WO2011027480A1

WO2011027480A1 - Scroll compressor, refrigerating cycle device, and heat pump water heater

Info

Publication number: WO2011027480A1
Application number: PCT/JP2009/070647
Authority: WO
Inventors: 雄幸野; 有吾向井; 和則津久井; 昌寛竹林; 敦大沼
Original assignee: 日立アプライアンス株式会社
Priority date: 2009-09-02
Filing date: 2009-12-10
Publication date: 2011-03-10
Also published as: KR20120040262A; CN102483060B; KR101410550B1; CN102483060A; JP5352386B2; JP2011052590A

Abstract

Insufficient oil sealing of a compression chamber leads to insufficient compression of refrigerant, and thereby leads to inefficiency. That is, oil sealing is an issue from the viewpoint of improvements in efficiency. To solve this, a back pressure control valve is disposed at a position where intermittent communication starts when both the volume of a suction chamber adjacent to the inner line of an orbiting scroll and that adjacent to the outer line of the orbiting scroll are increased. Moreover, oil is supplied to the compression chambers through a space formed at a position deeper than the tooth root of the fixed scroll for tooth tip lubrication.

Description

Scroll compressor, refrigeration cycle apparatus and heat pump water heater

The present invention relates to a scroll compressor or the like that compresses a refrigerant, and more particularly to a structure that improves the sealing performance by supplying lubricating oil to a compression chamber and reduces leakage loss.

Scroll compressors used in room air conditioners, heat pump water heaters, etc. control the back pressure, which is the pressure in the back pressure chamber provided on the opposite side of the orbiting scroll, by the back pressure control valve, and turn by the controlled back pressure. The scroll is urged toward the fixed scroll, and the refrigerant is compressed in a compression chamber formed by both scrolls. As a structure having a back pressure control valve, an intermittent communication structure is known in addition to a constant communication structure.

Lubricating oil is supplied to the compression chamber, improving the sealing performance of the compression chamber and reducing leakage loss. If the leakage loss can be kept as small as possible, the efficiency of the compressor can be increased accordingly. As a technique for reducing leakage loss, Patent Document 1, Patent Document 2, and the like are known.

The compressor disclosed in Patent Document 1 has a toothed oil supply structure, and a space (20) in which discharge pressure acts as an oil supply passage for supplying lubricating oil from an oil reservoir at the bottom of a sealed container, and an orbiting scroll. A passage (second communication passage) communicating with the tip of the lap is provided in the orbiting scroll. A pair of arc grooves that open to the compression chambers on both the inner and outer sides of the wrap are provided at the tip of the wrap of the orbiting scroll, and either one of the pair of arc grooves communicates with the second communication passage. Accordingly, it is disclosed that the lubrication of the sliding portion between the orbiting scroll and the fixed scroll can be maintained well even in the compression space where the pressure is higher than that of the suction chamber.

The compressor disclosed in Patent Document 2 is provided with a communication passage that communicates a high-pressure portion that is an inner region of a sliding partition ring and a compression chamber inside the orbiting scroll, and is provided on the compression chamber side of the opening of the communication passage. Is provided at the wrap tip of the orbiting scroll so as to face the discharge port at the center of the fixed scroll. Accordingly, it is disclosed that oil is supplied to a compression chamber that is relatively close to the end of compression, and seizure of the wrap tip of the fixed scroll and the end plate of the orbiting scroll is prevented. Further, it is disclosed that performance deterioration due to volumetric efficiency reduction due to suction heating is suppressed.

Further, in the conventional product, a back pressure control valve is disposed at approximately 11:00 when viewed in the direction of FIG. 3 of the present application (FIG. 22), from the back pressure chamber through the back pressure control valve. It was set as the structure which seals a compression chamber using the oil which flows in into the suction side.

JP 2009-062908 A JP 2009-052464 A

However, in Patent Document 1, when an arc groove is provided at the tip of the wrap, leakage occurs between the compression chambers on the inner and outer sides of the wrap, that is, between a swirling inner chamber and a swirling outer chamber described later. In order to prevent leakage, it is conceivable to increase the amount of oil supply to improve the sealing performance, but for this purpose, it is necessary to deepen the arc groove. When the arc groove is deepened, a contradictory phenomenon occurs such that leakage increases between the compression chambers.

Further, in Patent Document 2, it is not possible to expect an oil seal in a compression chamber on the outer diameter side far from the discharge port (region close to the suction portion), that is, a region near the start of compression.

If the oil seal in the compression chamber is not sufficient, the refrigerant cannot be sufficiently compressed, so the efficiency cannot be increased. In other words, from the viewpoint of improving efficiency, the issue is how to perform oil sealing.

In the conventional product, the R groove 5h is provided to lubricate the portion of the fixed scroll end plate surface having a large area and the end plate surface of the orbiting scroll. That is, the back pressure control valve is provided at the 11 o'clock position mainly from the viewpoint of lubrication. Therefore, the efficiency may be improved by devising the position of the back pressure control valve from the viewpoint of improving the sealing performance and efficiency of the compression chamber.

In view of the above problems, an object of the present invention is to provide a highly efficient scroll compressor. Another object of the present invention is to provide a highly efficient refrigeration cycle apparatus and heat pump water heater.

An object of the present invention is to provide a scroll compressor having an intermittent communication structure in which the orbiting scroll is urged to a fixed scroll by a back pressure controlled by a back pressure control valve, and the refrigerant is compressed in a compression chamber formed by both scrolls. Scroll compression in which the back pressure control valve is disposed at a position where intermittent communication is started when the volume of both the suction chamber on the inner line side of the orbiting scroll and the suction chamber on the outer line side of the orbiting scroll increases. Achieved by machine.

Another object of the present invention is to urge the orbiting scroll against the fixed scroll by the back pressure that is the pressure of the back pressure chamber provided on the opposite side of the orbiting scroll, and to supply the refrigerant in the compression chamber formed by both scrolls. This is achieved by a scroll compressor that performs tooth tip oiling into the compression chamber via a space provided at a position deeper than the tooth bottom of the fixed scroll.

Another object of the present invention is to control the back pressure, which is the pressure of the back pressure chamber provided on the side opposite to the orbiting scroll, by a back pressure control valve, and attach the orbiting scroll to the fixed scroll by the controlled back pressure. In a scroll compressor having an intermittent communication structure and a tooth tip oil supply structure that compresses refrigerant in a compression chamber formed by both scrolls, an inner suction side suction chamber and an outer suction side suction chamber of the orbiting scroll; When the volume of both increases, the back pressure control valve is disposed at a position where the communication start of intermittent communication is performed, and the space is provided through a space further deeper than the bottom of the fixed scroll. This is achieved by a scroll compressor that supplies the tip of the oil into the compression chamber.

According to the present invention, a highly efficient scroll compressor can be provided. In addition, a highly efficient refrigeration cycle apparatus and heat pump water heater can be provided.
Other objects, features and advantages of the present invention will become apparent from the following description of embodiments of the present invention with reference to the accompanying drawings.

The longitudinal cross-sectional view of a scroll compressor. Oiling structure. The figure with which the turning scroll and the fixed scroll are meshing. The figure showing a virtual turning extension room and a virtual turning outside room. The figure which shows the volume change of a suction chamber, and the communication area of the communication hole of a back pressure control valve. The figure which compared the graph of FIG. 5A with the graph of the 1st-order differentiation. The figure explaining the leak from a virtual turning extension room and a virtual turning outside room. The figure explaining the arrangement | positioning position of a back pressure control valve. The figure explaining the relationship between the amount of oil supply to suction space, and volumetric efficiency. Explanatory drawing (1) of tooth tip oiling. Explanatory drawing (2) of tooth tip oiling. Oil seal explanatory drawing of a compression chamber. The figure showing the other shape of a communicating hole. The figure which showed the pressure change at the time of starting. The figure showing mesh | engagement of a symmetrical wrap type scroll. A sectional view of a horizontal scroll compressor. The efficiency comparison of the compressor in this example and the 65 ° C. hot water storage condition of the prior art. The unit block diagram of a heat pump water heater. The figure explaining a hollow. The figure which connects a discharge pressure oil supply chamber and a turning outside line chamber. The figure explaining a shaft penetration type scroll type compressor. The figure explaining forced oiling. Position of the back pressure control valve of the conventional product.

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

The first embodiment will be described in detail below.

1 is a longitudinal sectional view of a scroll compressor, FIG. 2 is an oil supply structure, and FIG. 3 is a diagram in which a turning scroll and a fixed scroll are engaged with each other. Note that FIG. 2 is not an actual cross section, but a convenient cross section for explaining various configurations.

The basic configuration and operation of the scroll compressor 1 will be described. The scroll compressor 1 includes a compression mechanism unit 3, an electric motor 4 that drives the compression mechanism unit 3, an oil supply unit 50 for supplying lubricating oil to the compression mechanism unit 3, a compression mechanism unit 3, the electric motor 4, and an oil supply. A sealed container 2 for housing the portion 50.

The sealed container 2 is configured by welding a lid chamber 2b and a bottom chamber 2c vertically to a cylindrical case 2a. The lid chamber 2b is provided with a suction pipe 2d, and a discharge pipe 2e is provided on the side surface of the case 2a. The compression mechanism part 3 is arrange | positioned at the upper part in the airtight container 2, the electric motor 4 is arrange | positioned at the lower part, and the oil supply part 50 is arrange | positioned at the lower part. A lubricating oil 13 is stored at the bottom of the sealed container 2. The inside of the hermetic container 2 is a so-called high-pressure chamber type scroll compressor that serves as a discharge pressure chamber 2f.

The compression mechanism unit 3 includes a turning scroll 6 in which a spiral wrap 6a is erected on a base plate 6b, and a fixed scroll 5 in which a spiral wrap 5c is erected on a base plate 5d. A revolving scroll 6 is rotatably disposed opposite to the fixed scroll 5. An Oldham ring 12 is arranged between the lower surface side of the orbiting scroll 6 and the upper surface side of the frame 9, and each key formed on one surface and the other surface of the Oldham ring 12 is used for the orbiting scroll. 6 is fitted into a groove formed on the lower surface side of the frame 6 and a groove formed on the upper surface side of the frame 9 at a right angle.

The fixed scroll 5 is fixed to the frame 9 with bolts 8. The outer periphery of the frame 9 is fixed to the inner wall surface of the sealed container 2 by welding, so that the compression mechanism unit 3 is fixed to the sealed container 2. The frame 9 includes a main bearing 9a that rotatably supports the crankshaft 7. An eccentric portion 7 b of the crankshaft 7 is inserted on the lower surface side of the orbiting scroll 6. The orbiting scroll 6 is positioned between the fixed scroll 5 and the frame 9, and the orbiting scroll 6 is supported by the crankshaft 7.

The electric motor 4 has a stator 4a and a rotor 4b. The stator 4a is fixed to the sealed container 2 by press-fitting and / or welding. Further, the rotor 4b is fixed to the crankshaft 7 and is rotatably arranged in the stator 4a. The crankshaft 7 includes a main shaft 7 a and an eccentric portion 7 b and is supported by a main bearing 9 a and a lower bearing 17 provided on the frame 9. The eccentric portion 7 b is formed integrally with the main shaft 7 a of the crankshaft 7, and is fitted to a revolving bearing 6 c provided on the back surface of the orbiting scroll 6. The crankshaft 7 supports the orbiting scroll 6. .

The crankshaft 7 is driven by the electric motor 4, and the eccentric portion 7b is eccentrically rotated with respect to the main shaft 7a. The Oldham ring 12 transmits the eccentric rotation of the eccentric portion 7b of the crankshaft 7 without causing the orbiting scroll 6 to rotate, thereby causing the orbiting scroll 6 to revolve. Further, the crankshaft 7 is provided with an oil supply passage 7c that guides the lubricating oil 13 to the lower bearing 17, the main bearing 9a, and the slewing bearing 6c. The lubricating oil is provided on the lower side of FIG. An oil supply pipe 7d for sucking 13 and guiding it to the oil supply passage 7c is mounted. A mechanism for supplying lubricating oil to each part through the oil supply passage 7 c is an oil supply part 50.

