KR20070054662A - Rotary compressor - Google Patents

Abstract

The present invention discloses a rotary compressor having two compression capacities. To this end, the present invention is a clock and a counterclockwise rotatable drive shaft having a predetermined size of the eccentric; A cylinder forming an internal volume of a predetermined size; A roller rotatably installed on an outer circumferential surface of the eccentric portion so as to contact the inner circumferential surface of the cylinder, and having a rolling motion along the inner circumferential surface to form a fluid chamber together with the inner circumferential surface for suction and compression of the fluid; A vane elastically installed on the cylinder to continuously contact the roller; Upper and lower bearings respectively installed on upper and lower portions of the cylinder to rotatably support the driving shaft, and seal the internal volume; An oil channel configured to uniformly flow oil between the bearings and the drive shaft; Discharge ports communicating with the fluid chamber; Suction ports communicating with the fluid chamber and spaced apart from each other by a predetermined angle; And it provides a rotary compressor consisting of a valve assembly for selectively opening any one of the suction ports in accordance with the rotation direction of the drive shaft.

Description

Rotary compressors {ROTARY COMPRESSOR}

The present invention relates to a rotary compressor, and more particularly to a mechanism for changing the compression capacity of the compressor.

In general, a compressor is a machine that increases the pressure of a working fluid by receiving power from a power generating device such as an electric motor or a turbine and applying a compression work to a working fluid such as air or a refrigerant. Such compressors are widely used in general household appliances such as air conditioners and refrigerators to the plant industry.

Such compressors are classified into positive displacement compressors and dynamic compressors or turbo compressors according to the compression method. Among them, a volume compressor is widely used in the industrial field, and has a compression method of increasing pressure through volume reduction. The volumetric compressor is further classified into a reciprocating compressor and a rotary compressor.

The reciprocating compressor compresses the working fluid by a linear reciprocating piston inside the cylinder, and has the advantage of producing high compression efficiency with relatively simple mechanical elements. However, the reciprocating compressor has a limitation in the rotational speed due to the inertia of the piston, there is a disadvantage that a considerable vibration occurs due to the inertia force. The rotary compressor compresses the working fluid by a roller revolving in the cylinder eccentrically, and can produce a high compression efficiency at a low speed compared to the reciprocating compressor. Thus, the rotary compressor further has an advantage of less vibration and noise.

Recently, reciprocating compressors having at least two compression capacities have been developed. These dual displacement compressors have different compression capacities depending on the direction of rotation (ie clockwise or counterclockwise) using a partially modified compression mechanism. Since the dual capacity compressor can adjust the compression capacity according to the required load size, it is widely applied to increase the operating efficiency in various devices that require the compression of the working fluid, in particular, home appliances such as refrigerators. have.

However, the conventional rotary compressor has one inlet port and one discharge port communicating with the cylinder, and the roller compresses the working fluid while rolling along the inner circumferential surface of the cylinder from the inlet side to the outlet side. Therefore, when the roller rolls in the opposite direction (from the discharge port side to the suction port side), the working fluid is not compressed. That is, the conventional rotary compressor cannot have different compression capacities by changing the rotational direction. Therefore, there is a need for the development of a rotary compressor having not only the unique advantages described above but also a variable compression capacity.

In addition, in such a compressor, the driving units such as the motor and the drive shaft, etc., are placed in a harsh operating environment due to the change in compression capacity as well as the high speed movement. Therefore, in addition to the variable displacement mechanism, a suitable lubrication mechanism should be developed.

SUMMARY OF THE INVENTION The present invention has been made to solve the above problems, and an object of the present invention is to provide a rotary compressor capable of compression in both clockwise and counterclockwise rotation of the drive shaft.

Another object of the present invention is to provide a rotary compressor capable of varying the compression capacity.

Another object of the present invention is to provide a rotary compressor having a lubrication mechanism suitable for a variable displacement mechanism.

In order to achieve the above object, the present invention includes a drive shaft rotatable clockwise and counterclockwise and having an eccentric portion of a predetermined size; A cylinder forming an internal volume of a predetermined size; A roller rotatably installed on an outer circumferential surface of the eccentric portion so as to be in contact with the inner circumferential surface of the cylinder, for rolling along the inner circumferential surface and forming a fluid chamber together with the inner circumferential surface for suction and compression of the fluid; A vane elastically installed in the cylinder to continuously contact the roller; Upper and lower bearings respectively installed on upper and lower portions of the cylinder to rotatably support the driving shaft, and seal the internal volume; An oil channel configured to uniformly flow oil between the bearings and the drive shaft; Discharge ports communicating with the fluid chamber; Suction ports communicating with the fluid chamber and spaced apart from each other by a predetermined angle; And a valve assembly for selectively opening any one of the suction ports in accordance with the rotational direction of the drive shaft. Compression spaces having different sizes are formed in the fluid chamber in accordance with the rotational direction of the drive shaft. Disclosed is a rotary compressor having a capacity.

The discharge port includes first and second discharge ports positioned to face each other with respect to the vane.

The suction port includes a first suction port positioned near the vane and a second suction port spaced at a predetermined angle from the first suction port. The suction port further includes a third suction port positioned between the second suction port and the vane.

The valve assembly comprises a first valve rotatably installed between the cylinder and the bearing and a second valve for guiding the rotational movement of the first valve. First, the first valve is made of a disc member that contacts the eccentric portion of the drive shaft and rotates in the rotational direction of the drive shaft. The first valve includes a first opening in communication with the first suction port when the drive shaft is rotated in one direction and a second opening in communication with the second suction port when the drive shaft is rotated in another direction. Preferably, the first valve further includes a third opening that opens the third suction port simultaneously with the opening of the second suction port. On the other hand, the first valve includes a first opening that opens the third suction port simultaneously with opening the second suction port. Also, the second valve is fixed between the cylinder and the bearing and has a seat to receive the first valve.

The valve assembly further comprises means for controlling the angle of rotation of the first valve to accurately open the corresponding suction port in each direction of rotation. The control means also includes a curved groove of a predetermined length formed in the first valve and a stopper formed on the bearing and inserted into the groove.

On the other hand, the control means includes a projection projecting radially in the first valve and a groove formed in the second valve and movably receiving the projection. On the other hand, the control means includes a projection projecting radially in the second valve and a groove formed in the first valve and movably receiving the projection. On the other hand, the control means includes a protrusion that radially inwardly protrudes in the second valley portion and a cutout that movably houses the protrusion formed in the first valley portion.

The oil passage is preferably configured to flow oil between the bearings and the drive shaft both in clockwise and counterclockwise rotation. Specifically, the oil passage is formed in the bearing and consists of at least one linear groove for flowing oil regardless of the rotational directions of the drive shaft. On the other hand, the oil passage consists of first and second spiral grooves formed in the bearing and configured to flow oil in a corresponding direction of rotation of the drive shaft. The first and second spiral grooves extend in opposite directions and do not cross each other.

Further, the oil flow path is preferably provided in the bearing and is formed at a position where less eccentricity of the drive shaft occurs. The oil passage is formed in the bearing to be substantially spaced apart from the vane at a predetermined angle in a clockwise or counterclockwise direction with respect to the central axis of the drive shaft. In this case, the linear groove is preferably spaced 170 to 210 degrees clockwise or counterclockwise from the vane, more preferably 190 degrees clockwise or counterclockwise from the vane. Preferably, the first and second spiral grooves are spaced apart from the vanes by 130 ° -190 ° and 190 ° -250 ° in a clockwise or counterclockwise direction, respectively.

The oil passage substantially consists of a bearing passage formed in at least one of the bearings. The bearing flow path is formed at least in the upper bearing. In addition, the bearing flow path is formed on the inner circumferential surface of the bearing and extends continuously from the lower end to the upper end of the bearing.

