WO2013140912A9

WO2013140912A9 - Rotating compressor and freeze-cycle apparatus

Info

Publication number: WO2013140912A9
Application number: PCT/JP2013/053893
Authority: WO
Inventors: 富永健; 森嶋明; 加藤久尊; 長畑大志; 平山卓也
Original assignee: 東芝キヤリア株式会社
Priority date: 2012-03-23
Filing date: 2013-02-18
Publication date: 2014-08-07
Also published as: CN104081055B; JPWO2013140912A1; CN104081055A; WO2013140912A1

Abstract

4.6×10^-3≤dm/A≤6.5×10^-3(formula 1) and/or 4.6×10^-3≤ds/A≤6.5×10^-3(formula 2) are true, where A is an inner diameter area of a cylinder chamber (20a, 20b), dm is the inner diameter of a discharge port (22a) on a main bearing (16a), and ds is the inner diameter of a discharge port (22b) of a counter bearing (16b)). 1.2×10^-4≤Km/V≤3.5×10^-4 (formula 3) and/or 1.2×10^-4≤Ks/V≤3.5×10^-4 (formula 4) are true, where V is the volume of a cylinder chamber (20a, 20b), Km is the spring constant of a discharge valve (23a) on the main bearing (16a), and Ks is the spring constant of the discharge valve (23b) on the counter bearing (16b).

Description

Rotary compressor and refrigeration cycle apparatus

Embodiments of the present invention relate to a rotary compressor and a refrigeration cycle apparatus using the rotary compressor.

As a rotary compressor used in a refrigeration cycle apparatus such as an air conditioner, one having a configuration as shown in the invention described in Japanese Patent Application Laid-Open No. 2007-92534 is known. That is, a cylinder, a rotating shaft that penetrates the cylinder, a roller that is provided on the rotating shaft and moves eccentrically (eccentric rotation) in the cylinder, and rotatably supports the rotating shaft and closes the end face of the cylinder. It has two bearings (main bearing, sub bearing) which form a cylinder chamber in it. In such a rotary compressor, a discharge port for discharging the gas refrigerant compressed in the cylinder chamber and a discharge valve for opening and closing the discharge port are provided in the flange portion of the bearing.

The discharge valve opens when the gas refrigerant compressed in the cylinder chamber reaches a predetermined pressure. And it closes, after gas refrigerant is discharged from a discharge port.

Such rotary compressors are often operated in a medium-speed rotation range (operation frequency is 30 to 60 Hz) or a low-speed rotation range (operation frequency is 30 Hz or less) except during startup. Therefore, the dimensions of each part are set with an emphasis on increasing the compression efficiency in the medium speed rotation range or the low speed rotation range.

Therefore, the inner diameter of the discharge port is made as small as possible in order to reduce the dead volume and reduce the re-expansion loss. The ratio “d / A” of the inner diameter dimension d (mm) of the discharge port to the inner diameter area A (mm ² ) of the cylinder chamber is “3 × 10 ⁻³ <d / A <4 × 10 ⁻³ (mm / mm ^2). ) ”.

The discharge valve has a small spring constant so that the valve opens smoothly when the pressure in the cylinder chamber rises. The discharge valve spring constant K (N / mm) and the cylinder chamber volume V (mm ³ ) are used. The ratio “K / V” is set to “0.6 × 10 ⁻⁴ <K / V <0.7 × 10 ⁻⁴ (N / mm ⁴ )”.

JP 2007-92534 A

However, when such a rotary compressor is operated in a high-speed rotation region of 70 Hz or higher, which is 1.2 times or more of the commercial power supply frequency, the passage resistance of the discharge port increases. This is because the compression speed is increased while the inner diameter of the discharge port is small. And since the passage resistance of the discharge port is increased, the gas refrigerant in the cylinder chamber is excessively compressed, which causes a problem that loss due to overcompression increases and compression efficiency decreases. Further, the gas refrigerant load around the rotation axis increases due to excessive compression of the gas refrigerant, causing a problem that the sliding portion is easily deteriorated and the reliability of the sliding portion is lowered.

Also, when the rotary compressor is operated in the high speed rotation region when the spring constant of the discharge valve is small, a situation occurs in which the timing for closing the discharge valve is delayed. Then, when the timing at which the discharge valve closes is delayed, the compressed gas refrigerant discharged from the discharge port flows back into the cylinder chamber. The backflowed gas refrigerant is further re-expanded in the cylinder chamber, so that a loss due to the re-expansion occurs, and the compression efficiency of the rotary compressor is lowered.

An object of an embodiment of the present invention is to provide a rotary compressor capable of increasing compression efficiency when operated in a high-speed rotation region, and a refrigeration cycle apparatus including the rotary compressor.

In the rotary compressor according to the embodiment, an electric motor part and a compression mechanism part connected to the electric motor part via a rotary shaft having an eccentric part are accommodated in a sealed case, and the rotary shaft penetrates the compression mechanism part. A cylinder that is fitted to the eccentric part and moves eccentrically while part of the outer peripheral surface is in contact with the inner peripheral surface of the cylinder, and supports the rotating shaft and closes the end surface of the cylinder to close the cylinder in the cylinder. A discharge port for discharging the working fluid compressed in the cylinder chamber by a roller eccentrically moving to the main bearing and the sub-bearing into the sealed case, and the discharge port. And a discharge valve for opening and closing. When the inner diameter area of the cylinder chamber is A (mm ² ), the inner diameter dimension of the discharge port provided in the main bearing is dm (mm), and the inner diameter dimension of the discharge port provided in the auxiliary bearing is ds (mm). , And at least one of formula (1) and formula (2) is established,
The volume of the cylinder chamber is V (mm ³ ), the spring constant of the discharge valve provided in the main bearing is Km (N / mm), and the spring constant of the discharge valve provided in the auxiliary bearing is Ks (N / mm). At this time, at least one of Expression (3) and Expression (4) is set.

