WO2006046784A1

WO2006046784A1 - Rotary compressor

Info

Publication number: WO2006046784A1
Application number: PCT/KR2004/002719
Authority: WO
Inventors: Ji Young Bae; Chul Gi Roh; Kyoung Jun Park; Chang Yong Jang; Jong Bong Kim; Young Hwan Ko
Original assignee: Lg Electronics Inc.
Priority date: 2004-10-26
Filing date: 2004-10-26
Publication date: 2006-05-04
Also published as: EP1805419A1; EP1805419B1

Abstract

Disclosed is a rotary compressor having two compression capacities. The rotary compressor includes: a driving shaft (13) being rotatable clockwise and counterclockwise, and having an eccentric portion (13a) of a predetermined size; a cylinder (21) forming a predetermined inner volume; a roller (22) installed rotatably on an outer circumference of the eccentric portion (13a) so as to contact an inner circumference of the cylinder (21), performing a rolling motion along the inner circumference and forming a fluid chamber (29) to suck and compress fluid along with the inner circumference; a vane (23) installed elastically in the cylinder (21) to contact the roller (22) continuously; upper and lower bearings (24, 25) installed respectively in upper and lower portions of the cylinder (21), for supporting the driving shaft (13) rotatably and sealing the inner volume hermetically; an oil path configured to allow oil to uniformly flow between the bearings (24, 25) and the driving shaft (13); discharge ports (26a, 26b) communicating with the fluid chamber (29); suction ports (27a-c) communicating with the fluid chamber (29) and being spaced apart from each other by a predetermined angle; and a valve assembly (100) for selectively opening any one of the suction ports (27a-c) according to rotation direction of the driving shaft (13).

Description

ROTARY COMPRESSOR

Technical Field

The present invention relates to a rotary compressor, and more particularly, to a

mechanism for changing compression capacity of a rotary compressor.

Background Art

In general, compressors are machines that are supplied power from a power

generator such as electric motor, turbine or the like and apply compressive work to a

working fluid, such as air or refrigerant to elevate the pressure of the working fluid.

Such compressors are widely used in a variety of applications, from electric home

appliances such as air conditioners, refrigerators and the like to industrial plants.

The compressors are classified into two types according to their compressing

methods: a positive displacement compressor, and a dynamic compressor (a turbo

compressor). The positive displacement compressor is widely used in industry fields

and configured to increase pressure by reducing its volume. The positive displacement

compressors can be further classified into a reciprocating compressor and a rotary

compressor.

The reciprocating compressor is configured to compress the working fluid using a

piston that linearly reciprocates in a cylinder. The reciprocating compressor has an

advantage of providing high compression efficiency with a simple structure. However, the

reciprocation compressor has a limitation in increasing its rotational speed due to the inertia of the piston and a disadvantage in that a considerable vibration occurs due to the

inertial force. The rotary compressor is configured to compress working fluid using a

roller eccentrically revolving along an inner circumference of the cylinder, and has an

advantage of obtaining high compression efficiency at a low speed compared with the

reciprocating compressor, thereby reducing noise and vibration.

Recently, compressors having at least two compression capacities have been

^■developed. These compressors have compression capacities different from each other

according to the rotation directions (i.e., clockwise direction and counterclockwise

direction) by using a partially modified compression mechanism. Since compression

capacity can be adjusted differently according to loads required by these compressors,

such a compressor is widely used to increase an operation efficiency of several

equipments requiring the compression of working fluid, especially household electric

appliances such as a refrigerator which uses a refrigeration cycle.

However, a conventional rotary compressor has separately a suction portion and a

discharge portion which communicate with a cylinder. The roller rolls from the suction

port to the discharge portion along an inner circumference of the cylinder, so that the

working fluid is compressed. Accordingly, when the roller rolls in an opposite direction

(i.e.. from the discharge port to the suction port), the working fluid is not compressed.

In other words, the conventional rotary compressor cannot have different compression

capacities if the rotation direction is changed. Accordingly, there is a need for development of a rotary compressor having variable compression capacity as well as the

aforementioned inherent advantages.

Additionally, in such compressor having a variable capacity, driving elements such

as a motor, a driving shaft and the like are under a severe operational environment due to

a variation of the compression capacity as well as a high-speed rotation. Therefore, an

appropriate lubricating mechanism should be developed together with a variable capacity

mechanism.

Disclosure of Invention

Accordingly, the present invention is directed to a rotary compressor that

substantially obviates one or more problems due to limitations and disadvantages of the

related art.

An object of the present invention is to provide a rotary compressor in which the

compressing stroke is possibly performed to both of the clockwise and counterclockwise

rotations of a driving shaft.

Another object of the present invention is to provide a rotary compressor of which

compression capacity can be varied.

Still another object of the present invention is to provide a rotary compressor

having a lubricating mechanism adapted for a capacity variation mechanism.

Additional advantages, objects, and features of the invention will be set forth in

part in the description which follows and in part will become apparent to those having ordinary skill in the art upon examination of the following or may be learned from

practice of the invention. The objectives and other advantages of the invention may be

realized and attained by the structure particularly pointed out in the written description

and claims hereof as well as the appended drawings.

To achieve these objects and other advantages and in accordance with the purpose

of the invention, as embodied and broadly described herein, a rotary compressor includes:

a driving shaft being rotatable clockwise and counterclockwise, and having an eccentric

portion of a predetermined size; a cylinder forming a predetermined inner volume; a roller

installed rotatably on an outer circumference of the eccentric portion so as to contact an

inner circumference of the cylinder, performing a rolling motion along the inner

circumference and forming a fluid chamber to suck and compress fluid along with the

inner circumference; a vane installed elastically in the cylinder to contact the roller

continuously; upper and lower bearings installed respectively in upper and lower portions

of the cylinder, for supporting the driving shaft rotatably and sealing the inner volume

hermetically; an oil path configured to allow oil to uniformly flow between the bearings

and the driving shaft; discharge ports communicating with the fluid chamber; suction

ports communicating with the fluid chamber and being spaced apart from each other by a

predetermined angle; and a valve assembly for selectively opening any one of the suction

ports according to rotation direction of the driving shaft, wherein compression spaces that

have different volumes from each other are formed in the fluid chamber according to the

rotation direction of the driving shaft such that two different compression capacities are formed.

The discharge ports comprise a first discharge port and a second discharge port

that are positioned facing each other with respect to the vane.

The suction ports comprise a first suction port positioned in the vicinity of the

vane and a second suction port positioned spaced apart from the first suction port by a

predetermined angle. In addition, the suction ports further comprises a third suction port

positioned between the second suction port and the vane.

The valve assembly comprises a first valve installed rotatably between the

cylinder and the bearing and a second valve guiding a rotary motion of the first valve.

Firstly, the first valve comprises a disk member contacting the eccentric portion of the

driving shaft and rotating in the rotation direction of the driving shaft. The first valve

comprises a first opening communicating with the first suction port when the driving shaft

rotates in any one of the clockwise direction and the counterclockwise direction and a

second opening communicating with the second suction port when the driving shaft

rotates in the other of the clockwise direction and the counterclockwise direction.

Preferably, the first valve further comprises a third opening for opening the third suction

port simultaneously with opening the second suction port. Alternatively, the first valve

comprises the first opening for opening the third suction port simultaneously with

opening the second suction port. Additionally, the second valve is fixed between the

cylinder and the bearing and comprises a seat portion for receiving the first valve.

The valve assembly further comprises means for controlling a rotation angle of the first valve such that corresponding suction ports are opened accurately. The control means

comprises a curved groove formed at the first valve and having a predetermined length

and a stopper formed on the bearing and inserted into the curved groove.

Alternatively, the control means comprises a projection formed on the first valve

and projecting in a radial direction of the first valve and a groove formed on the second

valve, for receiving the projection movably. Alternatively, the control means comprises a

projection formed on the second valve and projecting in a radial direction of the second

valve and a groove formed on the first valve, for receiving the projection movably.

Alternatively, the control means comprises a projection formed on the second valve and

projecting toward a center of the second valve and a cut-away portion formed on the first

valve, for receiving the projection movably.

The oil path is configured to allow the oil to flow between the driving shaft and

the bearings in both of clockwise and counterclockwise rotations of the driving shaft.

Specifically, the oil path comprises a single straight groove formed at any one of the

bearings and allowing the oil to flow therein regardless of rotational directions of the

driving shaft. Alternatively, the oil path comprises first and second helical grooves formed

at any one of the bearings and each configured to allow the oil to flow therein in

corresponding rotation of the driving shaft. The first and second helical grooves extend in

opposite directions and do not intersect with each other.

