Hermetic compressor

TECHNICAL FIELD

The present invention relates to a hermetic compressor used in a refrigeration cycle system.

BACKGROUND ART

There is a conventionally proposed principle of a compressing mechanism which includes a rotary cylinder having a groove, and a piston slidable within the groove, so that the rotary cylinder is rotated in accordance with the movement of the piston to perform suction and compression strokes (for example, see German Patent No. 863,751 and British Patent No. 430,830).

The conventionally proposed principle of the compressing mechanism will be described below with reference to FIG.

26

.

The compressing mechanism is comprised of a rotary cylinder

101

having a groove

100

, and a piston

102

which is slidable within the groove

100

. The rotary cylinder

101

is provided for rotation about a point A, and the piston

102

is rotated about a point B.

The movements of the piston and the cylinder will be described as for a case where the rotational radius of the piston

102

is equal to the distance between the center A of rotation of the rotary cylinder

101

and the center B of rotation of the piston

102

.

When the rotational radius of the piston

102

is larger, or smaller than the distance between the rotational center A of the rotatable cylinder

101

and the rotational center B of the piston

102

, different movements are performed. The description of these different movements is omitted herein.

A broken line C in

FIG. 26

indicate a locus for the piston

102

.

FIGS. 26

a

to

26

i

show states in which the piston

102

has been rotated through every 90 degree.

First, the movement of the piston

102

will be described below.

FIG. 26

a

shows the state in which the piston lies immediately above the rotational center B.

FIG. 26

b

shows the state in which the piston

102

has been rotated through 90 degree in a counterclockwise direction from the state shown in

FIG. 26

a

.

FIG. 26

c

shows the state in which the piston

102

has been rotated through 180 degree in the counterclockwise direction from the state shown in

FIG. 26

a

.

FIG. 26

d

shows the state in which the piston

102

has been further rotated through 270 degree in the counterclockwise direction from the state shown in

FIG. 26

a

.

FIG. 26

e

shows the state in which the piston

102

has been rotated through 360 degree in the counterclockwise direction from the state shown in

FIG. 26

a

and has been returned to the state shown in

FIG. 26

a.

The movement of the rotary cylinder

102

will be described below. In the state shown in

FIG. 26

a

, the rotary cylinder

101

is located, so that the groove

100

is located vertically. When the piston

102

is moved through 90 degree in the counterclockwise direction from this state, the rotary cylinder

101

is rotated through 45 degree in the counter-clockwise direction, as shown in

FIG. 26

b

and hence, the groove

100

is likewise brought into a state in which it is inclined at 45 degree When the piston

102

is rotated through 180 degree in the counterclockwise direction from the state shown in

FIG. 26

a

, the rotary cylinder

101

is rotated through 90 degree in the counterclockwise direction, as shown in

FIG. 26

c

and hence, the groove

100

is likewise brought into a state in which it is inclined at 90 degree.

In this way, the rotary cylinder

101

is rotated in the same direction with the rotation of the piston

102

, but while the piston

102

is rotated through 360 degree, the rotary cylinder

101

is rotated through 180 degree.

The change in volume of the groove

100

defining the compressing space will be described below.

In the state shown in

FIG. 26

a

, the piston

102

lies at one end in the groove

100

and hence, only one space

100

exists. This space

100

is called a first space

100

a

herein. In the state shown in

FIG. 26

b

, the first space

100

a

is narrower, but a second space

100

b

is produced on the opposite side of the piston

102

. In the state shown in

FIG. 26

c

, the first space

100

a

is as small as half of the space in the state shown in

FIG. 26

a

, but a second space

100

b

of the same size as the first space

100

a

is defined on the opposite side of the piston

102

. The first space

100

a

is zero in volume in the state shown in

FIG. 26

e

in which the piston

102

has been rotated through 360 degree.

In this way, the two spaces

100

a

and

100

b

are defined by the piston

102

and repeatedly varied in volume from the minimum to the maximum and from the maximum to the minimum, whenever the piston

102

is rotated through 360 degree.

Therefore, the spaces defining the compressing chambers perform the compression and suction strokes by the rotation of the piston

102

through 720 degree.

It is a main object of the present invention to utilize the above-described compressing principle in the hermetic compressor.

The above-described compressing principle suffers from the following problem: When the piston

102

is at the center A of rotation of the rotary cylinder

101

, the direction of a force provided by the rotational force of the piston

102

is the same as the direction of the groove

100

and hence, this force does not serve a force for rotating the rotary cylinder

101

. Therefore, when the piston

102

is at the center A of rotation of the rotary cylinder

101

, the above-described movement is actually continuously not performed, if the rotational force is not applied to the rotary cylinder

101

.

A continuous movement is realized by using a plurality of compressing mechanisms synchronized with each other with different phases. More specifically, by using a plurality of compressing mechanisms synchronized with each other with different phases, the rotational force of one of the rotatable cylinders can be applied to the other rotatable cylinder. Therefore, even if either one of the rotatable cylinders is brought into a state in which it does not receive the rotational force from the piston, the other rotatable cylinder applies the rotational force to the one rotatable cylinder and hence, the rotation can be continuously maintained.

However, when the plurality of compressing mechanisms with different phases are used, the compressing strokes in the compressing chambers in the compressing mechanisms are different from each other. For this reason, a partition plate for isolating the adjacent compressing mechanisms is required. To ensure a smooth rotation, the synchronization of the plurality of compressing mechanisms must be made reliable.

Accordingly, it is an object of the present invention to provide a hermetic compressor using a plurality of compressing mechanisms with different phases, wherein the synchronization of the plurality of compressing mechanisms can be made reliable.

It is another object of the present invention to provide a hermetic compressor, wherein the reliable synchronization of the compressing mechanisms can be realized by a particular structure capable of being industrially produced.

It is a further object of the present invention to provide a hermetic compressor, wherein a high suction efficiency can be realized.

It is a yet further object of the present invention to provide a hermetic compressor, wherein a high compressing efficiency can be realized.

Further, it is an object of the present invention to provide a hermetic compressor, wherein a non-circular piston is employed, and the area of contact between a rotary cylinder and the piston is increased to enhance the sealability and to enhance the sucking and compressing efficiencies.

SUMMARY OF THE INVENTION

To achieve the above objects, according to a first aspect and feature of the present invention, there is provided a hermetic compressor comprising a plurality of compressing mechanisms each of which includes a rotary cylinder having a groove, and a piston which is slidable in the groove, so that a compressing stroke is carried out by rotation of the piston on a locus of a radius E about a point spaced apart at a distance E from the center of the rotary cylinder; a partition plate being interposed between the rotary cylinders of the adjacent compressing mechanisms, the partition plate being provided with a communication bore through which a shaft is passed, the shaft being provided with cranks on which the pistons can be mounted; and a motor mechanism for driving the pistons of the compressing mechanisms by the common shaft, at least one of the compressing mechanisms being different in phase in a compressing stroke from the other compressing mechanisms, the rotary cylinders of the adjacent compressing mechanisms and the partition plate sandwiched between these rotary cylinders being formed from different members, and relatively non-rotatably connected to each other.

With the above arrangement, two spaces are defined in the groove by the piston. The volumes of the spaces are varied by the sliding movement of the piston and hence, the compression and suction can be carried out. In this way, the compressing mechanism carries out the compression and suction only by the rotating movements of the rotary cylinder and the piston, and does not require a member moved diametrically such as a vane required in a rotary compressor and an Oldham's ring required in a scroll compressor. Therefore, it is possible to realize a hermetic compressor which produces only an extremely small amount of vibration, even if the compressor mechanism section is fixed within the shell. In the hermetic compress or according to the present invention, the piston rotated in the above manner about the above-described point is not necessarily requited to be rotated about its axis during the rotating movement about such point, and need be only slid along the groove. Therefore, the piston can be formed into a non-circular shape, whereby the area of contact of the piston with the groove can be increased to enhance the sealability, thereby enhancing the suction and compression efficiency.