4.6 × 10 ⁻³ ≦ dm / A ≦ 6.5 × 10 ⁻³ (mm / mm ² ) (1)
4.6 × 10 ⁻³ ≦ ds / A ≦ 6.5 × 10 ⁻³ (mm / mm ² ) (2)
1.2 × 10 ⁻⁴ ≦ Km / V ≦ 3.5 × 10 ⁻⁴ (N / mm ⁴ ) (3)
1.2 × 10 ⁻⁴ ≦ Ks / V ≦ 3.5 × 10 ⁻⁴ (N / mm ⁴ ) (4)
However, when there are a plurality of cylinders, V is the volume of one cylinder chamber.

FIG. 1 is a configuration diagram of a refrigeration cycle apparatus including a rotary compressor shown in cross section in the first embodiment. FIG. 2 is a plan view showing the main bearing. FIG. 3 is a plan view showing the discharge valve. FIG. 4 shows an experiment of the relationship between the ratio “dm / A” between the inner diameter area A (mm ² ) of the first cylinder chamber and the inner diameter dimension dm (mm) of the first discharge port and the compression efficiency of the rotary compressor. It is a graph which shows a result. FIG. 5 shows the ratio “K / V (N / mm ⁴ )” between the volume V (mm ³ ) of the cylinder chamber and the spring constant K (N / mm) of the discharge valve, and the compression efficiency of the rotary compressor. It is a graph which shows the experimental result of a relationship. FIG. 6 shows a rotary compressor that satisfies equations (1) to (4), a conventional rotary compressor that does not satisfy all of equations (1) to (4), and equations (1) and (2). It is a PV diagram measured when the rotary compressor of the comparative example which satisfy | fills (3) and (4) is satisfy | filled but is drive | operated in a high-speed rotation area. FIG. 7 shows a rotary compressor that satisfies equations (1) to (4), a conventional rotary compressor that does not satisfy all of equations (1) to (4), and equations (1) to (2). 6 is a graph showing the experimental results of the relationship between the rotational speed and the overall efficiency of the rotary compressor in the comparative rotary compressor that does not satisfy the expressions (3) and (4). FIG. 8 is a longitudinal front view showing the shape of the discharge port in the second embodiment. FIG. 9 is a graph showing experimental results of the relationship between the ratio “dm2 / dm1” between the minimum diameter dimension dm1 and the maximum diameter dimension dm2 of the tapered portion provided on the discharge side of the discharge port and the overall efficiency of the rotary compression. is there. FIG. 10 is a plan view showing a cylinder in the third embodiment. 11 is a cross-sectional view taken along line YY in FIG. FIG. 12 is a perspective view showing a part of the cylinder of FIG. 10 in cross section.

Hereinafter, embodiments of the present invention will be described with reference to the drawings.

(First embodiment)
1st Embodiment is described based on FIG. 1 thru | or FIG. A refrigeration cycle apparatus 1 shown in FIG. 1 includes a compressor body 2 and an accumulator 3, and includes a rotary compressor 4 that compresses a low-pressure gas refrigerant that is a working fluid into a high-pressure gas refrigerant. Further, the refrigeration cycle apparatus 1 is connected to the discharge side of the compressor body 2 to condense the high-pressure gas refrigerant into a liquid refrigerant, and an expansion device 6 connected to the condenser 5 to decompress the liquid refrigerant. And an evaporator 7 connected between the expansion device 6 and the accumulator 3 to evaporate the liquid refrigerant.

The compressor body 2 has a sealed case 8 formed in a cylindrical shape. Lubricating oil 9 is stored at the bottom of the sealed case 8. Further, in the sealed case 8, an electric motor part 10 located on the upper side and a compression mechanism part 11 located on the lower side are accommodated. The electric motor unit 10 and the compression mechanism unit 11 are connected via a rotating shaft 12. The rotating shaft 12 rotates about the center in the longitudinal direction of the compressor body 2 as a rotating shaft.

The electric motor unit 10 is a so-called motor, and rotates the rotary shaft 12. The electric motor unit 10 includes a rotor 13 and a stator 14. The rotor 13 is fixed to the rotating shaft 12 and is provided with a permanent magnet (not shown). The stator 14 is fixed to the sealed case 8 and disposed at a position surrounding the rotor 13, and a coil (not shown) is wound around the stator 14.

The compression mechanism unit 11 compresses the low-pressure gas refrigerant. The compression mechanism part 11 has the 1st cylinder 15a located in the upper part side, and the 2nd cylinder 15b located in the lower part side. A partition plate 17 is provided between the first cylinder 15a and the second cylinder 15b. A main bearing 16a that rotatably supports the rotary shaft 12 is fixed to the upper end surface of the first cylinder 15a, and a sub bearing 16b that rotatably supports the rotary shaft 12 is fixed to the lower end surface of the second cylinder 15b. Yes.

The rotary shaft 12 is disposed through the first cylinder 15a and the second cylinder 15b. The rotary shaft 12 is provided with a first eccentric portion 18a and a second eccentric portion 18b having the same diameter with a phase difference of 180 °. A first roller 19a is fitted to the first eccentric portion 18a, and a second roller 19b is fitted to the second eccentric portion 18b.