The oil passage preferably further includes an auxiliary passage formed in at least one of the journals of the drive shaft. The auxiliary flow passage is formed on the outer circumferential surface of the journal. The auxiliary flow passage is preferably configured to flow oil between the bearings and the drive shaft both in clockwise and counterclockwise rotation. In detail, the auxiliary flow passage includes at least one linear groove for flowing oil regardless of the rotation directions of the drive shaft. On the other hand, the auxiliary flow passage consists of first and second spiral grooves respectively configured to flow oil in a corresponding rotational direction of the drive shaft.

The compressor of the present invention preferably further comprises a suction plenum connected to the suction ports and preliminarily storing the fluid to be compressed. The suction plenum receives oil separated from the stored fluid and is mounted to the bottom of the bearing adjacent the suction port. The volume of the suction plenum is preferably 100% -400% of the fluid chamber volume.

The invention described above results in two different compression capacities in the rotary compressor. In addition, the drive is properly lubricated during operation for these two compression capacities.

The features and advantages of the present invention may be better understood with reference to the following accompanying drawings in conjunction with the following detailed description of embodiments of the invention, of which:

1 is a partial longitudinal sectional view showing a rotary compressor according to the present invention;

2 is an exploded perspective view showing a compression unit of the rotary compressor according to the present invention;

3 is a sectional view showing a compression unit of the rotary compressor according to the present invention;

4 is a cross-sectional view showing the inside of a cylinder of a rotary compressor according to the present invention;

5A and 5B are plan views showing the lower bearing of the rotary compressor of the present invention;

6 is a plan view showing a valve assembly of the rotary compressor of the present invention;

7A-7C are plan views showing modifications of the valve assembly;

8A-8B are plan views showing rotation limiting means of the valve assembly;

8C is a partial cross-sectional view of FIG. 8B;

9A and 9B are plan views showing modifications of the rotation limiting means of the valve assembly;

10A and 10B are plan views showing another modification of the rotation limiting means of the valve assembly;

11A and 11B are plan views showing yet another variation of the rotation limiting means of the valve assembly;

12 is an exploded perspective view showing a compression part of a rotary compressor according to the present invention including a suction plenum;

FIG. 13 is a sectional view of the compression unit of FIG. 12; FIG.

14A-14C are cross-sectional views sequentially showing the interiors of cylinders when the roller revolves counterclockwise in a rotary compressor according to the present invention;

15A-15C are transverse cross-sectional views sequentially showing the interiors of cylinders when the roller idles clockwise in a rotary compressor according to the present invention;

16 is a front view showing an oil channel of the rotary compressor according to the present invention;

FIG. 17A is a sectional view taken along line II of FIG. 16 showing a first embodiment of a bearing flow path; FIG.

17B is a partial plan view showing an inner circumferential surface of a bearing including the first embodiment of the bearing flow path;

17C is a graph showing an optimal setting angle of the first embodiment of the bearing flow path;

FIG. 18A is a sectional view taken along line II of FIG. 16 showing a second embodiment of a bearing flow path;

18B is a partial plan view showing an inner circumferential surface of a bearing including a second embodiment of a bearing flow path;

18C is a graph showing the optimum setting angle of the second embodiment of the bearing flow path; And

19A and 19B are partial front views of a drive shaft showing an auxiliary flow path.

BEST MODE FOR CARRYING OUT THE INVENTION

DESCRIPTION OF THE PREFERRED EMBODIMENTS Hereinafter, preferred embodiments of the present invention in which the above objects can be specifically realized are described with reference to the accompanying drawings. In describing the present embodiment, the same name and the same reference numerals are used for the same configuration and additional description thereof will be omitted below.

1 is a longitudinal sectional view showing the configuration of a rotary compressor according to the present invention. 2 is an exploded perspective view illustrating a compression unit of the rotary compressor according to the present invention, and FIG. 3 is a cross-sectional view illustrating the compression unit of the rotary compressor according to the present invention.

First, as shown in FIG. 1, the rotary compressor according to the present invention includes a case 1 and a power generating unit 10 and a compression unit 20 located inside the case 1. In FIG. 1, the power generator 10 is located at the top of the compressor, but the compression unit 20 is located at the bottom of the compressor, but their positions may be interchanged as necessary. The upper cap 3 and the lower cap 5 are respectively installed on the upper and lower portions of the case 1 to form a sealed inner space. A suction pipe 7 for sucking the working fluid is installed on one side of the case 1 and is connected to an accumulator 8 for separating the lubricant oil from the refrigerant. In addition, a discharge tube 9 through which compressed fluid is discharged is installed at the center of the upper cap 3. In addition, the lower cap 5 is filled with a certain amount of lubricant (0) for lubrication and cooling of the frictional member. At this time, the end of the drive shaft 13 is locked to the lubricating oil (0).

The power generator 10 includes a stator 11 fixed to the case 1, a rotor 12 rotatably supported inside the stator 11, and a driving shaft press-fitted into the rotor 12. (13). The rotor 12 rotates by electromagnetic force, and the drive shaft 13 transmits the rotational force of the rotor 12 to the compression unit 20. In order to supply external power to the stator 20, a terminal 4 is installed in the upper cap 3.

The compression unit 20 has a cylinder 21 fixed to the case 1 largely, a roller 22 positioned inside the cylinder 21, and upper and lower portions respectively installed at upper and lower portions of the cylinder 21. It consists of bearings 24 and 25. In addition, the compression unit 20 includes a valve assembly 100 installed between the lower bearing 24 and the cylinder 21. This compression unit 20 will be described in more detail with reference to FIGS. 2, 3 and 4 as follows.

The cylinder 21 has an internal volume of a predetermined size and has sufficient strength to withstand the pressure of the fluid to be compressed. The cylinder 21 also houses an eccentric 13a formed in the drive shaft 13 in the internal volume. The eccentric portion 13a is a kind of eccentric cam, and has a center spaced apart from the rotation center of the drive shaft 13 by a predetermined distance. The cylinder 21 is formed with a groove 21b extending to a predetermined depth from its inner circumferential surface. The grooves 21b are provided with vanes 23 to be described later. The groove 21b has a length sufficient to fully receive the vane 90.

The roller 22 is a ring member having an outer diameter smaller than the inner diameter of the cylinder 21. As shown in FIG. 4, the roller 22 is in contact with the inner circumferential surface of the cylinder 21 and rotatably coupled to the eccentric portion 13a. Accordingly, the roller 22 rotates on the inner circumferential surface of the cylinder 21 while rotating on the outer circumferential surface of the eccentric portion 13a when the drive shaft 13 rotates. In addition, during the rolling motion, the roller 22 is simultaneously spaced apart by a predetermined distance by the eccentric portion 13a with respect to the rotation center (0). Since the outer circumferential surface of the roller 22 is always in contact with the inner circumferential surface of the cylinder by the eccentric portion 13a, the outer circumferential surface and the inner circumferential surface of the roller 22 form a separate fluid chamber 29 in the inner volume. This fluid chamber 29 is used for suction and compression of fluid in a rotary compressor.

The vanes 23 are installed in the grooves 21b of the cylinder 21 as mentioned above. In addition, an elastic member 23a is installed in the groove 21b to elastically support the vane 23, and the vane 23 continuously contacts the roller 22. That is, one end of the elastic member 23a is fixed to the cylinder 21 and the other end is coupled to the vane 90 to push the vane 23 toward the roller 22. Thus, the vane 23 divides the fluid chamber 29 into two independent spaces 29a and 29b, as shown in FIG. During the rotation of the drive shaft 13, ie the idle of the roller 22, the sizes of the spaces 29a and 29b vary but are complementary. That is, when the roller 22 rotates in the clockwise direction, one space 29a gradually decreases while the other space 29b gradually increases. However, the sum of the spaces 29a and 29b is always constant and generally coincides with the size of the predetermined fluid chamber 29. These spaces 29a and 29b relatively act as suction chambers for sucking the fluid and compression chambers for compressing the fluid in either of the rotational directions of the drive shaft (ie, clockwise or counterclockwise), respectively. Accordingly, as described above, as the roller 22 rotates, the compression chamber of the spaces 29a and 29b gradually shrinks to compress the previously sucked fluid, and the suction chamber gradually expands to relatively suck the fluid. . If the rotation direction of the roller 22 is reversed, the functions of the respective spaces 29a and 29b are also changed. That is, when the roller 22 revolves counterclockwise, the right space 29b of the roller 22 becomes a compression chamber, and when the roller 22 revolves clockwise, the left space 29a is discharged. do.