Preferably, the oil path is provided to any one of the bearings and is positioned

where an eccentricity of the driving shaft is small. Substantially, the oil path is formed at any one of the bearings to be spaced apart from the vane in clockwise or

counterclockwise direction. In this case, the single straight groove is preferably spaced

apart from the vane by a range of 170°-210^° in clockwise or counterclockwise direction,

and more preferably spaced apart from the vane by 190° in clockwise or counterclockwise

direction. It is preferable that the first and second helical grooves are spaced apart from

the vane respectively by 130°-190° and 190^°-250° in clockwise or counterclockwise

direction.

The oil path substantially comprises a bearing path formed at any one of the

bearings. The bearing path is formed at the upper bearing al least. In addition, the bearing

path is formed on an inner circumferential surface of the bearing and continuously

extends from an upper end to a lower end of the bearing.

Preferably, the oil path further comprises an auxiliary path formed at any one of

journals in the driving shaft. The auxiliary path is formed on an outer circumferential

surface of the journal. The auxiliary path is preferably configured to allow the oil to flow

between the driving shaft and the bearings in both of clockwise and counterclockwise

rotations of the driving shaft. Specifically, the auxiliary path comprises a single straight

groove allowing the oil to flow therein regardless of rotational directions of the driving

shaft. Alternatively, the auxiliary path comprises first and second helical grooves each

configured to allow the oil to flow therein in corresponding rotation of the driving shaft.

The compressor of the present invention preferably further comprises a suction

plenum for preliminarily storing fluid to be compressed, the suction plenum being connected with the suction ports. The suction plenum accommodates oil extracted from

the stored fluid, and is installed at a lower portion of the bearing in the vicinity of the

suction port. It is desirable that the suction plenum has 100 - 400 % a volume as large as

the fluid chamber.

By the invention described above, two different compression capacities are

obtained in the rotary compressor. Also, corresponding driving parts are appropriately

lubricated while operating for the two different capacities.

It is to be understood that both the foregoing general description and the following

detailed description of the present invention are exemplary and explanatory and are

intended to provide further explanation of the invention as claimed.

Brief Description of Drawings

The accompanying drawings, which are included to provide a further

understanding of the invention and are incorporated in and constitute a part of this

application, illustrate embodiment(s) of the invention and together with the description

serve to explain the principle of the invention. In the drawings:

FIG. 1 is a partial longitudinal sectional view of a rotary compressor according to

the present invention;

FIG. 2 is an exploded perspective view of the compressing unit of the rotary

compressor according to the present invention;

FIG. 3 is a sectional view of the compressing unit of the rotary compressor according to the present invention;

FIG. 4 is a cross-sectional view of the cylinder of the rotary compressor according

to the present invention;

FIGS. 5 A and 5B are plan views of the lower bearing of the rotary compressor

according to the present invention;

FIG. 6 is a plan view of the valve assembly of the rotary compressor according to

the present invention;

FIGS. 7Ato 7C are plan views of exemplary modifications of the valve assembly;

FIGS. 8 A and 8B are plan views of the control means;

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

FIGS. 9 A and 9B are plan views illustrating exemplary modifications of a

revolution limitation means of the valve assembly;

FIGS. 1OA and 1OB are plan views illustrating exemplary modifications of the

control means of the valve assembly;

FIGS. 11A and HB are plan views illustrating exemplary modifications of the

control means of the valve assembly;

FIG. 12 is an exploded perspective view of a compressing unit of a rotary

compressor including a suction plenum according to the present invention;

FIG. 13 is a cross-sectional view of the compressing unit shown in FIG. 12;

FIGS. 14A to 14C are cross-sectional views illustrating an operation of the rotary

compressor when the roller revolves in the counterclockwise direction; FIGS. 15A to 15C are cross-sectional views illustrating an operation of the rotary

compressor when the roller revolves in the clockwise direction according to the present

invention;

FIG. 16 is a front view showing an oil path of the rotary compressor according to

the present invention;

FIG. 17A is a sectional view taken along a line I-I of FIG. 16 and showing a first

embodiment of a bearing path;

FIG. 17B is a partial sectional view showing an inner circumferential surface of a

bearing which includes the first embodiment of the bearing path;

FIG. 17C is a graph showing an optimal setting angle of the first embodiment of

the bearing path;

FIG. 18A is a sectional view taken along a line I-I of FIG. 16 and showing a

second embodiment of the bearing path;

FIG. 18B is a partial sectional view showing the inner circumferential surface of

the bearing which includes the second embodiment of the bearing path;

FIG. 18C is a graph showing an optimal setting angle of the second embodiment

of the bearing path; and

FIGS. 19A and 19B are partial front view each showing an auxiliary path.

Best Mode for Carrying Out the Invention

Reference will now be made in detail to the preferred embodiments of the present invention, examples of which are illustrated in the accompanying drawings. Wherever

possible, the same reference numbers will be used throughout the drawings to refer to the

same or like parts.

FIG. 1 is a partial longitudinal sectional view illustrating structure of a rotary

compressor according to the present invention. FIG. 2 is an exploded perspective view

illustrating a compressing unit of a rotary compressor according to the present invention.

As shown in FIG. 1, a rotary compressor of the present invention includes a case 1,

a power generator 10 positioned in the case 1 and a compressing unit 20. Referring to

FIG. 1, the power generator 10 is positioned on the upper portion of the rotary compressor

and the compressing unit 20 is positioned on the lower portion of the rotary compressor.

However, their positions may be changed if necessary. An upper cap 3 and a lower cap

5 are installed on the upper portion and the lower portion of the case 1 respectively to

define a sealed inner space. A suction pipe 7 for sucking working fluid is installed on a

side of the case 1 and connected to an accumulator 8 for separating lubricant from

refrigerant. A discharge tube 9 for discharging the compressed fluid is installed on the

center of the upper cap 3. A predetermined amount of the lubricant "0" is filled in the

lower cap 5 so as to lubricate and cool members that are moving frictionally. Here, an

end of a driving shaft 13 is dipped in the lubricant.

The power generator 10 includes a stator 11 fixed in the case 1, a rotor 12

rotatable supported in the stator 11 and the driving shaft 13 inserted forcibly into the rotor

12. The rotor 12 is rotated due to electromagnetic force, and the driving shaft 13 delivers the rotation force of the rotor to the compressing unit 20. To supply external

power to the stator 20, a terminal 4 is installed in the upper cap 3.

The compressing unit 20 includes a cylinder 21 fixed to the case I₃ a roller 22

positioned in the cylinder 21 and upper and lower bearings 24 and 25 respectively

installed on upper and lower portions of the cylinder 21. The compressing unit 20 also

includes a valve assembly 100 installed between the lower bearing 25 and the cylinder 21.

The compressing unit 20 will be described in more detail with reference to FIGS. 2, 3 and

4.

The cylinder 21 has a predetermined inner volume and strength enough to endure

the pressure of the fluid. The cylinder 21 accommodates an eccentric portion 13a

formed on the driving shaft 13 in the inner volume. The eccentric portion 13a is a kind

of an eccentric cam and has a center spaced by a predetermined distance from its rotation

center. The cylinder 21 has a groove 21b extending by a predetermined depth from its

inner circumference. A vane 23 to be described below is installed on the groove 21b.

The groove 21b is long enough to accommodate the vane 23 completely.

The roller 22 is a ring member that has an outer diameter less than the inner

diameter of the cylinder 21. As shown in FIG. 4, the roller 22 contacts the inner

circumference of the cylinder 21 and rotatably coupled with the eccentric portion 13 a.

Accordingly, the roller 22 performs rolling motion on the inner circumference of the

cylinder 21 while spinning on the outer circumference of the eccentric portion 13a when

the driving shaft 13 rotates. The roller 22 revolves spaced apart by a predetermined distance from the rotation center '0' due to the eccentric portion 13a while performing the

rolling motion. Since the outer circumference of the roller 22 always contacts the inner

circumference due to the eccentric portion 13 a, the outer circumference of the roller 22

and the inner circumference of the cylinder form a separate fluid chamber 29 in the inner

volume. The fluid chamber 29 is used to suck and compress the fluid in the rotary

compressor.

The vane 23 is installed in the groove 21b of the cylinder 21 as described above.

An elastic member 23 a is installed in the groove 21b to elastically support the vane 23.