In the first cylinder 15a, a first cylinder chamber 20a is formed in which both ends of the first cylinder 15a are closed by a main bearing 16a and a partition plate 17. Inside the second cylinder 15b, a second cylinder chamber 20b is formed in which both ends of the second cylinder 15b are closed by the partition plate 17 and the auxiliary bearing 16b. A first roller 19a fitted to the first eccentric part 18a is accommodated in the first cylinder chamber 20a, and a second roller 19b fitted to the second eccentric part 18b is accommodated in the second cylinder chamber 20b. Has been. The first roller 19a and the second roller 19b are moved eccentrically (eccentric rotation) while the outer peripheral surface thereof is in line contact with the inner peripheral surfaces of the first cylinder 15a and the second cylinder 15b when the rotary shaft 12 rotates. Has been placed.

In the first cylinder chamber 20a and the second cylinder chamber 20b, the two cylinder chambers are partitioned into two spaces whose volume and pressure change as the first roller 19a and the second roller 19b rotate. A blade (not shown) is accommodated. The blade is in contact with the outer peripheral surfaces of the first roller 19a and the second roller 19b at the tip. The main bearing 16a is provided with a first discharge valve mechanism 21a. The first discharge valve mechanism 21a includes a first discharge port 22a formed in the main bearing 16a, a first reed valve 23a, and a first valve stopper 24a. The first reed valve 23a is a first discharge valve that is screwed to the main bearing 16a to open and close the first discharge port 22a. The first valve stopper 24a is screwed to the main bearing 16a together with the first reed valve 23a to restrict the maximum opening of the first reed valve 23a. The first discharge valve mechanism 21a is covered with a first muffler 25a attached to the main bearing 16a. The first muffler 25a is formed with a discharge hole 26 that communicates the inside and outside of the first muffler 25a.

The second discharge valve mechanism 21b is provided in the auxiliary bearing 16b. The second discharge valve mechanism 21b has the same configuration as the first discharge valve mechanism 21b described above, and includes a second discharge port 22b formed in the auxiliary bearing 16b, a second reed valve 23b, and a second valve stopper 24b. Have. The second reed valve 23b is a second discharge valve that is screwed to the auxiliary bearing 16b to open and close the second discharge port 22b. The second valve stopper 24b is screwed to the auxiliary bearing 16b together with the second reed valve 23b to restrict the maximum opening of the second reed valve 23b. The second discharge valve mechanism 21b is covered with a second muffler 25b attached to the auxiliary bearing 16b. The inside of the second muffler 25b and the inside of the first muffler 25a are gas refrigerant guide passages (not shown) formed through the auxiliary bearing 16b, the second cylinder 15b, the partition plate 17, the first cylinder 15a, and the main bearing 16a. The gas refrigerant is movably connected.

The accumulator 3 has a cylindrical sealed case 27. The accumulator 3 and the evaporator 7 are connected so that the gas refrigerant vaporized by the evaporator 7 or the liquid refrigerant not vaporized by the evaporator 7 flows into the sealed case 27. In the sealed case 27, there are provided two suction pipes 28, one end of which opens at the upper side in the sealed case 27 and is arranged so that only the gas refrigerant in the sealed case 27 flows. The other ends of these suction pipes 28 extend from the lower end side of the sealed case 27 to the outside of the sealed case 27 and are connected to the first cylinder chamber 20 a and the second cylinder chamber 20 b of the compression mechanism unit 11. An oil return hole 29 into which lubricating oil accumulated at the bottom of the airtight case 27 flows is formed in a portion of the suction pipe 28 located on the lower side in the airtight case 27.

FIG. 2 is a plan view showing the main bearing 16a described above. As described above, the first discharge port 22a is formed in the main bearing 16a. The main bearing 16a is formed with a screw hole 30 in which a screw (not shown) for fixing the first reed valve 23a and the first valve stopper 24a described above to the main bearing 16a is fitted by a screw action. Yes. The inner diameter of the first discharge port 22a is set to dm (mm).

Although illustration by a plan view is omitted, the sub bearing 16b described above has the same structure as the main bearing 16a. Accordingly, the above-described second discharge port 22b is formed in the auxiliary bearing 16b, and the inner diameter of the second discharge port 22b is set to ds (mm).

FIG. 3 is a plan view showing the first reed valve 23a arranged at the mounting position on the main bearing 16a. The first reed valve 23a is formed of a plate-like member. The first reed valve 23 a has an arm portion 32 and a valve main body portion 33. The arm portion 32 has flexibility, and an attachment hole 31 into which a fixing screw is inserted is formed at one end. The valve body portion 33 is provided on the other end side of the arm portion 32 and is formed in a size capable of closing the first discharge port 22a in a disc shape. The outer dimension of the valve body 33 is set to R (mm), the width of the arm 32 is set to W (mm), and the ratio “R / W” is set to “R / W ≧ 2”. The connecting portion between the valve body 33 and the arm 32 is constricted from the valve body 33 toward the arm 32.

Although not shown in the plan view, the second reed valve 23b attached to the auxiliary bearing 16b is also formed in the same shape as the first reed valve 23a.

Here, in this rotary compressor 4, the inner diameter area of each of the first cylinder chamber 20a and the second cylinder chamber 20b is A (mm ² ), the inner diameter dimension of the first discharge port 22a is dm (mm), The inner diameter dimension of the discharge port 22b is ds (mm). At this time, the ratio “dm / A (mm / mm ² )” between the inner diameter area A (mm ² ) of the first cylinder chamber 20 a and the inner diameter dimension dm (mm) of the first discharge port 22 a, the second cylinder chamber 20 b The ratio “ds / A” (mm / mm ² ) between the inner diameter area A (mm ² ) and the inner diameter ds (mm) of the second discharge port 22b is respectively
4.6 × 10 ⁻³ ≦ dm / A ≦ 6.5 × 10 ⁻³ (mm / mm ² ) (1)
4.6 × 10 ⁻³ ≦ ds / A ≦ 6.5 × 10 ⁻³ (mm / mm ² ) (2)
Is set to hold.