The upper bearing 24 and the lower bearing 25 are installed in the upper and lower portions of the cylinder 21, as shown in Figure 2 by using a sleeve (sleeve) and through holes formed therein (24b, 25b) The drive shaft 13 is rotatably supported. More specifically, the upper and lower bearings 24 and 25 and the cylinder 21 include a plurality of fastening holes 24a, 25a and 21a formed to correspond to each other. The cylinder 21 and the upper and lower bearings 24 and 25 are firmly fastened to each other so that the cylinder internal volume, in particular the fluid chamber 29, is sealed using fastening members such as bolts and nuts.

Discharge ports 26a and 26b are formed in the upper bearing 24. The discharge ports 26a and 26b communicate with the fluid chamber 29 so that the compressed fluid can be discharged. The discharge ports 26a and 26b may be in direct communication with the fluid chamber 29, and on the other hand, the fluid chamber may be formed through a predetermined length flow path 21d formed in the cylinder 21 and the upper bearing 24. May be communicated with (29). Discharge valves 26c and 26d are installed in the upper bearing 24 to open and close the discharge ports 26a and 26b. The discharge valves 26c and 26d selectively open the discharge ports 26a and 26b only when the pressure in the chamber 29 is higher than or equal to a predetermined pressure. To this end, it is preferable that the discharge valves 26c and 26d are leaf springs whose one end is fixed near the discharge ports 26a and 26b and the other end is freely deformable. Although not shown, a retainer may be installed on the upper part of the discharge valves 26c and 26d to limit the deformation amount so that the valves operate stably. In addition, a muffler (not shown) may be installed above the upper bearing 24 to reduce noise generated when the compressed fluid is discharged.

Suction ports 27a, 27b, and 27c are formed in the lower bearing 25 to communicate with the fluid chamber 29. The suction ports 27a, 27b, and 27c serve to guide the fluid to be compressed to the fluid chamber 29. The suction ports 27a, 27b, and 27c are connected to the suction pipe 7 so that the fluid outside the compressor flows into the chamber 29. More specifically, the suction pipe 7 is branched into a plurality of auxiliary pipes 7a and connected to the suction ports 27, respectively. If necessary, the discharge ports 26a and 26b may be formed in the lower bearing 25 and the suction ports 27a, 27b and 27c in the upper bearing 24.

These suction and discharge ports 26 and 27 are important factors in determining the compression capacity of the rotary compressor and will be described in more detail below with reference to FIGS. 4 and 5. 4 shows the cylinder 21 combined with the lower bearing 25 without the valve assembly 100 to clearly show the suction port 27.

First, the compressor of the present invention includes at least two discharge ports 26a and 26b. As shown, even if the roller 22 revolves in any direction, a discharge port must exist between the suction port and the vane 23 positioned in the revolving path to discharge the compressed fluid. Therefore, one discharge port is required for each rotational direction, which enables the compressor of the present invention to discharge fluid regardless of the idle direction of the roller 22 (ie, the rotational direction of the drive shaft 13). Meanwhile, as described above, the compression chamber of the spaces 29a and 29b becomes smaller so that the fluid is compressed as the roller 22 approaches the vane 23. Accordingly, in order to discharge the compressed fluid as much as possible, the discharge ports 26a and 26b may be formed to face each other near the vanes 23. That is, as shown, the discharge ports 26a and 26b are located on the left and right sides of the vanes 23, respectively. Preferably, the discharge ports 26a and 26b are located as close to the vanes 23 as possible.

The suction port 27 is suitably positioned so that the fluid can be compressed between the discharge ports 26a and 26b and the roller 22. In practice, in a rotary compressor the fluid is compressed from any one suction port to any discharge port located within the idle path of the roller 22. That is, the relative position of the suction port relative to the corresponding discharge port determines the compression capacity, and accordingly, two compression capacities can be obtained by using different suction ports 27 according to the rotation direction. Therefore, the compressor of the present invention has two first and second suction ports 27a and 27b corresponding to the two discharge ports 26a and 26b, respectively, and these suction ports are different from each other with respect to the center (0). Spaced at a predetermined angle from each other for two compression capacities.

Preferably the first suction port 27a is located near the vane 23. Accordingly, the roller 22 compresses the fluid from the first suction port 27a to the second discharge port 26b located opposite the vane 23 in any one direction rotation (counterclockwise in the drawing). . By this first suction port 27a, the roller 22 compresses using the entire chamber 29, so that the compressor has a maximum compressive capacity in a counterclockwise rotation. That is, the refrigerant as much as the entire volume of the chamber 29 is compressed. The first suction port 27a is substantially spaced apart from the vane 23 at an angle θ1 of 10 ° in the clockwise or counterclockwise direction as shown in FIGS. 4 and 5A. In the drawings of the present invention, the first suction port 27a is shown spaced apart by the angle θ1 in the counterclockwise direction. At this separation angle θ1, the entire fluid chamber 29 may be used for compression without interference with the vanes 23.

The second suction port 27b is spaced apart from the first suction port 27a by a predetermined angle with respect to the center 0. The roller 22 compresses the fluid from the second suction port 27b to the first discharge port 26a during clockwise rotation. Since the second suction port 27b is spaced at a considerable angle clockwise from the vane 22, the roller 22 compresses using only a part of the chamber 29 and thus compresses less than the counterclockwise direction. Pay the capacity. That is, as much volume of the refrigerant of the chamber 29 is compressed. Preferably, the second suction port 27b is spaced from the vane 23 at an angle θ2 having a range of 90 ° -180 ° in a clockwise or counterclockwise direction. In addition, the second suction port 27b is more preferably positioned opposite the first suction port 27 for the difference in the compression capacity in each rotation direction and interference cancellation with each other.

As shown in Fig. 5A, the suction ports 27a and 27b are generally circular and their diameter is preferably 6-15 mm. In addition, the suction ports 27a and 27b may have various shapes, including a rectangle, to increase the suction amount of the fluid. Furthermore, the rectangular suction ports 27a and 27b may have a predetermined curvature, as shown in FIG. 5B, thereby minimizing interference with other adjacent parts, in particular the roller 22, during operation. have.

On the other hand, in order to obtain the desired compression capacity in each rotation direction, only one suction port valid in any one rotation direction should exist. If there are two suction ports in the rotation path of the roller 22, no compression occurs between these suction ports. In other words, when the first suction port 27a is opened, the second suction port 27b should be closed and vice versa. Therefore, the valve assembly 100 is installed in the compressor of the present invention to selectively open only one of the suction ports 27a and 27b according to the idle direction of the roller 22.

2, 3 and 6, the valve assembly 100 is installed between the cylinder 21 and the lower bearing 25 to be adjacent to the suction ports, the first and second valves 110, 120 It includes. If the suction ports 27a, 27b, 27c are formed in the upper bearing 24, the first and second valves 110, 120 are installed between the cylinder 21 and the upper bearing 24. .

First, as shown in FIG. 3, the first valve 110 is a disc member installed to contact the eccentric portion 13a more precisely than the drive shaft 13. Therefore, when the drive shaft 13 rotates (the roller 22 idles), the first valve 110 rotates in the same direction. Preferably, the first valve 110 has a diameter larger than the inner diameter of the cylinder 21. Accordingly, as shown in FIG. 3, a portion (ie, an outer circumferential portion) of the first valve 110 is defined in the cylinder ( 21) is supported to rotate stably. The thickness of the first valve 110 is appropriately 0.5mm-5mm.

