JP2009108739A - Rotary compressor - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To make the effective use of a resonance phenomenon including the effect of shielding a suction port by a roller of a rotary compressor, and to operate efficiently the compressor within a wide range of the number of revolutions. <P>SOLUTION: A rotary compressor includes an electric motor 4 within a hermetically-sealed container, an eccentric section 5 connected to the electric motor, a suction chamber 30 formed of a roller 11 and a cylinder 10 provided at the outer circumference of the eccentric section, an accumulator 16 provided at the suction side of the suction chamber, and a suction pipe 18 which opens toward the accumulator. The length of a suction flow passage from the open end of the suction pipe 18 to the open end of the suction port is represented by L [m], the maximum number of revolutions of the compressor 1 by fm [Hz], and the sound velocity of a gas coolant by C [m/s]. As viewed from a vane which separated the suction side and the discharge side of the suction chamber 30 from each other when the suction of the suction chamber is completed, a rotation angle formed of line segments connecting the center of the rotation of the roller 11 and the center of the eccentric section 5 is represented by θs [rad]. A relationship of L=(π-θs)/(4π×fm) is satisfied. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to an inverter-driven rotary compressor and a vertical two-cylinder rotary compressor used in an air conditioner having a refrigeration cycle.

  Conventionally, in a rotary compressor used in a refrigeration cycle, resonance at which volumetric efficiency (also referred to as volumetric efficiency or suction filling rate) is maximized at a specific rotation speed depending on the suction flow path length L [m]. The phenomenon is known. Here, the rotational speed that maximizes the volumetric efficiency is referred to as the resonant rotational speed fn [Hz]. The suction flow path length L is the length from the open end of the suction pipe to the open end of the suction port of the suction chamber.

  Regarding the above-described resonance phenomenon, for example, a structure disclosed in Patent Document 1 is known. As shown in FIG. 14, in the conventional compressor, the volumetric efficiency increases as the rotational speed is increased from the low speed side, and approaches a substantially constant value. Furthermore, the volumetric efficiency increases as the rotational speed approaches the resonant rotational speed fn, and the volumetric efficiency becomes maximum at the resonant rotational speed fn. At a rotational speed higher than the resonant rotational speed fn, the volumetric efficiency gradually decreases and approaches a constant value. In general, in a compressor, it is known that mechanical loss, noise, and vibration increase as the rotational speed increases. If the above resonance phenomenon is used, there is an advantage that the capacity of the air conditioner can be increased without increasing the rotational speed of the compressor.

  The relationship between the suction flow path length L that causes such a resonance phenomenon and the resonance rotational speed fn is described in, for example, Patent Document 1, Patent Document 2, and Patent Document 3.

The suction channel length L [m] described in Patent Document 2 uses the primary resonance of the air column resonance in the suction channel,
L = λC / (2π · fn) (1)
λ: π / 2. Here, C is the speed of sound of the refrigerant (the speed at which the pressure wave of the refrigerant propagates) [m / s]. Substituting λ = π / 2 into equation (1),
L = C / (4 · fn) (2)
Get.

In patent document 1, the influence of the pushing amount V [m ³ ] of the suction chamber and the flow passage cross-sectional area A [m ² ] of the suction pipe is corrected with respect to patent document 2,
fn = (2m−1) C / {4 (L + V / A)} (3)
m = 1, 2, 3,... For the primary resonance component that has a large impact on volumetric efficiency, that is, m = 1, the equation (3) is transferred and rearranged.
L = C / (4 · fn) − (V / A) (4)
Get.

In patent document 3, correction by experiment is performed with respect to patent document 2,
L = T · C / 4-0.2 (5)
It is. Here, T is the suction stroke cycle [s], and if T = 1 / fn is substituted into equation (5),
L = C / (4 · fn) −0.2 (6)
Get.
Japanese Patent Laid-Open No. 64-15489 (page 3, formula (1), page 5, FIG. 4) JP-A-4-103971 (page 2, formula (1)) JP-A-57-122192 (2nd page, formula (2))

  However, the conventional rotary compressor mainly focuses on the air column resonance of the suction flow path length L, and does not consider the phenomenon that the roller unique to the rotary compressor shields the suction port (the suction port in the roller). Does not take into account the reflection of refrigerant pressure waves due to the shielding of Therefore, as shown in the equations (2), (4), and (6), the relationship between the resonance rotational speed fn and the suction flow path length L is different, and there is a problem that application to an actual design is difficult. .