Thus, even if the piston is located at the center of the rotary cylinder in one of the compressor mechanism sections, it can be avoided that the driving force from the piston does not serve as a rotating force for the rotary cylinder, because the other compressing mechanism provides a rotating force.

According to a second aspect and feature of the present invention, in addition to the first feature, the rotary cylinder and the partition plate are formed of disks, respectively.

With the above arrangement, to make the groove in the rotary cylinder and the partition plate, the rotary cylinder and the partition plate can be machined easily and with a good accuracy without accompanying of a difficult operation.

According to a third aspect and feature of the present invention, in addition to the second feature, the rotary cylinder and the partition plate have through-bores defined therein, respectively, so that the rotary cylinder and partition plate are fixed by bolts inserted through the through-bores, the through-bores being disposed at locations where they are not aligned with an intake port and a discharge port for permitting a gas to flow into and out of the compressing mechanism.

With the above arrangement, a lower-pressure gas and a higher-pressure gas cannot flow into the through-bores in every rotation of the cylinder and hence, it is possible to prevent a reduction in efficiency due to the provision of the through-bores.

According to a fourth aspect and feature of the present invention, in addition to the third feature, the through-bore defined in the rotary cylinder is provided with a larger-diameter portion for receiving a head of the bolt.

With the above arrangement, the head of the bolt cannot protrude from the rotary cylinder and hence, to machine a member facing the bolt head, it is unnecessary to make a groove in this member for avoiding the interference with the bolt head and thus, the hermetic compressor can be produced at a lower cost.

According to a fifth aspect and feature of the present invention, in addition to the second feature, the rotary cylinder and the partition plate have through-bore defined therein, respectively, so that the rotary cylinder and the partition plate are fixed by pins fitted into the through-bores, the through-bores being disposed at locations where they are not aligned with an intake port and a discharge port for permitting a gas to flow into and out of the compressing mechanism.

With the above arrangement, a lower-pressure gas or a higher-pressure gas cannot flow into the through-bores in every rotation and hence, it is possible to prevent the reduction in efficiency due to the provision of the through-bores.

According to a sixth aspect and feature of the present invention, in addition to the second feature, the partition plate has pin-insertion bores defined therein, and each of the rotary cylinders located on opposite sides of the partition plate has bottomed pin-receiving bores defined therein, so that the relative rotation of the rotary cylinders of the adjacent compressing mechanisms is limited by pins inserted into the pin-receiving bores and the pin insertion bores.

With the above arrangement, a gas cannot flow into and out of the compressing mechanism through the bottomed pin receiving bores in the rotary cylinder. This provides an increased degree of freedom in design concerning the positions for disposition of and the sizes of the intake port and the discharge port. As a result, it is possible to select a port shape in which intake and discharge losses are small, and hence, it is possible to enhance the efficiency of the compressor.

According to a seventh aspect and feature of the present invention, in addition to the second feature, the rotary cylinder and the partition plate are fitted in a recess-projection manner with each other by a recess and a projection formed on opposed faces thereof.

With the above arrangement, the adjacent rotary cylinders can be separated from each other, while limiting the relative angle of the rotary cylinder by the fitting of the rotary cylinder and the partition plate. Therefore, a gas force applied to one of the rotary cylinders is not transmitted to the other rotary cylinder and as a result, the rotary cylinders cannot be inclined together during rotation thereof. Thus, it is possible to prevent the partial abutment of the rotary cylinder against the member which is sliding contact with the rotary cylinder to reduce the sliding wear of the outer peripheral surface of the rotary cylinder.

According to an eighth aspect and feature of the present invention, in addition to the second feature, the rotary cylinder and the partition plate are fixed to each other by welding.

With the above arrangement, a working technique commonly used in the machining can be utilized, thereby preventing the relative rotation between the adjacent rotary cylinders.

According to a ninth aspect and feature of the present invention, there is provided a hermetic compressor comprising a plurality of compressing mechanisms each of which includes a rotary cylinder having a groove, and a piston which is slidable in the groove, so that a compressing stroke is carried out by rotation of the piston on a locus of a radius E about a rotational center provided by a location spaced at a distance E apart from the center of the rotary cylinder; a partition plate being interposed between said rotary cylinders of the adjacent compressing mechanisms, the partition plate being provided with a communication bore through which a shaft is passed, the shaft being provided with cranks on which the pistons can be mounted; and a motor mechanism for driving the pistons of the compressing mechanisms by the common shaft, at least one of the compressing mechanisms being different in phase in a compressing stroke from the other compressing mechanism, the rotary cylinders of the adjacent compressing mechanisms and the partition plate sandwiched between these rotary cylinders being formed from an integrally formed piece.

With the above arrangement, it is unnecessary to provide a means for connecting the rotary cylinder and the partition plate which are separate from each other, as in the arrangement of any of the first to eighth features, and it is unnecessary to provide through-bores in the rotary cylinder, as in the arrangement of the third feature. This provides an increased freedom degree in design concerning the positions for disposition of and the sizes of an intake port and a discharge port. As a result, it is possible to select a port shape in which intake and discharge losses are small, and hence, it is possible to enhance the efficiency of the compressor.

According to a tenth aspect and feature of the present invention, there is provided a hermetic compressor comprising first and second compressing mechanisms each of which includes a rotary cylinder having a groove, and a piston which is slidable in the groove, so that a compressing stroke is carried out by rotation of the piston on a locus of a radius E about a rotational center provided by a location spaced at a distance E apart from the center of the rotary cylinder, all the rotary cylinders being connected together, all the pistons being driven by a common shaft, and the first and second compressing mechanisms being different in phase in a compressing stroke, the first and second compressing mechanisms being mounted between an upper bearing and a lower bearing, the upper bearing having an intake port and a discharge port provided therein for the first compressing mechanism, and the lower bearing having an intake port and a discharge port provided therein for the second compressing mechanism, the intake ports and the discharge ports being provided so that they do not communicate simultaneously with a compressive space defined by the rotary cylinder and the piston at all rotational angles of the shaft.

With the above arrangement, a high-pressure refrigerant gas cannot be leaked to the intake side through the compressive space during a compressing stroke and hence, a high suction efficiency (volume efficiency) can be realized.

According to an eleventh aspect and feature of the present invention, in addition to the tenth feature, the intake port is disposed so that it communicates with the compressive space which is in a volume-increasing course, at positions of all rotational angles excluding a suction starting point where the volume of the compressive space is smallest (minimum) and a suction completing point where the compressive space is largest (maximum).

With the above arrangement, the intake port cannot face the compressive space at the suction starting point and the suction completing point and hence, the intake port can be reliably cut off from the compressing stroke in the compressive space. Thus, a refrigerant gas cannot be leaked to the intake side during the compressing stroke and hence, a high suction efficiency can be realized. In addition, since the intake port communicates with the compressive space at all the suction stroke excluding the suction starting point and the suction completing point, the refrigerant gas is sucked through the intake port into the compressive space with a small pressure loss, when the volume of the compressive space is increased. As a result, a high suction efficiency can be realized.