The vane 23 continuously contacts the roller 22. In other words, the elastic member 23 a

has one end fixed to the cylinder 21 and the other end coupled with the vane 23, and

pushes the vane 23 to the side of the roller 22. Accordingly, the vane 23 divides the

fluid chamber 29 into two separate spaces 29a and 29b as shown in FIG. 4. While the

driving shaft 13 rotate or the roller 22 revolves, the volumes of the spaces 29a and 29b

change complementarily. In other words, if the roller 22 rotates clockwise, the space

29a gets smaller but the other space 29b gets larger. However, the total volume of the

spaces 29a and 29b is constant and approximately same as that of the predetermined fluid

chamber 29. One of the spaces 29a and 29b works as a suction chamber for sucking the

fluid and the other one works as a compression chamber for compressing the fluid

relatively when the driving shaft 13 rotates in one direction (clockwise or

counterclockwise). Accordingly, as described above, the compression chamber of the

spaces 29a and 29b gets smaller to compress the previously sucked fluid and the suction chamber expands to suck the new fluid relatively according to the rotation of the roller 22.

If the rotation direction of the roller 22 is reversed, the functions of the spaces 29a and

29b are exchanged. In the other words, if the roller 22 revolves counterclockwise, the

right space 29b of the roller 22 becomes a compression chamber, but if the roller 22

revolves clockwise, the left space 29a of the roller 22 becomes a discharge unit.

The upper bearing 24 and the lower bearing 25 are, as shown in FIG. 2, installed

on the upper and lower portions of the cylinder 21 respectively, and rotatably support the

driving shaft 12 using a sleeve and the penetrating holes 24b and 25b formed inside the

sleeve. More particularly, the upper bearing 24, the lower bearing 25 and the cylinder

21 include a plurality of coupling holes 24a, 25a and 21a formed to correspond to each

other respectively. The cylinder 21, the upper bearing 24 and the lower bearing 25 are

coupled with one another to seal the cylinder inner volume, especially the fluid chamber

29 using coupling members such as bolts and nuts.

The discharge ports 26a and 26b are formed on the first 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 can communicate

directly with the fluid chamber 29 or can communicate with the fluid chamber 29 through

a predetermined fluid passage 21 d formed in the cylinder 21 and the first bearing 24.

Discharge valves 26c and 26d are installed on the first bearing 24 so as 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 of the chamber 29 is greater than or equal to a predetermined pressure. To achieve this, it is desirable that the discharge

valves 26c and 26d are leaf springs of which one end is fixed in the vicinity of the

discharge ports 26 and 26b and the other end can be deformed freely. Although not

shown in the drawings, a retainer for limiting the deformable amount of the leaf spring

may be installed on the upper portion of the discharge valves 26c and 26d so that the

valves can operate stably. In addition, a muffler (not shown) can be installed on the

upper portion of the first bearing 24 to reduce a noise generated when the compressed

fluid is discharged.

The suction ports 27a, 27b and 27c communicating with the fluid chamber 29 are

formed on the lower bearing 25. The suction ports 27a, 27b and 27c guide the

compressed fluid to the fluid chamber 29. The suction ports 27a, 27b and 27c are

connected to the suction pipe 7 so that the fluid outside of the compressor can flow into

the chamber 29. More particularly, the suction pipe 7 is branched into a plurality of

auxiliary tubes 7a and is connected to suction ports 27 respectively. If necessary, the

discharge ports 26a, and 26b may be formed on the lower bearing 25 and the suction ports

27a, 27b and 27c may be formed on the upper bearing 24.

The suction and discharge ports 26 and 27 become the important factors in

determining compression capacity of the rotary compressor and will be described

referring to FIGS. 4 and 5. FIG. 4 illustrates a cylinder coupled with the lower bearing

25 without a valve assembly 100 to show the suction ports 27.

First, the compressor of the present invention includes at least two discharge ports 26a and 26b. As shown in the drawing, even if the roller 22 revolves in any direction, a

discharge port should exist between the suction port and vane 23 positioned in the

revolution path to discharge the compressed fluid. Accordingly, one discharge port is

necessary for each rotation direction. It causes the compressor of the present invention

to discharge the fluid independent of the revolution direction of the roller 22 (that is, the

rotation direction of the driving shaft 13). Meanwhile, as described above, the

compression chamber of the spaces 29a and 29b gets smaller to compress the fluid as the

roller 22 approaches the vane 23. Accordingly, the discharge ports 26a and 26b are

preferably formed facing each other in the vicinity of the vane 23 to discharge the

maximum compressed fluid. In other word, as shown in the drawings, the discharge

ports 26a and 26b are positioned on both sides of the vane 23 respectively. The

discharge ports 26a and 26b are preferably positioned in the vicinity of the vane 23 if

possible.

The suction port 27 is positioned properly so that the fluid can be compressed

between the discharge ports 26a and 26b and the roller 22. Actually, the fluid is

compressed from a suction port to a discharge port positioned in the revolution path of the

roller 22. In other words, the relative position of the suction port for the corresponding

discharge port determines the compression capacity and accordingly two compression

capacities can be obtained using different suction ports 27 according to the rotation

direction. Accordingly, the compression of the present invention has first and second

suction ports 27a and 27b corresponding to two discharge ports 26a and 26b respectively and the suction ports are separated by a predetermined angle from each other with respect

to the center 0 for two different compression capacities.

Preferably, the first suction port 27a is positioned in the vicinity of the vane 23.

Accordingly, the roller 22 compresses the fluid from the first suction port 27a to the

second discharge port 26b positioned across the vane 23 in its rotation in one direction

(counterclockwise in the drawing). The roller 22 compress the fluid due to the first

suction port 27a by using the overall chamber 29 and accordingly the compressor has a

maximum compression capacity in the counterclockwise rotation. In other words, the

fluid as much as overall volume of the chamber 29 is compressed. The first suction port

27a is actually separated by an angle θl of 10° clockwise or counterclockwise from the

vane 23 as shown in FIGS. 4 and 5 A. The drawings of the present invention illustrates

the first suction port 27a separated by the angle θl counterclockwise. At this separating

angle θl, the overall fluid chamber 29 can be used to compress the fluid without

interference of the vane 23.

The second suction port 27b is separated by a predetermined angle from the first

suction port 27a with respect to the center. The roller 20 compresses the fluid from the

second suction port 27b to the first discharge port 26a in its rotation in counterclockwise

direction. Since the second suction port 27b is separated by a considerable angle

clockwise from the vane 23, the roller 22 compresses the fluid by using a portion of the

chamber 29 and accordingly the compressor has the less compression capacity than that

of counterclockwise rotary motion. In other words, the fluid as much as a portion volume of the chamber 29 is compressed. The second suction port 27b is preferably

separated by an angle Θ2 of a range of 90 - 180° clockwise or counterclockwise from the

vane 23. The second suction port 27b is preferably positioned facing the first suction

port 27a so that the difference between compression capacities can be made properly and

the interference can be avoid for each rotation direction. In the middle of the night, power

of attorney will be

As shown in FIG. 5A₅ the suction ports 27a and 27b are generally in circular

shapes whose diameters are, preferably, 6 -15 mm. In order to increase a suction amount

of fluid, the suction ports 27a and 27b can also be provided in several shapes, including a

rectangle. Further, as shown in FIG. 5B, the suction ports 27a and 27b can be in

rectangular shapes having predetermined curvature. In this case, an interference with

adjacent other parts, especially the roller 22, can be minimized in operation.

Meanwhile, in order to obtain desired compression capacity in each rotation

direction, suction ports that are available in any one of rotation directions should be single.

If there are two suction ports in rotation path of the roller 22, the compression does not

occur between the suction ports. In other words, if the first suction port 27a is opened,

the second suction port 27b should be closed, and vice versa. Accordingly, for the

purpose of electively opening only one of the suction ports 27a and 27b according to the

revolution direction of the roller 22, the valve assembly 100 is installed in the compressor

of the present invention.

As shown in FIGS. 2, 3 and 6, the valve assembly 100 includes first and second valves 110 and 120, which are installed between the cylinder 21 and the lower bearing 25

so as to allow it to be adjacent to the suction ports. If the suction ports 27a, 27b and 27c

are formed on the upper bearing 24, the first and second valves 110 and 120 are installed

between the cylinder 21 and the upper bearing 24.

The first valve 110, as shown in FIG. 3, is a disk member installed so as to contact

the eccentric portion 13a more accurately than the driving shaft 13. Accordingly, if the

driving shaft 13 rotates (that is, the roller 22 revolves), the first valve 110 rotates in the

same direction. Preferably, the first valve 110 has a diameter larger than an inner

diameter of the cylinder 21. As shown in FIG. 3, the cylinder 21 supports a portion (i.e.,

an outer circumference) of the first valve 110 so that the first valve 110 can rotate stably.