As shown in FIG. 2, a back pressure chamber 14 is formed between the rear surface of the orbiting scroll 6 and the frame 9, that is, on the opposite side of the orbiting scroll 6. Lubricating oil to which discharge pressure, which is the pressure in the sealed container, is applied is introduced into the space between the rear surface of the orbiting scroll 6 and the upper end of the crankshaft 7 via the oil supply passage 7c. This space is referred to as a discharge pressure oil supply chamber 51. The discharge pressure oil supply chamber 51 is also formed on the side opposite to the wrapping scroll 6.

The lower part of the sealed container 2 in which the lubricating oil 13 is stored is the oil passage 7c → the discharge pressure oil chamber 51 → the gap between the slewing bearing 6c and the eccentric portion 7b → the back pressure chamber 14 → the back pressure control valve 16 → the path of the suction space 10 It communicates with. Further, the oil supply passage 7c → the hole 7z → the clearance between the main shaft 7a and the main bearing 9a → the notch 100 → the back pressure chamber 14 → the back pressure control valve 16 → the suction space 10 is communicated. The lubricating oil 13 tends to flow into the suction space 10 from the lower part of the sealed container 2 that is at the discharge pressure. When viewed from the back pressure chamber 14 at this time, the gap between the swivel bearing 6c and the eccentric portion 7b and the gap between the main shaft 7a and the main bearing 9a become a throttle on the oil inlet side, and the back on the oil outlet side. The pressure control valve 16 becomes a throttle, and the back pressure Pb, which is the pressure in the back pressure chamber 14, becomes an intermediate pressure between the suction pressure Ps and the discharge pressure Pd. Further, the lubricating oil 13 is supplied to the slewing bearing 6 c and the main bearing 9 a by a pressure difference between the discharge pressure in the compressor lower space and the back pressure in the back pressure chamber 14. This is a so-called differential pressure lubrication system.

When the orbiting scroll 6 revolves based on the rotation of the crankshaft 7 driven by the electric motor 4, the gas refrigerant is guided from the suction pipe 2d to the compression chamber 11 formed by the orbiting scroll 6 and the fixed scroll 5. The compressed gas refrigerant is discharged from the discharge port 5e provided substantially at the center of the base plate 5d of the fixed scroll 5 into the sealed container 2, that is, the discharge pressure chamber 2f, and flows out from the discharge pipe 2e to the outside. The refrigerant that has flowed out returns to the scroll compressor 1 via the suction pipe 2d via a first heat exchanger, an expansion device, and a second heat exchanger (not shown). A structure in which these are sequentially connected in a loop shape is referred to as a refrigeration cycle, and a device using this is referred to as a refrigeration cycle apparatus.

The fixed scroll 5 is provided with a release valve 15. The release valve 15 is for discharging from the compression chamber 11 to the discharge pressure chamber 2f when the pressure of the compression chamber 11 becomes equal to or higher than the pressure of the discharge pressure chamber 2f. For example, the release valve 15 works in the case of a liquid compression state or an overcompression state. A release valve hole 15a is provided between the release valve 15 and the compression chamber. It can be said that the release valve hole 15 a is a space provided at a deeper position than the tooth bottom of the fixed scroll 5.

In this embodiment, at least one release valve 15 is disposed in each compression chamber. This is because the compression chamber can be communicated with the release valve 15 in almost any crank angle compression chamber, so that the compression chamber does not become a completely sealed space and pressure can be released. Therefore, when the number of wraps is increased and the number of compression chambers is increased, the number of release valves 15 is preferably increased corresponding to the number of compression chambers.

Generally, the pressure in the compression chamber is expressed by the equation (1) and is determined by the ratio of the displacement volume and the compression chamber volume.

Pc = Ps · (V0 / Vc) ^γ (1)
Here, Pc is the compression chamber pressure, Ps is the suction pressure, V0 is the displacement volume, Vc is the compression chamber volume, and γ is the adiabatic index.

Depending on the operating pressure conditions, the pressure in the compression chamber may become higher than the pressure in the discharge pressure chamber 2f, and at this time, the gas refrigerant is discharged from the release valve 15. The release valve 15 located on the outer diameter side of the base plate does not open so much during steady operation because the pressure does not increase so much, but in order to avoid liquid compression when liquid refrigerant is sucked in such as immediately after startup. The implications provided are great.

FIG. 3 shows that the

scrolls

5 and 6 are cut at the end plate surface of the orbiting scroll 6 (the bottom surface of the orbiting scroll 6) or the end plate surface of the fixed scroll 5 (the tooth tip surface of the fixed scroll 5). The case where it sees toward the upper direction of FIG. 1, ie, FIG. 1, is represented. The lap of the orbiting scroll 6 is hatched. The center side is called the wrap winding start, and the outer diameter side is called the wrap winding end. In FIG. 3, the wrap is wound clockwise. It can be said that the lap is rewound counterclockwise.

The origin of the axis shown in FIG. 3 is the center of the sealed container 2. This coincides with the center of the base plate of the fixed scroll 5. The vertical axis is as follows, and the horizontal axis is perpendicular to the vertical axis and passes through the origin.

The vertical axis is based on the position of the winding end portion 6Xo of the outer line side wrap of the orbiting scroll 6 when the orbiting outer line chamber having the maximum volume is formed. The orbiting outer line chamber is a compression chamber on the outer diameter side of the wrap of the orbiting scroll 6. The turning outer line chamber having the largest volume is also the turning outer line chamber on the outermost diameter side (11a). A swirling outer line chamber is also formed on the inner diameter side from here, which is represented by reference numeral 11a '.

FIG. 3 is represented so that the point where the winding end portion 6Xo of the outer line side wrap of the orbiting scroll 6 contacts the fixed scroll 5 is on the vertical axis. The contact on the fixed scroll 5 side at this time is referred to as the winding end portion 5Xi of the extension-side wrap of the fixed scroll 5. Similarly, the winding end portions 5Xo of the outer line side wrap of the fixed scroll 5 are also defined, and these winding end portions 5X ride on the vertical axis in FIG. When the winding end portion 5Xo of the outer wrap of the fixed scroll 5 and the winding end portion 6Xi of the inner wrap of the orbiting scroll 6 are in contact with each other, a turning inner chamber, that is, a compression chamber on the inner diameter side of the wrap of the orbiting scroll 6 is formed. . The swivel extension chamber at this time is the swivel extension chamber having the largest volume, and is also the swivel extension chamber on the outermost diameter side. Although not shown in FIG. 3, depending on the rotation angle of the crankshaft 7, a swivel extension chamber is also formed on the inner diameter side. For example, it is represented as 11b 'in FIG. Further, although expressed as “contact”, more precisely, it means that the length of the virtual line AA connecting 6Xo-5Xi between the winding end portions and the virtual line BB connecting 6Xi-5Xo is minimized. Moreover, an inner line and an outer line mean the side surface of the tooth which is a spiral, ie, the wrap side surface.

The portion where the curve continues further in the clockwise direction than the winding end portions 5Xi and 5Xo of the fixed scroll 5 is called an extension portion. In FIG. 3, the winding end portions 5Xi and 5Xo are located at the 6 o'clock position indicated by the short hand of the watch, and the suction pipe 2d and the suction port 2d1 indicated by the broken line are located around the 7 o'clock position of the extension portion. Yes. Further, an R groove 5h is formed until about 11:00, and a back pressure control valve 16 is disposed at the 9 o'clock position. The conduction path 5i of the back pressure control valve 16 communicates with the R groove 5h. Since the contact area of the end plate surfaces of the

scrolls

5, 6 is larger in the second quadrant and the third quadrant than in the first quadrant and the fourth quadrant in FIG. 3, the R groove 5h is formed in such a portion. Is provided. This is a groove for introducing oil from the back pressure control valve 16 in order to lubricate the end plate surfaces of the

scrolls

5 and 6.

As shown in FIG. 3, a suction space 10 and a compression chamber 11 which are suction portions are formed between the wrap 5c and the wrap 6a. The suction space 10 refers to a region where the pressure becomes the suction pressure, and communicates with the suction pipe 2d. The compression chamber 11 is a region where communication with the suction pipe 2d is blocked, and is roughly classified into two types, a swirling outer chamber and a swirling inner chamber.

In general, there are four compression chamber boundaries: first the first boundary formed by the root of the fixed scroll, second the second boundary formed by the bottom of the orbiting scroll, and third the swirl It has four boundaries: a third boundary formed by the scroll inner line, and a fourth boundary formed by the outer line of the fixed scroll. For example, a compression chamber having such a boundary as a room indicated by 11b in FIG. 3 is referred to as a swivel extension chamber (or a fixed outer chamber). The first and second boundaries are the same as described above. Third, the third boundary formed by the outer line of the orbiting scroll, and fourth, the fourth boundary formed by the inner line of the fixed scroll, A compression chamber having four boundaries is referred to as a swirling outer chamber (or a fixed inner chamber), for example, a room indicated by 11a and 11a 'in FIG.

¡Lubricating oil is supplied between these boundaries to maintain the sealing performance. In any compression chamber, there is a minute gap (about 5 μm or less) between the wrap side surfaces, that is, between the third boundary and the fourth boundary. Among the minute gaps, in the gap closer to the discharge port 5e at the front end of the compression chamber, that is, closer to the winding start portion of the wrap, a compression chamber with higher pressure is formed further in front of the front end. Therefore, the gas refrigerant with higher pressure leaks from the minute gap between the third boundary and the fourth boundary. On the other hand, in the gap that is closer to the suction port 2d1 at the rear end of the compression chamber, that is, closer to the winding end of the wrap, a compression chamber having a lower pressure is formed further rearward of the rear end. Yes. Therefore, the gas refrigerant leaks from the minute gap between the third boundary and the fourth boundary into the compression chamber having the low pressure. It can be said that the leakage at the front end or the rear end is a leakage from the swirling extension chamber to the swiveling extension chamber or a leakage from the swirling outer chamber to the swirling outer chamber. This is referred to as type 1 leakage.

On the other hand, there is a finer gap (less than about 3 μm) between the tooth tip and the tooth bottom. When viewed from the compression chamber, both the swirl inner and outer line chambers have a minute amount between the first boundary and the third boundary, and between the second boundary and the fourth boundary. There is a gap. The compression chamber 11 is adjacent to a compression chamber having a higher pressure and a compression chamber having a lower pressure, with these gaps as a boundary. Naturally, the gas refrigerant leaks from the compression chamber having a high pressure, and the gas refrigerant leaks to the compression chamber having a low pressure. It can be said that the leak between the tooth tip and the tooth bottom is a leak from the swivel extension chamber to the swivel extension chamber, or a leak from the swivel extension chamber to the swivel extension chamber. This is referred to as a second type leak.

¡To reduce these leaks, supply oil to the compression chamber and fill the gap with this oil. Therefore, how to seal this part is important.

The winding end portion 6Xi of the inner side wrap of the orbiting scroll 6 moves so as to draw a counterclockwise locus as shown by a broken line with the position of FIG. 3 as the 6 o'clock position of the clock. The winding end portion 6Xo of the other outer line side wrap also draws a locus in the same manner, but is not shown. In FIG. 3, the swirling outer line chamber indicated by 11a is represented as a swirling outer line chamber having a crank angle of 0 °. Then, the turning outer line chamber indicated by 11a ′ can be expressed as a turning outer line chamber with a crank angle of 360 °. The volume of the swirling outer line chamber 11a with a crank angle of 0 ° is the largest of the volumes of the swirling outer line chamber. The turning extension chamber is formed when the crank angle is 180 °, and the volume of the turning extension chamber at that time is the largest of the volumes of the turning extension chamber (see FIG. 6B).