According to a twelfth aspect and feature of the present invention, in addition to the eleventh feature, the intake port has a crescent shape extending along a side edge of the groove at a position of a rotational angle of the rotary cylinder at the time when the compressive space is smallest or largest, an outer edge of the crescent shape being formed into an arc conforming with and extending along a locus of movement of an end edge of the groove.

With the above arrangement, the intake port can be formed into a shape free from any sufficiency and shortage, which can be employed for the mechanism at the suction stroke excluding the suction starting point and the suction completing point. As a result, a high suction efficiency can be realized.

According to a thirteenth aspect and feature of the present invention, in addition to the tenth or twelfth feature, the discharge port is comprised of a plurality of ports spaced apart from one another along a side edge of the groove at a position of a rotational angle of the rotary cylinder at the time when the compressive space is smallest or largest, the plurality of ports being provided with discharge valves, respectively and disposed at locations where they do not communicate with the compressive space at a compression starting point and a compression completing point in the compressive space.

With the above arrangement, it is possible to avoid a phenomenon of leakage of refrigerant gas on the high-pressure side to the compressive space. In addition, the refrigerant gas in the compressive space is discharged via the plurality of discharge ports to the higher-pressure side, as the compressing stroke is advanced, while the compressive space is rotated. Therefore, a phenomenon of over-compression can be avoided, and a high compression efficiency can be realized.

According to a fourteenth aspect and feature of the present invention, there is provided a hermetic compressor comprising first and second compressing mechanisms which are mounted within a casing and each of which includes a rotary cylinder having a groove, and a piston which is slidable in the groove, so that the suction and compression are carried out by rotation of the piston on a locus of a radius E about a center provided by a point spaced at a distance E apart from the center of the rotary cylinder, the two rotary cylinders of the first and second compressing mechanisms being connected to each other at a location where the first and second compressing mechanisms are different in phase in a compressing stroke, the two pistons being driven by a common crankshaft, the piston being formed into a shape such that its sectional contour is comprised of two arcs and two parallel straight lines having a length a, the groove in the rotary cylinder being formed into a shape such that it is comprised of arcs assuming the substantially same shape as the arcs forming the piston, and two parallel straight. lines having a length of 4 E+a.

With the above arrangement, the piston and the groove are in face contact with each other rather than in line contact. As a result, the leakage of the refrigerant from the higher-pressure compressing chamber to the lower-pressure compressing chamber is reduced and hence, the suction and compression efficiencies can be enhanced. In this case, the piston and the groove in the rotary cylinder can be finished easily and with a high accuracy by a simple working machine such as a drilling machine, a lathe and a milling machine.

According to a fifteenth aspect and feature of the present invention, the sectional contour of the piston is formed by cutting a cylindrical member in parallel.

With the above arrangement, the suction and compression efficiencies can be enhanced, and flat faces of the piston comprised of the parallel straight lines may be formed on the contour of the cylindrical member fabricated by a working machine such as lathe and hence, the piston can be made extremely easily and with a high accuracy. Thus, the manufacturing cost can be reduced.

According to a sixteenth aspect and feature of the present invention, the arc forming the sectional contour of the piston is semi-circular.

With the above arrangement, the suction and compression efficiencies can be enhanced, and any corner is not created at the connection between the semi-circular arc and the straight line, leading to a smooth connection, whereby the sliding movement of the piston can be conducted smoothly.

The above and other objects and advantages of the invention will become apparent from the following description of the preferred embodiments in conjunction with the accompanying drawings.

BEST MODE FOR CARRYING OUT THE PRESENT INVENTION

FIG. 1

is a vertical sectional view of a hermetic compressor according to an embodiment of the present invention;

FIG. 2

is a sectional view taken along a line II—II in

FIG. 1

;

FIG. 3

is a sectional view taken along a line III—III in

FIG. 1

;

FIGS. 4

a

to

4

h

are views for explaining the operation of a compressing mechanism in the embodiment;

FIG. 5

is a plan view of a first assembly as viewed from the side of a first rotary cylinder;

FIG. 6

is a vertical sectional view of the first assembly shown in

FIG. 5

;

FIG. 7

is a plan view of the first assembly as viewed from the side of a second rotary cylinder;

FIG. 8

is a plan view of a second assembly as viewed from the side of a first rotary cylinder;

FIG. 9

is a vertical sectional view of the second assembly shown in

FIG. 8

;

FIG. 10

is a plan view of the second assembly as viewed from the side of a second rotary cylinder;

FIG. 11

is a plan view of a third assembly as viewed from the side of a first rotary cylinder;

FIG. 12

is a vertical sectional view of the third assembly shown in

FIG. 11

;

FIG. 13

is an exploded perspective view of a fourth assembly;

FIG. 14

is a plan view of a fifth assembly as viewed from the side of a first rotary cylinder;

FIG. 15

is a vertical sectional view of the fifth assembly shown in

FIG. 14

;

FIG. 16

is a plan view of the fifth assembly as viewed from the side of a second rotary cylinder;

FIG. 17

is a plan view of a sixth assembly as viewed from the side of a first rotary cylinder;

FIG. 18

is a vertical sectional view of the sixth assembly shown in

FIG. 17

;

FIG. 19

is a plan view of the sixth assembly as viewed from the side of a second rotary cylinder;

FIG. 20

is a vertical sectional view of the entire structure of the compressor according to another embodiment of the present invention;

FIG. 21

is a sectional view taken along a line II—II in

FIG. 20

in the other embodiment;

FIG. 22

is a sectional view taken along a line III—III in

FIG. 20

in the other embodiment;

FIGS. 23

a

to

23

h

are views for explaining the operation of a compressing mechanism in the other embodiment;

FIG. 24

is a view similar to the sectional view taken along the line II—II in

FIG. 20

, but according to a second embodiment;

FIG. 25

is a view similar to the sectional view taken along the line III—III in

FIG. 20

, but according to the second embodiment;

FIGS. 26

a

to

26

i

are views for explaining the principle of the compressing mechanism.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

The present invention will now be described by way of embodiments with reference to the accompanying drawings.

FIG. 1

is a, vertical sectional view of a hermetic compressor according to an embodiment of the present invention;

FIG. 2

is a sectional view taken along a line II—II in

FIG. 1

;

FIG. 3

is a sectional view taken along a line III—III in

FIG. 1

; and

FIGS. 4

a

to

4

h

are views for explaining the movement of a compressor mechanism section in the embodiment.

Referring to

FIG. 1

, a hermetic compressor according to the embodiment includes a motor

30

and a compressor mechanism section

40

within a shell

10

constituting a hermetic container.

The shell

10

has a discharge pipe

11

at its upper portion, and two intake pipes

12

a

and

12

b

at a side of its lower portion.

The motor

30

comprises a stator

31

fixed to the shell

10

, and a rotor

32

which is rotated. The rotation of the rotor

32

, is transmitted to the compressor mechanism section

40

by a shaft

33

.

The compressor mechanism section

40

includes a first compressing mechanism

40

a

comprising a first rotary cylinder

41

a

and a first piston

42

a

, and a second compressing mechanism

40

b

comprising a second rotary cylinder

41

b

and a second piston

42

b

. The first rotary cylinder

41

a

has an elliptic groove

43

a

, and the second rotary cylinder

41

b

has an elliptic groove

43

b

. The first piston

42

a

is slidably provided in the groove

43

a

, and the second piston

42

b

is slidably provided in the groove

43

b

. The members constituting the first and second compressing mechanisms

40

a

and

40

b

are of the same size and shape.