Preferably, the first valve 110 is 0.5 - 5 mm thick.

Referring to FIGS. 2 and 6, the first valve 110 includes first and second openings

111 and 112 respectively communicating with the first and second suction ports 27a and

27b in specific rotation direction, and a penetration hole HOa into which the driving shaft

13 is inserted. In more detail, when the roller 22 rotates in any one of the clockwise and

counterclockwise directions, the first opening 111 communicates with the first suction

port 27a by the rotation of the first valve 110, and the second suction port 27b is closed by

the body of the first valve 110. When the roller 22 rotates in the other of the clockwise

and counterclockwise directions, the second opening 112 communicates with the second

suction port 27b. At this time, the first suction port 27a is closed by the body of the first

valve 110. These first and second openings 111 and 112 can be in circular or polygonal shapes. In case the openings 111 and 112 are the circular shapes, it is desired that the

openings 111 and 112 are 6 - 15 mm in diameter. Additionally, the openings 111 and

112 can be rectangular shapes having predetermined curvature as shown in FIG. 7 A, or

cut-away portions as shown in FIG. 7B. As a result, the openings are enlarged, such that

fluid is sucked smoothly. If these openings 111 and 112 are formed adjacent to a center

of the first valve 110, a probability of interference between the roller 22 and the eccentric

portion 13a becomes increasing. In addition, there is the fluid's probability of leaking

out along the driving shaft 13, since the openings 111 and 112 communicate with a space

between the roller 22 and the eccentric portion 13a. For these reasons, as shown in FIG.

7C, it is preferable that the openings 111 and 112 are positioned in the vicinity of the

outer circumference of the first valve. Meanwhile, the first opening 111 may open each

of the first and second suction ports 27a and 27b at each rotation direction by adjusting

the rotation angle of the first valve 110. In other words, when the driving shaft 13

rotates in any one of the clockwise and counterclockwise directions, the first opening 111

communicates with the first suction port 27a while closing the second suction port 27b.

When the driving shaft 13 rotates in the other of the clockwise and counterclockwise

directions, the first opening 111 communicates with the second suction port 27b while

closing the first suction port 27a. It is desirable to control the suction ports using such a

single opening 111, since the structure of the first valve 110 is simplified much more.

Referring to FIGS. 2, 3 and 6, the second valve 120 is fixed between the cylinder

21 and the lower bearing 25 so as to guide a rotary motion of the first valve 110. The second valve 120 is a ring-shaped member having a site portion 121 which receives

rotatably the first valve 110. The second valve 120 further includes a coupling hole

120a through which it is coupled with the cylinder 21 and the upper and lower bearings

24 and 25 by a coupling member. Preferably, the second valve 120 has the same

thickness as the first valve 110 in order for a prevention of fluid leakage and a stable

support. In addition, since the first valve 110 is partially supported by the cylinder 21,

the first valve 110 may have a thickness slightly smaller than the second valve 120 in

order to form a gap for the smooth rotation of the second valve 120.

Meanwhile, referring to FIG. 4, in the case of the clockwise rotation, the fluid's

suction or discharge between the vane 23 and the roller 22 does not occur while the roller

22 revolves from the vane 23 to the second suction port 27b. Accordingly, a region V

becomes a vacuum state. The vacuum region V causes a power loss of the driving shaft

13 and a loud noise. Accordingly, in order to overcome the problem in the vacuum

region V, a third suction port 27c is provided at the lower bearing 25. The third suction

port 27c is formed between the second suction port 27b and the vane 23, supplying fluid

to the space between the roller 22 and the vane 23 so as not to form the vacuum state

before the roller 22 passes through the second suction port 27b. Preferably, the third

suction port 27c is formed in the vicinity of the vane 23 so as to remove quickly the

vacuum state. However, the third suction port 27c is positioned to face the first suction

port 27a since the third suction port 27c operates at a different rotation direction from the

first suction port 27a. In reality, the third suction port 27c is positioned spaced by an angle (Θ3) of approximately 10° from the vane 23 clockwise or counterclockwise. In

addition, as shown in FIGS. 5 A and 5B, the third suction port 27c can be circular shapes

or curved rectangular shapes.

Since such a third suction port 27c operates along with the second suction port

27b, the suction ports 27b and 27c should be simultaneously, opened while the roller 22

revolves in any one of the clockwise and counterclockwise directions. Accordingly, 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.

According to the present invention, the third opening 113 can be formed independently,

which is represented with a dotted line in FIG. 6 A. However, since the first and third

suction ports 27a and 27c are adjacent to each other, it is desirable to open both the first

and third suction ports 27a and 27c according to the rotation direction of the first opening

111 by increasing the rotation angle of the first valve 110.

The first valve 110 may open the suction ports 27a, 27b and 27c according to the

rotation direction of the roller 22, but the corresponding suction ports should be opened

accurately in order to obtain desired compression capacity. The accurate opening of the

suction ports can be achieved by controlling the rotation angle of the first valve. Thus,

preferably, the valve assembly 100 further includes means for controlling the rotation

angle of the first valve 110, which will be described in detail with reference to FIGS. 8 to

11. ^' FIGS. 8 to 11 illustrate the valve assembly connected with the lower bearing 25 in

order to clearly explain the control means. As shown in FIGS. 8A and 8B, the control means includes a groove 114 formed at

the first valve and having a predetermined length, and a stopper 114a formed on the lower

^■ bearing 25 and inserted into the groove 114. The groove 114 and the stopper 114a are

illustrated in FIGS. 5A, 5B and 6. The groove 114 serves as locus of the stopper 114a

and can be a straight groove or a curved groove. If the groove 114 is exposed to the

chamber 29 during operation, it becomes a dead volume causing a re-expansion of fluid.

Accordingly, it is desirable to make the groove 114 adjacent to a center of the first valve

110 so that large portion of the groove 114 can be covered by the revolving roller 22.

Preferably, an angle (α) between both ends of the groove 114 is of 30 - 120° in the center

of the first valve 110. In addition, if the stopper 114a is protruded from the groove 114,

it is interfered with the roller 22. Accordingly, it is desirable that a thickness T2 of the

stopper 114a is equal to a thickness Tl of the valve 110, as shown in FIG. 8 C. Preferably,

a width L of the stopper 114a is equal to a width of the groove 114, such that the first

valve rotates stably.

In the case of using the control means, the first valve 110 rotates counterclockwise

together with the eccentric portion 13a of the driving shaft when the driving shaft 13

rotates counterclockwise. As shown in FIG. 8 A, the stopper 114a is then latched to one

end of the groove 114 to thereby stop the first valve 10. At this time, the first opening

111 accurately communicates with the first suction port 27a, and the second and third

suction ports 27b and 27c are closed. As a result, fluid is introduced into the cylinder

through the first suction port 27a and the first opening 111, which communicate with each other. On the contrary, if the driving shaft 13 rotates clockwise, the first valve 110 also

rotates clockwise. At the same time, the first and second openings 111 and 112 also

rotate clockwise, as represented with a dotted arrow in FIG. 8A. As shown in FIG. 8B, if

the stopper 114a is latched to the other end of the groove 114, the first and second

openings 111 and 112 are opened together with the third and second suction ports 27c and

27b. Then, the first suction port 27a is closed by the first valve 110. Accordingly, fluid

is introduced through the second suction port 27b/the second opening 112 and the third

suction port 27c/the first opening 111, which communicate with each other.

As shown in FIGS. 9 A and 9B, the control means can be provided with a

projection 115 formed on the first valve 110 and projecting in a radial direction of the first

valve, and a groove 123 formed on the second valve 220 and receiving the projection

movably. Here, the groove 123 is formed on the second valve 220 so that it is not

exposed to the inner volume of the cylinder 21. Therefore, a dead volume is not formed

inside the cylinder. In addition, as shown in FIGS. 1OA and 1OB, the control means can

be provided with a projection 124 formed on the second valve 120 and projecting in a

radial direction of the second valve 120, and a groove 116 formed on the first valve 110

and receiving the projection 124 movably.

In the case of using such a control means, as shown in FIGS. 9 A and 1OA, the

projections 115 and 124 are latched to one end of each groove 123 and 116 if the driving

shaft 13 rotates counterclockwise. Accordingly, the first opening 111 communicates with

the first suction port 27a so as to allow the suction of fluid, and the second and third suction ports 27b and 27c are closed. On the contrary, as shown in FIGS. 9B and 1OB, if

the driving shaft 13 rotates clockwise, the projections 115 and 124 are latched to the other

end of each groove 123 and 116, and the first and second openings- 111 and 112

simultaneously open the third and second suction ports 27c and 27b so as to allow the

suction of fluid. The first suction port 27a is closed by the first valve 110.