  The object of the present invention is to theoretically consider the effect of a roller specific to a rotary compressor to shield the suction port, to obtain the relationship between the resonance rotational speed fn and the suction flow path length L, and to determine the resonance phenomenon as a rotary compressor or 2 It is to make effective use of the cylinder rotary compressor.

In order to solve the above problems, the present invention mainly adopts the following configuration.
An electric motor disposed in an upper part of the sealed container; an eccentric portion having a rotating shaft coupled to the electric motor; a suction chamber formed by a roller and a cylinder provided on an outer periphery of the eccentric portion; and the suction chamber A rotary compressor comprising: an accumulator provided on the suction side; a suction pipe that opens in the accumulator; and a seal suction that connects the suction pipe and a suction port of the suction chamber;
The resonance speed of the rotary compressor causing a resonance phenomenon that maximizes the refrigerant suction filling rate of the suction chamber is limited to the maximum speed of the rotary compressor, and the suction of the suction chamber at the end of the suction of the suction chamber A suction flow from the open end of the suction pipe to the open end of the suction port is obtained by obtaining a rotation angle formed by a line connecting the rotation center of the roller and the center of the eccentric part as seen from the vane separating the side and the discharge side The path length is determined based on the limited maximum rotational speed of the rotary compressor and the rotation angle.

An electric motor disposed in an upper part of the sealed container; an eccentric portion having a rotating shaft connected to the electric motor; a suction chamber formed by a roller and a cylinder provided on an outer periphery of the eccentric portion; A rotary compressor comprising an accumulator provided on the suction side of the suction chamber, a suction pipe that opens in the accumulator, and a seal suction that connects the suction pipe and a suction port of the suction chamber,
The suction flow path length from the open end of the suction pipe to the open end of the suction port is L [m], the maximum rotational speed of the compressor is fm [Hz], and the sound speed of the gas refrigerant is C [m / m s], and a rotation angle formed by a line segment connecting the rotation center of the roller and the center of the eccentric portion when viewed from the vane separating the suction side and the discharge side of the suction chamber at the end of the suction of the suction chamber is θs [ rad], the relationship of L = (π−θs) / (4π · fm) is established.

In addition, an electric motor disposed in an upper portion of the sealed container, two eccentric portions having a rotating shaft connected to the electric motor, and the two eccentric portions provided via a partition plate through which the rotating shaft passes. Two suction chambers each formed by a roller and a cylinder respectively provided on the outer periphery, an accumulator provided on the suction side of the two suction chambers, and a substantially L-shape having a bent portion opened in the accumulator A vertical two-cylinder rotary compressor comprising two suction pipes, and two seal suctions connecting the two suction pipes and the suction ports of the two suction chambers,
The two suction pipes form different paths, and the suction flow path lengths from the open ends of the respective suction pipes to the open ends of the respective suction ports are made equal to L [m], respectively. The maximum number of revolutions is fm [Hz], the sound speed of the gas refrigerant is C [m / s], and each of the two suction chambers is viewed from the vane that separates the suction side and the discharge side at the end of the suction. When the rotation angle formed by the line segment connecting the rotation center of each roller and the center of each eccentric portion is θs [rad], the relationship L = (π−θs) / (4π · fm) is established. To do.

  According to the present invention, it is possible to effectively utilize the resonance phenomenon including the effect that the roller shields the suction port, and to efficiently operate the compressor in a wide range of rotation speeds.

  A rotary compressor according to an embodiment of the present invention will be described below in detail with reference to the drawings. The first embodiment will be described with reference to FIGS. 1 to 9, the second embodiment with reference to FIGS. 10 to 12, and the third embodiment with reference to FIG.