Further, in this rotary compressor 4, the volume of each of the first cylinder chamber 20a and the second cylinder chamber 20b is V (mm ³ ), the spring constant of the first reed valve 23a is Km (N / mm), the second Let the spring constant of the reed valve 23b be Ks (N / mm). At this time, the ratio “Km / V (N / mm ⁴ )” between the volume V (mm ³ ) of the first cylinder chamber 20a and the spring constant Km (N / mm) of the first reed valve 23a, the second cylinder chamber 20b. The ratio “Ks / V (N / mm ⁴ )” of the volume V (mm ³ ) of the second and the spring constant Ks (N / mm) of the second reed valve 23b is respectively
1.2 × 10 ⁻⁴ ≦ Km / V ≦ 3.5 × 10 ⁻⁴ (N / mm ⁴ ) (3)
1.2 × 10 ⁻⁴ ≦ Ks / V ≦ 3.5 × 10 ⁻⁴ (N / mm ⁴ ) (4)
Is set to hold.

The volumes “V” of the first cylinder chamber 20a and the second cylinder chamber 20b shown in the equations (3) and (4) mean the volume of one cylinder chamber when there are a plurality of cylinder chambers. .

Further, as shown in FIG. 3, in the first reed valve 23a and the second reed valve 23b, when the outer dimension of the valve body 33 is R (mm) and the width of the arm 32 is W (mm) The ratio “R / W” between the outer dimension R (mm) of the valve body 33 and the width W (mm) of the arm part 32 is:
R / W ≧ 2 ... (5)
Is set to hold.

The dimensions of each component of the rotary compressor 4 are set as follows, for example.

The inner diameters of the first cylinder chamber 20a and the second cylinder chamber 20b are 43 mm,
The height dimensions of the first cylinder chamber 20a and the second cylinder chamber 20b are each 18 mm,
The outer dimensions of the first roller 19a and the second roller 19b are each 35 mm,
The eccentric amounts of the 118th and second eccentric portions 18b of the rotating shaft 12 (the distances from the rotation center of the rotating shaft 12 to the centers of the 118a and second eccentric portions 18b) are 4 mm, respectively.
The inner diameter dimension (dm) of the first discharge port 22a is 8 mm,
The inner diameter dimension (ds) of the second discharge port 22b is 8 mm.

Here, the setting of the ranges of “dm / A (mm / mm ² )” and “ds / A (mm / mm ² )” described above will be described with reference to the graph showing the experimental results in FIG. FIG. 4 shows an experiment of the relationship between the ratio “dm / A” between the inner diameter area A (mm ² ) of the first cylinder chamber and the inner diameter dimension dm (mm) of the first discharge port and the compression efficiency of the rotary compressor. It is a graph which shows a result.

In this experiment, the rotation speed in the medium speed rotation range was 40 Hz, and the rotation speed in the high speed rotation range was 90 Hz. The reed valve used in this experiment is used in a conventional rotary compressor, and has a so-called small spring constant and is soft.

In the graph shown in FIG. 4, the horizontal axis is “dm / A (mm / mm ² )” and the vertical axis is the compression efficiency of the rotary compressor.

As can be seen from the graph shown in FIG. 4, when the rotary compressor is operated in the middle speed range, the value of “dm / A (mm / mm ² )” increases, that is, the discharge port. As the ratio of the inner diameter dimension “dm” increases, the compression efficiency decreases.

On the other hand, when the rotary compressor is operated in a high-speed rotation region, the value of “dm / A (mm / mm ² )” is “4.6 × 10 ⁻³ ≦ dm / A ≦ 6.5 ×. The compression efficiency was improved in the range of 10 ⁻³ ”, and the maximum compression efficiency was obtained when the value of“ dm / A (mm / mm ² ) ”was 5.5 × 10 ⁻³ . Therefore, by setting “dm / A (mm / mm ² )” in the range of “4.6 × 10 ⁻³ ≦ dm / A ≦ 6.5 × 10 ⁻³ ”, the rotational type in the high-speed rotation range is set. The compression efficiency of the compressor can be increased.

Note that the description has been given by taking the case of “dm / A (mm / mm ² )” as an example, but the same applies to the case of “ds / A (mm / mm ² )”.

Next, setting of the ranges of “Km / V (N / mm ⁴ )” and “Ks / V (N / mm ⁴ )” described above will be described using a graph showing the experimental results in FIG. FIG. 5 shows the ratio “K / V (N / mm ⁴ )” between the volume V (mm ³ ) of the cylinder chamber and the spring constant K (N / mm) of the discharge valve, and the compression efficiency of the rotary compressor. It is a graph which shows the experimental result of a relationship.

In the graph shown in FIG. 5, the horizontal axis is “Km / V (N / mm ⁴ )”, and the vertical axis is the compression efficiency of the rotary compressor. In this experiment, the value of “dm / A (mm / mm ² )” is fixed to 5.5 × 10 ⁻³ and the spring constant “Km” of the reed valve is varied. The experiment was conducted for operation in the high-speed rotation range.