In addition, as shown in FIGS. HA and 12B, the control means can be provided

with a projection 125 formed on the second valve 120 and projecting toward a center of

the second valve 120, and a cut-away portion 117 formed on the first valve 110 and

receiving the projection 125 movably. In such a control means, a gap between the

projection 125 and the cut-away portion 117 can open the first and second suction ports

27a and 27b by forming the cut-away portion 117 largely in a properly large size.

Accordingly, the control means decreases substantially in volume since the grooves of the

above-described control means are omitted.

In more detail, as shown in FIG. HA, if the driving shaft 13 rotates

counterclockwise, one end of the projection 125 contacts one end of the cut-away portion

17. Accordingly, a gap between the other ends of the projection 125 and the cut-away

portion 117 opens the first suction port 27a. In addition, as shown in FIG. HB, if the

driving shaft 13 rotates clockwise, the projection 125 is latched to the cut-away portion

117. At this time, the second opening 112 opens the second suction port 27b, and

simultaneously, the gap between the projection 125 and the cut-away portion 117 opens

the third suction port 27c as described above. In such a control means, preferably, the projection 125 has an angle βl of approximately 10° between both ends thereof and the

cut-away portion 117 has an angle β2 of 30 - 120° between both ends thereof.

Meanwhile, as described above with reference to FIG. 2, the suction ports 27a, 27b

and 27c are individually connected with a plurality of suction pipes 7a so as to supply

fluid to the fluid chamber 29 installed inside the cylinder 21. However, the number of

parts increases due to these suction pipes 7a, thus making the structure complicated. In

addition, fluid may not be properly supplied to the cylinder 21 due to a change in a

compression state of the suction pipes 7b separated during operation. Accordingly, as

shown in FIGS. 12 and 13, it is desirable to include a suction plenum 200 for

preliminarily storing fluid to be sucked by the compressor.

The suction plenum 200 directly communicates with all of the suction ports 27a,

27b and 27c so as to supply the fluid. Accordingly, the suction plenum 200 is installed

in a lower portion of the lower bearing 25 in the vicinity of the suction ports 27a, 27b and

27c. Although there is shown in the drawing that the suction ports 27a, 27b and 27c are

formed at the lower bearing 25, they can be formed at the upper bearing 24 if necessary.

In this case, the suction plenum 200 is installed in the upper bearing 24. The suction

plenum 200 can be directly fixed to the bearing 25 by a welding. In addition, a coupling

member can be used to couple the suction plenum 200 with the cylinder 21, the upper and

lower bearings 24 and 25 and the valve assembly 100. In order to lubricate the driving

shaft 13, a sleeve 25d of the lower bearing 25 should be soaked into a lubricant which is

stored in a lower portion of the case 1. Accordingly, the suction plenum 200 includes a penetration hole 200a for the sleeve. Preferably, the suction plenum 200 has 100 -

400 % a volume as large as the fluid chamber 29 so as to supply the fluid stably. The

suction plenum 200 is also connected with the suction pipe 7 so as to store the fluid. In

more detail, the suction plenum 200 can be connected with the suction pipe 7 through a

predetermined fluid passage. In this case, as shown in FIG. 12, the fluid passage

penetrates the cylinder 21, the valve assembly 100 and the lower bearing 25. In other

words, the fluid passage includes a suction hole 21c of the cylinder 21, a suction hole 122

of the second valve, and a suction hole 25c of the lower bearing.

Such a suction plenum 200 forms a space in which a predetermined amount of

fluid is always stored, so that a compression variation of the sucked fluid is buffered to

stably supply the fluid to the suction ports 27a, 27b and 27c. In addition, the suction

plenum 200 can accommodate oil extracted from the stored fluid and thus assist or

substitute for the accumulator 8.

Hereinafter, operation of a rotary compressor according to the present invention

will be described in more detail.

compressor when the roller revolves in the counterclockwise direction.

First, in FIG. 14A, there are shown states of respective elements inside the

cylinder when the driving shaft 13 rotates in the counterclockwise direction. First, the first

suction port 27a communicates with the first opening 111, and the remainder second

suction port 27b and third suction port 27c are closed. Detailed description on the state of the suction ports in the counterclockwise direction will be omitted since it has been

described with reference to FIGS. 8A₃ 9A, 1OA and HA.

In a state that the first suction port 27a is opened, the roller 22 revolves

counterclockwise with performing a rolling motion along the inner circumference of the

cylinder due to the rotation of the driving shaft 13. As the roller 22 continues to revolve,

the size of the space 29b is reduced as shown in FIG. 14B and the fluid that has been

sucked is compressed. In this stroke, the vane 23 moves up and down elastically by the

elastic member 23 a to thereby partition the fluid chamber 29 into the two sealed spaces

29a and 29b. At the same time, new fluid is continuously sucked into the space 29a

through the first suction port 27 so as to be compressed in a next cycle.

When the fluid pressure in the space 29b is above a predetermined value, the

second discharge valve 26d shown in FIG. 2 is opened. Accordingly, as shown in FIG.

14C, the fluid is discharged 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 the fluid is completely discharged, the second discharge valve

26d closes the second discharge port 26c by its self-elasticity.

Thus, after a single cycle is ended, the roller 22 continues to revolve

counterclockwise and discharges the fluid by repeating the same cycle. In the

counterclockwise cycle, the roller 22 compresses the fluid with revolving from the first

suction port 27a to the second discharge port 26b. As aforementioned, since the first

suction port 27a and the second discharge port 27b are positioned in the vicinity of the vane 23 to face each other, the fluid is compressed using the overall volume of the fluid

chamber 29 in the counterclockwise cycle, so that a maximal compression capacity is

obtained.

FIGS. 15A to 15C are cross-sectional views an operation sequence of a rotary

compressor according to the present invention when the roller revolves clockwise.

First, in FIG. 15 A, there are shown states of respective elements inside the

cylinder when the driving shaft 13 rotates in the clockwise direction. The first suction

port 27a is closed, and the second suction port 27b and third suction port 27c

communicate with the second opening 112 and the first opening 111 respectively. If the

first valve 110 has the third opening 113 additionally (refer to FIG. 6), the third suction

port 27c communicates with the third opening 113. Detailed description on the state of

the suction ports in the clockwise direction will be omitted since it has been described

with reference to FIGS. 8B, 9B, 1OB and HB.

In a state that the second and third suction ports 27b and 27c are opened, the roller

22 begins to revolve clockwise with performing a rolling motion along the inner

circumference of the cylinder due to the clockwise rotation of the driving shaft 13. In

such an initial stage revolution, the fluid sucked until the roller 22 reaches the second

suction port 27b is not compressed but is forcibly exhausted outside the cylinder 21 by the

roller 22 through the second suction port 27b as shown in FIG. 15 A. Accordingly, the

fluid begins to be compressed after the roller 22 passes the second suction port 27b as

shown in FIG. 15B. At the same time, a space between the second suction port 27b and the vane 23, i.e., the space 29b is made in a vacuum state. However, as aforementioned, as

the revolution of the roller 22 starts, the third suction port 27c communicates with the first

opening 111 (or third opening 113) and thus is opened so as to suck the fluid.

Accordingly, the vacuum state of the space 29b is removed by the sucked fluid, so that

generation of noise and power loss are constrained.

As the roller 22 continues to revolve, the size of the space 29a is reduced and the

fluid that has been sucked is compressed. In this compression stroke, the vane 23 moves

up and down elastically by the elastic member 23 a to thereby partition the fluid chamber

29 into the two sealed spaces 29a and 29b. Also, new fluid is continuously sucked into the

space 29b through the second and third suction ports 27b and 27c so as to be compressed

in a next stroke.

When the fluid pressure in the space 29a is above a predetermined value, the first

discharge valve 26c shown in FIG. 2 is opened and accordingly the fluid is discharged

through the first discharge port 26a. After the fluid is completely discharged, the first

discharge valve 26c closes the first discharge port 26a by its self-elasticity.

