JPH0223837Y2 - - Google Patents

Description

[Detailed explanation of the idea]

This proposal concerns the improvement of rotary compressors. To explain a refrigeration compressor, which is an example of a conventional rotary compressor, 1 is a main body of the compressor, and a discharge pipe 2 for discharging refrigerant inside the compressor main body 1 is led out to the outside. There is. Also, the same discharge pipe 2
A condenser 4, a constrictor 5, an evaporator 6, and an accumulator 7 are connected to the refrigerant supply pipe 3.
The accumulator 7 has a suction pipe 8 and a suction port 2.
1 to a cylinder 20 in the compressor main body 1. Further, in the same accumulator 7, the inlet portion 9 of the suction pipe 8 is provided at a sufficiently high position with respect to the bottom of the accumulator 7 in order to separate the liquid refrigerant. Therefore, the refrigerant 10 flowing into the suction pipe 8 toward the cylinder 20 is in a gaseous state. Further, a motor 11 is built in the compressor body 1, and the compressor is operated by driving the motor 11 and rotating the crankshaft 12, and the gas refrigerant sucked from the suction pipe 8 is 10 is adapted to be compressed within the cylinder 20. Further, the compressed gas refrigerant 10 is discharged into the discharge cavity 13 at the lower end of the compressor main body 1 via a discharge valve (not shown), and then led out to the space 14, passes around the motor 11, and passes through the discharge pipe 2. The air is discharged from the compressor body 1 to the outside of the compressor body 1. Note that the reference numeral 15 in the figure is lubricating oil stored in the lower part of the compressor main body 1. In the refrigeration system 16 including the refrigeration compressor, when the rotation speed of the compressor is made variable, the speed of the gas refrigerant 10 in the suction pipe 8 only increases in approximately direct proportion to the rotation speed of the compressor. The pressure pulsation of the gas refrigerant 10 changes in a complicated manner with the rotation speed of the compressor.
This point will be explained next. Figure 2 shows the inside of the cylinder 2.
2 shows an outline of the pressure change during the 0 suction stroke. FIG. 2 shows the compressor rotation speeds at low speed, medium speed, and high speed. In Fig. 2, the horizontal axis is the rotor rotation angle θ, and the vertical axis is the suction stroke pressure P in the cylinder 20, Po
is the average suction pressure in the refrigerant pipe 3 before the accumulator 7 in FIG. As described above, the P-V diagram is obtained from the suction stroke pressure P and cylinder volume V in the cylinder 20 shown in FIG. 2, and this and the average suction pressure
The power loss during suction can be calculated from Po. The power loss during suction determined in this way is influenced by the waveform of the pulsating flow inside the cylinder 20, but generally the accumulator volume is sufficiently large compared to the cylinder 20 volume, so the flow is slow. It may be considered to be divided by the compressor 7, and when a compressor with a certain cylinder volume is operating at a constant rotation speed, the pulsating flow waveform in the cylinder 20 is mainly determined by the length and diameter of the suction pipe 8. . However, in a variable speed compressor, the frequency of pulsation generated by the compressor changes, so in a certain rotation speed range, the suction pipe 8
The suction stroke pressure P in the cylinder 20 has various waveforms, such as large pressure fluctuations sometimes occurring within the cylinder 20 and small pressure fluctuations occurring in other rotational speed ranges. However, since the suction loss power is determined by the suction stroke pressure waveform in the cylinder 20, when the rotation speed of the compressor changes, the suction loss power may increase or decrease as a proportion of the gas compression power. Changes depending on the pulsation waveform. In this way, in a conventional compressor, the rotation speed of the compressor is changed while the length and diameter of the suction pipe 8 are constant, so especially in a certain high speed rotation speed range, the ratio of suction loss power to gas compression power is small. The problem was that the compressor's performance deteriorated due to the increase in the number of compressors. This proposal focuses on the fact that in a certain rotational speed range, the proportion of suction loss power in the gas compression power becomes extremely small. In a compressor equipped with a cavity such as an accumulator that is larger than the maximum stroke volume of the chamber, the suction chamber side of the suction pipe is branched into a plurality of suction pipes, and the static pressure in the cavity and the pressure in the suction pipe before branching are A rotary compressor characterized in that at least one of each suction pipe after branching is provided with an on-off valve that opens when the pressure difference with the static pressure exceeds a certain value, and the purpose of the rotary compressor is The purpose of this invention is to provide an improved rotary compressor that can reduce the ratio of suction power loss to gas compression power and improve the performance of the compressor. Next, the rotary compressor of the present invention will be explained using an embodiment shown in FIGS. 3 and 4. 8 is the suction pipe 8 in FIG.
