JPH0454838B2 - - Google Patents

Description

[Detailed description of the invention]

INDUSTRIAL APPLICATION FIELD The present invention relates to a rotary compressor used in an air conditioner. BACKGROUND ART A conventional rotary compressor of this type is the fourth rotary compressor as shown in Japanese Patent Application Laid-open No. 57-122192, for example.
The structure was as shown in the figure. In the figure, 1 is a closed container, and a compressor element 2,
An electric motor element 3 for driving this via a crankshaft 4 is 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. In addition, the length (m) of the suction connection pipe 6 is
It is determined by the following formula. = (T・a/4−0.2) ±0.1 ...(1) In the above equation, T is the suction stroke period of compressor element 2 (sec), and a is the sound velocity of the suction gas at the time of suction (m/sec) It is. 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 the volumetric efficiency (refrigerant: R22), but if the suction connecting pipe 6 is determined by equation (1), a supercharging effect can be obtained and the volumetric efficiency can be increased. Get better. 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. At the length of the suction connecting pipe 6 according to equation (1), the pulsation width of the pressure pulsation inside the cylinder 7 during the suction stroke becomes maximum. Looking at the phase, the pressure inside the cylinder 7 is at its lowest when the crank angle is about 180 degrees, 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 the crank angle of 180°, so the loss of work due to overexpansion increases more than necessary. the result,
The input increases more than the increase in refrigerant circulation due to the improvement in volumetric efficiency, which significantly reduces EER and causes pressure pulsation to be at its maximum.
Problems include worsening noise and vibration from the suction connection pipe and suction muffler. In view of the above problems, the present invention provides a rotary compressor with high EER and low noise and vibration. Means for Solving the Problems In order to solve the above problems, the present invention provides a suction starting crank angle of _s (deg),
The length of the suction pipe is L _s (m), and the shaft rotation speed is N
(rpm), and the sound velocity in the suction pipe is a _s (m/
s), and the feedback crank angle of the pressure wave generated at the cylinder side pipe end of the suction pipe at the start of suction is θ _SR (deg), that is, θ _SR = θ _s + 12LsN/a _s ...(2), 296 (deg) ≦θ _SR ≦406 (deg) ……(3). Effects The present invention has the above-described configuration, so that at the start of the suction stroke of the rotary compressor, a compression wave generated at the cylinder-side pipe end of the suction pipe and having a velocity flowing into the cylinder suction chamber is reflected at the accumulator-side open end. The crank angle at which the compressor returns is controlled within a certain range around the crank angle at the start of the next suction stroke, depending on the length of the suction line and the compressor rotation speed, thereby reducing the width of pressure pulsations in the cylinder suction chamber during the suction stroke. Embodiment Hereinafter, a rotary compressor according to an embodiment 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 archiver, and 10 is a cylinder. 1 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, and 16 is a partition vane. 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 vane 16. θ _s is the intake start crank angle (deg)
, and L _s is the length (m) of the suction pipe 13.
The suction start crank angle θ _s (deg), the length L _s of the suction pipe 13, the compressor rotation speed, that is, the rotation speed N of the shaft 14, and the sound velocity in the suction pipe are set to satisfy equation (3). There is. The operation will be explained below. The piston 15 rotates as the shaft 14 rotates, and when the crank angle exceeds θ _s (deg), the volume of the cylinder suction chamber 17 increases and a suction stroke begins. That is, the gas refrigerant that has flowed into the accumulator 9 is sucked into the cylinder suction chamber 17 via a suction pipe line 13 configured by a suction pipe 11 and a suction port 12 . When the suction stroke starts, suction line 1
3 and the cylinder suction chamber 17 . 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. dPcs/dt=K・Pcs(1/g _cs dg _cs /dt−1/VcsdVcs
/dt……(4) In the above equation (4), Pcs is the cylinder suction chamber 17
, t is the time, k is the adiabatic index, g _cs 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 pulsation in the suction line 13 is expressed by the continuity equation,
According to the law of conservation of momentum and the first law of thermodynamics, it can be written as follows. ∂ρ/∂t+ρ∂v/∂t+v∂ρ/∂v0 ...(5) ∂v/∂t+v∂v/∂x+1/ρ∂ρ/∂x+2f/Dv ²
v/|v|=0...(6) ∂/∂t{ρ(g/ACvT+v ² /2)}+∂/∂x{ρv
(g/A CvT+P/ρ+v ² /2)}=0...(7) Here, g/ACvT=a2 _s k(k-1) In the above equations (5), (6), and (7), ρ is density, v is flow velocity, p is pressure, D is pipe inner diameter, f is pipe friction coefficient, A
is the heat equivalent of work, g is the gravitational acceleration, Cv is the isovolumic specific heat, and a _s is the speed of sound. Further, the amount of refrigerant flowing into the open end of the suction pipe 13 on the cylinder 7 side is a change in the weight of the refrigerant in the cylinder suction chamber 17, and can be described by the state of the refrigerant at the open end of the suction port 12. dg _cs /dt=Sp・ρ _E・g・v _E ...(8) In the above equation (8), Sp is the cross-sectional area of the suction port 12, ρ _E is the density at the open end of the suction port 12, and v _E is This is the flow velocity at the open end of the suction port 12. Therefore, by solving equations (4), (5), (6), (7), and (8) simultaneously, the pressure pulsation in the cylinder suction chamber 17 can be obtained. For rotary compressors with a nominal output of 550W, JIS-A rated temperature conditions and power frequency of 60Hz
The analysis was carried out using The refrigerant used was R22, which is commonly used in air conditioners. FIG. 2 is a graph showing pressure pulsations in the cylinder suction chamber 17 with respect to the length Ls of the suction line 3.
Table 1 shows the relationship between the crank angle θ _b at the lowest pressure and the crank angle θ _SR at which the pressure wave that flows into the cylinder suction chamber 17 at a velocity that occurs at the start of suction is reflected at the open end on the accumulator 9 side and returns. It shows. It can be seen that there is a very good correlation. That is, θ _b is determined by θ _SR .