Thus, after a single stroke is ended, the roller 22 continues to revolve clockwise

and discharges the fluid by repeating the same stroke. In the counterclockwise stroke, the

roller 22 compresses the fluid with revolving from the second suction port 27b to the first

discharge port 26a. Accordingly, the fluid is compressed using a part of the overall fluid

chamber 29 in the counterclockwise stroke, so that a compression capacity smaller than

the compression capacity in the clockwise direction. In the aforementioned strokes (i.e., the clockwise stroke and the counterclockwise

stroke), the discharged compressed fluid moves upward through the space between the

rotator 12 and the stator 11 inside the case 1 and the space between the stator 11 and the

case 1. As a result, the compressed fluid is discharged through the discharge tube 9 out

of the compressor.

Meanwhile, during the operation as described above, mechanical elements such as

the motor 11, 12, the driving shaft 13, and the roller 22 are in a high-speed rotation.

Especially, as the driving shaft 13 alternately repeats the clockwise and counterclockwise

rotations, this driving shaft 13 is exposed to a severer operational condition in the

compressor of the present invention. Accordingly, a proper lubrication and a lubricating

mechanism for such lubrication are fairly important for a stable operation of the

compressor. The present invention provides an oil path as such lubricating mechanism

configured to supply oil, i.e. lubricant "O" to driving elements in the compressor, and this

oil path will be specifically described as follows, referring relevant drawings.

FIG. 16 is a front view showing the oil path of the rotary compressor according to

the present invention. FIGS. 17A-17C are drawings each illustrating a first embodiment

of a bearing path included in the oil path, and FIGS. 18A-18C are drawings each

illustrating a second embodiment of a bearing path.

As shown, the lubricating mechanism, i.e. the oil path 300 is formed along the

driving shaft 13 and the bearings 24, 25. Journals 13b, 13c of the driving shaft 13 are

surrounded by the upper and lower bearings 24, 25 respectively, and substantially form radial bearings supporting load normal to a center axis of the driving shaft. Additionally,

collars 13d, 13e form together with the bearings 24, 25 thrust bearings supporting load in

axial direction. The oil path 300 mainly comprises a shaft path 310 (hereinafter, referred

to as "a first path") formed within the driving shaft 13.

More specifically, the first path 310 extends from a lower end of the driving shaft

13 to an upper end, and thus substantially passes through the driving shaft in a length

direction thereof. In addition, at a lower end of the first path 310, an oil pump 311 is

mounted. This oil pump 310 is a sort of a centrifugal pump, and includes an oil pickup

311a and a propeller 311b inserted into the oil pickup 311a. The oil pump 311 is dipped in

the lubricant, i.e. the oil "O" in a bottom portion of the compressor (see FIG. 1), and thus

the oil can flows in the first path 310 through the oil pump 311. Then, the oil is pumped

along the first path 130 and is dispersed at the upper end of the driving shaft 13 in order to

be supplied to corresponding driving parts. Additionally, the first path 310 further

includes holes 312a, 312b formed at an upper portion and a lower portion of the eccentric

portion 13a respectively to communicate with the first path 310. The oil is first supplied

into the cylinder 21 though the holes 312a, 312b so as to lubricate the roller 22 and the

eccentric portion 13a. The holes 312a, 312b also allows the oil to be supplied to the upper

and lower bearings 24, 25 and the driving shaft 13, accurately the journals 13b, 13c.

However, as the journals 13b, 13c and the bearings 24, 25 form large frictional

surfaces as illustrated, the oil could not reach ends of the frictional surfaces only with a

small amount of supply through the holes 312a, 312b. That is, the oil could not be spread all over the junctional surfaces, and could not entirely form oil films thereon for a

prevention of abrasion. To solve such a problem, the oil path 300 in the present invention

has a bearing path 320 (hereinafter, referred to as "a second path") formed at any one of

the bearings 24, 25 as shown in FIG. 16, FIGS. 17A-17B, and FIGS. 18A-18B. The

second path 320 is substantially formed as a groove formed on an inner circumferential

surface in any one of the bearings. The second path 320 communicates the driving shaft

13, more accurately any one of the holes 312a, 312b adjacent thereto in order to be

provided with the oil by the first path 310. In addition, the second path 320 preferably

extends continuously between an upper end and a lower end of the inner circumferential

surface. Therefore, the oil is supplied to the second path 320 by any one of the holes 312a,

312b, and then flows between both ends of the inner circumferential surface along the

second path 320. Namely, due to the second path 320, the oil path 300 is configured to

allow the oil to uniformly flow between the bearings 24, 25 and the driving shaft 13. The

oil then spreads equally on the frictional surfaces, and forms the oil films entirely in order

to effectively prevent the abrasion. It is desirable that such second path 320 is formed at

the upper bearing 24 at least. This is because in. the lower bearing, the oil can flows

downward to some extent by gravity from the holes 312b. However, it is more desirable

for suitable lubrication that the second paths 320 are formed at both of the upper and

lower bearings 24, 25 respectively.

As described above, since the driving shaft 13 rotates in clockwise and

counterclockwise directions, the second path 320 should be able to allow the oil to flow therein in both rotational directions of the compressor. The second path 320 might be

formed as a helical groove. This helical groove expands a substantial flow passage and

enables a sufficient oil supply. However, the helical groove can allow the oil to flow

therein in any one direction of the rotations of the driving shaft 13 due to its geometric

characteristic. More specifically, the helical groove can allow the oil to flow and ascend

therein only when it extends in a direction opposite to the rotational direction of the

driving shaft 13. Accordingly, in a first embodiment, the second path 320 comprises a

single straight groove as shown in FIGS. 17A and 17B. The straight groove is not affected

by the geometrical characteristic contrary to the helical groove, and can allow the oil to

flow therein by centrifugal force generated by the driving shaft 13 regardless of the

rotational direction of the driving shaft. Alternatively, as a second embodiment, the

second path 320 comprises first and second helical grooves 320a, 320b as shown in FIGS.

18A and 18B. More specifically, the helical groove can allow the oil to flow therein only

in any one of the rotational directions of the driving shaft 13 as explain above. Therefore,

two helical grooves corresponding to the rotational direction respectively are applied to

the present invention, and these grooves extend in opposite directions (the clockwise and

counterclockwise directions). Also, if the first and second helical grooves 320a, 320b

intersect each other on the inner circumferential surfaces of the bearings 24, 25, the oil

flowing in one helical groove leaks into the other helical groove. As such a leakage causes

the bearings 24, 25 and the journals 13b, 13c not to be entirely lubricated, it is important

for an optimal lubrication that the helical grooves 320a, 320b do not intersect each other. Meanwhile, referring to FIGS. 17A and 18 A, clearances C with predetermined

sizes are formed between the bearings 24, 25 and the driving shaft 13 (more accurately,

the journals 13b, 13c), and the oil fills such clearances C using the second paths 320 to

form the oil films therein. The driving shaft 13 is subject to pressure from the compressed

working fluid during the operation of the compressor, and thus rotates eccentric from

centers O of the bearings 24, 25. In addition, as the second paths 320 damage the inner

circumferential surfaces of the bearings continuously along their length directions, the

clearances C are increased around the second paths 320 and the sufficient oil films are not

formed around the second paths 320 due to the increased clearances C. Accordingly, if the

second paths 320 are positioned where an eccentricity of the driving shaft 13 is greatly

generated , the driving shaft 13 may be in contact with the inner circumferential surfaces

of the bearings 24, 25. In this case, the abrasion between the bearings 24, 25 and the

driving shaft 13 may occur, and simultaneously, nose may be made during the operation

of the compressor. Also, the power loss of the driving shat 13 may occur due to the

excessive abrasion. Therefore, it is preferable for the second paths 320 to be positioned

where the eccentricity of the driving shaft 13 is small.

In the present invention, optimal positions of the second paths 320 were

determined by experiments, and FIGS. 17C and 18C show experimental results

considered for the optimal positions in the first and second embodiments of the second

path 320, respectively.

As illustrated, FIGS. 17C and 18C are graphs each showing change of an eccentricity ratio to an angle. First, the angle is set to be 0° at the vane 23 positioned

beneath and above the bearings 24, 25, and is also set to increase in the direction of

rotation in which the maximum compression capacity is generated. The compressor was

set to generate the maximum compression capacity in the counterclockwise rotation in the

experiment, and thus the angle was set to increase in the counterclockwise direction as

show. The eccentricity ratio is defined as a ratio of an eccentric distance (i.e. a distance

from the bearing center O to a center of the driving shaft) to the clearance C. This

eccentricity ratio is a dimensionless index showing how much the driving shaft 13 is close

to the inner circumferential surfaces of the bearings 24, 25. As the clearance C is constant,

the great eccentricity ratio means that the driving shaft has a great amount of the

eccentricity and is close to the inner circumferential surfaces of the bearings 24, 25. The

eccentricity ratios were also measured with regard to both of the maximum and minimum

compression capacities. The eccentricity ratio in the maximum capacity was measured in

the counterclockwise rotation of the driving shaft 13 as described above, and the

eccentricity ration in the minimum capacity was measured in the clockwise rotation. Due

to differences in various operational conditions as well as in compression capacities

according to the rotational directions, the eccentricity ratios in the maximum and

minimum capacities have different phases. As a result of the experiment, such eccentricity

ratios were not greatly changed and rather showed almost identical tendencies, with

regard to various specifications of the compressors subject to the experiment.