30 is a body provided in the same suction pipe 8 between the accumulator 7 and the suction port 21 in FIG. A pressure close to the total pressure of the suction pipe 8 acts from inside the accumulator 7 in the pipe 70 . 32 is the plunger of the on-off valve, and inside it,
A through hole 31 and a T-shaped ventilation hole 35 are provided, through which the static pressure in the suction pipe 8 acts on the space 34' in which the spring 34 is housed.
Further, the body 30 and the plunger 32 can slide vertically while maintaining airtightness. Further, the body 30 is provided with an inverted L-shaped through hole 33 so that the suction pipe 8 and the suction pipe 8b can communicate with each other through the through hole 33. Also, a stopper 3 for the plunger 32 is provided at the lower end of the body 30.
7 is provided. In addition, the suction pipe 8b is connected to the suction port 21 of the compressor in parallel with the suction pipe 8a, and compared to the cross-sectional area of the suction pipe 8 alone, the combined cross-sectional area of the suction pipes 8a and 8b becomes larger. . Although FIGS. 3 and 4 show the case where there are two suction pipes, the number of suction pipes may be three or more. Next, the operation of the rotary compressor will be explained. Even if the rotation speed of the compressor changes, in order to reduce the ratio of suction loss power to the gas compression power, even if the rotation speed of the compressor changes, the suction stroke pressure pulsation waveform in the cylinder 20 should be kept as small as possible. It is necessary to have a stable waveform. To explain this point specifically, the suction stroke pressure pulsation within the cylinder 20 is
If the volume of the cylinder 20 is V (V changes every moment depending on the rotor rotation angle), and the length and diameter of the suction pipe 8 is D, then this suction piping system will be as follows, which is explained in the so-called Helmholtz resonator. resonance frequency o
It may or may not resonate depending on the relationship between [Hz] and the compressor rotor rotational speed frequency (c). C: Sound velocity of refrigerant gas 10 in suction pipe 8 [m/
sec.] D: Inner diameter of suction pipe 8 [m]: Length [m] However, since V in equation (1) changes moment by moment depending on the rotor rotation angle θ, o also varies depending on the rotor rotation angle θ. However, according to experimental examples, in a compressor with the configuration shown in Figure 1, 100 < c < 200 Hz, and the rotation speed of a variable speed compressor is usually 20 Hz < c < 150 Hz.
That's about it. The further apart these frequencies c and o are, the less resonance occurs, and the cylinder 2
The suction stroke pressure waveform within 0 becomes stable, and as a result, the ratio of suction loss power to gas compression power becomes smaller. Here, the above equation (1) will be further explained. As shown in Figure 5, the resonant frequency _o of a Helmholtz resonant system with neck areas S ₁ , S ₂ . if, And, is known to be a special case. In addition, the method for deriving the resonant period T' of a Helmholtz resonator with neck dimensions S and cavity volume V is described in "Basics of Hydraulic Engineering" (written by Tsuyoshi Hisada, October 30, 1966).
Nikkan Kogyo Shimbun (1-1 Iida-cho, Higashiguto Chiyoda-ku)
Publication), pages 100-101. The result is the same as above. In the case of two pipes with neck dimensions S ₁ and S ₂ according to the idea of inducing o in this document, (i) Pressure changes within the cavity are transmitted to the two necks simultaneously. (ii) Both ends of the neck are connected to the same pressure source. Therefore, it can be assumed that the vibrations of the fluid in the two necks are in phase. Therefore, the formula (2.31) in the above document becomes as follows. ρS ₁ x¨＋ρωS ² / ₁ /λ・x〓＋ρC ² (S ₁ + S ² )
/V S ₁ x=0 (a)...(5) ρS ₂ x¨＋ρωS ² / ₂ /λ・x〓＋ρC ² (S ₁ + S ₂ )
/V S ₂ x=0 (b)...(5) In the above equation (5), if the term x〓 is omitted, x¨+C ² (S ₁ +S ₂ )/・Vx=0...(6) Become,