[Table] Figure 3 shows the relationship between θ _SR , cooling capacity Q, input W, and EER at θ _SR = 80 (deg) (EER is the highest)
It is a graph shown as a ratio based on . 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 is 1.80 (m). That is, θ _SR for an EER ratio of 0.99 to 1.0 is 42 (deg)≦θ _SR ≦108 (deg), 296 (deg)≦θ _SR ≦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 has a sufficient suction pipe 13 compared to the former.
The length of is obtained. For example, when θ _SR = 300 (deg), Ls
= 1.1m. means for suppressing vibrations as a result;
In other words, it becomes possible to route the piping and add a damping member to an appropriate position considering the mode of vibration, thereby making it possible to sufficiently reduce vibration. In addition, the range of θ _SR is wide, and EER does not deteriorate due to the difference in power frequency of 50Hz/60Hz. In addition, even for operation in a wide rotation speed range such as inverter drive, the rotation speed range must be set so that 296 (deg) ≦ θ _SR ≦ 406 (deg), or if the rotation speed is very frequently used. By adjusting the high rotational speed range to this range, the decrease in EER due to rotational speed can be sufficiently alleviated, and the SEER of the air conditioner can be increased, that is, the annual power consumption can be reduced. Effects of the Invention As described above, the present invention provides a suction pipe line consisting of a suction pipe connecting an accumulator and a cylinder, and a suction port punched in the cylinder, by setting the suction start crank angle to θ _S (deg). The length of is Ls (m), the shaft rotation speed is N (rpm), the sound velocity in the suction pipe is a _s (m/s), and the pressure generated at the cylinder side pipe end of the suction pipe at the start of suction is When the wave feedback crank angle is θ _SR (deg), that is, θ _SR = θ _S +12LsN/ _as , θ _SR (deg) should be in the range of 266 (deg) ≦ θ _SR ≦ 406 (deg). As a result, it is possible to achieve both high EER and low noise and vibration in rotary compressors, which not only solves the problem of lower EER due to power frequency, but also increases the SEER of inverter-driven air conditioners. It has the following effects.

[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, 1
1... Suction pipe, 12... Suction port, 13... Suction pipe line.

Claims (1)

[Claims] 1. In the suction line consisting of the suction pipe connecting the accumulator and the cylinder and the suction port drilled in the cylinder, the suction start crank angle is set to θ _s
(deg), the length of the suction pipe is Ls (m), the shaft rotation speed is N (rpm), the sound velocity in the suction pipe is a _s (m/s), and the suction pipe length is Ls (m) at the start of suction. When the return crank angle of the pressure wave generated at the cylinder side pipe end is θ _SR (deg), that is, θ _SR = θ _s +12LsN/a _s , θ _SR (deg) is 296 (deg) ≦ θ _SR ≦ 406 ( rotary compressor configured to have a range of deg).