The first and second compressing mechanisms

40

a

and

40

b

are partitioned from each other by a partition plate

44

. As will be described in detail hereinafter, the first rotary cylinder

41

a

, the second rotary cylinder

41

b

and the partition plate

44

are connected together and moved in unison with one another. However, the first and second rotary cylinders

41

a

and

41

b

are connected to each other with the grooves

43

a

and

43

b

offset from each other at 90 degree, so that the phases of compressing strokes are different at 180 degree from each other.

On the other hand, the first and second pistons

42

a

and

42

b

are fitted over first and second cranks

33

a

and

33

b

, respectively. The first and second cranks

33

a

and

33

b

are provided so that their eccentric directions are different at 180 degree from each other.

The first and second compressing mechanisms

40

a

and

40

b

are sandwiched from above and below by an upper bearing

50

a

and a lower bearing

50

b

and surrounded by a tubular casing

51

.

The upper bearing

50

a

is provided with an intake port

51

a

and a discharge port

52

a

for the first compressing mechanism

40

a

, and the lower bearing

50

b

is provided with an intake port

51

b

and a discharge port

52

b

for the second compressing mechanism

40

b

. The positions of disposition of the intake ports

51

a

and

51

b

and the discharge ports

52

a

and

52

b

will be described hereinafter. Provided in the discharge ports

52

a

and

52

b

are valves

53

a

and

53

b

which are opened by a predetermined pressure, and valve stops

54

a

and

54

b

for limiting the opening movements of the valves

53

a

and

53

b

. The intake port

51

a

communicates with the intake pipe

12

a

, and the intake port

51

b

communicates with the intake pipe

12

b

. The intake pipes

12

a

and

12

b

are connected to an accumulator

60

.

The flow of a refrigerant in the hermetic compressor having the above-described arrangement will be described below in brief.

The gas refrigerant within the accumulator

60

is introduced through the intake pipes

12

a

and

12

b

into the shell

10

and drawn through the intake port

51

a

and

51

b

into the first and second compressing mechanisms

40

a

and

40

b

. When the refrigerant compressed in the first and second compressing mechanisms

40

a

and

40

b

reaches a predetermined pressure, it pushes up the valves

53

a

and

53

b

and is discharged through the discharge ports

52

a

and

52

b

into the shell

10

. In this case, the discharging timings in the first and second compressing mechanisms

40

a

and

40

b

are not the same as each other, because the phases are different at 180 degree from each other. The refrigerant discharged into the shell

10

is passed through an area around the motor

30

and discharged to the outside of the shell

10

through the discharge pipe

11

provided at the upper portion of the shell

10

.

The relationship between the shaft

33

, the pistons

42

a

and

42

b

and the rotary cylinders

41

a

and

41

b

in the first and second compressing mechanisms

40

a

and

40

b

will be described below with reference to

FIGS. 2 and 3

.

The shaft

33

adapted to transmit the rotation of the motor

30

is rotated about a point B. The center C of the cranks

33

a

and

33

b

provided on the shaft

33

is eccentric by a distance E from the center B of rotation of the shaft

33

. The center C of the cranks

33

a

and

33

b

is also the center of rotation of the pistons

42

a

and

42

b

. Namely, the pistons

42

a

and

42

b

perform a rotating movement about the center C of the cranks

33

a

and

33

b

. On the other hand, the rotary cylinders

41

a

and

41

b

have the center of rotation provided by a position spaced apart at the distance E from the center B of rotation of the shaft

33

. Therefore, when the center C of the cranks

33

a

or the piston

42

a

is spaced to the maximum apart from the center A of rotation of the rotary cylinder

41

a

, the largest and smallest spaces are formed in the groove by the piston

42

a

, as shown in FIG.

2

. The second compressing mechanism

40

b

has a phase difference of 180 degree from the phase of the first compressing mechanism

40

a

and hence, when the first compressing mechanism

40

a

is in a state shown in

FIG. 2

, the center C of rotation of the crank

33

b

or the piston

42

b

in the second compressing mechanism

40

b

overlaps the center A of rotation of the rotary cylinder

41

b

, as shown in FIG.

3

. Therefore, the space section in the groove

43

b

is divided into two equal spaces by the piston

42

b

, as shown in FIG.

3

. The spaces defined in the groove

43

a

in the rotary cylinder

41

a

by the piston

42

a

and the spaces defined in the groove

43

b

in the rotary cylinder

41

b

by the piston

42

b

are called “compressive spaces” hereinafter.

The refrigerant gas sucking and compressing strokes will be described below with reference to FIG.

4

. The sucking and compressing strokes in the first compressing mechanism

40

a

will be described, but the second compressing mechanism

40

b

provides the same strokes, except that the phase in

FIG. 4

is different by 180 degree from that in the first compressing mechanism

40

a.

FIGS. 4

a

to

4

h

show states in which the shaft

33

has been rotated through every 90 degree, respectively.

When the shaft

33

is not rotated as shown in

FIG. 4

a

, the groove

43

a

is in a state in which the one of the compressive space I is largest, and the other compressive space II is smallest.

The volume of the one compressive space I is gradually decreased from the state shown in

FIG. 4

b

in which the shaft

33

has been rotated through 90 degree via the state shown in

FIG. 4

c

in which the shaft

33

has been rotated through 180 degree to the state shown in

FIG. 4

d

in which the shaft

33

has been rotated through 270 degree, whereby the compressed refrigerant is discharged from the discharge port

52

a

. In the compressive space I, the compressing stroke is finished in the state shown in

FIG. 4

e

in which the shaft

33

has been rotated through 360 degree.

The volume of the other compressive space II is gradually increased from the state shown in

FIG. 4

b

in which the shaft

33

has been rotated through 90 degree via the state shown in

FIG. 4

c

in which the shaft

33

has been rotated through 180 degree to the state shown in

FIG. 4

d

in which the shaft

33

has been rotated through 270 degree, whereby the compressed refrigerant is sucked from the intake port

52

a

. In the compressive space II, the sucking stroke is finished in the state shown in

FIG. 4

e

in which the shaft

33

has been rotated through 360 degree.

In the states shown in

FIG. 4

e

to

FIG. 4

h

, the sucking stroke is carried out in the one compressive space I, and the compressing stroke is carried out in the other compressing space II. When the shaft

33

is further rotated through 90 degree from the state shown in

FIG. 4

h

, the state shown in

FIG. 4

a

is obtained.

In this way, the compressing and sucking strokes are carried out in the two compressive spaces I and II defined in the groove

43

a

, respectively, while the shaft

33

is rotated through 720 degree.

With respect to the positions for disposition of the intake ports

51

a

and

51

b

and the discharge ports

52

a

and

52

b

, the intake port

51

a

and the intake port

51

b

as well as the discharge port

52

a

and the discharge port

52

b

are disposed in an axially opposed relation to each other. The intake port

51

a

and the discharge port

52

a

defined in the upper bearing

50

a

will be described representatively with reference to

FIG. 4

a

. The intake port

51

a

and the discharge port

52

a

are located to lie inside the locus of rotation of the end edge of the groove

43

a

and sideways of the elliptic groove

43

a

. More specifically, the intake port

51

a

has a crescent shape (see

FIGS. 4

a

and

4

e

showing the states of the shaft which is not rotated and has been rotated through 360 degree) having an inner edge extending one of side edges of the groove

43

a

when the compressive spaces I and II assume the largest or smallest volume. When the shaft is not rotated and has been rotated through 360 degree, the crescent-shaped intake port

51

a

does not communicate with both the compressive spaces I and II, and when the shaft

33

assumes a position of another angle, the crescent-shaped intake port

51

a

continuously communicates with the compressive space I or II, whereby an end edge of the crescent-shaped intake port

51

a

at a suction starting point and a suction completing point is formed to suck the refrigerant gas. Namely, the suction starting point and a suction completing point of the crescent-shaped intake port

51

a

are set at locations slightly offset from the groove

43

a

, when the compressive space I or II assume the largest or smallest volume. The outer edge of the crescent-shaped intake port

51

a

is set into an arc extending in substantial conformity to the locus of movement and along the locus of movement of the end edge of the groove

43

a

at an intermediate stroke between the suction starting point and the suction completing point.