As can be seen from the graph shown in FIG. 5, the value of “Km / V (N / mm ⁴ )” is “1.2 × 10 ⁻⁴ ≦ Km / V ≦ 3.5 × 10 ⁻⁴ (N / mm ⁴ ) By setting the spring constant “Km” of the reed valve so as to be in the range of “”, the compression efficiency of the rotary compressor is increased. The case of “Km / V (N / mm ⁴ )” has been described as an example, but the same applies to the case of “Ks / V (N / mm ⁴ )”.

In such a configuration, in the rotary compressor 4, when the motor unit 10 is energized, the first roller 19a and the second roller 19b rotate eccentrically around the center line of the rotary shaft 12, and the compression mechanism The unit 11 is driven.

When the compression mechanism 11 is driven, the volume and pressure of the two spaces in the first cylinder chamber 20a and the second cylinder chamber 20b change with the eccentric rotation of the first roller 19a and the second roller 19b. To do. By changing the volume and the pressure, a low-pressure gas refrigerant is sucked from the accumulator 3 through the suction pipe 28 into the first cylinder chamber 20a and the second cylinder chamber 20b. The sucked low-pressure gas refrigerant is compressed in the first cylinder chamber 20a and the second cylinder chamber 20b, and becomes a high-pressure gas refrigerant.

In the first cylinder 15a, the first reed valve 23a is opened at the timing when the pressure of the gas refrigerant in the first cylinder chamber 20a rises to a predetermined value. The high-pressure gas refrigerant in the first cylinder chamber 20a passes through the first discharge port 22a and is discharged into the first muffler 25a. The gas refrigerant discharged into the first muffler 25a is discharged into the sealed case 8 through the discharge hole 26 of the first muffler 25a.

In the second cylinder 15b, the second reed valve 23b is opened at the timing when the pressure of the gas refrigerant in the second cylinder chamber 20b rises to a predetermined value. The high-pressure gas refrigerant in the second cylinder chamber 20b passes through the second discharge port 22b and is discharged into the second muffler 25b. The gas refrigerant discharged into the second muffler 25b flows into the first muffler 25a through the gas refrigerant guide passage described above, and further discharged from the first muffler 25a through the discharge hole 26 into the sealed case 8. Is done.

The high-pressure gas refrigerant compressed in the first cylinder chamber 20a and the second cylinder chamber 20b and discharged into the sealed case 8 flows into the condenser 5 and is dissipated in the condenser 5 to be liquefied with the liquid refrigerant. Become. The liquid refrigerant flows into the expansion device 6 and is depressurized. After being depressurized, the liquid refrigerant flows into the evaporator 7 and absorbs heat to evaporate to become a gas refrigerant. The gas refrigerant evaporated in the evaporator 7 flows into the accumulator 3 and gas-liquid separation (separation of liquid components contained in the gas refrigerant) is performed. Of these, only the gas refrigerant passes through the suction pipe 28 of the accumulator 3 and is supplied into the first cylinder chamber 20a and the second cylinder chamber 20b of the compression mechanism 11, and is compressed again.

Here, FIG. 6 shows the rotary compressor 4 of the present embodiment that satisfies all of the expressions (1), (2), (3), and (4), and the expressions (1) and ( 2), the rotary compressor of the conventional example that does not satisfy all of the expressions (3) and (4), and the expressions (3) and (4) satisfying the expressions (1) and (2). It is a PV diagram measured when the rotary compressor of the comparative example which is not satisfied is operated in a high-speed rotation region.

Specifically, in the rotary compressor of the conventional example, dm / A is 3.5 × 10 ⁻³ (mm / mm ² ), ds / A is 3.5 × 10 ⁻³ (mm / mm ² ), Km / V is 0.8 × 10 ⁻⁴ (N / mm ⁴ ), and Ks / V is 0.8 × 10 ⁻⁴ (N / mm ⁴ ).

In the rotary compressor of the comparative example, dm / A is 5.5 × 10 ⁻³ (mm / mm ² ), ds / A is 5.5 × 10 ⁻³ (mm / mm ² ), and Km / V is 0. 8 × 10 ⁻⁴ (N / mm ⁴ ) and Ks / V is set to 0.8 × 10 ⁻⁴ (N / mm ⁴ ).

In the rotary compressor 4 of this embodiment, dm / A is 5.5 × 10 ⁻³ (mm / mm ² ), ds / A is 5.5 × 10 ⁻³ (mm / mm ² ), and Km / V. Is 1.7 × 10 ⁻⁴ (N / mm ⁴ ), and Ks / V is 1.7 × 10 ⁻⁴ (N / mm ⁴ ).

According to the PV diagram shown in FIG. 6, in the rotary compressor of the conventional example, since the inner diameter dimension of the discharge port is small, the passage resistance of the gas refrigerant passing through the discharge port is increased. Therefore, even after the reed valve is opened, loss due to overcompression occurs due to the compression of the gas refrigerant. In the conventional rotary compressor, a soft reed valve having a small spring constant is used, and therefore the timing at which the reed valve closes after the high-pressure gas refrigerant is discharged is delayed. When the reed valve closes late, the high-pressure gas refrigerant once discharged from the discharge port flows backward from the discharge port into the cylinder chamber. And the loss by re-expansion generate | occur | produces when this back-flowed gas refrigerant re-expands in a cylinder chamber.

On the other hand, in the rotary compressor of the comparative example, since the inner diameter of the discharge port is large, the passage resistance of the gas refrigerant passing through the discharge port is reduced, and the loss due to overcompression is reduced. Further, as the loss due to overcompression becomes smaller, the force acting in the direction to prevent the reed valve from closing is reduced, so that the delay in timing for closing the reed valve is suppressed, and the loss due to re-expansion is reduced.