First, as the working fluid is maximally compressed near the vane 23, both ^' eccentricity ratios of the maximum and minimum capacities have relatively great values

at 0° (360°), i.e. at the vane 23, as shown in FIGS. 17C and 18C. In addition, if the second

path 230 is positioned above or beneath the vane 23, the working fluid having the

maximum pressure near the vane 23 may leak into such second path 230. In view of these

conditions, it is preferable that the second paths 230 according to the first and second

embodiments are basically spaced apart from the vane 23 in the clock or

counterclockwise direction with reference to the center O.

More specifically, in the first embodiment of the second path 320, the eccentricity

ratios in the maximum and minimum capacities have relatively small values in a range of

170° - 210°, as shown in FIG. 17C. Accordingly, the single straight groove according to

the first embodiment is preferably spaced apart from the vane 23 by an angle A of 170° -

210° in the counterclockwise direction. In addition, the compressor of the present

invention could be designed to have the maximum compression capacity in the clockwise

rotation (i.e. the minimum compression capacity in the counterclockwise rotation). Even

in this case, it would be appreciated that the same result as FIG. 17C is obtained, when the

angle is set to increase in the clockwise direction in which the maximum capacity is

obtained, to be opposite to FIG 17 A. Therefore, the single straight groove could be spaced

apart from the vane 23 by the angle of 170° - 210° in the counterclockwise or clockwise

direction. Additionally, the eccentric ratios in the maximum and minimum capacities have

the same small value at 190°. That is, chances to be contact with shaft are minimized in

both of the clockwise and counterclockwise rotations. Accordingly, it is most preferable for the angle A to be 190°. In the second embodiment, as the second path 320 comprises

the first and second helical grooves 320a, 320b, it is important for these grooves 320a,

320b to be disposed respectively within angle ranges having relatively small eccentricity

ratios so as not to interfere with each other. Referring to FIG. 18C, the eccentricity ratios

have relatively small values in the ranges of angle in the vicinity of 190°. Accordingly, as

illustrated, the first and second helical grooves 320a, 320b are spaced apart from the vane

23 by a first angle Bl and a second angle B2 respectively in the clockwise or

counterclockwise directions, and these angles Bl, B2 have ranges of 130° - 190° and 190°

- 250°, respectively. Further, the second path 230 has appropriate width w and depth d in

order to allow the sufficient amount of the oil to flow therein as well as to reduce the

damage on the inner circumferential surfaces of the bearings interrupting a formation of

the oil films. These width w and depth d are preferably 3.8mm and 1.67mm respectively,

although they are slightly varied according to the specifications of the compressors.

Further, to allow the oil to flow more sufficiently between the bearings 24, 25 and

the driving shaft 13, the oil path 300 additionally includes an auxiliary path 330 as shown

in FIGS. 17A, 18A and FIGS. 19A-19B. This auxiliary path 330 comprises grooves

formed along the journals 13b, 13c, and preferably extends over entire lengths of the

journals 13b, 13c. Likewise, it is desirable that the auxiliary path 330 is configured to

allow the oil to flow therein in all the rotational directions of the driving shaft 13.

Accordingly, the auxiliary path 330 comprises a single straight groove as shown in FIGS.

17A and 19A, or two helical grooves 330a, 330b as shown in FIGS. 18A and 19B. It will be apparent to those skilled in the art that various modifications and

variations can be made in the present invention. Thus, it is intended that the present

invention covers the modifications and variations of this invention provided they come

within the scope of the appended claims and their equivalents.

Industrial Applicability

As described above, the rotary compressor of the present invention can compress

the fluid without regard to the rotation directions of the driving shaft and have the

compression capacity that is variable according to the rotation directions of the driving

shaft. Further, since the rotary compressor of the present invention have the suction and

discharge ports arranged properly, and a simple valve assembly for selectively opening

the suction ports according to the rotation directions, an overall designed refrigerant

chamber can be used to compress the fluid. In addition, the compressor of the present

invention has a lubricating mechanism supplying oil between a driving shaft and bearings

thereof under severest operational environment in order to support a capacity variation

mechanism. The rotary compressor as mentioned above provides following advantages.

First, according to the related art, several devices are combined in order to achieve

the dual-capacity compression. For example, an inverter and two compressors having

different compression capacities are combined in order to obtain the dual compression

capacities. In this case, the structure becomes complicated and the cost increases.

However, according to the present invention, the dual-capacity compression can be achieved using only one compressor. Particularly, the present invention can achieve the

dual-capacity compression by changing parts of the conventional rotary compressor to the

minimum.

Second, the conventional compressor having a single compression capacity cannot

provide the compression capacity that is adaptable for various operation conditions of air

conditioner or refrigerator. In this case, power consumption may be wasted

unnecessarily. However, the present invention can provide a compression capacity that

is adaptable for the operation conditions of equipments.

Third, according to the rotary compressor of the present invention, an overall

designed fluid chamber is used to provide the dual-compression capacity. It means that

the compressor of the present invention has at least the same compression capacity as the

conventional rotary compressor having the same cylinder and fluid chamber in size. In

other words, the rotary compressor of the present invention can substitute for the

conventional rotary compressor without modifying designs of basic parts, such as a size

of the cylinder. Accordingly, the rotary compressor of the present invention can be

freely applied to required systems without any consideration of the compression capacity

and any increase in unit cost of production.

Fourth, in the rotary compressor of the present invention, uniform oil films are

formed between the driving shaft and the bearings by the lubricating mechanism.

Accordingly, an abrasion of the driving shaft is effectively prevented under the severe

operational environment. Further, such lubricating mechanism allows the oil to flow therein in all the rotational directions of the driving shaft and is positioned where an

eccentricity of the driving shaft is small. Therefore, the lubrication for preventing the

^• abrasion becomes more stable and effective.

Claims

1, A rotary compressor comprising:

a driving shaft being rotatable clockwise and counterclockwise, and having an

eccentric portion of a predetermined size;

a cylinder forming a predetermined inner volume;

a roller installed rotatably on an outer circumference of the eccentric portion so as

to contact an inner circumference of the cylinder, performing a rolling motion along the

inner circumference and forming a fluid chamber to suck and compress fluid along with

the inner circumference;

a vane installed elastically in the cylinder to contact the roller continuously;

upper and lower bearings installed respectively in upper and lower portions of the

cylinder, for supporting the driving shaft rotatably and sealing the inner volume

hermetically;

an oil path configured to allow oil to uniformly flow between the bearings and the

driving shaft;

discharge ports communicating with the fluid chamber;

suction ports communicating with the fluid chamber and being spaced apart from

each other by a predetermined angle; and

a valve assembly for selectively opening any one of the suction ports according to

rotation direction of the driving shaft, wherein compression spaces that have different volumes from each other are

formed in the fluid chamber according to the rotation direction of the driving shaft such

that two different compression capacities are formed.

2. The rotary compressor of claim 1, wherein the roller compresses the fluid using

the overall fluid chamber only when the driving shaft rotates in any one of the clockwise

direction and the counterclockwise direction.

3. The rotary compressor of claim 1, wherein the roller compresses the fluid using

a portion of the fluid chamber when the driving shaft rotates in the other of the clockwise

direction and the counterclockwise direction.

4. The rotary compressor of claim 1, wherein the discharge ports comprise a first

discharge port and a second discharge port that are positioned facing each other with

respect to the vane.

5. The rotary compressor of claim 1, wherein the suction ports comprise:

a first suction port positioned in the vicinity of the vane; and

a second suction port positioned spaced apart from the first suction port by a

predetermined angle.

6. The rotary compressor of claim 5, wherein the suction ports are circular.

7. The rotary compressor of claim 5, wherein the suction ports are rectangles.

8. The rotary compressor of claim 7, wherein the suction ports have a

predetermined curvature.

9. The rotary compressor of claim 6, wherein the suction ports have diameters

ranged from 6 mm to 15 mm.

10. The rotary compressor of claim 5, wherein the first suction port is positioned

spaced by approximately 10° from the vane clockwise or counterclockwise.

