weight of the refrigerant in the cylinder suction chamber 17 and the state of the refrigerant at the open end of the suction port 12.

In the above equation (8), Sp is the cross-sectional area of the suction port 12, ρ6 is the density at the open end of the suction port 12, and vE is the flow velocity 2 at the open end of the suction port 12.

Therefore, the pressure pulsation in the cylinder suction chamber 17 can be obtained by simultaneously solving the equations (1), (6), (8).

Regarding rotary compressors with a nominal output of 550W, J
The analysis was performed under IS-A rated temperature conditions and power supply frequency 60)-1z. Note that R22, which is commonly used in air conditioners, was used as the refrigerant.

FIG. 2 is a graph showing pressure pulsations in the cylinder suction chamber 17 with respect to the length Ls of the suction pipe line 3. Table 1 shows the relationship between the minimum pressure crank angle θb and the crank angle θ8R at which the pressure wave generated at the start of suction and flowing into the cylinder suction chamber 17 at a velocity is reflected at the open end on the accumulator 9 side and returns. It is. It can be seen that there is a very good correlation. That is, Ob is determined by θSR.

Table 1 FIG. 3 is a graph showing the relationship between O8R, cooling capacity Q, input W, and EER as a ratio based on the value at θSR:80 (dog) (EER is the highest).

Here, the total sum of the suction connection pipe (not shown) on the upstream side of the accumulator 9 and the suction pipe line 13 was set to 1.80 (m).

That is, θSR for EER ratio of 0.99 to 1.0
is 42 (deg)≦O9R≦108 (deg)29108
(de≦θSR≦406 (, deg).

In the former case, the length of the suction pipe 13 is short and the vibrations of the rotary compressor 8 are directly propagated to the accumulator 9, so it is difficult to reduce the vibrations.

On the other hand, the latter provides a sufficient length of the suction pipe 13 compared to the former. For example, when θSR=300 (deg), Lg
: 1.1 m. As a result, it becomes possible to use means to suppress vibrations, such as the routing of piping and the addition of damping members at appropriate positions considering the mode of vibration, thereby making it possible to sufficiently reduce vibrations.

In addition, the range of O8H is wide, and there is no deterioration of EER due to the difference in power frequency of 50H2/601 (2).Also, even for operation in a wide rotation speed range such as inverter drive, 296 (deq) ≦ θSR≦406 (deg
), or by adjusting the rotation speed range that is very frequently used to this range, the decrease in EER due to rotation speed can be sufficiently alleviated and the 5EER of temple and shrine air conditioners can be increased. That is, annual power consumption can be reduced.

Effects of the Invention As described above, the present invention provides a suction starting crank angle Is in a suction pipe line that connects an accumulator and a cylinder, and a suction port formed in the cylinder.
(deg), the length of the suction pipe is Ls (m), the shaft rotation speed is N (mpm), the speed of sound in the suction pipe is as (m/-), and at the start of suction, the cylinder of the suction pipe is When the return crank angle of the pressure wave generated at the side pipe end is θSR(deq), θSR(d
Since the rotary compressor is configured so that eg) is in the range of 266 (deg)≦O8R≦406 (deck), it is possible to achieve both high EER and low noise and vibration in the rotary compressor, and the EER decreases depending on the power frequency. This not only solves the problem, but also has the effect of increasing the 5EER of an inverter-driven air conditioner.

[Brief explanation of the drawing]

Fig. 1 is a sectional view of a rotary compressor according to an embodiment of the present invention, Fig. 2 is a characteristic diagram showing pressure pulsations in the cylinder suction chamber during the suction stroke, and Fig. 3 is a basic characteristic of the pressure wave with respect to the feedback crank angle. FIG. 4 is a sectional view of a conventional rotary compressor, and FIG. 5 is an explanatory diagram showing volumetric efficiency versus suction connection pipe length. 9...Accumulator, 10...Cylinder, 11...Suction pipe, 12...Suction port, 13...Suction pipe line. Name of agent: Patent attorney Toshio Nakao and 1 other person9-
--Akiyum Shitata lθ-Cylinder U-m-Suction pipe 1 Figure I2- Suction port 1
3- Suction pipe 2nd cause Figure 3 θ5R (dε9)

Claims (1)

[Claims] In a suction pipe line consisting of a suction pipe connecting an accumulator and a cylinder, and a suction port drilled in the cylinder, the suction start crank angle is θ_s (deg), and the length of the suction pipe is Ls ( m), and the shaft rotation speed is N (rpm).
and the speed of sound in the suction pipe is a_s (m/s),
When the return crank angle of the pressure wave generated at the cylinder side pipe end of the suction pipe line at the start of suction is θ_S_R (deg), that is, θ_S_R=θ_s+(12LsN/a_s), θ_S_R (deg) is 296 (deg)≦θ_S_R A rotary compressor configured to be within the range of ≦406 (deg).