Furthermore, in the rotary compressor 4 of the present embodiment, the inner diameter dimensions of the discharge ports (the first discharge port 22a and the second discharge port 22b) are large. Therefore, the passage resistance of the gas refrigerant passing through these discharge ports (the first discharge port 22a and the second discharge port 22b) is reduced, and the loss due to overcompression is reduced. Further, since reed valves (first reed valve 23a and second reed valve 23b) having a large spring constant are used, the reed valves (first reed valve 23a and second reed valve) after high-pressure gas refrigerant is discharged are used. The timing for closing the valve 23b) is accelerated. Therefore, the backflow of the gas refrigerant from the discharge port (first discharge port 22a, second discharge port 22b) into the cylinder chamber (first cylinder chamber 20a, second cylinder chamber 20b) is suppressed, and loss due to re-expansion is small. Become.

When gas refrigerant is sucked into the cylinder chamber, a loss due to overexpansion occurs. This loss due to overexpansion is caused by the conventional rotary compressor, the comparative rotary compressor, and the present embodiment. The same occurs in the rotary compressor 4.

FIG. 7 shows the rotational speed and the overall efficiency (of the rotary compressor) in the rotary compressor of the conventional example described in FIG. 6, the rotary compressor of the comparative example, and the rotary compressor 4 of the present embodiment. It is the graph which showed the relationship with (total efficiency). From the graph of FIG. 7, it can be seen that in the rotary compressor of the conventional example, the overall efficiency is high in the medium speed rotation region and the overall efficiency is low in the high speed rotation region. On the other hand, in the rotary compressor of the comparative example, the overall efficiency is low in the medium speed rotation range and the total efficiency is high in the high speed rotation range, but the maximum value of the total efficiency is small. Furthermore, in the rotary compressor 4 of this embodiment, it turns out that total efficiency is low in a medium speed rotation area, and the total efficiency in a high speed rotation area is improving rather than the rotary compressor of a comparative example.

Therefore, the rotary compressor 4 satisfying the equations (1) to (4) can improve the overall efficiency in the high-speed rotation region and can reduce the power consumption.

In addition, when Formula (1) or Formula (2) is satisfied in any one of the first discharge port 22a and the second discharge port 22b, the first discharge port 22a in which Formula (1) or Formula (2) is satisfied, Loss due to overcompression can be reduced on the second discharge port 22b side. In addition, when the formula (3) or the formula (4) is established in any one of the first reed valve 23a and the second reed valve 23b, the first reed valve 23a that satisfies the formula (3) or the formula (4), Loss due to re-expansion can be reduced on the second reed valve 23b side. Therefore, in the rotary compressor 4 having the two discharge ports, the first discharge port 22a and the second discharge port 22b, in either one of the first discharge port 22a and the second discharge port 22b, the expression (1) or the expression ( If 2) is satisfied and either the first reed valve 23a or the second reed valve 23b satisfies the expression (3) or the expression (4), the overall efficiency in the high-speed rotation range can be improved. it can.

Next, as shown in FIG. 3, in the first reed valve 23a, the ratio of the outer dimension R (mm) of the valve body 33 and the width dimension W (mm) of the arm 32 is “R / W ≧ 2”. The connection portion between the valve main body 33 and the arm 32 is constricted. For this reason, when the first reed valve 23a is opened, the gas refrigerant discharged from the first discharge port 22a passes through the peripheral portion of the valve main body 33 and the peripheral portion of the arm portion 32 shown by hatching in FIG. Will come to do. Therefore, the passage area of the gas refrigerant discharged from the first discharge port 22a increases, and the flow rate of the discharged gas refrigerant decreases. Thereby, it is possible to reduce the force acting in the direction of preventing the gas refrigerant discharged from the first discharge port 22a from closing the first reed valve 23a that is about to close. Therefore, it is possible to prevent the delay due to the re-expansion by preventing the delay of the reed valve closing timing. This also applies to the second reed valve 23b formed in the same shape as the first reed valve 23a.

Further, since the ratio “R / W” between the outer dimension R (mm) of the valve main body 33 and the width dimension W (mm) of the arm 32 is set to “R / W ≧ 2”, the first The reed valve 23 a is easily twisted at the arm portion 32. For this reason, the valve body 33 is prevented from coming into contact with the valve seat portion of the first discharge port 22a, and the adhesion of the valve body 33 to the valve seat portion is improved. And the compression performance in the 1st cylinder chamber 20a can be improved because the adhesiveness of the valve main-body part 33 with respect to a valve-seat part improves. Furthermore, damage to the first discharge port 22a due to one-sided contact can be prevented, and a highly reliable valve mechanism can be obtained. This also applies to the second reed valve 23b formed in the same shape as the first reed valve 23a.

In the present embodiment, the rotary compressor in which the first roller 19a and the second roller 19b and the blade are separately formed has been described. However, in the present invention, the rotation in which the roller and the blade are integrally formed is described. It is also applicable to a type compressor (swing type compressor).

Further, in the present embodiment, there are two cylinder chambers, the first cylinder chamber 20a and the second cylinder chamber 20b, and the first discharge port 22a is formed in the main bearing 16a that forms the first cylinder chamber 20a. The case where the second discharge port 22b is formed in the auxiliary bearing 16b that forms the two-cylinder chamber 20b is described as an example. However, the rotary compressor to which the present invention can be applied may have a single cylinder chamber, and a discharge port may be formed in each of the main bearing and the sub-bearing that form the cylinder chamber. .