Similarly, when the shaft

33

is not rotated and has been rotated through 360 degree in which the compressive spaces I and II assume the smallest or largest volume, the discharge port

52

a

is constituted of a pair of circular port portions disposed at a distance along the other side edge of the groove

43

a

, so that it does not communicate with both the compressive spaces I and II. The pair of discharge port portions

52

a

and the crescent-shaped intake port

51

a

are disposed so that they do not communicate with each other through the compressive space I or II at locations of all rotational angles of the shaft

33

.

According to the embodiment, even if the piston is located at the center of the rotary cylinder in one of the compressing mechanisms, it can be avoided that the driving force from the piston does not serve as a rotating force for the rotary cylinder, because the other compressing mechanism provides a rotating force. In addition, the pistons can be disposed symmetrically by ensuring that the phase difference between the two compressing mechanisms is 180 degree, whereby the production of the hermetic compressor can be carried out easily. The freedom degree of setting of the positions of the intake port and the discharge port is increased by providing intake port and the discharge port in the upper and lower bearing, respectively. Therefore, it is possible to regulate the compression ratio and to prevent the over-compression by the positions of the intake port and the discharge port. Further, since the phases of the first and second compressing mechanisms are different from each other by 180 degree, and the intake port in the upper bearing and the intake port in the lower bearing are provided on the same axis, the position of mounting of the intake pipe can be the same side, and a piping cannot be drawn around for connection of the intake pipe to the accumulator or the like.

The shapes and positions of the intake port

51

a

(

51

b

) and the discharge port

52

a

(

52

b

) are determined, so that they do not simultaneously communicate with one of the compressive spaces at any rotational angle of the shaft

33

. Therefore, during compression, the high-pressure refrigerant gas is leaked toward the intake side through the compressive space and hence, a high suction efficiency (volume efficiency) can be realized. Further, the intake port

51

a

(

51

b

) is set into a shape such that it does not face the compressive space at the suction starting point and the suction completing point, leading to a construction in which the intake port

51

a

(

51

b

) is reliably cut away from the compressing stroke in the compressive space. As a result, the refrigerant gas cannot be leaked toward the intake side during compression and hence, the high suction efficiency can be realized. In addition, the intake port

51

a

(

51

b

) communicates with the compressive space at all the suction stroke excluding the suction starting point and the suction completing point by virtue of the shape of the crescent-shaped intake port

51

a

(

51

b

) and hence, when the volume of the compressive space is increased, the refrigerant gas is drawn or sucked from the intake port

51

a

(

51

b

) with a small pressure loss and consequently, the high suction efficiency can be realized.

In addition, since the outer edge of the crescent-shaped intake port

51

a

(

51

b

) is set into the arc extending in substantial conformity to the locus of movement and along the locus of movement of the end edge of the groove

43

a

(

43

b

) at the intermediate stroke between the suction starting point and the suction completing point, an affective suction efficiency can be realized by the crescent-shaped intake port

51

a

(

51

b

) free from any sufficiency and shortage. On the other hand, the pair of discharge ports

52

a

(

52

b

) including the discharge valve mechanisms

53

and

54

are provided at locations where they do not communicate with the compressive space at a compression starting point and a compression completing point and hence, a phenomenon of leakage of the refrigerant gas in the high-pressure space to the compressive space is not produced. In addition, the refrigerant gas in the compressive space is discharged via the plurality of discharge ports into the high-pressure space with advancing of the compressing stroke while permitting the compressive space to be rotated. Therefore, a phenomenon of over-compression cannot be produced and as a result, a high compressing efficiency can be realized.

FIGS. 5

to

7

show a first assembly

110

comprised of the first and second rotary cylinders

41

a

and

41

b

and the partition plate

44

.

FIG. 5

is a side view of the assembly

110

as viewed from the side of the first rotary cylinder

41

a

;

FIG. 6

is a vertical sectional view of the assembly; and

FIG. 7

is a side view of the assembly as viewed from the side of the second rotary cylinder

41

b

. A one-dot dashed line

111

in

FIG. 5

indicates a locus of rotation of the groove

43

a

with the rotation of the first rotary cylinder

41

a

, namely, a circumcircle of the groove

43

a

. Four bolt insertion bores

112

and

113

circumferentially spaced at equal distances apart from one another are defined respectively in the first rotary cylinder

41

a

and the partition plate

44

around the outer periphery of the locus of rotation (see FIG.

6

). Each of the bolt insertion bores

112

has a larger-diameter portion

112

a

for receiving ahead of a fastening bolt

114

at a location adjacent the outer surface of the first rotary cylinder

41

a

. Four threaded bores

115

are defined through the second cylinder

41

b

at locations corresponding to the bolt insertion bores

112

in the first rotary cylinder

41

a

, as shown in FIG.

7

.

The first assembly

110

is produced by disposing between the second rotary cylinders

41

a

and

41

b

and then inserting the fastening bolt

114

from the side of the first rotary cylinder

41

a

into the second rotary cylinder

41

b

to threadedly engage the bolt

114

into the threaded bore

115

in the second rotary cylinder

41

b

. With respect to the positions for disposition of the intake ports

51

a

and

51

b

and the discharge ports

52

a

and

52

b

, those of the intake port

51

a

and the discharge port

52

a

will be described representatively. The intake port

51

a

and the discharge port

52

a

are located to lie inside the locus

111

of rotation of the groove

43

a

and sideways of the elliptic groove

43

a.

The first assembly

110

is of a construction made by disposing the partition plate

44

between the first and second rotary cylinders

41

a

and

41

b

, and connecting the first and second rotary cylinders

41

a

and

41

b

to each other by the fastening bolt

114

in a state in which the partition plate

44

has been sandwiched between the first and second rotary cylinders

41

a

and

41

b

. Therefore, when each of the rotary cylinders

41

a

and

41

b

are to be made, they can be machined separately. Namely, each of the first and second rotary cylinders

41

a

and

41

b

is of a simple configuration in which the elliptic groove

43

a

,

43

b

is merely provided at a central portion of a disk. To form the rotary cylinders

41

a

and

41

b

, the grooves

43

a

and

43

b

can be machined with a good accuracy and easily by cutting or the like and hence, the cost for producing the rotary cylinders

41

a

and

41

b

can be reduced.

Since the first rotary cylinder

41

a

is provided with the larger-diameter portion

112

a

for receiving the head

114

a

of the fastening bolt

114

, the head

114

a

of the fastening bolt

114

cannot protrude from the first assembly

110

. Therefore, it is not required that a groove for avoiding the interference with the bolt head

114

a

is made by machining in the upper bearing

50

a

facing the bolt head

114

a

, whereby the cost due to the machining of the upper bearing

50

a

can be reduced. The bolt insertion bores

112

and the threaded bore

115

made through the first and second rotary cylinders

41

a

and

41

b

are disposed at locations where they cannot face the intake ports

51

a

and

51

b

and the discharge ports

52

a

and

52

b

. Therefore, the insertion bores

52

a

and

52

b

and the threaded bore

115

cannot be aligned with the intake ports

51

a

and

51

b

and the discharge ports

52

a

and

52

b

with rotation of the first and second rotary cylinders

41

a

and

41

b

. Thus, a lower-pressure gas or a higher-pressure gas cannot flow into the insertion bores

52

a

and

52

b

and the threaded bore

115

upon every rotation of the rotary cylinder and hence, it is possible to prevent a reduction in efficiency due to the flowing of the gas into the bores

112

and

115

.