2 and 6, the first valve 110 may be connected to the first and second openings 111 and 112 communicating with the first and second suction ports 27a and 27b in a specific rotational direction, respectively. It includes a through hole (110a) through which the drive shaft 13 passes. More specifically, the first opening 111 communicates with the first suction port 27a by the rotation of the first valve 110 when the roller 22 rotates in one direction. The second suction port 27b is closed by the body of the first valve 110. The second opening 112 communicates with the second suction port 27b when the roller 22 rotates in the other direction, wherein the first suction port 27a is connected to the first valve 110. Is closed by the body. The first and second openings 111 and 112 may be circular or polygonal. When the openings 111 and 112 are circular, their diameters are preferably 6-15 mm. In addition, the openings 111 and 112 may be rectangular shapes having a predetermined curvature as shown in FIG. 7A or cutouts as shown in FIG. 7B, and thus the sizes of the openings may be expanded to smoothly suck the fluid. Can be. If such openings 111 and 112 are formed adjacent to the center of the first valve 110, the possibility of interference with the roller 22 and the eccentric portion 13a is increased. In addition, the openings 111 and 112 may communicate with the space between the roller 22 and the eccentric portion 13a, so that fluids may leak out along the driving shaft 13. Therefore, in practice, the openings 111 and 112 are preferably located adjacent to the outer circumference of the first valve as shown in FIG. 7C. On the other hand, by adjusting the rotation angle of the first valve 110, the first opening 111 can open the first and second suction port 27a, 27b in each rotation direction. That is, the first opening 11 communicates with the first suction port 27a while closing the second suction port 27b in one direction of rotation of the drive shaft 13, and in the other rotational direction, the first suction port 27b is closed. As the first suction port 27a is closed, the first opening 111 may communicate with the second suction port 27b. The control of the suction port using the single opening 111 is preferable because the structure of the first valve 110 is further simplified.

2, 3 and 6, the second valve 120 is fixed between the cylinder 21 and the lower bearing 25 to guide the movement of the rotating first valve 110. The second valve 120 is a ring-shaped member having a seat portion 121 to rotatably receive the first valve 110. The second valve 120 also includes a fastening hole 120a to be fastened by a fastening member together with the cylinder 21 and the upper and lower bearings 24 and 25. The thickness of the second valve 120 is preferably the same as the thickness of the first valve 110 in order to prevent leakage of the fluid and to stably support the fluid. In addition, since the first valve 110 is partially supported by the cylinder 21, the thickness of the first valve 110 is formed to form a gap for smooth rotation of the second valve 120. It may be slightly smaller than the thickness of the two valves (120).

Meanwhile, referring to FIG. 4, in the clockwise rotation, while the roller 22 revolves from the vane 23 to the second suction port 27b, between the vane 23 and the roller 22. No suction or discharge of fluid occurs. Therefore, the region V becomes a vacuum state. Such a vacuum region (V) brings a power loss of the drive shaft 13 and generates a loud noise. Therefore, the third suction port 27c is formed in the lower bearing 25 to eliminate the vacuum region V. The third suction port 27c is formed between the second suction port 27b and the vane 23 so that the vacuum state is maintained before the roller 22 passes the second suction port 27b. It serves to supply fluid to the space between the roller 22 and the vanes 23 so as not to be formed. The third suction port 27c is preferably formed near the vane 23 so as to quickly eliminate the vacuum state. However, since the third suction port 27c operates in a rotational direction different from that of the first suction port 27a, the third suction port 27c is positioned to face the first suction port 27a. In practice, the third suction port 27c is spaced apart from the vane 23 at an angle θ3 of 10 ° in a clockwise or counterclockwise direction. In addition, as shown in FIGS. 5A and 5B, the third suction port 27c may have a circular or curved rectangle like the first and second suction ports 27a and 27b.

Since the third suction port 27c acts together with the second suction port 27b, these suction ports 27b and 27c must be opened at the same time during idle of the roller 22 in one direction. Therefore, the first valve 110 further includes a third opening configured to communicate with the third suction port 27c at the same time when the second suction port 27b is opened. In the present invention, the third opening 113 may be formed independently as shown by a dotted line in FIG. 6A. However, since the first and third suction ports 27a and 27c are adjacent to each other, the angle of rotation of the first valve 110 is increased so that the first and third suction ports 111 are rotated along the direction of rotation. It is desirable to open both ports 27a and 27c.

The first valve 110 may open the suction ports 27a, 27b, and 27c according to the rotational direction of the roller 22, but the suction ports should be opened correctly in order to obtain a desired compression capacity. And the correct opening of these suction ports can be obtained by controlling the rotation angle of the first valve. Thus, the valve assembly 100 preferably further comprises means for controlling the angle of rotation of the first valve 110, which means are described in detail with reference to FIGS. 8-11. 8-11 illustrate a valve assembly combined with a lower bearing 25 to better illustrate the function of the limiting means.

As shown in FIGS. 8A and 8B, the control means is first inserted into the groove 114 and the groove 114 of a predetermined length formed in the first valve 110 and on the lower bearing 25. It includes a stopper (114a) formed in. Such groove 114 and stopper 114a are also shown in FIGS. 5A, 5B and 6. The groove 114 serves as a guide of the stopper 114 and may be a straight groove or a curved groove. The groove 114 becomes a dead volume that causes re-expansion of the fluid when exposed to the chamber 29 during operation. Therefore, the groove 114 is preferably positioned adjacent to the center of the first valve 110 so that a large portion of the groove 114 is covered by the roller 22 to revolve. An angle α between both ends of the groove 114 may be about 30 ° to 120 ° with respect to the center of the first valve 110. In addition, when the stopper 114a protrudes from the groove 114, it interferes with the roller 22. Therefore, the thickness t2 of the stopper 114a is preferably equal to the thickness t1 of the valve 110 as shown in FIG. 8C. It is preferable that the width L of the stopper 114a is equal to the width of the groove 114 so that the first valve 110 can stably rotate.

When the limiting means is used, when the drive shaft 13 rotates counterclockwise, the first valve 110 rotates counterclockwise together with the eccentric portion 13a of the drive shaft. Thereafter, as shown in FIG. 8A, the first valve 110 is stopped while the stopper 114a is caught at one end of the groove 114. In this case, the first opening 111 is connected to the first suction port. Accurate communication with (27a) and the remaining second and third suction port (27b, 27c) is closed. Therefore, the fluid flows into the cylinder through the first suction port 27a and the first opening 11 communicated with each other. On the contrary, when the driving shaft 13 rotates clockwise, the first valve 110 also rotates clockwise. At the same time, the first opening 111 and the second opening 112 also move clockwise as indicated by the dotted arrows in FIG. 8A. As shown in FIG. 8B, the first opening 111 and the second opening 112 are connected to the third suction port 27c while the stopper 114a is caught by the other end of the groove 114. The second suction port 27b is opened together. The first suction port 27a is closed by the first valve 110. Therefore, the fluid flows through the second suction port 27b / second opening 112 and the third suction port 27c / first opening 111 which are in communication with each other.

As shown in FIGS. 9A and 9B, the restricting means are formed on the protrusion 115 and the second valve 220 protruding radially from the first valve 110 and move the protrusion 115. It may also be made of a groove 123 to accommodate. In this case, the groove 123 is formed in the second valve 220 so that the groove 123 is not exposed to the internal volume of the cylinder 21. 10A and 10B, the restricting means is formed on the first valve 110 and the protrusion 124 protruding radially inwardly from the second valve 120 and the protrusion 124. It may be made of a groove 116 to receive the movable.