It has a natural angular vibration frequency of [Formula]. This results in It can be seen that the above equation (2) can be obtained in general. Further, c is the operating frequency of the compressor, and if the number of poles is P, the synchronous rotational speed Ncvpm is expressed as Nc = 120c/P, and in the case of a two-pole motor, P = 2, so c = Nc/ 60c/s (8), and the power supply frequency and rotational frequency are equal if the motor slip is omitted. The power frequency c of the inverter drive is c=20 to 150Hz. Since the excitation source is assumed to be the volume change rate of the cavity V (the suction chamber volume of the compressor), resonance occurs when this excitation frequency matches the resonant frequency of the neck. The resonance characteristics at this time are shown in the same way as in Figures 2 and 40a and b on page 128 of the above-mentioned document. , increase o so that c/o<1
If this is the case, it can be seen that resonance can be prevented and at the same time the phase delay of the fluid flow in the neck can be reduced to 90° or less. The above is shown in Figure 2.
The figure shows how a high frequency component appears in the pressure inside the suction pipe during the progress of the suction stroke of the compressor (increase in V) and during the increase in rotational speed. The present invention has substantially the same effect as changing the cross-sectional area of the suction pipe described above. That is, FIG. 4 shows the operating state when the rotation speed of the compressor is high. Since the pipe 70 communicates with a position within the accumulator 7 where the gas flow rate is slow, the internal pressure within the pipe 70 is close to the total pressure within the suction pipe 8 as described above, and this is the pressure that is close to the total pressure within the suction pipe 8. and pushes the plunger 32 downward. On the other hand, the lower end of the plunger 32 and the space 3 where the spring is housed
The static pressure (less than the total pressure) of the flow rate in the suction pipe 8 acts on 4', pushing the plunger 32 upwards. The force for pushing this plunger 32 upward is provided by a spring 3.
4 is also added, and the upward force on the plunger 32 is the sum of the force due to the static pressure and the force of the spring 34.
In this case, the rotation speed of the compressor is high and the flow velocity in the suction pipe 8 is higher than when the rotation speed is low, so the static pressure is small, and therefore the downward pressing force against the plunger 32 is greater than the upward pressing force. As a result, the plunger 32 comes into contact with the stopper 37 as shown in FIG. In this state, the inverted L-shaped through hole 33 of the body 30 communicates with the suction pipe 8 and the suction pipe 8b via the through hole 31 of the plunger 32, so that the effective cross-sectional area of the suction pipe is
The size increases from suction pipe 8a+suction pipe 8b, and as a result, the resonance frequency increases. Fig. 3 shows the operating state of the compressor at low rotation speed, contrary to Fig. 4. Since the flow velocity in the suction pipe 8 is slow, the static pressure increases and the plunger 32 is pushed upward. , as shown in FIG. 3, the pipe 8b is blocked and no refrigerant flows, resulting in a decrease in the effective pipe cross-sectional area and a drop in the resonant frequency. The rotary compressor of the present invention is constructed as described above, and the effective cross-sectional area of the suction pipe increases or decreases depending on the number of rotations, thereby changing the resonance frequency.
Since the resonance phenomenon can be avoided, the ratio of the suction loss power to the gas compression power can be reduced, which has the effect of improving the performance of the compressor.

[Brief explanation of the drawing]

Fig. 1 is a vertical side view showing a conventional rotary compressor, Fig. 2 is an explanatory diagram showing an outline of pressure changes during the suction stroke in the cylinder of the compressor, and Figs. 3 and 4 are rotary compressors according to the present invention. FIG. 5 is a longitudinal side view showing one embodiment of the main part of the machine, and is an explanatory diagram illustrating the resonant frequency of the Helmholtz resonance system. 7... Accumulator, 8... Suction pipe, 8
a, 8b...branched suction pipe, 32...plunger of on-off valve.

Claims (1)

[Scope of utility model registration request] In a compressor that has a cavity such as an accumulator larger than the maximum stroke volume of the suction chamber of the compressor on the upstream side of a suction pipe that communicates with the suction chamber of the compressor, the suction chamber side of the suction pipe is connected to a plurality of suction chambers. At least one of the suction pipes after branching is provided with an on-off valve that opens when the pressure difference between the static pressure in the cavity and the static pressure in the suction pipe before branching exceeds a certain value. A rotary compressor characterized by: