KR20060086260A - Rotary compressor - Google Patents

Abstract

The present invention discloses a rotary compressor with an improved lubrication mechanism. To this end, the present invention is a drive shaft 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, and having a rolling motion along the inner circumferential surface; 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 support the driving shaft in an improved manner and to seal the internal volume; And a first flow passage extending in the drive shaft and configured to pump oil. And at least one second flow path formed in one of the bearings and configured to uniformly flow the pumped oil between the bearing and the drive shaft.

Description

Rotary compressors {ROTARY COMPRESSOR}

The present invention relates to a rotary compressor, and more particularly to a mechanism for supplying lubricant to each drive 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. On the other hand, 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. Therefore, the rotary compressor further has an advantage of generating less vibration and noise, and has recently been used more widely than the reciprocating type especially in home appliances.

In such rotary compressors, drive parts such as motors and drive shafts move at high speeds and are subject to harsh operating environments. Therefore, proper lubrication is important for such drive parts, and a lubrication mechanism is required to enable such lubrication.

The present invention has been made to solve the above-mentioned problems, and an object of the present invention is to provide a rotary compressor having an improved lubrication mechanism.

A drive shaft having an eccentric portion of a predetermined size to achieve the above object; 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, and having a rolling motion along the inner circumferential surface; 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; And a first flow passage extending in the drive shaft and configured to pump oil. And an oil passage formed in one of the bearings, the oil passage including at least one second flow passage configured to uniformly flow the pumped oil between the bearing and the drive shaft.

The second flow path may include at least one linear groove for flowing oil regardless of the rotational direction of the drive shaft. In addition, the second flow path may be formed of at least one spiral groove configured to flow oil therein in the rotational direction of the drive shaft, in which case the second flow path extends in a direction opposite to the rotational direction of the drive shaft. do.

Preferably, the second flow path is formed at a position where the eccentricity of the drive shaft is small, and is basically formed in the bearing to be spaced apart from the vane by a predetermined angle in a clockwise or counterclockwise direction.

Preferably, the second flow path is spaced apart 210 ° -270 ° clockwise or counterclockwise from the vane, and more preferably 240 ° spaced clockwise or counterclockwise from the vane. It is also appropriate that the width of the second flow path is 3.8 mm and the depth is 1.67 mm.

The second flow path is formed at least in the upper bearing. In addition, the second 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. In addition, the second flow path is configured to receive oil from the first flow path, and more specifically, communicates with the first flow path. To this end, the first flow path is formed in the drive shaft, and includes at least one hole connecting the first flow path and the second flow path.

The first flow path is configured to scatter oil to the drive of the compressor. Preferably, the first flow path is continuously formed from the lower end to the upper end of the drive shaft, and penetrates the drive shaft in the longitudinal direction thereof.

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. Specifically, the auxiliary flow passage consists of at least one straight groove or spiral groove.

By means of the invention described above the corresponding drive in the rotary compressor is properly lubricated during operation.

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 to which an oil passage according to the present invention is applied;

FIG. 2 is a cross-sectional view showing the inside of a cylinder obtained along the line I-I in FIG. 1; FIG.

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

4 is a cross-sectional view taken along line II-II of FIG. 3 showing a second flow path;

5A and 5B are partial plan views respectively showing an inner circumferential surface of a bearing including embodiments of the second flow path;

6 is a graph showing an optimum setting angle of a second flow path; And

7A and 7B 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 partial longitudinal cross-sectional view showing a dual displacement rotary compressor in which an oil channel according to the present invention is applied, and FIG. 2 is a cross-sectional view showing the inside of a cylinder obtained along the line I-I of FIG. 1.

First, as shown in FIG. 1, the rotary compressor according to the present invention includes a case 1, a power generation 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 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 is a stator 11 fixed to the case 1, a rotor 12 rotatably supported inside the stator 11, and the rotor 12 is press-fitted The drive shaft 13 is included. 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 is largely fixed to the cylinder 21 fixed to the case 1, the roller 22 (see Fig. 2) located inside the cylinder 21, and the upper and lower portions of the cylinder 21, respectively It consists of upper and lower bearings 24 and 25 installed.

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 inner volume as shown in FIG. 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 21a extending to a predetermined depth from its inner circumferential surface. The grooves 21a 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. 2, 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 center of rotation. 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 21a of the cylinder 21 as mentioned above. In addition, an elastic member 23a is installed in the groove 21a 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 23 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. For example, when the roller 22 rotates counterclockwise, one space 29b gradually shrinks while the other space 29a 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 serve as suction chambers for sucking fluid and compression chambers for compressing fluid, respectively, during rotation of the drive shaft 13. Therefore, 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 new suction of the fluid. do.

The upper bearing 24 and the lower bearing 25 are installed above and below the cylinder 21 as shown in FIG. 1 and rotatably support the drive shaft 13. By using fastening members such as bolts and nuts, the cylinder 21 and the upper and lower bearings 24 and 25 are firmly fastened to each other so that the cylinder inner volume, in particular the fluid chamber 29, is sealed.

In addition, although not shown, the compression unit 20 is provided with one suction port and one discharge port. The suction port is connected to the suction pipe 7 shown in FIG. 1 and is formed in either the cylinder 21 or the bearing, and a working fluid is supplied to the fluid chamber 29 through the suction port. In addition, the discharge port is generally formed in any one of the bearings, and is opened by a valve to discharge the compressed working fluid. Therefore, once the working fluid is supplied through the suction port, the roller 22 compresses the working fluid while rolling along the inner circumferential surface of the cylinder from the suction port side to the discharge port side. Thereafter, the compressed working fluid is discharged through the discharge port by the operation of the valve, and this series of processes are repeatedly performed during the rotation of the drive shaft 13 to continuously generate the compressed working fluid.

On the other hand, during operation of the compressor, mechanical elements such as the motors 11, 12, the drive shaft 13, and the rollers 22 move at high speed and are placed in a harsh operating environment. 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.

3 is a front view showing an oil channel of the rotary compressor according to the present invention. 4 and 5B are views showing a second flow path included in the oil flow path, and FIG. 6 is a graph showing an optimum setting angle of the second flow path.

As shown, the lubrication mechanism of the present invention, ie the oil passage 100, is formed along the drive shaft 13 and the 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 100 mainly consists of an axis flow path 110 (hereinafter referred to as a first flow path) formed in the driving shaft 13.

More specifically, the first flow path 110 is formed from the lower end to the upper end of the drive shaft 13 to substantially penetrate the drive shaft 13 along its length direction. In addition, an oil pump 111 is mounted at the lower end of the first flow path 110. The oil pump 110 is a kind of centrifugal pump, and includes an oil pick up 111a and a propeller 111b inserted into the pick up 111a. The oil pump 111 is immersed into the lubricating oil, that is, oil 0, accommodated in the bottom of the compressor (see FIG. 1), so that oil may flow into the first flow path 110 through the oil pump 111. do. Thereafter, the oil is pumped along the first flow path 130 and scattered so as to be supplied to the eastern part at the upper end of the drive shaft 13. In addition, the first flow path 110 further includes holes 112a and 112b formed in upper and lower portions of the eccentric portion 13a to communicate with the first flow path 110. The oil is first supplied into the cylinder 21 through the holes 112a and 112b to lubricate the roller 22 and the eccentric portion 13a. The holes 112a and 112b 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 is only supplied through the holes 112a and 112b. It does not reach the ends. That is, the oil does not spread evenly on the friction surfaces, and does not form an entire oil film for preventing abrasion. In order to solve this problem, in the present invention, the oil passage 100 is formed in at least one of the bearings 24 and 25 as shown in FIGS. 3 and 4-5B. (Hereinafter referred to as a second flow path). The second flow path 120 is actually formed as a groove formed in the inner circumferential surface of any one of the bearings. The second flow path 120 communicates with the drive shaft 13, either one of the holes 112a and 112b adjacent to receive oil from the first flow path 110. In addition, the second flow path 120 preferably extends continuously between the upper end and the lower end of the inner circumferential surface. Accordingly, the oil is supplied to the second flow path 120 from one of the holes 112a and 112b and flows between both ends of the inner circumferential surface along the second flow path 120. That is, due to the second flow path 120, the oil flow path 100 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 120 to form an oil film as a whole to effectively prevent abrasion. The second flow path 120 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 112b by gravity. However, it is more preferable that the second flow paths 120 are formed in both the upper and lower bearings 24 and 25 for more stable lubrication.

The second flow path 120 may be formed as a spiral groove as shown in FIG. 5B, which spiral groove expands the flow path to allow 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. For this reason, the second flow path 120 preferably has the form of a single straight groove, as shown in FIG. 5A. 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.

Meanwhile, referring to FIG. 4, 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 formed in the first oil. The two flow paths 120 are used to form an oil film between the gaps C. During operation of the compressor, the drive shaft 13 is pressurized from the compressed working fluid and rotates eccentrically a predetermined distance from the center O of the bearings 24 and 25. In addition, since the second flow path 120 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 120 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 120 is positioned where the eccentricity of the driving shaft 13 is large, the driving 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 120 is preferably formed at a position where the eccentricity of the drive shaft 13 is less generated. More specifically, the second flow path 120 is formed on a portion 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 120 has been determined through experiments, and FIG. 6 shows experimental results considered for the optimal position of the second flow path 120.

As shown, Figure 6 is a graph showing the change in the eccentricity with respect to the angle from the vanes. The angle is first set at 0 ° at the position of the vanes 23 beneath the bearings 24, 25, and increases in the direction of rotation of the drive shaft 13. In this experiment the compressor was set to rotate counterclockwise, so that 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 O to the center of the drive shaft). 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. As a result, the eccentricity did not change much according to the specification of the compressor and showed almost the same tendency.

First, since the working fluid is maximally compressed near the vane 23 and exerts a large pressure on the drive shaft 13, the eccentricities have a relatively high value at 0 ° (360 °), ie the vane 23. In addition, when the second flow path 120 is formed directly on the vane 23, the working fluid may leak into the second flow path 120 due to its high pressure near the vane 23. In view of these conditions, it is preferable that the second flow path 120 is basically spaced clockwise or counterclockwise from the vane 23 with respect to the center O.

More specifically, as shown, the eccentricities have relatively low values at 210 ° -270 °. Accordingly, the second flow path 120 is preferably spaced apart from the vane 23 at an angle A having a range of 210 ° to 270 ° in the counterclockwise direction. In addition, the compressor may be designed to compress the working fluid in a clockwise direction as necessary. Even in this case, it is understood that the same experimental result as in FIG. 4 is obtained when the angle is set to increase clockwise. Therefore, the second flow path 120 may be spaced apart from the vanes 23 at 210 ° -270 ° in a clockwise or counterclockwise direction. In addition, the eccentricities have the lowest value at 240 °. In other words, the possibility of contact in the clockwise or counterclockwise rotation at this angle is minimized. Therefore, it is most preferable that the said angle A is 240 degrees. In addition, the second flow path 120 should have a suitable width (w) and depth (d) to allow a sufficient amount of oil to flow while reducing breakage of the inner circumferential surface of the bearing which prevents the formation of the oil film. 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.

On the other hand, in order to allow the oil to flow more sufficiently between the bearings 24 and 25 and the drive shaft 13, the oil channel 100 is an auxiliary channel 130 as shown in Figures 4 and 7A-7B. It may further include. This auxiliary flow path 130 consists substantially of grooves formed along at least one of the journals 13b and 13c, and preferably extends over the length of the journals 13b and 13c. Similar to the second flow path 120, as shown in FIG. 7A, the auxiliary flow path 130 may be formed as a single straight groove or as a spiral groove as shown in FIG. 7B.

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.

In 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 (24)

A drive shaft 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, and having a rolling motion along the inner circumferential surface; 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; And A first flow passage extending in the drive shaft and configured to pump oil; And at least one second flow path formed in one of the bearings and configured to uniformly flow the pumped oil between the bearing and the drive shaft. The method of claim 1, And the second flow passage comprises at least one linear groove for flowing oil regardless of the rotational direction of the drive shaft. The method of claim 1, And the second flow passage comprises at least one spiral groove configured to flow oil therein in the rotational direction of the drive shaft. The method of claim 3, wherein And the second flow passage extends in a direction opposite to the rotational direction of the drive shaft. The method of claim 1, And the second flow path is formed at a position where the eccentricity of the drive shaft is less generated. The method of claim 1, And the second flow path is formed at a position where the drive shaft does not approach close. The method of claim 1, And the second flow path is formed in the bearing to be spaced apart from the vane by a predetermined angle in a clockwise or counterclockwise direction. The method of claim 1, And the second flow path is spaced apart from the vane in a clockwise or counterclockwise direction at 210 ° -270 °. The method of claim 1, And the second flow path is spaced apart from the vane in a clockwise or counterclockwise direction by 240 °. The method of claim 1, The width of the second flow path is a rotary compressor, characterized in that 3.8 mm. The method of claim 1, And the depth of the second flow path is 1.67 mm. The method of claim 1, And the second flow passage is formed in an upper bearing installed at least in the upper cylinder. The method of claim 1, And the second flow path is formed on an inner circumferential surface of the bearing. The method of claim 1, And the second flow passage extends continuously from the lower end to the upper end of the bearing. The method of claim 1, And the second flow path is configured to receive oil from the first oil. The method of claim 1, And the second flow passage communicates with the first flow passage. The method of claim 1, And the first flow path is formed in the drive shaft and includes at least one hole connecting the first flow path and the second flow path. The method of claim 1, And the first flow path is configured to scatter oil toward the driving part of the compressor. The method of claim 1, The first flow path is a rotary compressor, characterized in that extending continuously from the lower end to the upper end of the drive shaft. The method of claim 1, And the first flow passage penetrates the drive shaft in the longitudinal direction thereof. The method of claim 1, The oil channel further comprises a secondary channel formed in at least one of the journals of the drive shaft. The method of claim 21, The auxiliary passage is formed on the outer circumferential surface of the journal. The method of claim 21, And the auxiliary flow passage comprises at least one linear groove. The method of claim 21, And the auxiliary flow passage comprises at least one spiral groove.

Images