When such a limiting means is used, when the driving shaft 13 rotates counterclockwise, the protrusions 115 and 124 are caught at one end of the grooves 123 and 116 as shown in FIGS. 9A and 10A. Accordingly, the first opening 111 communicates with the first suction port 27a to suck the fluid, and the remaining second and third suction ports 27b and 27c are closed. On the contrary, when the driving shaft 13 rotates in the clockwise direction, as shown in FIGS. 9B and 10B, the protrusions 115 and 124 are caught by the other end of the grooves 123 and 116, and thus the first opening ( 111 and the second opening 112 open the third suction port 27c and the second suction port 27b together to suck the fluid. The first suction port 27a is closed by the first valve 110.

11A and 12B, the limiting means is formed in the protrusion 125 and the first valve 110 protruding radially inward from the second valve 120 and the protrusion 125. ) May be made of a cutout 117 to movably receive. By forming the cutout 117 appropriately large in such a limiting means, the gap formed between the protrusion 125 and the cutout 117 may allow the first suction port 27a and the second suction port 27b. Can open Therefore, since the grooves of the above-described limiting means are omitted, the limiting means substantially reduces the dead volume.

More specifically, when the drive shaft 13 rotates counterclockwise, one end of the protrusion 125 abuts one end of the cutout 117, as shown in Figure 11A. Therefore, the gap between the protrusion 125 and the other ends of the cutout 117 opens the first suction port 27a. In addition, when the drive shaft 13 rotates in the clockwise direction, as shown in FIG. 11B, the protrusion 125 is engaged with the cutout 117. At this time, the second opening 112 opens the second suction port 27b and at the same time, the gap formed between the protrusion 125 and the cutout 117 is the third suction port 27b as described above. Open). In this limiting means, the angle β1 between both ends of the protrusion 125 is about 10 °, and the angle β2 between both ends of the cutout 117 is preferably 30 ° to 120 °. Do.

On the other hand, as described with reference to Figure 2, the suction ports 27a, 27b, 27c is separate from the plurality of suction pipes (7a) for supplying a fluid in the fluid chamber 29 in the cylinder 21 Is connected. However, these suction pipes 7a increase the number of parts and complicate the structure. In addition, since the pressure states inside the suction pipes 7b separated from each other during operation may be different from each other, fluid may not be properly supplied into the cylinder 21. Therefore, in the present invention, as shown in Figs. 12 and 13, it is preferable that the compressor has a suction plenum 200 for preliminarily storing the fluid to be sucked.

The suction plenum 200 is in direct communication with all of the suction ports 27a, 27b, 27c to supply fluid. Thus, the suction plenum 200 is mounted to the lower portion of the lower bearing 25 adjacent to the suction ports 27a, 27b, 27c. In the drawing, the suction ports 27a, 27b, and 27c are formed in the lower bearing 25, but may be formed in the upper bearing 24 as necessary. In this case, the suction plenum 200 is formed in the upper bearing ( 24). The plenum 200 may be directly fixed to the bearing 25 by welding, and may be fastened together with the cylinder 21, upper and lower bearings 24 and 25, and the valve assembly 100 by using a fastening member. It may be. The sleeve 25d of the lower bearing 24 should be immersed in the lubricating oil under the case 1 to lubricate the drive shaft 13. Thus, the suction plenum 200 includes a through hole 200a for the sleeve. The volume of the plenum 200 is preferably 100% -400% of the volume of the fluid chamber 29 to stably supply the fluid. The suction plenum 200 is also connected with the suction tube 7 to store the fluid. In more detail, the suction plenum 200 may be connected to the suction pipe 7 through a predetermined flow path. In this case, the flow path may be formed through the cylinder 21, the valve assembly 100, and the lower bearing 25, as shown in FIG. 12. That is, the flow path includes the suction hole 21c of the cylinder 21, the suction hole 122 of the second valve, and the suction hole 25c of the lower bearing.

Such a suction plenum 200 forms a space for always storing a certain amount of fluid to buffer the pressure change of the suction fluid and to stably supply the fluid to the suction ports 27a, 27b, and 27c. The intake plenum 200 may also contain oil that separates from the stored fluid and thus assist or replace the accumulator 8.

Hereinafter, the operation of the rotary compressor according to the present invention in detail.

14A to 14C are cross-sectional views sequentially showing the action when the roller rotates counterclockwise in the rotary compressor according to the present invention.

First, in FIG. 14A, the state of each component inside the cylinder when the drive shaft 13 rotates counterclockwise is shown. First, the first suction port 27a communicates with the first opening 111, and the remaining second suction port 27b and the second suction port 27c are closed. The state of the suction ports in the counterclockwise direction has been described above with reference to FIGS. 8A, 9A, 10A, and 11A, and thus a detailed description thereof will be omitted.

In the state where the first suction port 27a is opened, the roller 22 revolves counterclockwise while rolling along the inner circumferential surface of the cylinder 21 due to the rotation of the drive shaft 13. As the roller 22 continues to revolve, as shown in FIG. 4B, the size of the space 29b is reduced and the fluid already sucked is compressed. During this process, the vane 23 divides the fluid chamber 29 into two spaces 29a and 29b to be hermetically moved up and down elastically by the elastic member 23a. At the same time, new fluid continues to be sucked into the space 29a through the first suction port 27 to be compressed in the next stroke.

When the fluid pressure in the space 29b becomes a predetermined value or more, the second discharge valve 26d (see Fig. 2) is opened. Therefore, as shown in Fig. 14C, the discharge is performed through the second discharge port 26b. As the roller 22 continues to revolve, all the fluid in the space 29b is discharged through the second discharge port 26b. After all of the fluid is discharged, the second discharge valve 26d closes the second discharge port 26c by its elasticity.

After this one stroke is finished, the roller 22 continues to revolve counterclockwise, repeating the same stroke and discharging the fluid. In the counterclockwise stroke, the roller 22 compresses the fluid while revolving from the first suction port 27a to the second discharge port 26b. As described above, since the first suction port 27a and the second discharge port 27b are positioned near the vanes 23 to face each other, the volume of the entire fluid chamber 29 during the counterclockwise stroke is used. The fluid is compressed so that a maximum compressive capacity is obtained.

15A to 15C are cross-sectional views sequentially showing the action when the roller rotates clockwise in the rotary compressor according to the present invention.

First, in FIG. 15A, the states of the respective components inside the cylinder are shown when the drive shaft 13 rotates clockwise. The first suction port 27a is closed and the second suction port 27b and the third suction port 27c communicate with the second opening 112 and the first opening 111, respectively. If the first valve 110 additionally has a third opening 113 (see FIG. 6), the third suction port 27c is in communication with the third opening 113. The state of the suction ports in this clockwise direction has been described above with reference to FIGS. 8B, 9B, 10B and 11B, and thus a detailed description thereof will be omitted.

In the state in which the second and third suction ports 27b and 27c are opened, the roller 22 is clocked while rolling along the inner circumferential surface of the cylinder 21 due to the clockwise rotation of the drive shaft 13. Start to rotate in the direction. During this initial stage of idle, the fluids sucked in until the roller 22 reaches the second suction port 27b are not compressed and are removed by the roller 22 as shown in Fig. 15A. 2 is pushed out of the cylinder 21 through the suction port 27b. Thus, the fluids begin to compress after the roller 22 passes the second suction port 27b as shown in FIG. 15B. At the same time, the space between the second suction port 27b and the vane 23, that is, the space 29b, becomes a vacuum state. However, as described above, when the idle of the roller 22 is started, the third suction port 27c communicates with the first opening 111 (or the third opening 113) to suck the fluid. Open. Therefore, the vacuum state is eliminated by the sucked fluid, and the generation of noise and power loss are suppressed.

As the roller 22 continues to revolve, the size of the space 29a is reduced and the fluid already sucked is compressed. During this compression process, the vane 23 divides the fluid chamber 29 into two spaces 29a and 29b to be hermetically moved up and down elastically by the elastic member 23a. The new fluid continues to be sucked into the space 29b through the second suction port 27b and the third suction port 27c to be compressed in the next stroke.