Thus, a compressor of a type in which the compression start timing of the turning inner and outer line chambers is shifted by 180 ° by the rotation angle of the crankshaft 7 is referred to as an asymmetric wrap type. The maximum volume swirl extension chamber is shown in FIG. 6B, but is not shown in FIG. In FIG. 3, the turning extension chamber indicated by 11b is a turning extension chamber whose crank angle is advanced by 180 ° from the maximum volume turning extension chamber, and becomes a turning extension chamber with a crank angle of 360 °. FIG. 3 shows a swirling outer chamber 11a with a crank angle of 0 °, a swirling outer chamber 11a ′ with a crank angle of 360 °, a swirling inner chamber 11b with a crank angle of 360 °, a swirling inner chamber (11b ′) with a crank angle of 720 °, A total of four compression chambers are shown. Since the swivel extension chamber (11b ') with a crank angle of 720 ° is open to the discharge port 5e, it cannot be strictly called a compression chamber, but is expressed in this way for easy understanding.

Next, the back pressure control valve 16 that is a mechanism for adjusting the back pressure Pb that is the pressure in the back pressure chamber 14 will be described. The orbiting scroll 6 is urged toward the fixed scroll 5 by the back pressure Pb. That is, the orbiting scroll 6 receives a force that is pressed against the fixed scroll 5 by the back pressure Pb. If the back pressure is large, the urging force increases, and the frictional force generated between the two scrolls also increases, which is not preferable. The back pressure control valve 16 is a valve that controls so that the back pressure does not become too large.

The fixed scroll 5 has a spring housing hole 5f. A through hole 5g is formed on the back pressure chamber 14 side of the spring housing hole 5f, and a piece 16a is press-fitted into the through hole 5g. The piece 16a is formed with a communication hole 16b that allows the spring housing hole 5f and the back pressure chamber 14 to communicate with each other. A valve body 16c is disposed in the spring housing hole 5f, and the valve body 16c is biased by a spring 16d so as to close the communication hole 16b. The spring 16d is attached to the seal member 16e, and the seal member 16e is press-fitted into the fixed scroll 5 so as to partition the spring housing hole 5f and the discharge pressure chamber 2f. On the side surface of the spring housing hole 5f, a conduction path 5i is formed which communicates with the R groove 5h formed in the extension of the end plate surface of the fixed scroll 5. Since the R groove 5h communicates with the suction pipe 2d, the pressure in the spring housing hole 5f eventually becomes the suction pressure Ps.

The operation of the back pressure control valve 16 will be described. The lubricating oil 13 stored in the lower part of the sealed container 2 is supplied to each bearing through the oil supply pipe 7d and the oil supply passage 7c due to the pressure difference between the pressure in the closed container 2 and the pressure in the back pressure chamber 14. The lubricating oil 13 reaching the end of the upper crankshaft 7, that is, the discharge pressure oil supply chamber 51, enters the back pressure chamber 14 through the throttle. There is also lubricating oil 13 that enters back pressure chamber 14 through hole 7z and notch 100. The refrigerant dissolved in the lubricating oil 13 is foamed in the back pressure chamber 14.

The upward force applied to the valve body 16c (the upward force in FIGS. 1 and 2) due to the pressure difference between the pressure in the back pressure chamber 14 and the pressure in the spring housing hole 5f, that is, the suction pressure Ps is the downward force by the spring 16d. When larger than that, the valve body 16c is opened, and the lubricating oil 13 in the back pressure chamber 14 is supplied to the suction space 10 through the conduction path 5i and the R groove 5h. This is because not only the gas refrigerant but also the oil passes through the back pressure control valve 16. This oil is considered to be oil that has adhered to the hole or the wall surface or oil that has become a mist. The pressure in the back pressure chamber 14 is adjusted by the spring force in this way, and becomes the suction pressure Ps + a constant value. This constant value is determined by the spring force of 16d, and the back pressure Pb can be increased as the amount of spring deflection when the valve is closed, which is the initial displacement, and as the spring constant k increases, that is, the spring is less likely to bend.

In a so-called EcoCute (registered trademark) heat pump water heater, carbon dioxide CO ₂ is used as a refrigerant in the refrigeration cycle apparatus. This refrigeration cycle is called a supercritical refrigeration cycle in which the pressure on the high pressure side exceeds the critical point of CO ₂ . This high pressure is produced by, for example, a scroll compressor as in this embodiment. A scroll compressor using CO ₂ as a refrigerant has an operating pressure of 3 to 5 times that of a conventional fluorocarbon refrigerant scroll compressor, and a differential pressure controlled by a back pressure control valve is also 3 to 5 times. When the back pressure determined by the spring force (= suction pressure Ps + constant value) is reached, the back pressure control valve opens. Under such a high pressure difference environment, the amount of gas refrigerant flowing from the back pressure chamber to the suction space and the amount of oil Since the amount is large, there is a difference between the back pressure when the back pressure control valve is opened and the back pressure after the back pressure control valve is opened. The back pressure after opening is lower than the back pressure at the moment of opening. Since the back pressure during the steady operation has an appropriate pressure in terms of efficiency, the design of the back pressure control valve is performed in accordance with the back pressure during the steady operation. For this reason, it is difficult to bend the spring so that the back pressure does not decrease excessively even if the back pressure control valve operates. Then, I was able to notice the problem that the back pressure control valve did not open even when necessary, such as at startup.

As a means for solving this problem, there is a so-called intermittent communication structure in which the revolving motion of the orbiting scroll 6 is used to close or communicate the communication hole 16b of the back pressure control valve 16 with its end plate surface. This is because the end surface of the base plate 6b of the orbiting scroll 6 intermittently passes through the communication hole 16b of the back pressure control valve 16 when the orbiting scroll 6 revolves with respect to the fixed scroll 5 based on the rotation of the crankshaft 7. In other words, the back pressure control valve 16 and the back pressure chamber 14 are intermittently communicated with each other.

The valve body 16c and the piece 16a are merely in metal contact with each other by a spring force, and a slight gap exists due to the surface roughness of the member and the leakage is not completely reduced to zero. While the base plate 6b closes the communication hole 16b, the gas refrigerant in the communication hole 16b leaks into the spring housing hole 5f through a minute gap between the valve body 16c and the piece 16a. At this time, since the back pressure control valve 16 is closed, the back pressure does not change. Now, the volume of the communication hole 16b is small. For example, if the diameter of the communication hole 16b is 2 mm, the volume is about 0.03 cm ^3. Depressurize. When the pressure in the communication hole 16b decreases, the gas refrigerant and oil in the back pressure chamber 14 flow into the communication hole 16b at the moment when the base plate 6b of the orbiting scroll 6 passes through the communication hole 16b and communicates with the back pressure chamber 14. Inertia force F expressed by the following equation acts on the valve body 16c.

F∝A ^* ρ ^* V ² (2)
Here, F is the inertial force, A is the cross-sectional area of the communication hole 16b, ρ is the fluid density, and V is the flow velocity.

As described above, the intermittent communication structure causes the inertial force of the fluid to act on the valve body 16c in addition to the pressure difference between the back pressure chamber 14 and the spring storage chamber 5f, thereby making it easy to open the back pressure control valve 16. In addition, since such an effect is recognized in the scroll compressor used in the supercritical refrigeration cycle, even if it is a scroll compressor for a fluorocarbon refrigerant operating at a lower pressure, the degree may be small, but the same effect It is thought that there is.

If the intermittent communication structure is adopted, the amount of oil distributed from the back pressure chamber 14 through the back pressure control valve 16 to the turning outer line chamber 11 a and the turning extension chamber 11 b changes depending on the position of the back pressure control valve 16. The oil distribution to the swirling outer chamber and the swirling inner chamber will be described with reference to FIGS. 4 is a diagram showing a virtual swirl extension chamber and a virtual swirl outer chamber, FIG. 5A is a diagram showing the volume change of the suction chamber and the communication section of the communication hole of the back pressure control valve, and FIG. 5B is a graph of FIG. It is the figure which compared with the graph of the order differential, arranges in the same size so that the mutual relationship is understood. FIG. 6 is a diagram for explaining leakage from the virtual swirl extension chamber and the virtual swirl extension chamber, and FIG. 7 is a diagram for explaining the position of the back pressure control valve.

FIG. 4 shows

virtual rooms

11A and 11B into which gas refrigerant and oil flow in the suction stroke. These regions are called suction chambers and are a part of the suction space 10. These

virtual rooms

11A and 11B are referred to as a virtual swirl outer line room 11A and a virtual swirl line room 11B. Therefore, the suction chamber is the virtual swirl outer chamber 11A or the virtual swirl inner chamber 11B. The virtual swirl outer chamber 11A and the virtual swirl chamber 11B are both at suction pressure. This is because the above-described virtual lines AA and BB are always in communication with the suction pipe 2d. Note that the virtual swirl extension chamber 11A and the virtual swirl extension chamber 11B are defined as follows.

The virtual orbiting outer line chamber 11A includes an imaginary line AA connecting the winding end portion 5Xi of the inner line side wrap of the fixed scroll 5 and the winding end portion 6Xo of the outer line side wrap of the orbiting scroll 6, and the outer line side wrap of the orbiting scroll 6 An area surrounded by the extension side wrap of the fixed scroll 5.

The virtual orbiting extension chamber 11B includes an imaginary line BB connecting the winding end portion 5Xo of the outer line side wrap of the fixed scroll 5 and the winding end portion 6Xi of the inner line side wrap of the orbiting scroll 6, and the inner line side wrap of the orbiting scroll 6 The area surrounded by the outer line side wrap of the fixed scroll 5.

Although the

rooms

11A and 11B are not partitioned, the spaces targeted by these rooms will be the swirling extension room 11a and the turning extension room 11b later. That is, the room 11A at the completion of the suction is the swirling outer chamber 11a having the maximum volume, and the room 11B at the completion of the suction is the swirling inner chamber 11b having the maximum volume. Thereafter, the volume is reduced as the crankshaft 7 rotates, and the gas refrigerant is compressed.

FIG. 5A shows a change in the volume of each suction chamber in each suction stroke with respect to the rotation angle of the crankshaft 7 and a communication section of the communication hole 16b of the back pressure control valve 16 at the back pressure control valve position θb. In the scroll compressor 1, the type of spiral is an involute curve, but it is known that an algebraic spiral also exhibits a similar volume change. Here, the volume change of the suction chamber is shown as a ratio with the suction completion time being 1. Accordingly, the peak of the suction volume ratio is larger in the virtual swirl extension chamber 11B than in the virtual swirl extension chamber 11A. The vertical axis intercept and the rotation angle of 0 ° are the start of suction of the virtual swirling outer line chamber 11A, and are the states shown in FIG. 3 and FIG. 6 (a).

The back pressure control valve position θb is shown as an angle with the minus side of the vertical axis in FIG. 7 being 0 °. It can be considered that it corresponds to the crank angle described above. In terms of the short hand of the watch, θb = 0 ° is the 6 o'clock direction, θb = 210 ° is the 11 o'clock direction, and θb = 270 ° is the 9 o'clock direction. The virtual orbiting outside line chamber 11A gradually increases in volume while the orbiting scroll 6 makes one rotation, exceeds the volume when it is fully closed (α ₁ ), reaches a peak on the way (α ₂ ), and then decreases. It disappears at a rotation angle of 360 ° (α ₃ ). Then, the next virtual swirling outer line chamber 11A defined as described above is newly formed and exhibits the same volume change.

As shown in FIG. 6B, the virtual swirl extension chamber 11B shows a volume change that is 180 ° shifted from the virtual swirl extension chamber 11A. In FIG. 3 and FIG. 6A, a virtual swirl extension chamber 11B having a certain volume is shown, which is about 60% of the maximum swirl extension chamber as shown in FIG. 5A. Thereafter, the volume of the virtual swirl extension chamber 11B gradually increases, exceeds the volume when fully closed (β ₁ ), reaches a peak in the middle (β ₂ ), then decreases, and disappears at a rotation angle of 180 °. (Β ₃ ). Then, the next virtual swivel extension chamber 11B defined as described above is newly formed and exhibits the same volume change.

However, if the space which is the target of the volume of the virtual swirling outer chamber 11A and the virtual swirling inner chamber 11B which are suction chambers is followed, the same space becomes a compression chamber which is a closed space from the suction chamber, and the designation is given. It changes with the turning outside line room 11a and the turning inside line room 11b. In other words, when the winding end portion that is an element of the imaginary line of each suction chamber is matched by the revolving motion of the orbiting scroll 6, strictly speaking, when the imaginary line becomes the minimum length, the suction stroke is completed, Since the virtual swirling outer chamber 11A and the virtual swirling inner chamber 11B, which are the target spaces of the suction chamber, are disconnected from the suction pipe 2d, the word “virtual” is taken and the sign is changed. The compression chambers are closed spaces, and the names of the swirling outer chamber 11a and the swirling inner chamber 11b are changed.

If only the volume changes in the virtual swirl outer chamber 11A and the virtual swirl chamber 11B are observed, the same volume change is repeated every 180 °. For example, when the back pressure control valve 16 is provided at the position θb = 30 °, that is, at the 5 o'clock position, the communication section of the communication hole 16b is in the range of 130 ° to 290 ° in rotation angle. At the back pressure control valve position, the volume of the virtual turning extension chamber 11B begins to decrease (β ₂ to β ₃ ) as the crankshaft 7 rotates in the early stage of the communication hole 16b. When the back pressure control valve 16 is provided at the position θb = 210 ° shifted by 180 °, that is, at the 11 o'clock position, the communication section of the communication hole 16b is in the range of 310 ° to 470 ° in rotation angle. At the back pressure control valve position, the volume of the virtual turning outer line chamber 11A starts to decrease with the rotation of the crankshaft 7 at the beginning of the communication hole 16b (α ₂ to α ₃ ).

Further, the communication section of the communication hole 16b, which is a communication section for intermittent communication, is a portion between two oblique broken lines shown in FIG. 5A. For the purpose of explanation, the characteristic installation position of the back pressure control valve 16 is indicated by a white arrow. Of course, no matter where the back pressure control valve 16 is disposed, the angle range in which the back pressure control valve 16 and the back pressure chamber 14 communicate with each other by intermittent communication is the same at 160 °.

The oil that has passed through the back pressure control valve 16 from the back pressure chamber 14 passes through the R groove 5h, and is supplied to the virtual swirl outer chamber 11A and the virtual swirl chamber 11B, and then the swirl outer chamber 11a and the swirl chamber 11b. And is used to seal the compression chamber. However, in the section where the volume is reduced (α ₂ to α ₃ , β ₂ to β ₃ ), the gas refrigerant escapes from the suction chamber and flows back in the direction of the suction port 2b. It will be done against this. That is, it is difficult to refuel in this section, and refueling efficiency is poor. It can be said that refueling is hindered. This interval can be seen in FIG. 5B. The section in which the volume is reduced is a portion where the first derivative graph of FIG. 5B (b) becomes negative.

That is, when θb = 30 °, it is difficult to lubricate the virtual swirl extension chamber 11B, and when θb = 210 °, it is difficult to lubricate the virtual swirl outer chamber 11A. In particular, a large amount of oil passing from the back pressure chamber 14 through the back pressure control valve 16 may come out at the moment when the back pressure control valve 16 is opened. Cannot be supplied. Accordingly, the oil distribution in the swirling extension chamber 11a and the swirling extension chamber 11b is biased to either one. In the compression chamber in which less oil is taken in, the sealing performance of the compression chamber is lowered and leakage loss may occur.

Here, in the conventional product, the back pressure control valve 16 is provided at approximately the 11:00 position, that is, at the position of θb≈210 °, so that oil can be supplied to the end of the R groove 5h. This is because the end plate surfaces of the

scrolls

5 and 6 can be lubricated over a wide range.

However, as described above, since the back pressure control valve 16 is provided at the 11 o'clock position, it is considered that it is difficult to supply oil to the virtual turning outer line chamber 11A. Generally, extra oil is supplied, and the excess oil is discharged from the discharge port 5e to the discharge pressure chamber 2f. Previously, it was thought that both the virtual swirl outer chamber 11A and the virtual swirl chamber 11B were well balanced and could supply more than the appropriate amount of oil from the viewpoint of volumetric efficiency. I thought that it was, but when I examine it individually, there is an imbalance in the amount of oil supply with respect to the sealing performance of the compression chamber 11, and in some cases the amount is insufficient to the appropriate amount from the viewpoint of volume efficiency. It is thought that the oil was supplied to a certain extent, and some of the oil seemed to be in excess or very excessively. Even if the back pressure control valve 16 is disposed at the 10 o'clock position (α ₂ to α ₃ ), the imbalance is the same. Of course, this unbalance is caused by intermittent communication, and no imbalance will occur if it is always connected as described later.

On the other hand, for example, if the back pressure control valve position θb is set to a position of 270 °, that is, the 9 o'clock position, the bias can be eliminated. At θb = 270 °, both the virtual swirl outer chamber 11A and the virtual swirl chamber 11B have no portion where the suction chamber volume is reduced and there is no backflow (α ₄ ). 11A, it is easy to refuel the virtual turning extension chamber 11B. That is, if the back pressure control valve is provided at the position where the intermittent communication start is performed when the volumes of both the virtual swirl outer chamber 11A and the virtual swirl chamber 11B are not reduced, the crankshaft 7 rotates. The volume of each suction chamber increases, and the suction chamber takes in oil by itself. This is considered to be a phenomenon similar to that of natural aspiration in an automobile reciprocating engine. As described above, the back pressure control valve position θb where the suction chamber volume does not decrease (β ₃ to α ₂ , α ₃ to β ₂ ) overlaps with the communication opening stage is 90 ° to 210 °, 270 ° to It is a position of 390 °, which is an 11 o'clock to 3 o'clock position, and an o'clock to 9 o'clock position.

When considering the comparison with gasoline injection, it takes time for the oil injected from the back pressure control valve 16 to reach the suction chamber. This time may be an extremely short time and may not necessarily be as follows, but it is considered that the crankshaft 7 is rotating and the volume is changing during the arrival time. It is more desirable to inject oil from a position where the volume does not decrease when reaching the virtual swirl inner /

outer line chambers

11A, 11B. That is, it seems preferable to avoid injection immediately before the bottom dead center of the piston. Accordingly, preferable ranges from such points are 90 ° to 150 °, 270 ° to 330 °, 1 o'clock to 3 o'clock, and 7 o'clock to 9 o'clock. This is a point-symmetrical position as can be seen from FIG. This is because the volume change of the suction chamber is repeated every 180 °, and since it is an asymmetric wrap type scroll compressor, such a position is preferable. However, at the 1 o'clock to 3 o'clock position, the R groove 5h must be extended from 11 o'clock to 12 o'clock → 1 o'clock to 3 o'clock, so it is necessary to change the outer diameter or change the inner spiral. End up.

In this embodiment, since the wrap is wound clockwise, there is no bottom of the fixed scroll 5 at the position of 270 ° to 330 °, that is, from 7 o'clock to 9 o'clock, and the back pressure control valve 16 is provided. There is enough space. Therefore, it is more preferable to arrange the back pressure control valve 16 at this position. Although it seems that there is almost no volume of the virtual swirling outer chamber 11A around the 9 o'clock position (α ₄ ), when compared with the gasoline injection, oil reaches the virtual swirling outer chamber 11A. It seems that sometimes it enters the area where the volume increases. In this respect, at the 1 o'clock to 3 o'clock position, the distance from the back pressure control valve 16 to the suction chamber becomes long, and it is difficult to adjust the timing of oil injection from the back pressure control valve 16.

Accordingly, it is more preferable to dispose the back pressure control valve 16 at the 7 to 9 o'clock position from the point of injection, and the closer to the 7 o'clock position, the shorter the distance between the back pressure control valve 16 and the suction chamber. As a result, the time for the oil to reach the suction chamber can be shortened. Compared with setting the back pressure control valve 16 to the 1 to 3 o'clock position, the volume change of the suction chamber in this short time is small, so the rotation angle from the start of intermittent communication until the oil reaches the suction chamber And the rotation angle at which the volume of the suction chamber changes can be examined with little time delay. If the back pressure control valve 16 is disposed at the 7 to 9 o'clock position, the oil in the R groove 5h from that position to the 11 o'clock position is likely to stay. However, in reality, the supplied high-pressure oil leaks through a gap between the end plate surfaces, so that the oil can be circulated. Here, the high pressure is the back pressure Pb with respect to the suction pressure Ps.

Further, considering a more preferable position, a space is required to dispose the suction pipe 2d, so that the workability and assemblability are improved by removing that portion. In the present embodiment, as shown in FIG. 7, it is considered that there is no margin at the position of 7:30, and therefore, a position of approximately 7:40 to 9 o'clock, that is, a position of 270 ° to 310 ° is more preferable.

Furthermore, it is considered that the range (α ₅ to β ₂ ) in which the volume of both the suction chambers increases is more preferable. Within this range, the back pressure control valve 16 and the back pressure chamber 14 communicate with each other when the volumes of both the suction chambers increase, and the injection starts while there is a flow sucked into the suction chamber. Because it can. However, if the volume of the suction chamber exceeds 1, the volume decreases and eventually the ratio becomes 1. That is, the virtual swirl inner / outer line chamber is larger than the volume when it becomes a compression chamber, which is a closed space, and is finally closed by reducing its volume. Considering the time for the oil to reach the suction chamber by injection, both the back pressure control valve 16 and the back pressure chamber 14 are within a range (α ₅ to β ₁ ) in which the volume of both the suction chambers increases until the volume ratio becomes 1. It is preferable to communicate. A position of approximately 7 o'clock to 8:30 o'clock, that is, a position of 285 ° to 330 ° is more preferable.

In other words, the volume of the space that is the target of the volume of the virtual swirl inner / outer line chamber that is the suction chamber increases to the volume when it becomes the respective compression chamber and swirl inner / outer line chamber that are closed spaces. It is preferable to arrange the back pressure control valve at a position where the start of intermittent communication is performed.

Therefore, the range considered as the best position is a position of 285 ° to 310 °, that is, a position of 7:40 to 8:30.

As described above, by supplying oil in a well-balanced manner to both the turning inner line side and the turning outer line side before forming the compression chamber, both can maintain the sealing performance from the start of the compression stroke and improve the efficiency. .

The lubricating oil 13 stored in the lower part of the sealed container 2 passes through the oil supply pipe 7d and the oil supply passage 7c due to the pressure difference between the pressure in the closed container 2 and the pressure in the back pressure chamber 14. However, the amount of oil supply is closely related to the volumetric efficiency. Here, the bearing oil supply amount will be described. The bearing oil supply amount is the amount of oil flowing into the back pressure chamber 14, and the amount of oil flowing into the back pressure chamber 14 through the gap between the slewing bearing 6c and the eccentric portion 7b, and between the main shaft 7a and the main bearing 9a. It is the sum total of the amount flowing into the back pressure chamber 14 through the gap. That is, the bearing oil supply amount is an amount of oil mainly for lubricating the bearing.

This oil is supplied from the back pressure control valve 16 to the suction space 10. Basically, the amount of oil supplied to the suction space 10 may be considered to be the same as the amount of oil supplied to the bearing. However, the amount of oil supplied from the back pressure control valve 16 to the suction space 10 is different from the amount of oil actually supplied into the compression chamber 11. A part of the amount of oil supplied to the suction space 10 is used for lubricating the end plate surface, and a part is taken into the suction chamber and supplied into the compression chamber 11.

If the amount of oil supplied to the suction space 10 is small, the amount of oil supplied to the compression chamber 11 is reduced, and the oil cannot be sealed, resulting in increased leakage loss and reduced volume efficiency. However, even if the amount of oil supplied to the suction space 10 is too large, the volume efficiency is lowered. The reason is as follows. Since the oil supplied to the suction space 10 through the back pressure control valve 16 has a temperature higher than that of the suction gas, the suction gas is heated. As a result, the gas density of the suction gas is lowered, and the refrigerant circulation amount of the gas refrigerant flowing into the suction chamber and then the compression chamber 11 is reduced. Therefore, the volumetric efficiency decreases from the later-described equation (3). This is called heating loss of the suction gas.

That is, the amount of oil supplied to the suction space 10 has an appropriate range from the viewpoint of volume efficiency. FIG. 8 schematically shows the relationship between the amount of oil supplied to the suction space and the volume efficiency. Here, the range in which the volumetric efficiency is a certain value or more is appropriate. However, if the amount of oil supplied to the suction space 10 is made appropriate from the viewpoint of volume efficiency, the required amount of bearing oil cannot be provided. When the amount of bearing oil supply is too small, problems such as seizure, which are more serious than a decrease in volumetric efficiency, may occur. Therefore, generally, extra oil is supplied, and excess oil is discharged from the discharge port 5e to the discharge pressure chamber 2f. As described above, there is a required amount of oil supply from the viewpoint of the reliability of the bearing and the sliding portion, and the amount of oil supply required for the entire scroll compressor must be an appropriate amount. If it becomes so, the amount of oil supply to the suction space 10 seen from volume efficiency will be more than the appropriate amount of oil supply, and will become excessive. That is, the volumetric efficiency is reduced due to the heating loss of the suction gas.

In the present embodiment described below, the heating loss of the suction gas can be reduced, which will be described below with reference to FIGS. 9 to 13 and FIG. 9 is an explanatory diagram (1) of the tooth tip lubrication, FIG. 10 is an explanatory diagram (2) of the tooth tip lubrication, FIG. 11 is an explanatory diagram of the oil seal of the compression chamber, and FIG. FIG. 13 is a view showing a pressure change at the time of activation.

As shown in FIG. 9, oil is supplied from the back pressure chamber 14 to the compression chamber 11 (the swirling outer line chamber 11a) through the communication hole 18 and the release valve hole 15a1. A structure in which oil is supplied from the tooth tip of the orbiting scroll 6 in this way is referred to as a tooth tip oil supply structure. In FIG. 9, in addition to the tooth tip oil supply structure, oil is supplied to the compression chamber 11 (the swirling outer line chamber 11a) using a release valve hole 15a1 which is a space provided deeper than the tooth bottom of the fixed scroll 5. Supply.

The orbiting scroll 6 has a communication hole 18 in the wrap, and the first opening is provided at a tooth tip that is an end surface of the wrap, and the first opening with respect to the base plate of the orbiting scroll 6 is on the back side. That is, the second opening is provided on the side opposite to the wrap. The first opening is referred to as a wrap tip side opening or a tooth tip opening, and the second opening is referred to as an anti-wrap side opening. The tooth tip opening communicates with the release valve hole 15 a 1, and the anti-wrap side opening communicates with the back pressure chamber 14 formed on the anti-wrap side of the orbiting scroll 6, where the pressure is a back pressure space.

FIG. 10 shows a state in which the communication hole 18 communicates with the release valve hole 15a1. By the revolving motion of the orbiting scroll, the back pressure chamber 14 and the orbiting outer line chamber 11a communicate with each other through the tooth tip opening of the communication hole 18 and the release valve hole 15a1. This release valve hole 15a1 is a hole corresponding to the release valve 15 located on the outer diameter side of the base plate shown in FIG. 3, and is for the swirling outer line chamber 11a formed on the outermost diameter side.

FIG. 10A is the same as the positional relationship between the

scrolls

5 and 6 shown in FIG. 3, and the angle of the crankshaft 7 is 0 °. Here, the communication hole 18 is not in communication with the release valve hole 15 a 1, and the lap tip side opening of the communication hole 18 is blocked by the wrap bottom surface of the fixed scroll 5.

FIG. 10B shows a case where the angle of the crankshaft 7 is about 80 °. Here, the communication hole 18 communicates with the release valve hole 15a1. As shown in FIG. 9, at this timing, the back pressure chamber 14 and the swirling outer line chamber 11a communicate with each other via the communication hole 18 and the release valve hole 15a1, so that the lubricating oil 13 from the back pressure chamber 14 is swirled. 11a. The back pressure chamber 14 and the swirling outer line chamber 11a communicate with each other through the communication hole 18 and the release valve hole 15a1 in this way when the angle of the crankshaft 7 is in the range of about 45 ° to about 90 °. It can be said that they communicate intermittently.

FIG. 10C shows the case where the angle of the crankshaft 7 is about 120 °. Here, the communication hole 18 is not in communication with the release valve hole 15 a 1, and the tooth tip opening of the communication hole 18 is again blocked by the bottom surface of the fixed scroll 5.

FIG. 11 shows a schematic diagram of a state in which the gap in the compression chamber is sealed with lubricating oil. The compression chamber pressure is P1 <P2 <P3. The lubricating oil 13 supplied to the compression chamber 11 adheres to the wrap wall surface, seals between the tooth tip and the tooth bottom, and suppresses the second type of leakage. Although not shown in this figure, naturally, the oil that has entered the compression chamber 11 seals the gap between the wraps and suppresses the first type leakage.

In FIG. 11, since there is a pressure difference between adjacent compression chambers, the lubricating oil 13 flows into the gap 191 and the gap 192 due to this pressure difference. If the lubricating oil 13 supplied to the compression chamber 11 is small, the

gaps

191 and 192 are not filled with the lubricating oil 13 and the seal is broken. As a result, the sealing performance cannot be maintained, so that the gas refrigerant is blown out, the leakage loss is increased, and the efficiency is lowered.

10B, when the pressure in the back pressure chamber 14 is higher than the pressure in the swirling outer line chamber 11a, the lubricating oil 13 flows from the back pressure chamber 14 into the swirling outer line chamber 11a. At this time, not only the lubricating oil 13 but also the gas refrigerant flows into the swirling outer line chamber 11a. During this time, the crankshaft 7 rotates about 45 °, so that the pressure in the orbiting outer chamber 11a rises. This rise is the rise from the lower broken line to the upper broken line shown in FIG. Inflow of oil or gas refrigerant also contributes to the increase in pressure, but the volume change of the swirling outer line chamber 11a is the main factor.

When the communication hole 18 is blocked by the bottom surface of the fixed scroll 5, the minimum seal length of the gap 192 is half the value obtained by subtracting the diameter of the communication hole 18 from the thickness t of the wrap. Even if the seal length at other portions is sufficiently maintained, if the minimum seal length is short and insufficient, an unfavorable situation as described above may occur.

Therefore, the minimum seal length has a lower limit value from the viewpoint of the strength of the wrap, the viewpoint of the sealing performance, and the viewpoint of the supply amount of the lubricating oil 13. From the viewpoint of supply amount, it is desirable that the diameter of the communication hole 18 is as large as possible. However, in terms of sealing performance and strength, it is desirable to reduce the diameter of the communication hole 18 and increase the seal length as much as possible.

Suppose that the thickness of the wrap is 3.0 mm, and if the minimum dimension of the wall thickness is secured from the viewpoint of strength, 0.5 mm, the diameter of the communication hole 18 is 2.0 mm at the maximum. Further, the value determined from the size of the tool is the minimum dimension of the communication hole 18, which is, for example, 0.6 mm. Accordingly, the diameter of the communication hole 18 is, for example, about 0.6 mm to 2.0 mm (1/5 · t to 2/3 · t). When the supply amount of the lubricating oil 13 is insufficient, such as when the diameter of the communication hole 18 is 0.6 mm or less, a further device as shown in FIG. The ratio of the wrap thickness t and the minimum seal length at these times is 1/6 · t to 2/5 · t, that is, the minimum seal length is 17% or more and 40% or less of the wrap thickness t. That would be desirable. However, this is a case where the center of the communication hole 18 having a circular cross section is aligned with the center line of the lap. As described above, the actual minimum seal length should be expressed by the length, regardless of the ratio to the tooth thickness. When considering this scroll compressor, the tooth thickness is at most within a range of about 1.5 to 6.0 mm, which is half the maximum, so there is no particular inconvenience in expressing the minimum seal length by the ratio as described above.

Since the diameter of the release valve hole 15a is 1.8 mm, such a minimum seal length allows the release valve hole 15a1 to straddle the communication hole 18, as shown in FIG. Therefore, oil can be supplied from the back pressure chamber 14 to the swirling outer line chamber 11a through the communication hole 18 and the release valve hole 15a1.

Besides, as shown in FIG. 12, the communication hole 18 may be formed in an oval shape. FIG. 12 shows a movement locus of the communication hole 18 when the orbiting scroll 6 revolves. By making the communication hole 18 into an oval shape or the like, a section where the release valve hole 15a1 and the communication hole 18 communicate with each other can be lengthened. Further, the area where the communication hole 18 opens into the release valve hole 15a1 can be increased while keeping the minimum seal length large, and there is an advantage that the amount of oil supply from the back pressure chamber 14 to the swirling outer line chamber 11a can be increased.

On the contrary, when the pressure in the swirling outer line chamber 11a is higher than the pressure in the back pressure chamber 14, the flow of the lubricating oil 13 opposite to that described with reference to FIG. If a reverse flow occurs, a large amount of gas in the swirling outer line chamber 11a flows into the back pressure chamber 14 and excessively increases the pressure in the back pressure chamber 14. However, since the back pressure control valve 16 opens, the back pressure Pb settles to a predetermined value. In this case, even if it is compressed halfway, the energy required for it is wasted and the efficiency is reduced. Therefore, it is necessary to take a sufficient seal length in order to maintain the sealing performance. Here again, as described above, it is desirable that the minimum seal length is 17% to 40% of the wrap thickness t.

Fig. 13 shows the pressure change after starting the scroll compressor. Three lines of suction pressure Ps, back pressure Pb, and discharge pressure Pd are experimental results. The portion indicated by Pcom is the pressure in the swirling outer line chamber 11a in the section where the communication hole 18 and the release valve hole 15a1 shown in FIG. 10 communicate with each other. When the compressor is started, the differential pressure between the discharge pressure Pd and the suction pressure Ps increases.

The virtual swirling outer chamber 11A changes to the swirling outer chamber 11a at the moment when it is closed as it follows the same space. The pressure in the turning outer line chamber 11a at that moment is Ps. Thereafter, as the crankshaft 7 rotates, the pressure in the turning outer line chamber 11a increases. The movement of both scrolls when going up is shown in FIG.

13 represents a pressure Pcom in the swirling outer line chamber 11a when the communication hole 18 and the release valve hole 15a1 communicate with each other. That is, the pressure in the swirling outer line chamber 11a described with reference to FIG. However, the pressure after n rotations is represented continuously, not following the closed one-turn outer chamber 11a.

Before starting, the suction pressure Ps, back pressure Pb, and discharge pressure chamber pressure P _2f are of course the same and there is no difference. _Therefore , after the start-up, the pressure Pc in the compression chamber immediately becomes the pressure P _2f in the discharge pressure chamber. As a result, the release valve 15 opens. As shown in the equation (1), the compression chamber pressure Pc is determined by raising the ratio of the displacement volume V0 and the compression chamber volume Vc to a constant value, so that the discharge pressure Pd and the suction pressure Ps are immediately after starting. when the ratio Pd / Ps is low, the pressure Pc of the compression chamber 11 is because would reach the pressure _{P 2f} of immediately discharge chamber 2f. Then the pressure in the swirling external chamber 11a is the same as the pressure _{P 2f} discharge chamber, Pcom in FIG. 13 will be represented as the discharge pressure Pd. This is region A. The pressure _P2f in the discharge pressure chamber has the same meaning as the discharge pressure Pd.

Accordingly, the ratio Pd / Ps is low in the A section immediately after the activation, Pcom becomes P _2f or more, and the release valve 15 opens. When Pcom is higher than the back pressure Pb immediately after the start, the back refrigerant chamber 14 and the swirling outer line chamber 11a communicate with each other through the communication hole 18 and the release valve hole 15a1, whereby the gas refrigerant in the swirling outer line chamber 11a. Flows back into the back pressure chamber 14 and tries to increase the pressure in the back pressure chamber 14. Since the difference between the back pressure Pb and the suction pressure Ps at this time is small, the urging force by the spring 16d of the back pressure control valve 16 cannot be overcome yet, and the back pressure control valve 16 does not open. Therefore, the back pressure Pb is increased by the back flow, and the orbiting scroll 6 is reliably lifted when the scroll compressor 1 is started, so that the gap between the tooth tip and the tooth bottom can be reduced. That is, the efficiency at startup can be improved.

When a certain amount of time has elapsed after startup, the difference between the suction pressure Ps and the discharge pressure Pd increases. This is region B. In the initial stage of the region B, the release valve 15 is opened, and Pcom becomes the discharge pressure Pd. Up to this point, since Pcom is higher than the back pressure Pb, the gas refrigerant flows from the tooth tip opening to the non-wrap side opening and raises the back pressure Pb.

After that, Pcom decreases with the suction pressure Ps. In the formula (1), when Vc is the volume of the swirling outer line chamber 11a when the communication hole 18 and the release valve hole 15a1 are in communication and Pc is Pcom, Pcom decreases as the suction pressure Ps decreases. I understand that. The back pressure Pb is controlled by the spring force of the back pressure control valve 16 so as to be a suction pressure Ps + a constant value. However, unless the valve is opened by overcoming the spring force, the behavior of Pb> Ps as shown in FIG. Show. When the spring force is overcome and the valve is opened, the back pressure Pb suction pressure Ps + a constant value. As a result, Pcom at the beginning of the B section is higher than the back pressure Pb, but thereafter becomes lower than the back pressure Pb.

When Pcom becomes lower than the back pressure Pb due to a steady operation or the like, that is, when the back pressure Pb becomes higher than Pcom, the lubrication of the back pressure chamber 14 is performed so that the gas refrigerant flows from the counter lap side opening to the tooth tip opening. Oil 13 is supplied to the swirling outer line chamber 11a through the communication hole 18 and the release valve hole 15a1. As described above, in the conventional product, since the back pressure control valve 16 is provided at the 11 o'clock position, it is considered that it was difficult to supply oil to the virtual swirling outer line chamber 11A. Therefore, even after the formation of the swirling outer line chamber 11a, it is possible to increase the sealing performance of the swirling outer line chamber 11a and improve the efficiency of the compressor by refueling the swirling outer line chamber 11a, which is a closed space. it can. Further, this tooth tip oiling acts in a direction in which the above-mentioned unbalance is eliminated.

Further, the compression chamber is closed without supplying a part of the lubricating oil 13 supplied to the back pressure chamber 14 through the back pressure control valve 16 from the suction space 10 to the suction chamber via the back pressure control valve 16. Since it is directly supplied to the swirling outer line chamber 11a, the heating efficiency of the suction gas can be reduced and the volume efficiency can be improved. Details are as follows.

When oil is supplied to the virtual swirl outer chamber 11A and the virtual swirl chamber 11B, which are suction chambers, via the back pressure control valve 16 and the suction space 10, the suction gas is heated and the density of the gas refrigerant is reduced, so that the amount of refrigerant circulation is reduced. The volumetric efficiency is lowered. However, if the oil supplied to the suction space 10 is reduced, it is possible to suppress a decrease in the circulation rate of the refrigerant and to suppress a decrease in volume efficiency. The reduced amount of oil is supplied from the tooth tip to the compression chamber 11 through the communication hole 18. Since the compression chamber 11 is a closed space, the refrigerant circulation amount does not change and the volumetric efficiency does not decrease.

That is, the amount of oil supply from the indirect oil supply path for supplying oil indirectly to the compression chamber 11 via the suction space 10 to the direct oil supply path for supplying oil directly to the compression chamber 11 without passing through the suction space 10. Even if a part of the suction gas is moved, the heat loss of the suction gas does not move from the indirect oil supply path to the direct oil supply path, so the heat loss of the suction gas is reduced as much as the amount of oil supply to the suction space 10 is reduced. be able to. Therefore, the amount of oil supplied to the compression chamber 11 from the back pressure chamber 14 via the back pressure control valve 16 and the suction space 10 is reduced from the viewpoint of volume efficiency, and the volume efficiency is reduced by the reduced amount. The volume efficiency can be improved as a whole. This corresponds to shifting to the left in the “excess” range of FIG. 8, that is, approaching the “appropriate” range. This is because if it falls within the “appropriate” range, the bearing oil supply amount will be insufficient.

For example, even if the amount of oil supply to the virtual swirl inner / outer line chamber is already balanced, such as by always communicating or arranging the back pressure control valve 16 at the 7-9 o'clock position, An efficiency improvement effect can be obtained. The constant communication structure is a configuration in which the back pressure chamber 14 and the back pressure control valve 16 are always in communication, and is used when the pressure in the back pressure chamber 14 is kept relatively small. That is, the back pressure control valve 16 is easy to open without intermittent communication. Mainly used in the above-mentioned CFC refrigerant scroll compressor. Further, even in a scroll compressor using CO ₂ as a refrigerant used in a so-called Ecocute (registered trademark) heat pump water heater, a discharge pressure out of the projected area in the vertical axis direction on the side opposite to the wrapping scroll 6 is shown. If the area in which the above acts is larger than that of the present embodiment, the back pressure can be reduced, and the continuous communication structure can be adopted.

As described above, immediately after startup, the pressure scroll 14 functions to increase the pressure of the back pressure chamber 14 to positively urge the orbiting scroll 6 to the fixed scroll 5. ) To improve the sealing performance in the compression chamber and improve the efficiency of the compressor.

Here, the method of using the lubricating oil 13 for the gap seal differs depending on whether the gap in the compression chamber 11 is adjacent to a high-pressure room or a low-pressure room.

First, it will be specifically described that the first type leakage is suppressed.

The compression chamber formed as a certain chamber has a lower pressure than the compression chamber formed 360 ° earlier in crank angle than the compression chamber. Accordingly, the oil from the room preceding the 360 ° leaks from the gap at the front end of the compression chamber. In addition, oil leaks from the gap at the rear end of the compression chamber into the room that is followed by 360 °.

The swirling outer line chamber 11a 'shown in FIGS. 6 (a) and 6 (b) starts to compress 360 degrees ahead of the swirling outer line chamber 11a in terms of crank angle, so that the swirling outer line chamber 11a' and the swirling outer line chamber 11a Comparing the pressure, the pressure in the swirling outer line chamber 11a 'is higher. Therefore, the lubricating oil 13 in the swirling outer line chamber 11a ′ leaks through the gap into the swirling outer line chamber 11a in the compression stroke, and the leaked lubricating oil 13 seals the gap in the swirling outer line chamber 11a. Further, the lubricating oil 13 supplied to the swirling outer line chamber 11a leaks into the virtual swirling outer line chamber 11A, thereby sealing the swirling outer line chamber.

6 (a) and 6 (b), the swivel extension chamber 11b 'starts to compress 360 degrees ahead of the swivel extension chamber 11b in terms of the crank angle. Comparing the pressure, the pressure in the swivel extension chamber 11b 'is higher. Therefore, the lubricating oil 13 in the turning extension chamber 11b ′ leaks through the clearance into the turning extension chamber 11b in the compression stroke, and this leaked lubricating oil 13 seals the clearance in the turning extension chamber 11b. Further, the lubricating oil 13 in the turning extension chamber 11b leaks into the virtual turning extension chamber 11B, thereby sealing the turning extension chamber.

Next, it will be specifically explained that the second type of leakage is suppressed.

In FIG. 11, the swirling outer line chamber 11a supplied with the lubricating oil 13 through the communication hole 18 and the release valve hole 15a1 is defined as a P2 room. The lubricating oil 13 in the swirling outer line chamber 11 a leaks out of the gap 191 by leaking into the P1 room having a low pressure through the gap 191. The lubricating oil 13 leaks from the gap 192 on the side adjacent to the P3 room where the pressure is high, so that the sealing property of the room is maintained.

The swirling outer chamber 11a 'shown in FIG. 6 (a) starts compression with a crank angle of 180 ° ahead of the swirling inner chamber 11b, so that the outer swirling chamber 11a' and the swirling inner chamber 11b having the same volume are started. Is compared, the pressure in the swirling outer line chamber 11a 'is higher. Accordingly, the lubricating oil 13 in the swirling outer chamber 11a ′ leaks through the clearance into the swirling inner chamber 11b in the compression stroke, and this leaked lubricating oil 13 seals the clearance in the swirling inner chamber 11b. Further, the lubricating oil 13 supplied to the swirling outer chamber 11a ′ leaks into the virtual swirling inner chamber 11B, and thus the swirling inner chamber is sealed.

In addition, the swirl extension chamber 11b shown in FIG. 6A starts compression with a crank angle of 180 ° ahead of the swirl extension chamber 11a, so the pressures in the swirl extension chamber 11a and the swirl extension chamber 11b are compared. Then, the pressure in the swivel extension chamber 11b is higher. Therefore, the lubricating oil 13 in the swirling extension chamber 11b leaks through the gap into the swirling outer line chamber 11a in the compression stroke, and the leaked lubricating oil 13 seals the gap in the turning outer line chamber 11a.

In addition, since the swirling outer line chamber 11a shown in FIG. 6A has started to be compressed in advance of the virtual swirling inner line chamber 11B, the pressure in the swirling outer line chamber 11a is higher. Therefore, the lubricating oil 13 in the swirling extension chamber 11a leaks through the clearance into the virtual swirling extension chamber 11B in the compression stroke, and eventually the swiveling extension chamber is sealed.

Further, the swirling outer line chamber 11a 'shown in FIG. 6B communicates with the discharge port 5e. Therefore, the swirling outer line chamber 11a' is no longer strictly a compression chamber, but is related to the front and rear crank angles. In order to facilitate understanding, the above is described as above. Since the swirling outer chamber 11a 'starts to be compressed 180 degrees ahead of the swirling inner chamber 11b' at the crank angle, the pressures of the swirling outer chamber 11a 'and the swirling inner chamber 11b' are compared. The pressure is higher. Further, the swirling outer chamber 11a 'starts to be compressed 360 ° ahead of the swirling

inner chamber

11b and 180 ° ahead of the swirling inner chamber 11b', so that the pressure in the outer swirling chamber 11a 'is higher. It is high. In the swirling outer line chamber 11a ′ itself shown in FIG. 6 (b), the pressure is not limited to being high, and the pressure in the swirling outer line chamber 11a ′ is the discharge pressure. This is because, as described above, the swirling outer line chamber 11a 'communicates with the discharge port 5e. Accordingly, since the pressure in the swirling outer chamber 11a ′ is higher as described above, the lubricating oil 13 in the swirling outer chamber 11a ′ has a gap between the swirling inner chamber 11b ′ and the swirling inner chamber 11b in the compression stroke. The leaked lubricating oil 13 seals the gap between the swivel extension chamber 11b 'and the gap between the swivel extension chamber 11b.

In addition, the swivel extension chamber 11b 'shown in FIG. 6B starts to be compressed 180 degrees ahead of the swivel outer stroke chamber 11a in terms of crank angle, so that the pressure in the swirl extension chamber 11a and the swirl extension chamber 11b' is increased. When compared, the pressure in the swivel extension chamber 11b 'is higher. Accordingly, the lubricating oil 13 in the swirling extension chamber 11b 'leaks through the gap into the swirling outer line chamber 11a in the compression stroke, and the leaked lubricating oil 13 seals the gap in the turning outer line chamber 11a.

In addition, the swirling outer chamber 11a shown in FIG. 6B starts compression with a crank angle of 180 ° ahead of the swirling inner chamber 11b, so that the outer swirling chamber 11a and the swirling inner chamber 11b having the same volume are started. When the pressures of these are compared, the pressure of the swirling outer line chamber 11a is higher. Accordingly, the lubricating oil 13 in the swirling outer chamber 11a leaks through the clearance into the swirling inner chamber 11b during the compression stroke, and the leaked lubricating oil 13 seals the clearance of the swirling inner chamber 11b.

Further, since the swirl extension chamber 11b shown in FIG. 6B starts to be compressed in advance of the virtual swirl outer chamber 11A, the pressure in the swirl extension chamber 11b is higher. Therefore, the lubricating oil 13 in the turning extension chamber 11b leaks through the clearance into the virtual turning outer line chamber 11A in the compression stroke, and eventually seals the turning outer line chamber.

In the above description, the release valve 15 has been described as not opening. However, the release valve 15 may be opened depending on actual operating conditions. When the release valve 15 is opened, the pressure in the compression chamber 11 exposed thereto becomes the same as the discharge pressure. There is no leakage between the compression chambers having the same pressure, and between the compression chambers having different pressures, the lubricating oil 13 leaks from the compression chamber having a high pressure to the compression chamber having a low pressure.

In addition, the suction completed in the virtual swirling extension chamber 11A is defined as the swirling outer stroke chamber a, and the swirling outer stroke chamber a 'whose phase is advanced by 360 ° from the swirling outer stroke chamber 11A. This is a swivel extension chamber b, and a swirl extension chamber b 'whose phase is 360 ° higher than that of the swivel extension chamber. When the number of wraps is increased to increase the number of compression chambers, it can be similarly described as a ″, a ′ ″,..., B ″, b ′ ″,.

Thus, the lubricating oil 13 supplied through the communication hole 18 and the release valve hole 15a1 is used to seal the gap until the end of discharge, and the surplus lubricating oil is discharged from the discharge port 5e to the discharge pressure chamber 2f. It is.

Considering the above leakage and remainder, it is more advantageous to supply oil to the compression chamber on the outer diameter side as much as possible before and after forming the compression chamber so that the compression can be reliably performed from the start of compression. In addition, the higher the pressure in the compression chamber 11, the smaller the amount of oil supplied by the tooth tip lubrication. When the pressure in the compression chamber 11 becomes higher than the back pressure, the tip pressure oil cannot be supplied from the back pressure chamber 14. Therefore, it is more advantageous to supply oil to the compression chamber on the outer diameter side as much as possible from this point. By performing the oil supply to the suction chamber by the back pressure control valve 16 and the oil supply to the swirling outer line chamber 11a by the tip oil supply, a more efficient compressor can be obtained.

Here, in this embodiment, the back pressure chamber 14 and the swirling outer line chamber 11a are communicated with each other via the communication hole 18 and the release valve hole 15a1 for the swirling outer line chamber 11a, and the tooth tip oil supply is performed only to the swirling outer line chamber 11a. Although explained in the form, further high efficiency can be achieved if the tooth tip lubrication is performed also in the swivel extension chamber 11b. In this case, similarly to the communication hole 18, a communication hole (18-2) is provided in the wrap of the orbiting scroll 6, and the tooth tip is also provided to the orbiting extension chamber 11 b via the release valve hole 15 a 2 for the orbiting extension chamber 11 b. A structure that supplies oil may be used. At this time, the non-wrap side opening of the communication hole (18-2) may be shared with that of the communication hole 18. The turning inner /

outer line chambers

11a and 11b that perform tooth tip lubrication are the outermost turning inner / outer line chambers.

The configuration in which the tooth tip oil supply is performed to both the swirling outer line chamber 11a and the swirling inner line chamber 11b has a specific effect particularly in the case of a symmetrical wrap type. FIG. 14 is a diagram of the timing at which suction is completed for both the swirling outer chamber 11a and the swirling inner chamber 11b of the symmetrical wrap. In the case of the symmetrical lap type, the swirling outer chamber 11a and the swirling inner chamber 11b start to be compressed at the same timing, so that the pressure is the same if the volume of the swirling outer chamber 11a and the volume of the swirling inner chamber 11b are the same. become. Therefore, in the case of the symmetrical wrap type, the leakage of the lubricating oil 13 between the swirling inner and outer line chambers having the same volume, that is, the second type of leakage between the swirling inner and outer line chambers having the same volume is eliminated. It is preferable to provide communication holes 18 and 18-2 communicating with the release valve holes 15a1 and 15a2 in each of the swivel extension chambers 11b.

If the aforementioned imbalance is such that the suction toward the virtual swivel extension chamber 11B is small, the tooth refueling may be performed only to the swivel extension chamber 11b.

Although the above description has been made on the assumption that the vertical scroll compressor 1 shown in FIG. 1 is used, the same effect can be obtained even if the horizontal scroll compressor is assumed as shown in FIG.

FIG. 16 shows the efficiency of the scroll compressor 1 of this embodiment. In both (a) and (b), the upper diagram represents volumetric efficiency as a ratio, and the lower diagram represents compressor efficiency as a ratio. (A) is the figure which compared the left figure showing (theta) b = 270 degrees (9 o'clock position) and the right figure showing the conventional (theta) b = 210 degrees (11 o'clock position) of a present Example. (B) is the figure which compared the left figure which has the communicating hole 18, and the right figure which does not have the communicating hole 18 at (theta) b = 270 degrees. Here, (a) shows the ratio of the right figure representing θb = 210 ° as 100%, and (b) shows the ratio of the right figure without communication hole 18 as 100%. Yes. The operating conditions are conditions for storing hot water of 65 ° C. when the scroll compressor 1 is mounted on Ecocute (registered trademark). Volumetric efficiency can be expressed by equation (3), and compressor efficiency can be expressed by equation (4).

η _v = Γ / (V0 · ρ _s · f) (3)
η _c = Γ · Δh / w (4)
Here, eta _v is volumetric efficiency, eta _c compressor efficiency, gamma refrigerant circulation amount, V0 is押除volume, [rho _s is the suction gas density, f is the motor rotation frequency, Delta] h is the suction gas and discharge gas enthalpy The difference, w, represents the motor input.

As can be seen from (a), the back pressure control valve position θb = 270 ° (9 o'clock position) has a volume efficiency and a compressor efficiency of 1.5-2% relative to θb = 210 ° (11 o'clock position). The degree has improved. Further, as can be seen from (b), the volume efficiency and the compressor efficiency of the one having the communication hole 18 are improved by 2 to 2.5% compared to the one having no communication hole 18. In particular, it can be seen that since the volumetric efficiency is increased, the sealing performance of the compression chamber is improved, and the refrigerant circulation amount of the equations (3) and (4) is increased.

As described above, the virtual swirl outer chamber 11A and the virtual swirl chamber 11B can be supplied with good balance, the sealing performance of each compression chamber can be improved, and the leakage loss can be reduced. Further, by connecting the communication hole 18 and the release valve hole 15a1 and supplying oil from the back pressure chamber 14 to the swirling outer line chamber 11a, the amount of oil flowing into the suction space 10 and the suction chamber through the back pressure control valve 16 is reduced. The heat loss of the suction gas can be reduced.

Next, the case where the scroll compressor 1 is mounted on a heat pump water heater, Ecocute (registered trademark) to form a unit will be described. FIG. 17 is a unit configuration diagram. The same reference numerals as those in the above embodiment have the same operational effects, and thus the description thereof is omitted.

At a certain time (for example, 3:00 am) at midnight, the scroll compressor 1 is activated and the high-temperature and high-pressure refrigerant compressed from the discharge pipe 2e is discharged. The discharged refrigerant is heat-exchanged with water in the hot water storage tank 32 by the water-refrigerant heat exchanger 29 and cooled. The first heat exchanger described above can be used as the water-refrigerant heat exchanger 29. The refrigerant exiting the water-refrigerant heat exchanger 29 is depressurized by the expansion valve 33 and enters the evaporator 34 to absorb the heat of the atmosphere and evaporate. The refrigerant exiting the evaporator 34 is sucked into the scroll compressor 1 from the suction pipe 2d and is compressed again here. Such a refrigeration cycle apparatus equipped with the scroll compressor 1 of this embodiment also becomes a highly efficient refrigeration cycle apparatus by the amount that the efficiency of the scroll compressor 1 is increased.

On the other hand, the water in the hot water storage tank 32 is conveyed by the water circulation pump 31 and led to the water-refrigerant heat exchanger 29. Water guided from the lower part of the hot water storage tank 32 is heated by the water-refrigerant heat exchanger 29, and the heated water is returned to the upper part of the hot water storage tank 32.

Next, the control method will be described. The remote controller 30 sets the temperature of hot water stored in the hot water storage tank 32 by the user. Signals from the hot water temperature sensor 35 and the discharge gas temperature sensor 36 are input to the control unit 25. When the temperature detected by the hot water temperature sensor 35 or the discharge gas temperature sensor 36 is lower than the hot water temperature set by the remote controller 30, the rotation speed of the scroll compressor 1 is increased to increase the refrigerant circulation amount, Control is performed such that the valve 33 is throttled to increase the discharge pressure, thereby increasing the temperature of the hot water.

By the above method, the refrigeration cycle is controlled so that the temperature of the hot water in the hot water storage tank 32 becomes a desired temperature. For example, the operation is stopped at 7 o'clock in the morning. In the daytime, hot water in the hot water storage tank 32 and tap water from the water pipe are mixed, and hot water is supplied from the shower 27 and the faucet 28 which are used terminals according to the user's request. When the hot water in the bathtub 24 is replenished, the hot water in the bathtub and the hot water in the hot water storage tank 32 are heat-exchanged by the reheating heat exchanger 26 provided in the hot water storage tank 32.

Such scroll compressors are installed in room air conditioners, commercial packaged air conditioners, heat pump water heaters, and the like. There is a year-round energy consumption efficiency (annual performance factor) as an index indicating the year-round performance of room air conditioners and heat pump water heaters. For example, in the case of a heat pump water heater, this APF is determined by how much power the device consumes with respect to the hot water supply load for each outside air temperature defined in the standard, and is expressed by hot water supply load / power consumption. . Here, the hot water supply load is expressed by the following equation.

Lw = (θo−θi) · Cw · v · d (5)
Here, Lw is the hot water supply load, θo is the hot water supply temperature, θi is the incoming water temperature, Cw is the specific heat of water, v is the amount of hot water supply, and d is the number of days.

Here, the hot water supply temperature θo and the incoming water temperature θi are determined by the outside air temperature. The number of days d is determined by how many days in the year the outside temperature is. When the hot water supply load Lw is integrated over the year, the annual hot water supply load is calculated. An improvement in compressor efficiency means a reduction in power consumption, and an APF is improved in a device equipped with the scroll compressor of this embodiment. That is, energy saving can be achieved. Alternatively, when the same power consumption as before can be used, the heating capacity can be increased. For example, since the heating capacity can be increased even in a cold region, the temperature for storing hot water can be increased, and the amount of hot water that can be used can be substantially increased without changing the capacity of the hot water storage tank 32.

FIG. 18 shows a second embodiment. The scroll compressor shown in FIG. 18 has substantially the same configuration as that of the first embodiment, and those having the same name and code have the same operational effects. The difference between the second embodiment and the first embodiment is that the communication hole 18 communicates not with the release valve hole 15a but with a recess 20 formed at a position deeper than the bottom surface of the fixed scroll 5, that is, the tooth bottom. It is to be. In other words, the tooth tip opening communicates with the recess 20, and the anti-wrap side opening communicates with the back pressure chamber 14. It can be said that the recess 20 is also a space provided at a position deeper than the tooth bottom of the fixed scroll 5.

As described in the first embodiment, the main purpose of the release valve 15 is to operate when the liquid refrigerant is sucked, such as when the pressure in the compression chamber 11 becomes equal to or higher than the pressure P2f of the discharge pressure chamber 2f or immediately after startup. Therefore, the installation position is defined to some extent. However, as in this embodiment, the installation position can be made free by using the recess 20, and the degree of freedom in setting the timing at which the back pressure chamber 14 and the compression chamber 11 communicate with each other via the communication hole 18 and the recess 20 is increased. .

FIG. 19 shows a third embodiment. The scroll compressor 1 shown in FIG. 19 has substantially the same configuration as that of the first embodiment, and those having the same name and code have the same operational effects. The difference from the first embodiment is that the communication hole 18 communicates with the upper space of the crankshaft 7 in the slewing bearing 6 c, that is, the discharge pressure oil supply chamber 51. The tooth tip opening communicates with the release valve hole 15a1, and the anti-wrap side opening communicates with the discharge pressure oil supply chamber 51 that is formed on the anti-wrap side of the orbiting scroll 6 and has a pressure higher than the back pressure. is there.

Since the inside of the discharge pressure oil supply chamber 51 is substantially at the discharge pressure Pd, it is possible to supply oil from the discharge pressure oil supply chamber 51 to the compression chamber 11 using the differential pressure by connecting the communication hole 18 and the release valve hole 15a. is there. However, as compared with the first embodiment, since the differential pressure to be supplied is increased, it is necessary to suppress the supply amount of the lubricating oil 13 by shortening the communication section between the communication hole 18 and the release valve hole 15a. Therefore, it can be considered that the cross-sectional area of the communication hole 18 is made smaller than that of the first and second embodiments. Here, the communication hole 18 penetrates a hole from the outer peripheral surface of the base plate 6b of the orbiting scroll 6 toward the discharge pressure oil supply chamber 51, and the hole is machined from the tooth tip of the orbiting scroll 6 toward the through hole. It can be formed by plugging a stopper into a hole penetrating the outer peripheral surface of 6b by press fitting or screwing.

FIG. 20 shows a fourth embodiment. The refrigerant flow and lubricating oil flow in this embodiment are almost the same as those in the embodiment shown in FIG. The difference from the embodiment of FIG. 1 is that the orbiting bearing 6 c is a so-called shaft-through scroll compressor in which the orbiting scroll 6 penetrates. The gas compression load due to the pressure in the compression chamber 11 acts on the central portion of the lap height. This gas compression load acts in the direction of the slewing bearing 6c and acts as a bearing load on the slewing bearing 6c. Therefore, the action points of the gas compression load and the bearing load coincide with each other, and the moment for overturning the orbiting scroll 6 is eliminated.

FIG. 21 shows a fifth embodiment. The flow of the refrigerant in this embodiment is almost the same as that in the embodiment shown in FIG. A different point from the embodiment of FIG. 1 is an oil supply system, which is a so-called forced oil supply system. An oil supply pump 103 such as a trochoid pump is provided at the lower end of the crankshaft 7. This oil pump 103 is interlocked with the rotation of the crankshaft 7. Oil is supplied to the slewing bearing 6 c and the main bearing 9 a by an oil supply pump 103. The space around the crankshaft 7 and the back pressure chamber 14 are partitioned by a seal ring 102. Oil is supplied to the back pressure chamber 14 by an oil pocket 101 that travels between the space around the crankshaft 7 and the back pressure chamber 14. The traffic uses the revolving motion of the orbiting scroll 6. When oil is supplied by an oil pump, there is an advantage that oil can be supplied by the volume of the oil pump regardless of pressure conditions, and the amount of oil supplied to the bearing can be reduced when the pressure difference between the discharge pressure and the suction pressure is large.

As described above, the efficiency of the compressor, the refrigeration cycle apparatus, and the like can be increased by the technology described in each embodiment. It should be noted that, in addition to the structure itself described in these embodiments, the position and tip of the back pressure control valve, which are characteristic features, can be obtained even if the vertical scroll compressor is replaced with the horizontal scroll compressor. If the portion for refueling is not changed, a similar effect can be obtained even in a configuration in which the respective configurations are appropriately combined.
While the above description has been made with reference to exemplary embodiments, it will be apparent to those skilled in the art that the invention is not limited thereto and that various changes and modifications can be made within the spirit of the invention and the scope of the appended claims.

DESCRIPTION OF SYMBOLS 1 Scroll compressor 2 Airtight container 2a Case 2b Cover chamber 2c Bottom chamber 2d Suction pipe 2d1 Suction port 2e Discharge pipe 2f Discharge pressure chamber 3 Compression mechanism part 4 Electric motor 4a Stator 4b Rotor 5 Fixed scroll 5c Lap 5d Base plate 5e Discharge Exit 5f Spring storage hole 5g Through hole 5h R groove 5i Conducting path 5Xi End of winding of inner line side wrap of fixed scroll 5 5Xo End of winding of outer side wrap of fixed scroll 5 6 Orbiting scroll 6a Wrap 6b Base plate 6c Orbiting bearing 6Xi Winding end portion 6Xo of the inner line side wrap of the orbiting scroll 6 7Crank shaft 7a Main shaft 7b Eccentric part 7c Oil supply passage 7d Oil supply pipe 7z hole 8 Bolt 9 Frame 9a Main bearing 10 Suction space 11 Compression chamber 11A Virtual Rotating outer chamber 11a Swivel outer chamber 11a 'Swivel outer chamber 11B Virtual swirl inner chamber 11b Swivel inner chamber 11b' Swivel inner chamber 12 Oldham ring 13 Lubricating oil 14 Back pressure chamber 15 Release valve 15a Release valve hole 16 Back pressure control valve 16a Piece 16b Communication hole 16c Valve body 16d Spring 16e Seal member 17 Lower bearing 18 Communication hole 191, 192 Clearance 20 Recess 25 Control unit 26 Reheating heat exchanger 29 Water-refrigerant heat exchanger 30 Remote control 31 Water circulation pump 32 Hot water storage tank 33 Expansion Valve 34 Evaporator 35 Hot water temperature sensor 36 Discharge gas temperature sensor 50 Oil supply section 51 Discharge pressure oil supply chamber 101 Oil pocket 102 Seal ring 103 Oil supply pump

Claims

In the scroll compressor of the intermittent communication structure that urges the orbiting scroll to the fixed scroll by the back pressure controlled by the back pressure control valve and compresses the refrigerant in the compression chamber formed by both scrolls,
The back pressure control valve is disposed at a position where intermittent communication starts when the volumes of the suction chamber on the inner line side of the orbiting scroll and the suction chamber on the outer line side of the orbiting scroll are increased. A scroll compressor characterized by that.
Surrounded by an imaginary line connecting the winding end portion of the inner scroll side wrap of the fixed scroll and the winding end portion of the outer scroll side wrap of the orbiting scroll, the outer scroll side wrap of the orbiting scroll, and the inner scroll side of the fixed scroll The back pressure control valve and the back pressure chamber communicate with each other when the volume of the virtual swirl outer line chamber that is one of the suction chambers is increased,
Surrounded by an imaginary line connecting the winding end portion of the outer line side wrap of the fixed scroll and the winding end portion of the inner line side wrap of the orbiting scroll, the inner line side wrap of the orbiting scroll, and the outer line side wrap of the fixed scroll The back pressure control valve is disposed at a position where the back pressure control valve communicates with the back pressure chamber when the volume of the virtual swirl extension chamber, which is one of the suction chambers, is increased. The scroll compressor according to claim 1, wherein the scroll compressor is provided.
The volume of both the suction chamber on the inner line side of the orbiting scroll and the suction chamber on the outer line side of the orbiting scroll is up to the respective volume when the respective compression chambers are spaces in which the respective suction chambers are closed. 2. The scroll compressor according to claim 1, wherein the back pressure control valve is disposed at a position where the communication start of intermittent communication is performed when increasing.
The compression chamber is a swirling extension chamber and a swirling outer line chamber formed on an inner line side of the orbiting scroll and an outer line side of the orbiting scroll, and the suction chamber on the inner line side of the orbiting scroll becomes the orbiting extension chamber. When the volume of the suction scroll on the outer line side of the orbiting scroll increases to the same volume as that of the orbiting outer line chamber. The scroll compressor according to claim 1, wherein the back pressure control valve is disposed at a position where the communication start is performed.
The scroll compressor according to claim 1, wherein the back pressure control valve is disposed at a position of 270 ° to 330 ° in a direction in which the wrap is rewound from a winding end portion of the wrap of the fixed scroll.
A scroll having a toothed oil supply structure in which the orbiting scroll is urged to a fixed scroll by a back pressure that is a pressure of a back pressure chamber provided on the opposite side of the orbiting scroll, and the refrigerant is compressed in a compression chamber formed by both scrolls. In the compressor, a scroll compressor that supplies tooth tip oil to the compression chamber via a space deeper than a tooth bottom of the fixed scroll.
The orbiting scroll has a communication hole provided with a tooth tip opening provided at a tooth tip which is an end surface of the wrap and an anti-lap side opening provided on the anti-lap side with respect to the base plate of the orbiting scroll. A space that is provided in the lap side of the orbiting scroll and has a pressure equal to or higher than the back pressure; The scroll compressor according to claim 6, wherein the anti-wrap side opening communicates with each other and tooth tip oil is supplied to the compression chamber from a space having a pressure equal to or higher than the back pressure.
The fixed scroll has a release valve that releases the pressure of the compression chamber into the sealed container of the scroll compressor, and the space provided at a position deeper than the bottom of the fixed scroll has a release valve The scroll compressor according to claim 7, wherein a space that is a valve hole and has a pressure equal to or higher than the back pressure is the back pressure chamber.
The space provided at a position deeper than the tooth bottom of the fixed scroll is a recess provided in the fixed scroll, and the space having a pressure higher than the back pressure is the back pressure chamber. Scroll compressor described in 1.
The space which becomes the pressure more than the back pressure is a discharge pressure oil supply chamber into which the oil of the pressure in the sealed container of the scroll compressor is introduced on the side opposite to the orbiting scroll. Scroll compressor.
In a sealed container,
A fixed scroll having a spiral wrap;
A rotating scroll that has a spiral wrap and forms a compression chamber that meshes with the wrap of the fixed scroll to compress refrigerant, and revolves with respect to the fixed scroll based on rotation of a crankshaft;
A back pressure control valve for controlling the pressure of the back pressure chamber formed on the side opposite to the orbiting scroll;
A release valve disposed in the fixed scroll, the release valve for discharging the refrigerant in the compression chamber into the sealed container when the pressure in the compression chamber becomes larger than the pressure in the sealed container;
In the scroll compressor in which the back pressure control valve and the back pressure chamber communicate intermittently,
Surrounded by an imaginary line connecting the winding end part of the inner line side wrap of the fixed scroll and the winding end part of the outer line side wrap of the orbiting scroll, the outer line side wrap of the orbiting scroll, and the inner line side wrap of the fixed scroll. When the volume of the virtual swirling outer line chamber increases as the crankshaft rotates, the back pressure control valve and the back pressure chamber communicate with each other, and the fixed scroll A virtual swirl surrounded by a virtual line connecting a winding end part of the outer line side wrap and a winding end part of the inner line side wrap of the orbiting scroll, the inner line side wrap of the orbiting scroll, and the outer line side wrap of the fixed scroll. The back pressure control valve is disposed at a position where the back pressure control valve and the back pressure chamber communicate with each other when the volume of the extension chamber increases as the crankshaft rotates.
The orbiting scroll has a communication hole provided with a tooth tip opening provided at a tooth tip which is an end surface of the wrap and an anti-lap side opening provided on the anti-lap side with respect to the base plate of the orbiting scroll. In the wrap,
By the revolving motion of the orbiting scroll, the tooth tip opening and the release valve hole of the release valve communicate intermittently,
The back pressure chamber communicates with the anti-wrap side opening,
The scroll compressor, wherein the back pressure chamber and the compression chamber communicate with each other.
The scroll compressor according to claim 8 or 11, wherein the compression chamber is a turning outer line chamber which is a compression chamber formed on an outer diameter side of the wrap of the turning scroll.
The fixed scroll has a second release valve for letting the pressure of the swivel extension chamber formed on the inner diameter side of the wrap of the orbiting scroll out of the compression chamber into the sealed container of the scroll compressor,
The orbiting scroll includes a second tooth tip opening provided at a tooth tip which is an end surface of the wrap, and a second anti-wrap side opening provided on the anti-wrap side with respect to the base plate of the orbiting scroll. A second communication hole in the wrap;
By the revolving motion of the orbiting scroll, the second tooth tip opening and the release valve hole of the second release valve communicate intermittently,
The back pressure chamber communicates with the second anti-wrap side opening,
The scroll compressor according to claim 12, wherein the back pressure chamber and the swivel extension chamber communicate with each other.
In the refrigeration cycle apparatus in which the discharge pipe of the scroll compressor, the first heat exchanger, the expansion device, the second heat exchanger, and the suction pipe of the scroll compressor are sequentially connected,
The scroll compressor according to any one of claims 1, 6, and 11 is used as the scroll compressor,
A refrigeration cycle apparatus comprising a supercritical refrigeration cycle in which the refrigerant is carbon dioxide.
A refrigeration cycle apparatus according to claim 14, a hot water storage tank, a water-refrigerant heat exchanger, and a water circulation pump,
The first heat exchanger is used as the water-refrigerant heat exchanger, the water circulation pump is operated to guide water from the hot water storage tank, and water is heated by the water-refrigerant heat exchanger. The heated water is returned to the tank,
A heat pump water heater in which hot water stored in the hot water storage tank is supplied to a use terminal.