(Second Embodiment)
A second embodiment will be described with reference to FIGS. Note that, in the second embodiment and other embodiments, the same components as those described in the first embodiment are denoted by the same reference numerals, and redundant descriptions are omitted.

The basic structure of the second embodiment is the same as that of the first embodiment, and the second embodiment is different from the first embodiment in that the valve seat side, which is the outlet side of the first discharge port 22a. A taper portion 34 that gradually expands in the gas refrigerant discharge direction is formed. A similar taper portion is formed on the outlet side of the second discharge port 22b.

In the first discharge port 22a, when the minimum diameter of the tapered portion 34 is dm1 (mm) and the maximum diameter of the tapered portion 34 is dm2 (mm), the ratio “dm2 / dm1” is
1.1 ≦ dm2 / dm1 ≦ 1.35 (6)
Is set to hold.

Although not shown, when the minimum diameter dimension of the tapered portion formed on the outlet side of the second discharge port 22b is ds1 (mm) and the maximum diameter dimension is ds2 (mm), the ratio “ ds2 / ds1 is
1.1 ≦ ds2 / ds1 ≦ 1.35 (7)
Is set to hold.

Here, the above-described setting of the range of “dm2 / dm1” will be described using the graph showing the experimental results in FIG. In the graph shown in FIG. 9, the horizontal axis is a ratio “dm2 / dm1” between the minimum diameter dm1 (mm) of the tapered portion 34 and the maximum diameter dm2 (mm) of the tapered portion 34, and the vertical axis is This is the compression efficiency of the rotary compressor.

As can be seen from the graph shown in FIG. 9, the overall efficiency of the rotary compressor is improved by 3% to 5% in the range of “1.1 ≦ dm2 / dm1 ≦ 1.35” shown in the equation (6). Turned out to be.

The same applies to the tapered portion of the discharge port 22b formed in the second discharge port 22b. In the range of “1.1 ≦ ds2 / ds1 ≦ 1.35” shown in the equation (7), the rotary compressor Overall efficiency is improved by 3% to 5%.

In such a configuration, the tapered portion 34 formed on the outlet side of the first discharge port 22a satisfies the formula (6), so that the gas discharged from the first discharge port 22a during operation in the high-speed rotation region. The refrigerant flows along the tapered portion 34 as shown by the arrow in FIG. As a result, the passage resistance of the discharged gas refrigerant is further reduced, the loss due to overcompression can be further reduced, and as a result, the overall efficiency of the rotary compressor 4 can be improved. The same can be said when Expression (7) is satisfied on the second discharge port 22b side.

In addition, when Formula (6) or Formula (7) is satisfied in any one of the first discharge port 22a and the second discharge port 22b, the first discharge port 22a in which Formula (6) or Formula (7) is satisfied, Loss due to overcompression can be reduced on the second discharge port 22b side. Therefore, in the rotary compressor 4 having two discharge ports such as the first discharge port 22a and the second discharge port 22b, in either one of the first discharge port 22a and the second discharge port 22b, the expression (6) or the expression Even when (7) is satisfied, the overall efficiency in the high-speed rotation region can be improved.

Although FIG. 8 shows an example in which the valve seat on the outlet side of the first discharge port 22a has a flat shape, the shape of the valve seat may be an arc shape. In this case, the maximum diameter dimension “dm2” of the tapered portion 34 is a dimension between the apexes of the arcuate shape.

(Third embodiment)
A third embodiment will be described with reference to FIGS.

The basic structure of the third embodiment is the same as that of the first embodiment, and the third embodiment is different from the first embodiment in that a relief portion 35 is provided on the inner peripheral surface of the first cylinder 15a. It is a point that is formed. The escape portion 35 is formed in a recessed shape with respect to the inner peripheral surface of the first cylinder 15a. Although not shown, a similar relief portion is formed on the inner peripheral surface of the second cylinder 15b.

As shown in FIG. 10, a reference line connecting the center point “O” of the first cylinder 15a and the blade accommodation chamber 36 formed in the first cylinder 15a is “X”. An angle from the reference line “X” to the center of the first discharge port 22a formed in the main bearing 16a (see FIG. 1) fixed to the first cylinder 15a is “θ1”. Furthermore, when the angle to the point where the first discharge port 22a and the inner peripheral surface of the first cylinder 15a intersect and intersect on the side where the angle is larger than the angle “θ1” is “θ2”, the relief portion 35 is formed in the range of the angle “θ3 (θ3> θ2)” from the reference line “X”.

FIG. 11 is a cross-sectional view taken along line YY in FIG. As shown in FIG. 11, when the height dimension of the first cylinder 15a is “H”, the height dimension of the escape portion 35 is “h” from the side where the main bearing 16a is fixed. It is formed in the range. The height dimension “H” of the first cylinder 15a and the height dimension “h” of the relief portion 35 are set to “H / 2 ≦ h <H”.

FIG. 12 is a perspective view showing a part of the first cylinder 15a in cross section. On the inner peripheral surface of the first cylinder 15a, an escape portion 35, a blade accommodating chamber 36, and a suction hole 37 into which the gas refrigerant is sucked into the first cylinder 15a are formed.

In such a configuration, during the compression stroke, the first roller 19a (see FIG. 1) contacts the inner peripheral surface of the first cylinder 15a and rotates eccentrically. In the vicinity of the end point of the compression stroke in which the eccentric rotating first roller 19a approaches the first discharge port 22a, the space in which the gas refrigerant is compressed is narrowed, and the pressure in the space is increased. Further, when the first roller 19a contacts the inner peripheral surface of the first cylinder 15a and rotates eccentrically, the lubricating oil in the first cylinder 15a is carried into a narrow space.