When the fluid pressure in the space 29a is greater than or equal to a predetermined value, the first discharge valve 26c (see FIG. 2) is opened as shown in FIG. 15C, and is discharged through the first discharge port 26a. . After all the fluid is discharged, the first discharge valve 26c closes the first discharge port 26a by its elasticity.

After this one stroke is finished, the roller 22 continues to idle clockwise, repeats the same stroke and discharges the fluid. In the counterclockwise stroke, the roller 22 compresses the fluid while revolving from the second suction port 27b to the first discharge port 26a. Thus, the fluid is compressed using only a portion of the entire fluid chamber 29 during the counterclockwise stroke and a compression capacity less than the clockwise compression capacity is obtained.

In each of the strokes described above (ie, clockwise and counterclockwise strokes), the discharged compressed fluid is spaced between the rotor 12 and the stator 11 inside the case 1 and the stator 11 and the case 1. It is moved upward through the space between the and finally discharged to the outside of the compressor through the discharge pipe (9).

On the other hand, during the operation of such a compressor, mechanical elements such as motors 11 and 12, drive shaft 13 and roller 22 move at high speed. In particular, as described above, the drive shaft 13 is repeated in alternating clockwise and counterclockwise rotations and thus is placed in a more severe operating environment in the rotary compressor of the present invention. Therefore, proper lubrication and lubrication mechanism for this are very important for the smooth operation of the compressor. The present invention provides, as the lubrication mechanism, an oil flow path configured to supply oil, ie, lubricating oil (0), to each of the drive elements described above, and is described in detail below with reference to the accompanying drawings.

16 is a front view showing an oil channel of the rotary compressor according to the present invention. 17A to 17C are views showing a first embodiment of a bearing passage included in the oil passage, and FIGS. 18A to 18C are views showing a second embodiment of the bearing passage.

As shown, the lubrication mechanism of the present invention, ie oil passage 300, is formed along drive shaft 13 and bearings 24, 25. Journals 13b and 13c of the drive shaft 13 are respectively enclosed by the upper and lower bearings 24 and 25 and substantially support radial bearings that support the load in a direction perpendicular to the central axis. Form. The collars 13d and 13e together with the bearings 24 and 25 form a thrust bearing which supports an axial load during operation. The oil flow path 300 mainly consists of an axis flow path 310 (hereinafter referred to as “first flow path”) formed in the driving shaft 13.

More specifically, the first flow path 310 is formed from the lower end to the upper end of the drive shaft 13 to substantially penetrate the drive shaft 13 along its longitudinal direction. In addition, an oil pump 311 is mounted at a lower end of the first flow passage 310. The oil pump 310 is a kind of centrifugal pump and includes an oil pickup 311a and a propeller 311b inserted into the pickup 311a. The oil pump 310 is immersed in the lubricating oil, that is, oil 0, received on the bottom of the compressor (see FIG. 1), so that the oil can flow into the first flow path 130 through the oil pump 150. do. Thereafter, the oil is pumped along the first flow path 130 to be scattered to be supplied to each of the driving units at the upper end of the driving shaft 13. In addition, the first flow passage 310 further includes holes 312a and 312b formed in upper and lower portions of the eccentric portion 13a to communicate with the first flow passage 310. The oil is first supplied into the cylinder 21 through the holes 312a and 312b to lubricate the roller 22 and the eccentric portion 13a. The holes 312a and 312b also allow oil to be supplied between the upper and lower bearings 24 and 25 and the drive shaft, precisely the journals 13b and 13c.

However, as shown, the journals 13b and 13c and the bearings 24 and 25 form fairly large friction surfaces, so that the oil can be supplied with the holes 312a and 312b only by the supply of the friction surfaces. It does not reach the ends. That is, the oil does not spread evenly on the friction surfaces, and does not form an overall oil film for preventing abrasion. In order to solve this problem, in the present invention, the oil passage 300 is formed in at least one of the bearings 24 and 25 as shown in FIGS. 16, 17A-17B, and 18A-18B. It has a bearing flow path 320 (hereinafter referred to as "second flow path"). The second flow path 320 is actually formed as a groove formed in the inner circumferential surface of any one of the bearings. In addition, the second flow path 320 communicates with the driving shaft 13, one of the holes 312a and 312b adjacent to receive oil from the first flow path 310. In addition, the second flow path 320 preferably extends continuously between the upper end and the lower end of the inner circumferential surface. Therefore, the oil is supplied to one of the holes 312a and 312b into the second flow path 320, and flows between both ends of the inner circumferential surface along the second flow path 320. That is, due to the second flow path 320, the oil flow path 300 is configured to uniformly flow oil between the bearings 24 and 25 and the drive shaft 13. The oil then spreads evenly on the friction surfaces from the second flow path 320 to form an oil film as a whole to effectively prevent abrasion. The second flow path 320 is preferably formed at least in the upper bearing 24. This is because in the lower bearing 24, the oil can flow down to some extent from the hole 312b by gravity. However, it is more preferable that the second flow paths 320 are formed in both the upper and lower bearings 24 and 25 for more stable lubrication.

As described above, since the drive shaft 13 rotates clockwise and counterclockwise, the second flow path 320 should be able to flow oil in both rotation directions of the drive shaft 13. The second flow path 320 may be formed as a spiral groove, which extends the flow path to enable sufficient oil supply. However, the spiral groove can flow oil only in one rotational direction of the drive shaft 13 due to its geometric characteristics. That is, the spiral groove may flow and raise oil therein only when the spiral groove extends in a direction opposite to the rotational direction of the drive shaft 13. Therefore, in the first embodiment, the second flow path 320 has the form of a single straight groove, as shown in FIGS. 17A and 17B. Unlike the spiral grooves, the linear grooves are not affected by the geometric characteristics, and may flow oil by centrifugal force generated by the drive shaft 13 regardless of the rotation direction. On the other hand, as a second embodiment, the second flow path 320 may be composed of two first and second spiral grooves (320a, 320b) as shown in Figs. 18A and 18B. More specifically, as described above, the spiral groove can flow oil only for one direction rotation of the drive shaft 13. Thus, two separate spiral oil grooves corresponding to each direction of rotation are applied to the present invention, which extend in opposite directions (clockwise and counterclockwise). In addition, when the first and second spiral grooves 320a and 320b cross each other on the inner circumferential surfaces of the bearings 24 and 25, oil flows out of the other spiral groove while flowing to one spiral groove. This outflow prevents the bearings 24 and 25 and the journals 13b and 13c from being entirely lubricated, so it is important for optimum lubrication that the oil grooves 243a and 243b do not cross each other.

Meanwhile, referring to FIGS. 17A and 18A, a gap C having a predetermined size is formed between the bearings 24 and 25 and the driving shaft 13, and precisely the journals 13b and 13c, and the oil Is filled to form an oil film between the gaps C using the second flow path 320. During operation of the compressor, the drive shaft 13 is subjected to pressure from the compressed working fluid and rotates eccentrically a predetermined distance from the center 0 of the bearings 24, 25. In addition, since the second flow path 320 continuously breaks the inner circumferential surfaces of the bearings 24 and 25 in the longitudinal direction, the gap C is increased in the second flow path 320 and the increased gap ( Due to C), a sufficient oil film is not formed around the second flow path. Therefore, when the second flow path 320 is located where the eccentricity of the drive shaft 13 is large, the drive side 13 may contact the inner circumferential surfaces of the bearings 24 and 25. In this case, wear of the bearings 24 and 25 and the drive shaft 13 may occur, and noise may be generated during operation of the compressor. In addition, power loss of the drive shaft 13 may also occur due to excessive friction. Therefore, the second flow path 320 is preferably formed at a position where the eccentricity of the drive shaft 13 is less generated. More specifically, the second flow path 320 is formed on a part of the inner circumferential surfaces of the bearings 24 and 25 facing the position where the eccentricity of the drive shaft 13 is less generated.

In the present invention, the optimum position of the second flow path 320 has been determined through experiments, and FIGS. 17C and 18C show the optimum positions for the first and second embodiments of the second flow path 320, respectively. The experimental results considered are shown.

As shown, FIGS. 17C and 18C are graphs showing the change in the eccentricity with respect to the angle from the vanes. First the angle is set to 0 ° at the position of the vane 23 beneath the bearings 24, 25, and is increased in the direction of rotation which produces the maximum compressive capacity. In the above experiment, the compressor was set such that the maximum compression capacity was generated at the counterclockwise rotation, and thus the angle was also set to increase along the counterclockwise direction as shown. The eccentricity ratio is defined as the ratio of the eccentric distance of the drive shaft with respect to the gap C (ie, the distance from the bearing center 0 to the drive shaft center). This eccentricity is substantially a dimensionless value indicating how close the drive shaft 13 is to the inner circumferential surfaces of the bearings 24 and 25 during rotation. That is, since the gap C has a constant value, the large eccentricity means that the drive shaft 13 has a large eccentricity and is close to the inner circumferential surfaces of the bearings 24 and 25. The eccentricities were also measured for both maximum and minimum compression capacities as shown. The eccentricity of the maximum capacity was measured at the counterclockwise rotation of the drive shaft 13 as mentioned above, and the eccentricity of the minimum capacity was measured at the clockwise rotation of the drive shaft 13. The eccentricity of the maximum capacity and the minimum capacity has different phases due to the difference in the compression capacity, the difference in the compression capacity according to the direction of rotation, and various operating conditions. As a result, the eccentricity did not change much according to the specification of the compressor and showed almost the same tendency.

First, as shown in Figs. 17C and 18C, since the working fluid is compressed to maximum near the vane 23, the eccentricities are 0 ° (360 °), i.e., for both the maximum and minimum compression capacities. The vanes 23 have a relatively high value. In addition, when the second flow path 320 is formed directly on the vane 23, a working fluid having the highest pressure near the vane 23 may leak into the second flow path 320. In view of these conditions, it is preferable that the second flow path 320 is basically spaced clockwise or counterclockwise from the vane 23 with respect to the center 0.

More specifically, in the first embodiment of the second flow path 320, as shown in FIG. 17C, the eccentricities of the maximum and minimum capacities have a relatively small value at 170 ° -210 °. Thus, the single straight groove according to the first embodiment is preferably spaced apart from the vane 23 by an angle A having a range of 170 ° -210 ° in the counterclockwise direction. In addition, the compressor may be designed to have a maximum compression capacity in a clockwise direction as needed. (Minimum compression capacity in a counterclockwise direction.) In such a case, the angle is a clockwise direction in which the angle is the direction of the maximum compression capacity as opposed to FIG. 17A. It can be understood that the same experimental result as in Fig. 17C is obtained when it is set to increase by. Thus, the single straight groove may be spaced 170 ° -210 ° clockwise or counterclockwise from the vane 23. In addition, the eccentricities have a low value equal to each other at 190 °. That is, the possibility of contact is minimized in both the clockwise and counterclockwise rotation. Therefore, it is most preferable that the said angle A is 190 degrees. In the second embodiment, since the second flow path 320 is formed of the first and second spiral grooves 320a and 320b, the second flow path 320 has relatively low eccentricity so as not to interfere with each other. It is important to place within angular ranges. Referring to Fig. 18C, the eccentricities have relatively low values in the angular ranges around 190 °. Accordingly, as shown, the first and second spiral grooves 320a and 320b are spaced apart from the vane 23 at an angle B1 and an angle B2 in a clockwise or counterclockwise direction, respectively. The angles B1 and B2 have ranges of 130 ° -190 ° and 190 ° -250 °, respectively. In addition, the second flow path 320 should have an appropriate width (w) and depth (d) so as to allow a sufficient amount of oil to flow while reducing breakage of the inner circumferential surface of the bearing which prevents oil film formation. This width w and depth d may vary slightly depending on the specifications of the compressor, but are preferably 3.8 mm and 1.67 mm, respectively.

Meanwhile, in order to flow the oil more sufficiently between the bearings 24 and 25 and the drive shaft, the oil channel 300 may include the auxiliary channel 330 as shown in FIGS. 17A, 18A, and 19A-19B. It may further include. This auxiliary flow path 330 consists of grooves formed substantially along the journals 13b and 13c, and preferably extends over the length of the journals 13b and 13c. Similarly, it is preferable that the auxiliary flow passage 330 can flow oil in all directions of rotation of the drive shaft 13. Thus, as shown in FIGS. 17A and 19A, the auxiliary flow path 330 may consist of a single straight groove or two spiral grooves 330a and 330b as shown in FIGS. 18A and 19B.

Although several embodiments have been described above, it will be apparent to those skilled in the art that the present invention may be embodied in many other forms without departing from the spirit and scope thereof. Accordingly, the described embodiments are to be considered as illustrative and not restrictive, and all embodiments within the scope of the appended claims and their equivalents are included within the scope of the present invention.

The rotary compressor according to the present invention can compress the fluid in any direction in which the drive shaft rotates and has compression capacities that vary according to the rotational direction of the drive shaft. Furthermore, the rotary compressor of the present invention has a properly arranged suction and discharge ports and a simple valve assembly that selectively opens the suction ports in the rotational direction to compress the fluid using the entire designed fluid chamber. have. In addition, the compressor of the present invention has a lubrication mechanism for uniformly supplying oil between the drive shaft and its bearing in the harshest kinetic environment to support the aforementioned variable capacity mechanism. This rotary compressor of the present invention provides the following effects.

First, in the prior art, various devices were combined to realize double capacity compression. For example, two compressors and inverters with different compression capacities were combined for double compression capacities. In this case, the structure is considerably complicated and the unit price has to rise. However, the present invention can implement double capacity compression with only one compressor. In particular, the present invention can realize double capacity compression by changing only minimal components in the conventional rotary compressor.

Second, conventional compressors having a single compression capacity could not produce a compression capacity suitable for various operating conditions such as an air conditioner or a refrigerator. In this case, power consumption was inevitably wasted more than necessary. However, the present invention can produce a suitable compression capacity corresponding to the operating conditions of the device.

Third, the rotary compressor of the present invention uses the entire predesigned fluid chamber in producing the double compression capacity. This means that the compressor of the invention has at least the same compression capacity as a conventional rotary compressor having the same cylinder size, ie the same fluid chamber size. That is, the rotary compressor of the present invention can replace the conventional rotary compressor without changing the design of the basic parts such as the cylinder size. Therefore, the rotary compressor of the present invention can be freely applied to the required system without considering the compression capacity and increasing the production cost.

Fourth, in the rotary compressor of the present invention, a uniform oil film is formed on the drive shaft and the bearing by the lubrication mechanism described above. Thus, wear of the drive shaft is effectively prevented even under severe operating environments. This lubrication mechanism also flows oil in all rotational directions of the drive shaft, and is formed at a location where less eccentricity of the drive shaft occurs. Therefore, the lubrication for preventing the drive shaft wear becomes more stable and more efficient.

Claims (71)

A drive shaft rotatable clockwise and counterclockwise and having an eccentric portion of a predetermined size; A cylinder forming an internal volume of a predetermined size; A roller rotatably installed on an outer circumferential surface of the eccentric portion so as to be in contact with the inner circumferential surface of the cylinder, for rolling along the inner circumferential surface and forming a fluid chamber together with the inner circumferential surface for suction and compression of the fluid; A vane elastically installed on the cylinder to continuously contact the roller; Upper and lower bearings respectively installed on upper and lower portions of the cylinder to rotatably support the driving shaft, and seal the internal volume; An oil channel configured to uniformly flow oil between the bearings and the drive shaft; Discharge ports communicating with the fluid chamber; Suction ports communicating with the fluid chamber and spaced apart from each other by a predetermined angle; And It consists of a valve assembly for selectively opening any one of the suction ports according to the rotational direction of the drive shaft, the compression spaces of different sizes are formed in the fluid chamber in accordance with the rotational direction of the drive shaft two different compression capacity Rotary compressor having a. The method of claim 1, And the roller compresses the fluid using the entire fluid chamber only when the drive shaft is rotated in any one direction. The method of claim 1, And the roller compresses the fluid using a portion of the fluid chamber when the drive shaft rotates in the other direction. The method of claim 1, And the discharge port includes first and second discharge ports positioned to face each other with respect to the vane. The method of claim 1, And said suction port comprises a first suction port positioned near said vane and a second suction port spaced at a predetermined angle from said first suction port. The method of claim 5, The suction port is a circular compressor, characterized in that the circular. The method of claim 5, The suction port is a rotary compressor, characterized in that the rectangular. The method of claim 7, wherein And said suction port has a predetermined curvature. The method of claim 6, The suction port has a diameter of 6mm-15mm rotary compressor. The method of claim 5, And the first suction port is spaced approximately 10 ° clockwise or counterclockwise from the vane. The method of claim 5, And the second suction port is located within a range of 90 ° -180 ° from the vane to face the first suction port. The method according to claim 1 or 5, The valve assembly comprises a first valve rotatably installed between the cylinder and the bearing and a second valve for guiding the rotational movement of the first valve. The method of claim 12, And the first valve is formed of a disc member which contacts the eccentric portion of the drive shaft and rotates in the rotational direction of the drive shaft. The method of claim 13, The diameter of the first valve is larger than the inner diameter of the cylinder, the rotary compressor. The method of claim 13, The thickness of the first valve is a rotary compressor, characterized in that 0.5mm-5mm. The method of claim 12, The first valve may include a first opening communicating with the first suction port when the driving shaft is rotated in one direction and a second opening communicating with the second suction port when the driving shaft is rotated in another direction. Rotary compressor. The method of claim 16, And the first and second openings are circular or polygonal. The method of claim 16, And the first and second openings are cutouts. The method of claim 16, And the first and second openings are rectangular with a predetermined curvature. The method of claim 17, The diameter of the first and second openings is 6mm-15mm rotary compressor. The method of claim 16, The first and second openings are positioned adjacent to an outer circumferential surface of the first valve. The method of claim 12, The first valve comprises a through hole through which the drive shaft is inserted. The method of claim 12, And the second valve is fixed between the cylinder and the bearing and has a seat to receive the first valve. The method of claim 23, And the thickness of the second valve is equal to the thickness of the first valve. The method of claim 16, The suction port further comprises a third suction port located between the second suction port and the vane. The method of claim 25, And the third suction port is spaced apart from the vane by 10 ° in a clockwise or counterclockwise direction opposite the first suction port. The method of claim 25, The first valve further comprises a third opening for opening the third suction port at the same time as the opening of the second suction port. The method of claim 25, And the first valve includes a first opening that opens the third suction port simultaneously with opening the second suction port. The method of claim 12, And said valve assembly further comprises means for controlling the rotational angle of said first valve to accurately open said suction port in each direction of rotation. The method of claim 29, And the control means includes a curved groove having a predetermined length formed in the first valve and a stopper formed on the bearing and inserted into the groove. The method of claim 30, And the groove is located adjacent to the center of the first valve. The method of claim 30, The thickness of the stopper is a rotary compressor, characterized in that the same as the thickness of the first valve. The method of claim 30, The width of the stopper is equal to the width of the groove. The method of claim 30, The angle between the two ends of the groove is a rotary compressor, characterized in that 30 °-120 °. The method of claim 29, And said control means comprises a projection projecting radially on said first valve and a groove formed in said second valve and movably receiving said projection. The method of claim 29, The control means includes a rotary protrusion protruding in the radial direction to the second valve and a groove formed in the first valve, the groove for receiving the protrusion movably. The method of claim 29, The control means is a rotary compressor, characterized in that the cut-out portion for movably receiving the projection formed in the radial direction inwardly in the second valley portion and the projection formed in the first valley portion. The method of claim 37, A gap formed between the protrusion and the cutout opens the first suction port or the third suction port in a rotational direction. The method of claim 37, An angle between both sides of the protrusion is 10 °-90 ° rotary compressor, characterized in that. The method of claim 37, The angle between the two ends of the incision is a rotary compressor, characterized in that 30 °-120 °. The method of claim 1, And a plurality of suction tubes which supply fluid to be compressed to the cylinder and are individually connected to the suction ports. The method of claim 1, And a suction plenum connected to the suction ports and preliminarily storing fluid to be compressed. The method of claim 42, And the suction plenum contains oil separated from the stored fluid. The method of claim 42, And said suction plenum is mounted to a lower portion of said bearing adjacent said suction port. The method of claim 42, The volume of the suction plenum is 100% -400% of the volume of the fluid chamber. The method of claim 42, The suction plenum is a rotary compressor, characterized in that connected to the suction pipe for supplying the fluid to be compressed through a predetermined flow path. The method of claim 46, And the flow passage passes through the cylinder, the valve assembly and the lower bearing. The method of claim 12, And the first valve includes a single opening communicating with the first suction port when the drive shaft is rotated in one direction and communicating with the second suction port when the drive shaft is rotated in another direction. The method of claim 1, Wherein the oil flow path is configured to flow oil between the bearings and the drive shaft in both clockwise and counterclockwise rotation. The method of claim 1, And the oil flow path is formed in the bearing and includes at least one linear groove for flowing oil regardless of rotation directions of the drive shaft. The method of claim 1, And the oil flow path comprises first and second spiral grooves formed in the bearing and configured to flow oil in a corresponding rotational direction of the drive shaft. The method of claim 51, wherein And the first and second spiral grooves extend in opposite directions. The method of claim 51, wherein Rotary compressor, characterized in that the first and second spiral grooves do not cross each other. The method of claim 1, And the oil flow path is provided at the bearing and is formed at a position in which the eccentricity of the drive shaft is less generated. The method of claim 1, And the oil flow path is formed in the bearing so as to be spaced apart from the vane by a predetermined angle in a clockwise or counterclockwise direction with respect to the central axis of the drive shaft. 51. The method of claim 50, The linear groove is spaced apart from the vane in a clockwise or counterclockwise direction by 170 ° -210 °. The method of claim 56, wherein And the straight groove is spaced 190 degrees clockwise or counterclockwise from the vane. The method of claim 51, wherein And the first and second spiral grooves are spaced apart from the vane in a clockwise or counterclockwise direction by 130 ° -190 ° and 190 ° -250 °, respectively. The method of claim 50 or 51, And the width of the oil passage is 3.8 mm. The method of claim 50 or 51, The depth of the bearing flow path is a rotary compressor, characterized in that 1.67 mm. The method of claim 1, The oil flow path is a rotary compressor, characterized in that consisting of a bearing flow path formed in at least one of the bearings. 62. The method of claim 61, And the bearing passage is formed at least in the upper bearing. 62. The method of claim 61, The bearing passage is formed on the inner circumferential surface of the bearing. 62. The method of claim 61, The bearing passage is a rotary compressor, characterized in that extending continuously from the bottom to the top of the bearing. 62. The method of claim 61, The bearing passage is a rotary compressor, characterized in that the oil is supplied from the drive shaft. The method of claim 1, The oil flow path further comprises a shaft flow path formed in the drive shaft and configured to supply oil to the driving portions of the compressor. The method of claim 1, The oil flow path further comprises a secondary flow path formed in at least one of the journals of the drive shaft. The method of claim 67 wherein The auxiliary passage is formed on the outer circumferential surface of the journal. The method of claim 67 wherein And the auxiliary flow passage is configured to flow oil between the bearings and the drive shaft in both clockwise and counterclockwise rotation. The method of claim 67 wherein And the auxiliary flow passage comprises at least one linear groove for flowing oil regardless of the rotational directions of the drive shaft. The method of claim 67 wherein And the auxiliary flow passage comprises first and second spiral grooves configured to flow oil in a corresponding rotational direction of the drive shaft.

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