“First Embodiment”
FIG. 1 is a longitudinal sectional view of a rotary compressor according to a first embodiment of the present invention. The first embodiment is a compressor for an air conditioner using a refrigerant R410A. The compressor 1 includes a sealed container 3 including a bottom part 6, a lid part 7, and a body part 8. An electric motor 4 driven by an inverter and having a stator and a rotor is provided above the inside of the sealed container 3. The rotating shaft 2 connected to the electric motor 4 includes an eccentric portion 5 and is pivotally supported by a main bearing 9 having an end plate portion and a concave sub-bearing 19 having an end plate portion. The main shaft 9, the substantially cylindrical cylinder 10, the sub-bearing 19, and the sub-bearing 19 are separated from the rotary shaft 2 in this order from the electric motor 4 side. The plates 20 are laminated and integrated with fastening elements (not shown) such as bolts. The main bearing 9 is fixed to the inner wall of the body portion 8 by welding.

  The suction chamber 30 includes a cylinder 10, an end plate portion of the main bearing 9, a cylindrical roller 11 fitted to the outer periphery of the eccentric portion 5, and an end plate portion of the sub main bearing 19. As shown in FIGS. 2 to 6, the suction chamber 30 has a flat vane 28 connected to an urging force applying means (not shown) such as a coil spring that rotates in accordance with the eccentric motion of the eccentric portion 5. The suction chamber 30 is separated from the discharge-side space by advancing and retreating while making contact with the outer periphery of the roller 11.

  As shown by the arrows in FIG. 1, the gas refrigerant that is a working fluid includes an inflow pipe 25, an inflow hole 24 provided in the mesh seat 22 in the accumulator 16, and a suction pipe 18 (the refrigerant that has flowed into the inflow pipe 5 flows into the inflow hole 24. And is sucked into the suction chamber 30 through the seal suction 15 and the suction port 13. The suction pipe 18 has a substantially L shape having a bent portion 32. Further, on the upstream side of the mesh seat 22, a mesh 21 for capturing foreign substances in the refrigerant is provided. The screen seat 22 is welded in the container 17 of the accumulator 16 that is long in the height direction.

  Further, the gas refrigerant compressed in the cylinder 10 is discharged into the internal space 14 including the auxiliary bearing 19 and the closing plate 20, and flows into the sealed container 3 through the through hole 12 and the soundproof cover 27. Thereafter, the liquid is discharged from the outflow pipe 26 through the gap of the electric motor 4.

  Next, the resonance phenomenon of the present embodiment will be described with reference to FIGS. 2 to 6 as an interaction of air column resonance in the roller 11 and the suction flow path. FIG. 2 is a diagram illustrating a suction state at time t = 0 in the rotary compressor according to the first embodiment. FIG. 3 is a diagram illustrating a suction state at time t = L / C in the rotary compressor according to the first embodiment. FIG. 4 is a diagram illustrating a suction state at time t = 2L / C in the rotary compressor according to the first embodiment. FIG. 5 is a diagram illustrating a suction state at time t = 3 L / C in the rotary compressor according to the first embodiment. FIG. 6 is a diagram illustrating a suction state at time t = 4 L / C in the rotary compressor according to the first embodiment.

  In FIG. 2 to FIG. 6, for simplification, the pressure wave in the suction channel (pressure fluctuation wave generated in the suction pipe) is represented by a periodic rectangular wave synchronized with the rotation speed. Furthermore, the bending part 32 and the accumulator 16 which have little influence on the phase of the pressure wave are omitted. Let time be t, and let the suction start (suction end on one side), that is, the rotation angle of the eccentric part 5 where the roller 11 shields the suction port 13 be θs. Here, the rotation angle is an angle formed from the direction of the vane 28. In other words, θs shown in FIG. 2 refers to an angle formed by a line segment connecting the rotation center of the roller 11 and the center of the eccentric portion 5 when viewed from the vane 28 at the end of the suction of the suction chamber (this θs is defined as the suction chamber). This is referred to as the rotation angle of the eccentric portion at the end of suction). The rotation angle is θs and the time t is 0. The direction in which the pressure wave propagates is positive in the direction from the open end of the suction pipe 18 to the open end of the suction port 13 and negative in the reverse direction. The speed at which the pressure wave propagates is the sonic speed C of the refrigerant, which is 173 [m / s] in the embodiment.