In place of the threaded bore

115

defined in the second rotary cylinder

41

b

, a bolt insertion bore may be made in the second rotary cylinder

41

b

, and a larger-diameter portion for receiving a nut adapted to be threadedly engaged with the fastening bolt

114

may be provided in the bolt insertion bore.

FIGS. 8

to

10

shows a second assembly

120

comprised of first and second rotary cylinders

41

a

and

41

b

and a partition plate

44

.

FIG. 8

is a side view of the assembly

120

as viewed from the side of the first rotary cylinder

41

a

;

FIG. 9

is a vertical sectional view of the assembly

120

; and

FIG. 10

is a side view of the assembly as viewed from the side of the second rotary cylinder

41

b

. The second assembly

120

corresponds to a modification to the first assembly

110

. As can be understood from

FIG. 8

, the positions for disposition of the intake port

51

a

and the discharge port

52

a

are similar to those in the first assembly

110

. However, in place of the bolt insertion bores

112

and

113

and the threaded bore

115

in the first assembly

110

, pin insertion bores

121

,

122

and

123

are defined in corresponding elements

41

a

,

44

and

41

b

, respectively, and the first and second rotary cylinders

41

a

and

41

b

located with the partition plate

44

interposed therebetween are integrally connected together by inserting a pin

124

through the pin insertion bores

121

,

122

and

123

.

With the second assembly

120

, when the rotary cylinders

41

a

and

41

b

are to be produced, they can be machined separately, as in the first assembly

110

. To form the rotary cylinders

41

a

and

41

b

, grooves

43

a

and

43

b

can be made by machining such as cutting with a good accuracy and easily. The pin insertion bores

121

and

123

made through the first and second rotary cylinders

41

a

and

41

b

are disposed at locations where they cannot face the intake ports

51

a

and

51

b

and the discharge ports

52

a

and

52

b

. Therefore, the pin insertion bores

121

and

123

cannot be aligned with the intake ports

51

a

and

51

b

and the discharge ports

52

a

and

52

b

with rotation of the first and second rotary cylinders

41

a

and

41

b

. Thus, it is possible to prevent a reduction in efficiency due to the flowing of a gas into the bores

121

and

123

.

FIGS. 11 and 12

show a third assembly

130

comprised of first and second rotary cylinders

41

a

and

41

b

and a partition plate

44

.

FIG. 11

is a side view of the third assembly as viewed from the side of the first rotary cylinder

41

a

; and

FIG. 12

is a vertical sectional view of the third assembly. In the third assembly

130

, four bottomed pin receiving bores

131

a

and

131

b

circumferentially spaced at equal distances apart from one another are defined respectively in opposed inner surfaces of the first and second rotary cylinders

41

a

and

41

b

(not shown in

FIG. 9

) around the outer periphery of the locus

111

of rotation. Pin insertion bores

132

are defined in the partition plate

44

at locations corresponding to the pin receiving bores

131

a

and

131

b.

The third assembly

130

is made by superposing the first or second rotary cylinder

41

a

or

41

b

and the partition plate

44

one onto another, inserting a pin

133

into each of the bores, and superposing the remaining first or second rotary cylinder

41

a

or

41

b

. In the third assembly

130

, the relative rotation of the first and second rotary cylinders

41

a

and

41

b

located with the partition plate sandwiched therebetween is prohibited.

With the third assembly

130

, when the rotary cylinders

41

a

and

41

b

are to be produced, they can be machined separately, as in the first and second assemblies

110

and

120

. Therefore, to form the rotary cylinders

41

a

and

41

b

, grooves

43

a

and

43

b

can be made with a good accuracy and easily by machining such as cutting. Since the bores

131

a

and

131

b

for receiving the pins

133

provided in the first and second rotary cylinders

41

a

and

41

b

are bottomed, a gas cannot flow into and out of the intake ports

51

a

and

51

b

and the discharge ports

52

a

and

52

b

through the pin receiving bores

131

a

and

131

b

. This provides an increased degree of freedom in design concerning the positions for disposition of and the sizes of the intake ports

51

a

and

51

b

and the discharge ports

52

a

and

52

b

. As a result, it is possible to select a port shape in which intake and discharge losses are small, and from this viewpoint, it is possible to enhance the efficiency of the compressor.

FIG. 13

is an exploded perspective view of a fourth assembly

140

comprised of first and second rotary cylinders

41

a

and

41

b

and a partition plate

44

. In the fourth assembly

140

, recesses and projections are formed on opposed surface of the first rotary cylinder

41

a

and the partition plate

44

and opposed surfaces of the second rotary cylinder

41

b

and the partition plate

44

, so that the relative rotation of the elements is prohibited by fitting of the projections and recesses with each other. More specifically, two recesses

131

are formed at a distance of 180° in diametrical portions of the first rotary cylinder

41

a

, and two projections

132

corresponding to the recesses are formed on the partition plate

44

. In addition, two recesses

133

are formed at a distance of 180° in diametrical portions of the second rotary cylinder

41

b

, and two projections

134

corresponding to the recesses are formed on the partition plate

44

. Alternatively, a recess may be provided in each of the first and second rotary cylinders

41

a

and

41

b

, and a recess may be provided in the partition plate

44

.

With the fourth assembly

140

, the two rotary cylinders

41

a

and

41

b

can be separated from each other, while limiting the relative angle of the first and second rotary cylinders

41

a

and

41

b

by recess-protrusion fitting of the first and second rotary cylinders

41

a

and

41

b

with the partition plate

44

. Therefore, a gas force applied to one of the rotary cylinders is not transmitted to the other rotary cylinder and as a result, the rotary cylinders

41

a

and

41

b

cannot be inclined together during rotation of the fourth assembly. Thus, it is possible to prevent the partial abutment of the rotary cylinders

41

a

and

41

b

against the upper and lower bearings

50

a

and

50

b

to reduce the sliding wear of outer peripheral portions of the rotary cylinders

41

a

and

41

b.

FIGS. 14

to

16

show a fifth assembly comprised of first and second rotary cylinders

41

a

and

41

b

and a partition plate

44

.

FIG. 14

is a side view of the assembly

150

as viewed from the side of the first rotary cylinder

41

a

;

FIG. 15

is a vertical sectional view of the assembly

150

; and

FIG. 16

is a side view of the assembly

150

as viewed from the side of the second rotary cylinder

41

b

. Reference character

151

in

FIG. 15

indicates a weld zone. As can be understood from

FIG. 15

, the first and second rotary cylinders

41

a

and

41

b

and the partition plate

44

are integrally connected together by welding. In this case, as can be seen from

FIG. 15

, the partition plate

44

may have a diameter considerably smaller than those of the first and second rotary cylinders

41

a

and

41

b

, or may have a diameter substantially equal to those of the first and second rotary cylinders

41

a

and

41

b.

With the fifth assembly

150

, when the rotary cylinders

41

a

and

41

b

are to be produced, they can be machined separately, as in the first and second assemblies

110

and

120

, and to form the rotary cylinders

41

a

and

41

b

, grooves

43

a

can be made with a good accuracy and easily by machining such as cutting. When the diameter of the partition plate

44

is considerably smaller than those of the first and second rotary cylinders

41

a

and

41

b

, the area of contact between the outer peripheral portion of the partition plate

44

and the first and second rotary cylinders

41

a

and

41

b

can be increased, and hence, they can be fixed with a higher strength at a small number of welded points.

FIGS. 17

to

19

show a sixth assembly

160

comprised of first and second rotary cylinders

41

a

and

41

b

and a partition plate

44

.

FIG. 17

is a side view of the assembly

160

as viewed from the side of the first rotary cylinder

41

a

;

FIG. 18

is a vertical sectional view of the assembly

160

; and

FIG. 19

is a side view of the assembly

160

as viewed from the side of the second rotary cylinder

41

b

. As can be understood from

FIG. 18

, the first and second rotary cylinders

41

a

and

41

b

and the partition plate

44

are formed integrally with each other.

With the sixth assembly

160

, a means for mechanically fastening the two first and second rotary cylinders

41

a

and

41

b

, e.g., a member such as a bolt and a pin, is not required, and a means such as recess-projection fitting for limiting the relative rotation between the first and second rotary cylinders

41

a

and

41

b

is not required. It is unnecessary to define through-bores in the rotary cylinders, as in the first and second assemblies

110

and

120

and hence, the flowing-out of a gas through the through-bores cannot be produced. This provides an increased degree of freedom in design concerning the positions for disposition of and the sizes of the intake ports

51

a

and the discharge ports

52

a

and

52

b

. As a result, it is possible to select a port shape in which intake and discharge losses are small, and from this viewpoint, it is possible to enhance the efficiency of the compressor.

The phase difference between the two compressing mechanisms is 180 degree in the embodiment, but is not limited to this angle and may be 90 degree, 270 degree or any angle other than these angles. The embodiment has been described as being provided with the two compressing mechanisms, but three or more compressing mechanisms may be provided.

Another embodiment of a compressor according to the present invention will now be described with reference to the drawings.

FIG. 20

is a vertical sectional view of a hermetic compressor having first and second compressing mechanisms according to the present embodiment;

FIG. 21

is a sectional view taken along a line II—II in

FIG. 20

;

FIG. 22

is a sectional view taken along a line III—III in

FIG. 20

; and

FIG. 23

is a view for explaining the operation of the compressing mechanism in this embodiment.

In

FIGS. 20

to

23

, members or portions having the same function as those in the embodiment shown in

FIGS. 1

to

4

are designated by like reference characters.

As shown in

FIG. 20

, a hermetic compressor in this embodiment includes a motor

30

and a compressor mechanism section

40

within a shell

10

constituting a hermetic container.

The shell

10

includes a discharge pipe

11

at its upper portion, and two intake pipes

12

c

and

12

d

on a side of its lower portion.

The motor

30

comprises a stator

31

fixed to the shell

10

, and a rotor

32

which is rotated. The rotation of the rotor

32

is transmitted to the compressor mechanism section

40

by a shaft

33

.

The compressor mechanism section

40

includes a first compressing mechanism

40

c

comprised of a first rotary cylinder

41

c

and a first piston

42

c

, and a second compressing mechanism

40

d

comprised of a second rotary cylinder

41

d

and a second piston

42

d

. The first rotary cylinder

41

c

has a first groove

43

c

, and the second rotary cylinder

41

d

has a second groove

43

d

. The first piston

42

c

is slidably provided in the first groove

43

c

, and the second piston

42

d

is slidably provided in the second groove

43

d

. The members constituting the first and second compressing mechanisms

40

c

and

40

d

are of the same size and shape.

As shown in

FIGS. 21 and 22

, each of the first and second pistons

42

c

and

42

d

is formed by cutting a cylindrical member in parallel, so that the contour of its section is comprised of two arcs

70

,

70

and two parallel straight lines

71

,

71

having a length a. Namely, flat faces

72

,

72

having the length a are formed in areas provided by the straight lines

71

,

71

. On the other hand, each of the first and second grooves

43

c

and

43

d

in the first and second rotary cylinders

41

c

and

41

d

having the first and second pistons

42

c

and

42

d

slidably retained therein is formed by arcs

73

,

73

having the substantially same shape as the arcs

70

,

70

, and two parallel straight lines

74

,

74

having a length 4 E+a. Namely, flat faces

75

,

75

having the length 4 E+a are formed in areas provided by the straight lines

74

,

74

.

As shown in

FIGS. 21 and 22

, the first and second pistons

42

c

and

42

d

having the above-described shape are slidably retained in the first and second grooves

43

c

and

43

d

with their flat faces

72

,

72

being in abutment against the flat faces

75

,

75

of the first and second grooves

43

c

and

43

d

in the first and second rotary cylinders

41

c

and

42

d

, respectively. The first and second pistons

42

c

and

42

d

are slid within the grooves

43

c

and

43

d

while being maintained in such retained states, respectively.

As shown in

FIG. 20

, the first and second compressing mechanisms

40

c

and

40

d

are partitioned from each other by the partition plate

44

. The first rotary cylinder

41

c

, the second rotary cylinder

41

d

and the partition plate

44

are connected together and moved in the same manner. However, the first and second rotary cylinders

41

c

and

41

d

are connected to each other with the first and second grooves

43

c

and

43

d

offset at 90 degree from each other, so that the phases in the compressing strokes are different by 180 degree from each other.

On the other hand, the first and second pistons

42

c

and

42

d

are fitted over first and second cranks

33

c

and

33

d

provided on the shaft

33

. The first and second cranks

33

c

and

33

d

are provided so that their eccentric directions are different by 180 degree from each other.

The first and second compressing mechanisms

40

c

and

40

d

are sandwiched from above and below by an upper bearing

50

c

and a lower bearing

50

d

and surrounded by a tubular casing

51

.

The upper bearing

50

c

is provided with an intake port

51

c

and a discharge port

52

c

for the first compressing mechanism

40

c

, and the lower bearing

50

d

is provided with an intake port

51

d

and a discharge port

52

d

for the second compressing mechanism

40

d

. Provided in the discharge ports

52

c

and

52

d

are valves

53

c

and

53

d

which are opened by a predetermined pressure, and valve stops

54

c

and

54

d

for limiting the opening movements of the valves

53

c

and

53

d

. The intake port

51

c

communicates with the intake pipe

12

c

, and the intake port

51

d

communicates with the intake pipe

12

d

. The intake pipes

12

c

and

12

d

are connected to an accumulator

60

.

The flow of a refrigerant in the hermetic compressor having the above-described arrangement will be described below in brief.

The gas refrigerant in the accumulator

60

is introduced into the shell

10

through the intake pipes

12

c

and

12

d

. The refrigerant passed through the intake ports

51

c

and

52

d

and compressed in the first and second compressing mechanisms

40

c

and

40

d

, when it reaches a predetermined pressure, pushes up the valves, and is then discharged through the discharge ports

52

c

and

52

d

into the shell

10

. At this time, the discharging timings are not the same as each other, because the first and second compressing mechanisms

40

c

and

40

d

are different in their phases by 180 degree from each other. The refrigerant discharged into the shell is passed through an area around the motor

30

and discharged to the outside of the shell

10

through the discharge pipe

11

mounted at the upper portion of the shell

10

.

The relationship between the shaft

33

, the first and second pistons

42

c

and

42

d

and the first and second rotary cylinders

41

c

and

41

d

in the first and second compressing mechanisms

40

c

and

40

d

will be described below with reference to

FIGS. 21 and 22

.

The shaft

33

adapted to transmit the rotation of the motor

30

is rotated about a point B. The center C of the cranks

33

c

and

33

d

provided on the shaft

33

is eccentric by a distance E from the center B of rotation of the shaft

33

. The center C of the cranks

33

c

and

33

d

is also the center of rotation of the pistons

42

c

and

42

d

. On the other hand, the rotary cylinders

41

c

and

41

d

have the center of rotation provided by a position spaced apart at the distance E from the center B of rotation of the shaft

33

. Therefore, when the center C of the cranks

33

c

or the piston

42

c

is spaced to the maximum apart from the center A of rotation of the rotary cylinder

41

c

, the largest and smallest spaces are formed in the groove

43

c

, as shown in FIG.

21

. The second compressing mechanism

40

d

has a phase difference of 180 degree from the phase of the first compressing mechanism

40

c

and hence, when the first compressing mechanism

40

c

is in a state shown in

FIG. 21

, the center C of rotation of the crank

33

d

or the piston

42

d

in the second compressing mechanism

40

d

overlaps the center A of rotation of the rotary cylinder

41

d

, as shown in FIG.

22

. Therefore, the space section in the groove

43

b

is divided into two equal spaces, as shown in FIG.

3

.

The refrigerant gas sucking and compressing strokes will be described below with reference to

FIG. 23

, but the second compressing mechanism

40

b

provides the same strokes, except that the phase in

FIG. 4

is different by 180 degree from that in the first compressing mechanism

40

a.

FIGS. 23

a

to

23

h

show states in which the shaft

33

has been rotated through every 90 degree, respectively.

When the shaft

33

is not rotated as shown in

FIG. 23

a

, the first groove

43

c

is in a state in which the one of the space D is largest, and the space F is smallest.

The volume of the space D is gradually decreased from the state shown in

FIG. 23

b

in which the shaft

33

has been rotated through 90 degree via the state shown in

FIG. 23

c

in which the shaft

33

has been rotated through 180 degree to the state shown in

FIG. 23

d

in which the shaft

33

has been rotated through 270 degree, whereby the compressed refrigerant is discharged from the discharge port

52

c

. In the space D, the compressing stroke is finished in the state shown in

FIG. 23

e

in which the shaft

33

has been rotated through 360 degree.

On the other hand, the volume of the space F is gradually increased from the state shown in

FIG. 23

b

in which the shaft

33

has been rotated through 90 degree via the state shown in

FIG. 23

c

in which the shaft

33

has been rotated through 180 degree to the state shown in

FIG. 23

d

in which the shaft

33

has been rotated through 270 degree, whereby the compressed refrigerant is sucked from the intake port

51

c

. In the space F, the sucking stroke is finished in the state shown in

FIG. 23

e

in which the shaft

33

has been rotated through 360 degree.

In the states shown in

FIG. 23

e

to

FIG. 23

h

, the sucking stroke is carried out in the space D, and the compressing stroke is carried out in the space F. When the shaft

33

is further rotated through 90 degree from the state shown in

FIG. 23

h

, the state shown in

FIG. 23

a

is obtained.

In this way, the compressing and sucking strokes are carried out in the two spaces D and F defined in the first groove

43

c

, respectively, whenever the shaft

33

is rotated through 720 degree.

According to this embodiment, even if the piston is located at the center of the cylinder, it can be avoided that the driving force from the piston does not serve as a rotating force for the rotary cylinder, because the other compressing mechanism provides a rotating force. In addition, the pistons can be disposed symmetrically by ensuring that the phase difference between the two compressing mechanisms is 180 degree, whereby the production of the hermetic compressor can be carried out easily. The freedom degree of setting of the positions of the intake port and the discharge port is increased by providing intake port and the discharge port in the upper and lower bearing, respectively. Therefore, it is possible to regulate the compression ratio and to prevent the over-compression by the positions of the intake port and the discharge port. Further, since the phases of the first and second compressing mechanisms are different from each other by 180 degree, and the intake port in the upper bearing and the intake port in the lower bearing are provided on the same axis, the position of mounting of the intake pipe can be the same side, and a piping cannot be drawn around for connection of the intake pipe to the accumulator or the like.

The phase difference between the two compressing mechanisms is 180 degree in the embodiment, but is not limited to this angle and may be 90 degree, 270 degree or any angle other than these angles.

The embodiment has been described as being provided with the two compressing mechanisms, but three or more compressing mechanisms may be provided.

FIGS. 24 and 25

show a second embodiment of a compressor mechanism section according to the present invention. In this embodiment, the structure of the compressor mechanism section is only different from that in the first embodiment, and the structures of the other components are the same as those in the first embodiment and hence, the duplicated description thereof is omitted.

FIG. 24

shows a first compressing mechanism

40

e

, and

FIG. 25

shows a second compressing mechanism

40

f

. The phases of the first and second compressing mechanisms

40

e

and

40

f

in the compressing stroke are different from each other by 180 degree.

The first compressing mechanism

40

e

in the present embodiment comprises a first rotary cylinder

41

e

and a first piston

42

e

, and the second compressing mechanism

40

f

comprises a second rotary cylinder

41

f

and a second piston

42

f

. First and second grooves

43

c

and

43

f

are defined in the first and second rotary cylinders

40

e

and

40

f

, respectively. The first and second compressing mechanisms

40

e

and

40

f

are of the same structure, and hence, only the structure of the first compressing mechanism will be described, and the duplicated description is omitted.

The first piston

42

e

is formed, so that the contour of its section is comprised of two arcs

70

′,

70

′ and two parallel straight lines

71

′,

71

′ having a length a. Namely, flat faces

72

′,

72

′ having the length a are formed in areas provided by the straight lines

71

′,

71

′.

On the other hand, the first groove

43

e

in the first rotary cylinder

41

e

is formed by arcs

73

′,

73

′ having the substantially same shape as the arcs

70

′,

70

′ of the first piston

42

e

, and two parallel straight lines

74

′,

74

′ having a length 4 E+a. Namely, flat faces

75

′,

75

′ having the length 4 E+a are formed in areas provided by the straight lines

74

′,

74

′.

The first piston

42

e

having the above-described structure is of a shape in which the semi-circular arcs

70

′,

70

′ are connected to each other by the two straight lines

71

′,

71

′, as described above and hence, any corner is not produced at such connection area. The first piston

42

e

is slid within the first groove

43

e

with its flat faces

72

′,

72

′ being in abutment against the flat faces

75

′,

75

′ of the first groove

43

e

. In this case, the smooth sliding movement is carried out, because any corner is not present in the first piston

40

e

, as described above. In addition, the flat faces

72

′,

72

′ and the flat faces

75

′,

75

′ are in close contact with each other and hence, the sealability can be enhanced, and the sucking and compressing efficiency can be enhanced, as in the first embodiment.

Number	Date	Country
10-049017	Feb 1998	JP
10-049019	Feb 1998	JP
10-140605	May 1998	JP

Number	Name	Date
3883273	King	May 1975
4207736	Loo	Jun 1980
4764097	Hirahara et al.	Aug 1988
5666015	Uchibori et al.	Sep 1997

Number	Date	Country
863751	Jan 1953	DE
430830	Jun 1935	GB

Hermetic compressor

Information

Patent Number

Date Filed

Date Issued

Inventors

Original Assignees

Examiners

Agents

CPC

US Classifications

Field of Search

US

International Classifications

Abstract

Description

Claims

Priority Claims (3)

US Referenced Citations (4)

Foreign Referenced Citations (2)