11. The rotary compressor of claim 5, wherein the second suction port is

positioned in a range of 90 - 180" from the vane to face the first opening.

12. The rotary compressor of claim 1 or claim 5, wherein the valve assembly

comprises:

a first valve installed rotatably between the cylinder and the bearing; and

a second valve guiding a rotary motion of the first valve.

13. The rotary compressor of claim 12, wherein the first valve comprises a disk

member contacting the eccentric portion of the driving shaft and rotating in the rotation

direction of the driving shaft.

14. The rotary compressor of claim 13, wherein the first valve has a diameter

larger than an inner diameter of the cylinder.

15. The rotary compressor of claim 13, wherein the first valve is 0.5 - 5 mm thick.

16. The rotary compressor of claim 12, wherein the first valve comprises:

a first opening communicating with the first suction port when the driving shaft

rotates in any one of the clockwise direction and the counterclockwise direction; and

a second opening communicating with the second suction port when the driving

shaft rotates in the other of the clockwise direction and the counterclockwise direction.

17. The rotary compressor of claim 16, wherein the first opening and the second

opening are circular or polygonal.

18. The rotary compressor of claim 16, wherein the first opening and the second

opening are cut-away portions.

19. The rotary compressor of claim 16, wherein the first opening and the second

opening are rectangles each having a predetermined curvature.

20. The rotary compressor of claim 17, wherein the first opening and the second

opening have diameters ranged from 6 mm to 15 mm.

21. The rotary compressor of claim 16, wherein the first opening and the second

opening are positioned in the vicinity of the outer circumference of the first valve.

22. The rotary compressor of claim 12, wherein the first valve comprises a

penetration hole into which the driving shaft is inserted.

23. The rotary compressor of claim 12, wherein the second valve is fixed between

the cylinder and the bearing and comprises a seat portion for receiving the first valve.

24. The rotary compressor of claim 23, wherein the second valve has the same

thickness as the first valve.

25. The rotary compressor of claim 16, wherein the suction port further comprises

a third suction port positioned between the second suction port and the vane.

26. The rotary compressor of claim 25, wherein the third suction port is positioned

spaced by approximately 10° from the vane clockwise or counterclockwise.

27. The rotary compressor of claim 25, wherein the first valve further comprises a

third opening a third opening for opening the third suction port simultaneously with

opening the second suction port.

28. The rotary compressor of claim 25, wherein the first valve comprises a first

opening for opening the third suction port simultaneously with opening the second

suction port.

29. The rotary compressor of claim 12, wherein the valve assembly further

comprises means for controlling a rotation angle of the first valve such that corresponding

suction ports are opened accurately.

30. The rotary compressor of claim 29, wherein the control means comprises:

a curved groove formed at the first valve and having a predetermined length; and

a stopper formed on the bearing and inserted into the curved groove.

31. The rotary compressor of claim 30, wherein the curved groove is positioned in

the vicinity of a center of the first valve.

32. The rotary compressor of claim 30, wherein the stopper has the same thickness

as the first valve.

33. The rotary compressor of claim 30, wherein the stopper has the same width as

the curved groove.

34. The rotary compressor of claim 30, wherein the curved groove has an angle of

30 - 120° between both ends thereof.

35. The rotary compressor of claim 29, wherein the control means comprises:

a projection formed on the first valve and projecting in a radial direction of the

first valve; and

a groove formed on the second valve, for receiving the projection movably.

36. The rotary compressor of claim 29, wherein the control means comprises:

a projection formed on the second valve and projecting in a radial direction of the

second valve; and

a groove formed on the first valve, for receiving the projection movably.

37. The rotary compressor of claim 29, wherein the control means comprises: a projection formed on the second valve and projecting toward a center of the

second valve; and

a cut-away portion formed on the first valve, for receiving the projection movably.

38. The rotary compressor of claim 37, wherein the projection and the cut-away

portion form a gap therebetween and the gap opens the first suction port or the third

suction port according to the rotation direction of the driving shaft.

39. The rotary compressor of claim 37, wherein the projection has an angle of 10 -

90° between both side surfaces.

40. The rotary compressor of claim 37, wherein the cut-away portion has an angle

of 30 - 120° between both ends thereof.

41. The rotary compressor of claim 1, further comprising a plurality of suction

pipes supplying the cylinder with fluid to be compressed, the suction pipes being

individually connected with the suction ports.

42. The rotary compressor of claim 1, further comprising a suction plenum for

preliminarily storing fluid to be compressed, the suction plenum being connected with the

suction ports.

43. The rotary compressor of claim 42, wherein the suction plenum accommodates

oil extracted from the stored fluid.

44. The rotary compressor of claim 42, wherein the suction plenum is installed at a

lower portion of the bearing in the vicinity of the suction port.

45. The rotary compressor of claim 42, wherein the suction plenum has 100 -

400 % a volume as large as the fluid chamber.

46. The rotary compressor of claim 42, wherein the suction plenum is connected

with a suction pipe through a predetermined fluid passage, the suction pipe supplying the

fluid to be compressed.

47. The rotary compressor of claim 46, wherein the fluid passage penetrates the

cylinder, the valve assembly and the lower bearing.

48. The rotary compressor of claim 12, wherein the first valve comprises a single

opening which communicates with the first suction port when the driving shaft rotates in

any one of the clockwise direction and the counterclockwise direction, and communicates

with the second suction port when the driving shaft rotates in the other of the clockwise direction and the counterclockwise direction.

49. The rotary compressor of claim 1, wherein the oil path is configured to allow

the oil to flow between the driving shaft and the bearings in both of clockwise and

counterclockwise rotations of the driving shaft.

50. The rotary compressor of claim 1, wherein the oil path comprises a single

straight groove formed at any one of the bearings and allowing the oil to flow therein

regardless of rotational directions of the driving shaft.

51. The rotary compressor of claim 1, wherein the oil path comprises first and

second helical grooves formed at any one of the bearings and each configured to allow the

oil to flow therein in corresponding rotation of the driving shaft.

52. The rotary compressor of claim 51, wherein the first and second helical

grooves extend in opposite directions.

53. The rotary compressor of claim 51, wherein the first and second helical

grooves do not intersect each other.

54. The rotary compressor of claim 1, wherein the oil path is provided to any one of the bearings and is positioned where an eccentricity of the driving shaft is small.

55. The rotary compressor of claim 1, wherein the oil path is formed at any one of

the bearings to be spaced apart from the vane in clockwise or counterclockwise direction.

56. The rotary compressor of claim 50, wherein the single straight groove is

spaced apart from the vane by a range of 170°-210° in clockwise or counterclockwise

direction.

57. The rotary compressor of claim 56, wherein the single straight groove is

spaced apart from the vane by 190° in clockwise or counterclockwise direction.

58. The rotary compressor of claim 51, wherein the first and second helical

grooves are spaced apart from the vane respectively by 130°-190° and 190°-250° in

clockwise or counterclockwise direction.

59. The rotary compressor of claim 50 or 51, wherein a width of the oil path is

3.8mm.

60. The rotary compressor of claim 50 or 51, wherein a depth of the oil path is

1.67mm.

61. The rotary compressor of claim 1, wherein the oil path comprises a bearing

path formed at any one of the bearings.

62. The rotary compressor of claim 61, wherein the bearing path is formed at the

upper bearing al least.

63. The rotary compressor of claim 61, wherein the bearing path is formed on an

inner circumferential surface of the bearing.

64. The rotary compressor of claim 61, wherein the bearing path continuously

extends from an upper end to a lower end of the bearing.

65. The rotary compressor of claim 61, wherein the bearing path is provided with

the oil from the driving shaft.

66. The rotary compressor of claim 1, wherein the oil path further comprises a

shaft path formed in the driving shaft and configured to supply the oil to driving parts of

the compressor.

67. The rotary compressor of claim 1, wherein the oil path further comprises a auxiliary path formed at any one of journals in the driving shaft.

68. The rotary compressor of claim 67, wherein the auxiliary path is formed on an

outer circumferential surface of the journal.

69. The rotary compressor of claim 67, wherein the auxiliary path is configured to

allow the oil to flow between the driving shaft and the bearings in both of clockwise and

counterclockwise rotations of the driving shaft.

70. The rotary compressor of claim 67, wherein the auxiliary path comprises a

single straight groove allowing the oil to flow therein regardless of rotational directions of

the driving shaft.

71. The rotary compressor of claim 67, wherein the auxiliary path comprises first

and second helical grooves each configured to allow the oil to flow therein in

corresponding rotation of the driving shaft.