  FIG. 2 shows a state where the rotation angle is θs and the time t is zero. At this time, the suction volume of the suction chamber 30 is substantially zero. The magnitude of the pressure wave propagating in the positive direction reaching the closed end of the suction port 13 is defined as (+ P). Here, the symbol “+” represents the sign of the pressure wave with respect to the time average value of the absolute pressure. In FIG. 2, the pressure wave (+ P) is reflected in the same phase by the roller 11 and propagates in the negative direction (the open end direction shown in FIG. 2).

  FIG. 3 shows a state where the time t is L / C. Here, when θ = 2πf · t and t = L / C, the rotation angle is θs + (2πf) L / C. The pressure wave attenuated by viscosity reaches the open end of the suction pipe 18, and since the open end side capacity is large, it is reflected in an opposite phase and a negative pressure wave propagates in the positive direction. Here, since the volume of the accumulator 16 is sufficiently larger than the suction volume of the suction chamber 30 at the open end, the net pressure wave is 0 (the pressure waves (+) and (−) shown in FIG. 3 are canceled at the open end). 0).

  FIG. 4 shows a state where the time t is 2 L / C. The rotation angle is θs + (4πf) L / C. The attenuated pressure wave (the pressure wave (−) shown in FIG. 2 attenuates as shown by the dotted line in FIG. 4 and becomes a solid pressure wave at the suction port 13) reaches the suction chamber 30. At the same time, the pressure wave is reflected in the same phase by the roller 11 and the cylinder 10, and the negative pressure wave (the pressure wave component indicated by the oblique line shown in FIG. 4) propagates in the negative direction.

  On the other hand, with the time change of the suction volume, the negative expansion wave shown in FIG. 4 is generated and propagates in the negative direction together with the negative pressure wave described above. Here, the expansion wave generates a negative pressure in the suction chamber 30 when the rotation angle is increased as shown in FIG. 4. This negative pressure is called an expansion wave, and can be said to be a kind of pressure wave. .

  As will be described later with reference to FIG. 7, when the rotation angle from the vane 28 is π, the rate of change of the suction volume of the suction chamber 30 is maximized, so the rotation angle shown in FIG. In this example, a deviation occurs from π in the example of the rotation angle shown in FIG. 4, but this deviation is determined depending on the rotation speed of the compressor. In other words, since the period of the pressure wave (+ P) from FIG. 2 to FIG. 6 is determined by the suction flow path length L and the refrigerant sound speed C, whether or not the state shown in FIG. Depending on the rotation speed f.

  FIG. 5 shows a state where the time t is 3 L / C. The rotation angle at this time is θs + (6πf) L / C. The attenuated negative pressure wave is reflected in the opposite phase at the open end of the suction pipe 18, and the positive pressure wave propagates in the positive direction.

  FIG. 6 shows a state at the end of the suction at time t of 4 L / C. The attenuated pressure wave reaches the suction chamber end 13. Since the pressure wave fluctuates periodically, the rotation angle is θs and the magnitude of the pressure wave is (+ P) as in FIG.

  From the above, it can be seen that the pressure wave (+ P) at the suction port 13 at the rotation angle θs is determined by the expansion wave generated by the suction volume change of the suction chamber 30 and the viscous damping of the pressure wave. In general, since the viscous damping is large in the high rotation speed region, the influence of the expansion wave becomes dominant (for the pressure wave (+ P)) as compared with the pressure wave reciprocated by the suction flow path length L with the large damping. . Therefore, the pressure wave (+ P) increases with the magnitude of the expansion wave. Thus, when the pressure wave (+ P) affected by the magnitude of the expansion wave has a maximum value, the density of the refrigerant increases, and the refrigerant mass in the suction chamber at the end of the suction reaches a maximum. That is, the volumetric efficiency (suction filling rate) is maximized, resulting in a resonance phenomenon.

On the other hand, the magnitude of the expansion wave becomes a maximum when the time change of the suction volume is a maximum from the law of conservation of momentum. FIG. 7 is a diagram showing the relationship between the rotation angle of the rotary compressor according to the first embodiment and the time change rate of the suction volume. As shown in FIG. 7, when the rotation angle is π (the angle from the vane 28 is π), the time change rate of the suction volume becomes a maximum. Therefore, the relationship of π = θs + (4π · fn) L / C is established between the resonance speed fn and the suction flow path length L from FIG. If this is transformed,
L = (π−θs) / (4π · fn) (7)
It is.

  As shown in FIG. 8, the characteristic of the present embodiment (7) is different from that of the conventional expressions (2), (4), and (6). The actual machine of the rotary compressor according to the present embodiment is in good agreement with the equation (7) shown in FIG. 8, and is designed to take advantage of the effect that the roller 11 unique to the rotary compressor shields the suction port 13. . FIG. 8 is a diagram illustrating the relationship between the resonance rotational speed fn and the suction flow path length L of the rotary compressor according to the first embodiment. In the experiment shown in FIG. 8, the pushing amount V is 13 [mL], the inner diameter of the suction pipe 18 is 11 [mm], the suction flow path length L is 0.37 [m], and the suction pressure is 0.61 [MPa. The suction temperature is −5 [° C.] and the discharge pressure is 2.8 [MPa]. Further, the minimum inner diameter of the seal suction 15 is 10 [mm], which is a minute difference of about 1 [mm] with respect to the inner diameter of the suction pipe 18. Therefore, the fluid loss at this minimum inner diameter and the difference in acoustic impedance are negligible. The radius of curvature of the bent portion 32 is 25 [mm].

  Next, FIG. 9 shows the relationship between the rotational speed of the compressor of this embodiment and the efficiency of the compressor (total adiabatic efficiency × motor efficiency). According to FIG. 9, the total adiabatic efficiency (compressor energy efficiency) sharply decreases when the rotational speed is equal to or higher than the resonant rotational speed fn. That is, the gradient of the total adiabatic efficiency with respect to the rotational speed is inflected at the resonant rotational speed fn. Generally, the total heat insulation efficiency is divided into refrigerant leakage loss, fluid loss, and friction loss between sliding parts. Since the volumetric efficiency increases as the rotational speed increases up to the resonance rotational speed fn, the increase in friction loss is offset and the decrease in the total adiabatic efficiency is suppressed. When the rotational speed is equal to or higher than the resonant rotational speed fn, the effect of friction loss becomes significant as the volumetric efficiency decreases, so the gradient of the total adiabatic efficiency increases.

  In this embodiment, since the resonance rotational speed fn is set to the maximum rotational speed fm, the compressor can be operated in a region where the gradient in FIG. 9 is gentle and the efficiency is high. Here, the maximum number of revolutions fm is the maximum number of revolutions that operates stably except for a transient response at the time of startup. Specifically, it is the maximum number of revolutions during the low-temperature heating operation of the air conditioner. The low temperature heating operation is a condition in which the dry bulb temperature of the outside air is 2 [° C.], the wet bulb temperature is 1 [° C.], the indoor dry bulb temperature is 20 [° C.], and the wet bulb temperature is 10 [° C.].

  From the above, in this embodiment, since the suction flow path length L that generates a resonance phenomenon at the maximum rotational speed fm of the rotary compressor is set, high volumetric efficiency can be obtained with a smaller rotational speed.

“Second Embodiment”
A vertical two-cylinder rotary compressor according to a second embodiment of the present invention will be described with reference to FIGS. FIG. 10 is a longitudinal sectional view of a vertical two-cylinder rotary compressor according to the second embodiment of the present invention. In the second embodiment, the first suction chamber 30a is provided on the opposite side of the motor 3 and the second suction chamber 30b is provided on the motor 3 side, and the structure of the first embodiment is applied to each suction chamber 30. is there. The first suction chamber 30a includes a cylinder 10a, an end plate portion of the auxiliary bearing 19, a cylindrical roller 11a fitted on the outer periphery of the eccentric portion 5a, and a partition plate 31. The second suction chamber 30b includes a cylinder 10b, an end plate portion of the main bearing 9, a cylindrical roller 11b fitted on the outer periphery of the eccentric portion 5b, and a partition plate 31.

  The cylinder 10a, the roller 11a, and the eccentric portion 5a that constitute the first suction chamber 30a have the same shape as the cylinder 10b, the roller 11b, and the eccentric portion 5b that constitute the second suction chamber 30b. The suction rotation angles θs of the suction chambers 30a and 30b and the inner diameters and arrangements of the suction ports 13a and 13b were set to the same value. Therefore, the same member can be manufactured with the same processing tool. However, as shown in FIG. 10, the eccentric part 5a and the eccentric part 5b were eccentric in opposite directions so that the suction strokes differed by 180 °. This is because the load torque applied to the rotary shaft 2 in the compression stroke is dispersed in the opposite phase to suppress the loss and vibration of the compressor.

  Each suction chamber 30 is connected to two independent suction pipes 18. The first suction chamber was on the lower side of the second suction chamber, and the first suction pipe 18a was disposed on the outer peripheral side of the compressor with respect to the second suction pipe 18b. For each suction chamber 30, the suction flow path length L was set based on the formula (7) so that resonance occurred at the maximum compressor speed fm [Hz]. In the present embodiment, the suction end rotation angle θs is 0.6 [rad] and the resonance rotational speed fn is 120 [Hz] in both the first and second suction chambers. Based on 7), it was set to 0.3 [m]. Moreover, the curvature radius of the bending part 32 of the suction pipe 18 was made into the same value, respectively. Since the suction channel length L is the same, the open end height of the first suction pipe 18a is lower than the open end height of the second suction pipe 18b. The difference is about 40 [mm], which is about four times as large as the inner diameter of the suction pipe 13.

  In the conventional vertical two-cylinder rotary compressor, the open ends of the two suction pipes have the same height for ease of processing, and therefore the length of the suction flow path to each suction chamber is also different.

  According to this embodiment, since the suction flow path length L is the same, the resonance phenomenon can be utilized at the same rotation speed, and the volumetric efficiency of the entire compressor can be improved. FIG. 11 is a diagram showing the relationship between the rotational speed and the volumetric efficiency of the vertical two-cylinder rotary compressor according to the second embodiment. Conventionally, since the open ends of the two suction pipes have the same height, the first suction pipe is longer than the second suction pipe, and the suction flow path lengths are different. Therefore, from equation (7), the resonance speed fn is different between the first suction chamber and the second suction chamber, so that the volumetric efficiency of the entire compressor is lowered. In the present embodiment, the volume efficiency of the first and second suction chambers is simultaneously maximized according to the equation (7), so that the volume efficiency of the entire compressor can be further increased.

  FIG. 12 is a diagram showing the relationship between the rotational speed of the vertical two-cylinder rotary compressor according to the second embodiment and the total adiabatic efficiency. From the characteristics shown in FIG. 9, in the conventional two-cylinder rotary compressor, since the suction flow path lengths of the respective suction chambers are different, the total adiabatic efficiency is reduced at the resonance rotational speed fn of the entire compressor. In this embodiment, since the suction flow path length L is the same, the total heat insulation efficiency at the resonance rotational speed fn can be increased. Further, since the resonance rotational speed fn is set to the maximum rotational speed fm, the compressor can be operated in a rotational speed range in which the total adiabatic efficiency is high as shown in FIG.

“Third Embodiment”
A vertical two-cylinder rotary compressor according to a third embodiment of the present invention will be described with reference to FIG. Since the third embodiment has substantially the same configuration as the second embodiment, the description will focus on the main differences.

  In the third embodiment, the curvature radius Ra of the bent portion 32a of the first suction pipe 13a is set to 50 [mm], and the curvature radius Rb of the bent portion 32b of the second suction pipe 13b is set to 12 [mm]. The suction flow path length L was set to the same value based on the formula (7). Since the first curvature radius Ra is larger than the second curvature radius Rb, when setting the same suction flow path length L, in this embodiment, the open ends of the first and second suction pipes The difference in height can be reduced. Specifically, the difference between the two is about 20 [mm], and the difference is about 20 [mm] smaller than the structure of the second embodiment.

  In the second embodiment, as shown in FIG. 10, the distance from the inflow hole 24 is large because the height of the open end of the first suction pipe 18a is low. Therefore, when the second embodiment is applied to an air conditioner with a large amount of refrigerant filled, when the liquid refrigerant and the gas refrigerant flow from the inlet pipe 25 when the compressor is started, the liquid refrigerant is directly discharged from the suction pipe 18a. There was a possibility of inhalation.

  Note that a plurality of inflow holes 24 through which the refrigerant flows toward the suction pipe 18 are provided near the outer peripheral side of the flat screen seat 22 fixed to the inner wall of the accumulator 16, and the upper open end of the suction pipe 18 is It is a normal arrangement configuration that is arranged on the inner peripheral side out of the inflow hole 24. Assuming this arrangement, in particular, the first suction pipe 18a shown in FIG.

  On the other hand, in the third embodiment, the radius of curvature Ra of the bent portion 32a is increased and the height of the open end of the first suction pipe 18a is increased (compared to FIG. 10). It has a structure that suppresses inhalation. Furthermore, in the third embodiment, as shown in FIG. 13, the projection 23 of the mesh seat 22 is enlarged to lower the height of the mesh seat, and the structure closer to the first suction pipe 18a. Suppress inhalation.

  On the other hand, since the second suction pipe 18b is opened in the protrusion 23, when the gas refrigerant staying in the accumulator 16 is sucked, no fluid loss occurs when the suction pipe 18b sucks due to the enlarged protrusion shape. The structure. In FIG. 13, the open end of the second suction pipe 18b is shown so as to be accommodated in the projection (convex shape portion) of the flat screen seat. However, the present invention is not limited thereto, and the first suction pipe 18a is opened. The ends may also be accommodated in the convex portion. Therefore, in the third embodiment, it is possible to improve the reliability with respect to the suction of the liquid refrigerant while making use of the effects of the second embodiment.

It is a longitudinal section of the rotary compressor concerning a 1st embodiment of the present invention. It is a figure showing the suction state when time t = 0 in the rotary compressor concerning a 1st embodiment. It is a figure showing the suction state when time t = L / C in the rotary compressor which concerns on 1st Embodiment. It is a figure showing the suction state at the time of t = 2L / C in the rotary compressor which concerns on 1st Embodiment. It is a figure showing the suction state at the time t = 3L / C in the rotary compressor which concerns on 1st Embodiment. It is a figure showing the suction state at the time t = 4L / C in the rotary compressor which concerns on 1st Embodiment. It is a figure showing the relationship between the rotation angle of the rotary compressor which concerns on 1st Embodiment, and the time change rate of a suction volume. It is a figure showing the relationship between the resonant rotation speed fn and the suction flow path length L of the rotary compressor which concerns on 1st Embodiment. It is a figure showing the relationship between the rotation speed of the rotary compressor which concerns on 1st Embodiment, and total heat insulation efficiency. It is a longitudinal cross-sectional view of the vertical 2 cylinder rotary compressor which concerns on the 2nd Embodiment of this invention. It is a figure which shows the relationship between the rotation speed and volumetric efficiency of the vertical 2 cylinder rotary compressor which concerns on 2nd Embodiment. It is a figure showing the relationship between the rotation speed of the vertical type 2 cylinder rotary compressor which concerns on 2nd Embodiment, and total heat insulation efficiency. It is a longitudinal cross-sectional view of the vertical type 2 cylinder rotary compressor which concerns on the 3rd Embodiment of this invention. It is a figure showing the relationship between the rotation speed and volumetric efficiency of the rotary compressor regarding a prior art.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Compressor 2 Rotating shaft 3 Sealed container 4 Electric motor 5 Eccentric part 10 Cylinder 11 Roller 12 Through hole 13 Suction port 14 Internal space 15 Seal suction 16 Accumulator 18 Suction pipe 20 Closure plate 21 Net 22 Net seat 23 Projection (convex shape part)
24 Inflow hole 30 Suction chamber 31 Partition plate 32 Suction pipe bending part

Claims (5)

An electric motor disposed in an upper part of the sealed container; an eccentric portion having a rotating shaft coupled to the electric motor; a suction chamber formed by a roller and a cylinder provided on an outer periphery of the eccentric portion; and the suction chamber A rotary compressor comprising: an accumulator provided on the suction side; a suction pipe that opens in the accumulator; and a seal suction that connects the suction pipe and a suction port of the suction chamber;
Limiting the rotational speed of the rotary compressor that causes a resonance phenomenon that maximizes the refrigerant suction filling rate of the suction chamber to the maximum rotational speed of the rotary compressor;
When the suction of the suction chamber is completed, a rotation angle formed by a line segment connecting the rotation center of the roller and the center of the eccentric portion as seen from the vane separating the suction side and the discharge side of the suction chamber is obtained.
The length of the suction flow path from the open end of the suction pipe to the open end of the suction port is determined based on the maximum number of rotations of the limited rotary compressor and the rotation angle. Machine. An electric motor disposed in an upper part of the sealed container; an eccentric portion having a rotating shaft coupled to the electric motor; a suction chamber formed by a roller and a cylinder provided on an outer periphery of the eccentric portion; and the suction chamber A rotary compressor comprising: an accumulator provided on the suction side; a suction pipe that opens in the accumulator; and a seal suction that connects the suction pipe and a suction port of the suction chamber;
The suction flow path length from the open end of the suction pipe to the open end of the suction port is L [m], the maximum rotational speed of the compressor is fm [Hz], and the sound speed of the gas refrigerant is C [m / m s], and a rotation angle formed by a line segment connecting the rotation center of the roller and the center of the eccentric portion when viewed from the vane separating the suction side and the discharge side of the suction chamber at the end of the suction of the suction chamber is θs [ rad]
L = (π−θs) / (4π · fm)
The rotary compressor is characterized in that the relationship is established. On the outer periphery of the two eccentric parts provided via an electric motor arranged at the upper part in the sealed container, two eccentric parts having a rotating shaft connected to the electric motor, and a partition plate through which the rotating shaft passes. Two suction chambers each formed by a roller and a cylinder provided, an accumulator provided on the suction side of the two suction chambers, and two substantially L-shaped openings opened in the accumulator and having bent portions. A vertical two-cylinder rotary compressor comprising a suction pipe, and two seal suctions connecting the two suction pipes and the suction ports of the two suction chambers,
The two suction pipes form different paths, and the suction flow path lengths from the open ends of the respective suction pipes to the open ends of the respective suction ports are made equal to L [m], respectively. The maximum number of revolutions is fm [Hz], the sound speed of the gas refrigerant is C [m / s], and each of the two suction chambers is viewed from the vane that separates the suction side and the discharge side at the end of the suction. When the rotation angle formed by the line segment connecting the rotation center of each roller and the center of each eccentric portion is θs [rad],
L = (π−θs) / (4π · fm)
The vertical two-cylinder rotary compressor is characterized in that the relationship is established. In claim 3,
The curvature radius of the bent portion of the first suction pipe communicating with the first suction chamber located on the opposite side of the motor is the bending of the second suction pipe communicating with the second suction chamber located on the motor side. A vertical two-cylinder rotary compressor characterized in that the difference in height between the open ends of the first and second suction pipes on the accumulator side is reduced. In claim 3,
Provided with a mesh seat having a plurality of inflow holes for allowing the coolant to flow down at the upper part of the accumulator, and provided with a convex shape portion for accommodating the open end of the suction pipe on the accumulator side,
The heights of the open ends of the two suction pipes on the accumulator side are different between the first and second suction pipes, and the height of the open ends of the at least one suction pipe is lower than that of the screen seat. A vertical two-cylinder rotary compressor characterized by:

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