The lubricating oil carried by the first roller 19a enters the escape portion 35 in the vicinity of the end point of the compression stroke. Accordingly, the narrow space in which the gas refrigerant in the first cylinder 15a is compressed is prevented from being excessively increased due to the liquid compression state.

Here, in this embodiment, the angle at which the escape portion 35 is formed is “θ3”, and the range of “θ3” is “θ3> θ2”, so the timing at which the lubricating oil enters the escape portion 35 is reached. Get faster. Therefore, even when the rotary compressor 4 is operated in the high-speed rotation range, the start of liquid compression is accelerated, so that the timing at which the lubricating oil enters the escape portion 35 is accelerated and the occurrence of liquid compression can be prevented. It is possible to prevent the rotary compressor 4 from being damaged by liquid compression.

As mentioned above, although embodiment of this invention was described, these embodiment is shown as an example and is not intending limiting the range of invention. These embodiments can be implemented in various other forms, and various omissions, replacements, and changes can be made without departing from the spirit of the invention. These embodiments and their modifications are included in the scope and gist of the invention, and are also included in the invention described in the claims and the equivalents thereof.

Industrial applicability

The present invention is used for a rotary compressor.

Claims

An electric motor part and a compression mechanism part connected to the electric motor part via a rotating shaft having an eccentric part are accommodated in a sealed case,
The compression mechanism portion supports the rotation shaft, a cylinder through which the rotation shaft passes, a roller that is fitted to the eccentric portion and moves eccentrically while bringing a part of the outer peripheral surface into contact with the inner peripheral surface of the cylinder, and A main bearing and a sub-bearing that close the end face of the cylinder and form a cylinder chamber in the cylinder;
The main bearing and the sub-bearing are provided with a discharge port for discharging the working fluid compressed in the cylinder chamber into the sealed case by the eccentric movement of the roller, and a discharge valve for opening and closing the discharge port. In the rotary compressor
The inner diameter area of the cylinder chamber is A (mm 2 ), the inner diameter dimension of the discharge port provided in the main bearing is dm (mm), and the inner diameter dimension of the discharge port provided in the auxiliary bearing is ds (mm). Is set so that at least one of formula (1) and formula (2) holds,
The volume of the cylinder chamber is V (mm 3 ), the spring constant of the discharge valve provided in the main bearing is Km (N / mm), and the spring constant of the discharge valve provided in the sub-bearing is Ks (N / mm), the rotary compressor is set so that at least one of formula (3) and formula (4) is established.
4.6 × 10 −3 ≦ dm / A ≦ 6.5 × 10 −3 (mm / mm 2 ) (1)
4.6 × 10 −3 ≦ ds / A ≦ 6.5 × 10 −3 (mm / mm 2 ) (2)
1.2 × 10 −4 ≦ Km / V ≦ 3.5 × 10 −4 (N / mm 4 ) (3)
1.2 × 10 −4 ≦ Ks / V ≦ 3.5 × 10 −4 (N / mm 4 ) (4)
However, when there are a plurality of cylinders, V is the volume of one cylinder chamber.
The discharge valve has a flexible arm portion with one end fixed to the main bearing or the sub-bearing, and a disc-shaped valve body portion provided on the other end side of the arm portion to close the discharge port; The valve body is set so that equation (5) is established when the outer dimension of the valve body is R (mm) and the width of the arm is W (mm). The rotary compressor according to claim 1.
R / W ≧ 2 ... (5)
A taper portion that gradually expands in the discharge direction is provided on the outlet side of the discharge port, and the minimum diameter dimension of the taper portion in the discharge port provided in the main bearing is dm1 (mm). When the diameter dimension is dm2 (mm), the minimum diameter dimension of the tapered portion in the discharge port provided in the auxiliary bearing is ds1 (mm), and the maximum diameter dimension of the tapered portion is ds2 (mm), The rotary compressor according to claim 1, wherein at least one of 6) and formula (7) is established.
1.1 ≦ dm2 / dm1 ≦ 1.35 (6)
1.1 ≦ ds2 / ds1 ≦ 1.35 (7)
A taper portion that gradually expands in the discharge direction is provided on the outlet side of the discharge port, the minimum diameter of the taper portion in the discharge port provided in the main bearing is dm1 (mm), and the maximum of the taper portion is When the diameter dimension is dm2 (mm), the minimum diameter dimension of the tapered portion in the discharge port provided in the auxiliary bearing is ds1 (mm), and the maximum diameter dimension of the tapered portion is ds2 (mm), The rotary compressor according to claim 2, wherein at least one of 6) and formula (7) is established.
1.1 ≦ dm2 / dm1 ≦ 1.35 (6)
1.1 ≦ ds2 / ds1 ≦ 1.35 (7)
A rotary compressor according to claim 1;
A condenser connected to the rotary compressor;
An expansion device connected to the condenser;
An evaporator connected between the expansion device and the rotary compressor;
A refrigeration cycle apparatus comprising:
A rotary compressor according to claim 2;
A condenser connected to the rotary compressor;
An expansion device connected to the condenser;
An evaporator connected between the expansion device and the rotary compressor;
A refrigeration cycle apparatus comprising:
A rotary compressor according to claim 3;
A condenser connected to the rotary compressor;
An expansion device connected to the condenser;
An evaporator connected between the expansion device and the rotary compressor;
A refrigeration cycle apparatus comprising:
A rotary compressor according to claim 4;
A condenser connected to the rotary compressor;
An expansion device connected to the condenser;
An evaporator connected between the expansion device and the rotary compressor;
A refrigeration cycle apparatus comprising: