Positive displacement pump rotatable in opposite directions

BACKGROUND OF THE INVENTION

The invention generally relates to hermetic compressors and more particularly to positive displacement oil pumps for hermetic compressors.

Oil pumps of various types are typically employed in hermetic compressors to provide sufficient lubrication to a multitude of interfacing bearing surfaces within the compressor. These types of pumps may be, for example, impeller pumps, centrifugal pumps or positive displacement pumps, the present invention related to the lattermost type. Positive displacement pumps are considered by many in the field to be the preferred type of pump for compressor applications, in part for the reason that these pumps can generate higher oil pressure than other types of pumps.

Previous positive displacement pumps include designs which cannot effectively be interchanged between compressor applications which have crankshafts which rotate in opposite directions, or in compressor applications having a reversibly rotating crankshaft. Such a pump design is intended to pump lubrication to the various interfacing bearing surfaces of a compressor only when the compressor crankshaft is rotating in a single, given direction.

Many compressors driven by an electric motor are intended to rotate only in a single direction (referred to hereinbelow as “unidirectionally-rotating”), but may, due to miswiring of the electric motor during assembly, be caused to run in a reverse direction. Under such circumstances, some previous, unidirectionally-rotating positive displacement pumps will not operate to provide lubrication to the interfacing bearing surfaces, and the compressor may seize or experience excessive wear during the reverse rotation.

Further, many unidirectionally-rotating compressors are subject to unintended reverse rotation upon shutdown of the compressor, as discharge pressure gases within the compressor, or within the refrigerant system into which the compressor is incorporated, expand through the compression mechanism thereof. This phenomenon is well known, particularly in scroll compressors. As discharge gases expand on shutdown of the compressor, they backflow into the discharge port of the interleaved scroll wraps, and cause the orbiting scroll to orbit in the direction opposite that in which the gases were initially compressed. Thus, on shut down, the compressor may behave like an expansion motor, the compressed gases causing rotation of the crankshaft in a direction opposite that in which the electric motor drives the shaft. Objectionable noise and vibration usually accompany such reverse rotation of the orbiting scroll, and are well known problems. Much effort has been made to prevent of reverse rotation of the orbiting scroll; these efforts may, for example, include the provision of check valves over the discharge port to prevent reversely flowing discharge gases from reentering the space between the interleaved scroll wraps. Indeed, a scroll compressor embodiment described hereinbelow includes such a check valve. Where reverse rotation of a compressor having a previous, unidirectionally operable positive displacement pump is not entirely prevented, however, sufficient lubrication to the interfacing bearing surfaces of the compressor may not be achieved during the period of reverse rotation. During such reverse rotation, even for brief periods on shutdown of the compressor, the interfacing bearing surfaces, which remain in sliding contact with each other, may not be provided with adequate lubrication, and may be subject to excessive wear or seizure.

Moreover, in some unidirectionally-rotating compressors, during periods of brief power interruption during which the compressor is caused to be reversely rotated by expanding discharge gases, the compressor may continue rotation in the reverse direction, driven by the motor, if power is restored to the motor while the compressor is still reversely rotating under influence of the expanding discharge gases. In such situations, the compressor may run in the reverse direction for quite some time and, if no provision is made for pumping lubricant to its interfacing bearing surfaces during reverse rotation, the compressor will likely seize.

Positive displacement pumps are often at least partially submerged in the oil located in the oil sump provided in the lower portion of the compressor housing, and are driven by the rotating crankshaft coupled to the rotor of the electric motor, the end of the shaft disposed in, and rotatable relative to a pump body. Oil is forced by the pump through an axial passageway provided through the crankshaft, the passageway in fluid communication with points of lubrication in the compression mechanism. In previous pumps, a radially-extending passage communicating with the axial oil passageway in the crankshaft is provided to lubricate the interface between the shaft and the pump body. The pump body may, in some compressors, also serve as a bearing which rotatably and/or axially supports the shaft relative to the compressor housing. Here, too, a radially-extending passage communicating with the axial oil passageway in the crankshaft is provided to lubricate the interface between the shaft and the pump body. The tolerance between the peripheral surface of the crankshaft and the pump body must be held to rather close tolerances, and the provision of the radially-extending passage requires additional machining and cost.

A positive displacement pump which provides lubrication to the interfacing bearing surfaces of a compressor which rotates in two directions, whether by design (hereinafter referred to as “bidirectionally-rotating”) or a unidirectionally-rotating compressor caused to rotate in the reverse direction due to reexpansion of discharge gases, miswiring of the motor or a brief power interruption as described above, is highly desirable.

Further, a means of accommodating tolerances between the crankshaft and pump body of a compressor, and providing lubrication between the crankshaft and the pump body and/or a crankshaft-supporting bearing which comprises a pump body without requiring the additional machining associated with a radially-extending oil passage in the shaft, is also highly desirable.

SUMMARY OF THE INVENTION

Although the compressor described hereinbelow is a unidirectionally-rotating scroll compressor, it is to be understood that the positive displacement pump of the present invention has applications in other types of compressors or expansion motors, such as, for example, unidirectionally or bidirectionally-rotating rotary or reciprocating piston compressors, or bidirectionally-rotating scroll machines. To better facilitate understanding of the compressor embodiment described hereinbelow, however, U.S. Pat. No. 5,306,126 (Richardson), issued to the assignee of the present invention, is incorporated herein by reference and provides a detailed description of the operation of a typical scroll compressor.

The present invention, as it relates to the below-described embodiment, provides a positive displacement type oil pump which is provided at the lower end of a crankshaft and extends into an oil sump defined by a compressor housing. Two embodiments of the inventive oil pump are disclosed hereinbelow and in the figures. In the first embodiment, the positive displacement pump is supported by an outboard shaft bearing. In the second embodiment, the pump is supported by an anti-rotational spring that is attached to the compressor housing or some other support. The pump is comprised of an oil pump body, a shaft extension (second embodiment), a vane, a reversing port plate, a retention pin, a wave washer, a retainer plate and a snap ring. The outboard bearing of the first embodiment and the anti-rotational spring of the second embodiment respectively serve as the oil pump body. A slot is formed at the lower end of the crankshaft to receive the rotary vane which is caused to rotate by the rotation of the crankshaft during compressor operation.

With the pump submerged in the oil sump and with the crankshaft rotating during compressor operation, the pump collects oil via at least one passage and the rotary vane, much like a wiper or rotary piston, acts upon the collected oil in combination with the enclosed area formed by the oil pump body and reversing port plate to force the oil into and through an anchor-shaped oil passage provided in the reversing plate. The oil travels upward into an inner axial bore formed in the crankshaft and the crankshaft extension. The axial oil passage extends to the uppermost portion of the crankshaft to deliver lubricating oil thereto.

Various parts of the compressor mechanism, such as rotational or thrust bearings, associated with the scroll compressor are lubricated via lateral or radially-extending openings and passages or grooves formed in and/or along the crankshaft. The oil pump of the present invention may provide a certain amount of leakage to permit the communication of oil to lower bearing surfaces without detracting from the primary oil flow of the pump or the need for radially-extending passages in the lower end of the shaft. The rotary vane of the present invention may be a spring loaded rotary vane to provide a more positive contact between both ends of the vane member and the inner surface of the oil pump body so as to decrease leakage and improve the efficiency of the oil pump.

The present invention provides a compressor assembly including a compression mechanism, a rotating crankshaft operably coupled to the compression mechanism, the crankshaft provided with a longitudinally-extending oil conveyance passageway, the oil conveyance passageway in fluid communication with relatively moving interfacing bearing surfaces of the compression mechanism, and an oil pump assembly. The oil pump assembly includes an oil pump body having an interior surface and being rotatable relative to the crankshaft, a vane disposed within the pump body and rotating with the crankshaft, the vane having at least one end in sliding engagement with the interior surface of the oil pump body, and a port plate disposed within the pump body and having rotatably opposite first and second positions. The vane is in sliding engagement with an adjacent surface of the port plate, and the port plate is provided with an inlet and an outlet. The pump body receives oil from a source of oil, the oil received in the pump body directed by the vane into the port plate inlet, the port plate outlet in fluid communication with the oil conveyance passageway, the oil directed into the port plate inlet urged toward the port plate outlet in response to relative movement between the vane and the port plate, whereby oil is pumped from the source of oil through the oil conveyance passageway.

The present invention also provides a pump assembly including a rotating shaft provided with a longitudinally extending passageway, a pump body disposed about a shaft and having an interior surface, relative rotation existing between the shaft and the pump body, a vane disposed within the pump body, the vane rotating with the shaft, the vane having at least one end in sliding engagement with the interior surface of the pump body, and a port plate disposed within the pump body. The vane is in sliding engagement with an adjacent surface of the port plate. The port plate is provided with an inlet and an outlet, the port plate inlet receiving liquid directed thereinto by the vane from a source of liquid, the outlet in fluid communication with the shaft passageway. Liquid urged from the port plate inlet toward the port plate outlet in response to relative movement between the vane and the port plate, whereby liquid is pumped from the source of liquid through the passageway. The shaft has a surface surrounded by a surface of the pump body, providing an interface between the shaft surface and the surrounding pump body surface. The pump assembly has means for providing liquid leaked from the pump assembly along a surface of the shaft to this interface, whereby the interface is lubricated by the leaked liquid.

The present invention also provides a compressor assembly including a compression mechanism, a rotating crankshaft operably coupled to the compression mechanism, the crankshaft provided with a longitudinally-extending oil conveyance passageway, the oil conveyance passageway in fluid communication with relatively moving interfacing bearing surfaces of the compression mechanism, and an oil pump assembly. The oil pump assembly includes a pump body disposed about the crankshaft and having an interior surface, relative rotation existing between the crankshaft and the pump body, a vane disposed within the pump body, the vane rotating with the crankshaft, the vane having at least one end in sliding engagement with the interior surface of the pump body, and a port plate disposed within the pump body, the vane in sliding engagement with an adjacent surface of the port plate. The port plate is provided with an inlet and an outlet, the port plate inlet receiving oil directed thereinto by the vane from a source of oil, the outlet in fluid communication with the crankshaft oil conveyance passageway. Oil is urged from the port plate inlet toward the port plate outlet in response to relative movement between the vane and the port plate, whereby oil is pumped from the source of oil through the oil conveyance passageway. The crankshaft has a surface surrounded by a surface of the pump body, an interface is thus provided between the crankshaft surface and the surrounding pump body surface. The pump assembly has means for providing oil leaked from the pump assembly along a surface of the crankshaft to the interface, whereby the interface is lubricated by the leaked oil.

BRIEF DESCRIPTION OF THE DRAWINGS

The above-mentioned and other features and objects of this invention, and the manner of attaining them, will become more apparent and the invention itself will be better understood by reference to the following description of an embodiment of the invention taken in conjunction with the accompanying drawings, wherein:

FIG. 1

is a scroll sectional view of the scroll compressor of the present invention;

FIG. 2

is a top view looking inside the housing of the scroll compressor of

FIG. 1

;

FIG. 3

is an enlarged, fragmentary sectional view of a first embodiment of a sealing structure between the fixed scroll member and the frame member of the compressor of

FIG. 1

;

FIG. 4

is a bottom view of the fixed scroll member of the scroll compressor of

FIG. 1

;

FIG. 5

is a top view of the fixed scroll member of

FIG. 4

;

FIG. 6

is a fragmentary sectional view showing the mounting feature of the fixed scroll member of

FIG. 4

;

FIG. 7

is a fragmentary sectional view of the fixed scroll member of

FIG. 4

;

FIG. 8

is a sectional side view of the fixed scroll member taken along line

8

—

8

of

FIG. 5

;

FIG. 9

is an enlarged fragmentary bottom view of the innermost position of the involute scroll wrap of the fixed scroll member of

FIG. 4

;

FIG. 10

is a bottom view of the orbiting scroll member of the scroll compressor of

FIG. 1

;

FIG. 11

is a top view of the orbiting scroll member of

FIG. 10

;

FIG. 12

is a fragmentary sectional side view of the orbiting scroll member of

FIG. 10

showing the inner hub portion with an axial oil passage;

FIG. 13

is an enlarged fragmentary top view of the innermost portion of the scroll wrap of the orbiting scroll member of

FIG. 10

;

FIG. 14

is a sectional side view of the orbiting scroll member of

FIG. 10

taken along line

14

—

14

of

FIG. 11

;

FIG. 15

is an enlarged fragmentary sectional side view of the orbiting scroll member of

FIG. 10

showing an axial oil passage;

FIG. 16

is an enlarged fragmentary sectional side view of a first embodiment of a seal disposed intermediate the orbiting scroll member and the main bearing or frame of the scroll compressor of

FIG. 1

;

FIG. 17

is an enlarged fragmentary sectional side view of a second embodiment of a seal disposed intermediate the orbiting scroll member and the main bearing or frame of the scroll compressor of

FIG. 1

;

FIG. 18

is a top view of one embodiment of a one piece seal located intermediate the outer peripheries of the fixed scroll member and the main bearing or frame of a scroll compressor;

FIG. 19

is an enlarged, fragmentary sectional side view illustrating an alternative to the sealing structure embodiment depicted in

FIG. 3

;

FIG. 20

is a top perspective view of a first embodiment of the Oldham ring of the scroll compressor of

FIG. 1

;

FIG. 21

is a bottom perspective view of the Oldham ring of

FIG. 20

;

FIG. 22

is a top view of the Oldham ring of

FIG. 20

;

FIG. 23

is a first side view of the Oldham ring of

FIG. 20

;

FIG. 24

is a second side view of the Oldham ring of FIG.

20

:

FIG. 25

is a top view of a second embodiment of the Oldham ring of the scroll compressor of

FIG. 1

;

FIG. 26

is a sectional top view of the compressor assembly of

FIG. 1

along line

26

—

26

, its Oldham coupling and the fixed scroll member recess in which is disposed shown shaded;

FIG. 27

is a top view of a first embodiment of a discharge valve member for use in the discharge check valve assembly of the scroll compressor of

FIG. 1

;

FIG. 28

is a left side view of the discharge valve member of

FIG. 27

;

FIG. 29

is a front view of a first embodiment of a discharge valve retaining member for use in the discharge check valve assembly of the compressor of

FIG. 1

;

FIG. 30

is a top view of the discharge valve retaining member of

FIG. 29

;

FIG. 31

is a left side view of the discharge valve retaining member of

FIG. 29

;

FIG. 32

is an end view of a roll spring pin used in one embodiment of the discharge check valve assembly;

FIG. 33

is a front view of the roll spring pin of

FIG. 32

;

FIG. 34

is a side view of a bushing for use in said one embodiment of the discharge check valve assembly;

FIG. 35

is a top view of a second embodiment of a discharge valve member for use with the discharge check valve assembly,

FIG. 36

is a rear view of the discharge valve member of

FIG. 35

;

FIG. 37

is a right side view of the discharge valve member of

FIG. 35

;

FIG. 38

is a top view of a third embodiment of a discharge valve member for use in the discharge check valve assembly;

FIG. 39

is a rear view of the discharge valve member of

FIG. 38

;

FIG. 40

is a right side view of the discharge valve member of

FIG. 38

;

FIG. 41

is a sectional side view of the fixed scroll member of the compressor of

FIG. 1

with one embodiment of a discharge check valve assembly;

FIG. 42

is a sectional side view of the fixed scroll member of the compressor of

FIG. 1

with an alternative embodiment of the discharge check valve assembly;

FIG. 43

is a front view of a second embodiment of a discharge valve retaining member for use in the discharge check valve assembly of the compressor of

FIG. 1

;

FIG. 44

is a left side view of the discharge valve retaining member of

FIG. 43

;

FIG. 45

is a top view of the discharge valve retaining member of

FIG. 43

;

FIG. 46

is a side view of a first embodiment of a discharge gas flow diverting mechanism;

FIG. 47

is a top view of the discharge gas flow diverting mechanism of

FIG. 46

;

FIG. 48

is a front view of the discharge gas flow diverting mechanism of

FIG. 46

;

FIG. 49

is a side view of a second embodiment of a discharge gas flow diverting mechanism;

FIG. 50

is a top view of the discharge gas flow diverting mechanism of

FIG. 49

;

FIG. 51

is a front view of the discharge gas flow diverting mechanism of

FIG. 49

;

FIG. 52

is a side view of a third embodiment of a discharge gas flow diverting mechanism;

FIG. 53

is a top view of the discharge gas flow diverting mechanism of

FIG. 52

;

FIG. 54

is a front view of the discharge gas flow diverting mechanism of

FIG. 52

;

FIG. 55

is a side view of the crankshaft of the scroll compressor of

FIG. 1

;

FIG. 56

is a sectional side view of the crankshaft of

FIG. 55

along line

56

—

56

;

FIG. 57

is a bottom view of the crankshaft of

FIG. 55

;

FIG. 58

is a top view of the crankshaft of

FIG. 55

;

FIG. 59

is an enlarged fragmentary side view of the crankshaft of

FIG. 55

showing the toroidal shaped oil channel or gallery associated with the bearing lubrication system of the compressor of

FIG. 1

;

FIG. 60

is an enlarged fragmentary sectional side view of the upper portion of the crankshaft of

FIG. 55

;

FIG. 61A

is a bottom view of the eccentric roller of the scroll compressor of

FIG. 1

;

FIG. 61B

is a side view of the eccentric roller of

FIG. 61A

;

FIG. 61C

is a side view of the eccentric roller of

FIG. 61B

from line

61

C—

61

C;

FIG. 62

is a sectional side view of the eccentric roller of

FIG. 61A

along line

62

—

62

;

FIG. 63A

is a first enlarged, fragmentary sectional side view of the compressor assembly of

FIG. 1

;

FIG. 63B

is a second enlarged, fragmentary sectional side view of the compressor assembly of

FIG. 1

;

FIG. 64

is a fragmentary sectional end view of the compressor assembly of

FIG. 63A

along line

64

—

64

;

FIG. 65

is a first fragmentary sectional side view of the lower portion of the scroll compressor of

FIG. 1

showing a first embodiment of a positive displacement oil pump;

FIG. 66

is a second fragmentary sectional side view of the positive displacement oil pump of

FIG. 65

;

FIG. 67

is a bottom view of the scroll compressor of

FIG. 1

illustrated with the lower bearing and oil pump removed;

FIG. 68

is an exploded lower view of the lower bearing and positive displacement oil pump assembly of

FIG. 65

;

FIG. 69

is a sectional side view of the lower bearing and pump housing of the positive displacement oil pump assembly of

FIG. 65

;

FIG. 70

is an enlarged fragmentary sectional side view of the lower portion of the pump housing of

FIG. 69

;

FIG. 71

is an enlarged fragmentary sectional side view of the upper portion of the lower bearing of

FIG. 69

;

FIG. 72

is an enlarged fragmentary sectional side view of the oil pump housing of

FIG. 69

showing the oil pump inlet;

FIG. 73

is a bottom view of the lower bearing and oil pump housing of

FIG. 69

;

FIG. 74

is a top view of the pump vane or wiper of the oil pump of

FIG. 68

;

FIG. 75

is a side view of the pump vane of

FIG. 74

;

FIG. 76

is a top view of the reversing port plate of the oil pump of

FIG. 68

;

FIG. 77

is a right side view of the reversing port plate of

FIG. 76

;

FIG. 78

is a bottom view of the reversing port plate of

FIG. 76

;

FIG. 79

is a top perspective view of the reversing port plate of

FIG. 76

;

FIG. 80

is an exploded side view of a second embodiment of a positive displacement oil pump;

FIG. 81

is a sectional side view of the oil pump of

FIG. 80

, assembled;

FIG. 82

is a force diagram for a swing link radial compliance mechanism;

FIG. 83

is a graph showing the values of flank contact force versus orbiting radius variation due to fixed scroll to crankshaft center offset for tangential gas forces varying from 100 to 1000 lbf.;

FIG. 84

is a graph showing the values of flank sealing force versus crankshaft angle for several values of tangential gas force for a fixed scroll to crankshaft center offset of 0.010 inch;

FIG. 85

is a graph showing the values of tangential gas force variation versus crankshaft angle for a highly loaded compressor;

FIG. 86

is a graph showing the flank sealing force versus the crankshaft angle for a fixed scroll to crankshaft center offset of 0.020 inch and a tangential gas force variation as shown in

FIG. 85

;

FIG. 87

is a graph showing the calculated values of peak to peak crankshaft torque load variation versus crankshaft angle for various fixed scroll to crankshaft center offset values;

FIG. 88

is a graph showing the calculated values of peak to peak crankshaft torque load variation versus radial compliance angle for various fixed scroll to crankshaft center offset values;

FIG. 89

is a top view of the compressor shown in

FIG. 1

, along line

89

—

89

thereof, showing crankshaft center axis to fixed scroll centerline offset;

FIG. 90

is a top view of the compressor shown in

FIG. 1

, along line

90

—

90

thereof, showing the axial centerline of the fixed scroll member;

FIG. 91

is a bottom view of the compressor shown in

FIG. 1

, along line

91

—

91

thereof, showing the axial centerline of the fixed scroll member;

FIG. 92

is a greatly enlarged fragmentary bottom view of the compressor as shown in

FIG. 91

, showing the crankshaft center axis to fixed scroll centerline offset;

FIG. 93

is a side view of the lower bearing and pump housing of the positive displacement oil pump of

FIG. 65

;

FIG. 94

is partial sectional view of the lower bearing and pump housing of

FIG. 93

along line

94

—

94

, showing the orientation of the reversing port plate therein when the compressor shaft is rotated in a first direction;

FIG. 95

is partial sectional view of the lower bearing and pump housing of

FIG. 93

along line

95

—

95

, showing the orientation of the reversing port plate therein when the compressor shaft is rotated in a second direction; and

FIG. 96

is a sectional view of the lower bearing and pump housing of

FIG. 93

along line

96

—

96

, showing the components of the inventive positive displacement oil pump therein.

Corresponding reference characters indicate corresponding parts throughout the several views. The exemplifications set out herein illustrate a preferred embodiment of the invention, in one form thereof, and such exemplifications are not to be construed as limiting the scope of the invention in any manner.

DETAILED DESCRIPTION OF THE INVENTION

In an exemplary embodiment of the invention as shown in the drawings, scroll compressor

20

is shown in one vertical shaft embodiment. This embodiment is only provided as an example to which the invention is not limited.

Referring now to

FIG. 1

, scroll compressor

20

is shown having housing

22

consisting of upper portion

24

, central portion

26

and lower portion

28

. In an alternative form central portion

26

and lower portion

28

may be combined as a unitary lower housing member. Housing portions

24

,

26

, and

28

are hermetically sealed and secured together by such processes as welding or brazing. Lower housing member

28

also serves as a mounting flange for mounting compressor

20

in a vertical upright position. The present invention is also applicable in horizontal compressor arrangements. Within housing

22

is electric motor

32

, crankshaft

34

, which is supported by lower bearing

36

, and scroll mechanism

38

. Motor

32

includes stator

40

and rotor

42

which has aperture

44

into which is received crankshaft

34

. Oil collected in oil sump or reservoir

46

provides a source of oil and is drawn into positive displacement oil pump

48

at inlet

50

and is discharged from oil pump

48

into lower oil passageway

52

. Lubricating oil travels along passageways

52

and

54

, whereby it is delivered to bearings

57

,

59

and between the intermeshed scroll wraps as described further below.

Scroll compressor mechanism

38

generally comprises fixed scroll member

56

, orbiting scroll member

58

, and main bearing frame member

60

. Fixed scroll member

56

is fixably secured to main bearing frame member

60

by a plurality of mounting bolts or members

62

. Fixed scroll member

56

comprises generally flat end plate

64

, having substantially planar face surface

66

, sidewall

67

and an involute fixed wrap element

68

which extends axially downward from surface

66

. Orbiting scroll member

58

comprises generally flat end plate

70

, having substantially planar back surface

72

and substantially planar top face surface

74

, and involute orbiting wrap element

76

, which extends axially upward from top surface

74

. With compressor

20

in a de-energized mode, back surface

72

of orbiting scroll plate

70

engages main bearing member

60

at thrust bearing surface

78

.

Scroll mechanism

38

is assembled with fixed scroll member

56

and orbiting scroll member

58

intermeshed so that fixed wrap

68

and orbiting wrap

76

operatively interfit with each other. To insure proper compressor operation, face surfaces

66

and

74

and wraps

68

and

76

are manufactured so that when fixed scroll member

56

and orbiting scroll member

58

are forced axially toward one another, the tips of wraps

68

and

76

sealingly engage with respective opposite face surfaces

74

and

66

. During compressor operation, back surface

72

of orbiting scroll member

58

becomes axially spaced from thrust surface

78

in accordance with strict machining tolerances and the amount of permitted axial movement of orbiting scroll member

58

toward fixed scroll member

56

. Situated on the top of crankshaft

34

about offset crankpin

61

is cylindrical roller

82

, which comprises swinglink mechanism

80

. Referring to

FIG. 61A

, roller

82

is provided with offset axial bore

84

which receives crankpin

61

and offset axial bore

618

which receives limiting pin

83

, which is interference-fitted into and extends from hole

620

provided in the upper axial surface of crankshaft journal portion

606

(FIG.

56

). Roller

82

is allowed to pivot slightly about crankpin

61

, its motion relative thereto limited by limiting pin

83

, which fits loosely in roller bore

618

(FIG.

61

C). When crankshaft

34

is caused to rotate by motor

32

, cylindrical roller

82

and Oldham ring

93

cause orbiting scroll member

58

to orbit with respect to fixed scroll member

56

. In this manner swinglink mechanism

80

functions as a radial compliance mechanism to promote sealing engagement between the flanks of fixed wrap

68

and orbiting wrap

76

.

With compressor

20

in operation, refrigerant fluid at suction pressure is introduced through suction tube

86

(FIG.

2

), which is sealingly received into counterbore

88

(

FIG. 4

,

8

) in fixed scroll member

56

. The sealing of suction tube

86

with counterbore

88

is aided by the use of O-ring

90

(FIG.

8

). Suction port

88

provided in fixed scroll member

56

receives suction tube

86

and annular O-ring

90

in a groove for proper sealing of suction tube

86

with fixed scroll

56

. Suction tube

86

is secured to compressor

20

by suction tube adapter

92

which is brazed or soldered to suction tube

86

and opening

94

of housing

22

(FIG.

2

). Suction tube

86

includes suction pressure refrigerant passage

96

through which refrigerant fluid is communicated from a refrigeration system (not shown), or other such system, to suction pressure chamber

98

which is defined by fixed scroll member

56

and frame member

60

.

Suction pressure refrigerant travels along suction passage

96

and enters suction chamber

98

for compression by scroll mechanism

38

. As orbiting scroll member

58

is caused to orbit with respect to fixed scroll member

56

, refrigerant fluid within suction chamber

98

is captured and compressed within closed pockets defined by fixed wrap

68

and orbiting wrap

76

. As orbiting scroll member

58

continues to orbit, pockets of refrigerant are progressed radially inwardly towards discharge port

100

. As the refrigerant pockets are progressed along scroll wraps

68

and

76

towards discharge port

100

their volumes are progressively decreased, thereby causing an increase in refrigerant pressure. This increase in pressure internal the scroll set results in an axial force which acts outwardly to separate the scroll members. If this axial separating force becomes excessive, it may cause the tips of the scroll wraps to become spatially removed from the adjacent scroll plates, resulting in leakage of compressed refrigerant from the pockets and loss of efficiency. At least one axial biasing force, discussed hereinbelow, is applied against the back of the orbiting scroll member to overcome the axial separating force within the scroll set to maintain the pockets of compression. However, should the axial biasing force become excessive, further inefficiencies will result. Accordingly, all forces which act upon the scroll set must be considered and taken into account when designing an effective compressor design which effects a sufficient, yet not excessive, axial biasing force.

Upon completion of the compression cycle within the scroll set, refrigerant fluid at discharge pressure is discharged upwardly through discharge port

100

, which extends through face plate

64

of fixed scroll

56

, and discharge check valve assembly

102

. To more readily exhaust the high pressure refrigerant from between the scroll wraps, surface

66

of fixed scroll member

56

may be provided with kidney shaped recess

101

as shown in

FIG. 9

, within which discharge port

100

is located. Alternatively, and for the same purpose, surface

74

of orbiting scroll member

58

′ may be provided with kidney shaped recess

101

′ as shown in FIG.

11

. The refrigerant is expelled from between the scroll wraps through discharge port

100

into discharge plenum chamber

104

, which is defined by the interior surface of discharge gas flow diverting mechanism

106

and top surface

108

of fixed scroll member

56

. The compressed refrigerant is introduced into housing chamber

110

where it exits through discharge tube

112

(

FIG. 2

) into the refrigeration or air-conditioning system into which compressor

20

is incorporated.

To illustrate the relationship between the various fluids at varying pressures which occur inside compressor

20

during normal operation, we shall examine the example of the compressor in a typical refrigeration system. When refrigerant flows through a conventional refrigeration system during the normal refrigeration cycle, the fluid drawn into the compressor at suction pressure undergoes changes as the load associated with the system varies. As the load increases, the suction pressure of the entering fluid increases, and as the load decreases, the suction pressure decreases. Because the fluid which enters the scroll set, and eventually the pockets of compression formed therein, is at suction pressure, as the suction pressure varies, so varies the pressure of the fluid within the pockets of compression. Accordingly, the intermediate pressure of the refrigerant within the pockets of compression correspondingly increases and decreases with the suction pressure. The change in suction pressure results in a corresponding change in the axial separating forces within the scroll set. As the suction pressure decreases the axial separating force within the scroll set decreases and the requisite level of axial biasing force needed to maintain scroll set integrity decreases. Clearly this is a dynamic situation in which the operating envelope of the compressor may vary with the suction pressure. Because the axial compliance force is derived from the pockets of compression and therefore tracks the fluctuations in the suction pressure, an effective operating envelope for compressor

20

is maintained. The actual magnitude of the axial compliance force is in part determined by the location of aperture

85

(

FIG. 12

) and the volume of chamber

81

.

Annular chamber

81

is defined by back surface

72

of orbiting scroll

58

and the upper surface of bearing

60

. Annular chamber

81

forms an intermediate pressure cavity that is in communication, via aperture

85

, with fluid contained in pockets of compression formed in the scroll set. The fluid in the pockets of compression is at a pressure intermediate discharge and suction pressures. Although, oil and/or the natural sealing properties of contact surfaces may provide sufficient sealing, in the embodiment shown, continuous seals

114

and

116

, which may each be annular as shown, isolate intermediate pressure cavity

81

from radially adjacent volumes, which are respectively at suction and discharge pressure. Seal

114

is substantially longer in circumference than seal

116

.

As shown in

FIG. 12

, aperture, passage or conduit

85

is provided in plate portion

70

of orbiting scroll member

58

and provides fluid communication between the pockets of compression and intermediate pressure cavity

81

. Although this particular arrangement is described herein, it is by way of example only and not limitation. O-ring seal

118

is provided between the fixed scroll member

56

and flame

60

which separates the discharge and suction sides of the compressor. Referring to

FIG. 3

, it is shown that fixed scroll member

56

and frame

60

are provided with abutting axial surfaces

120

,

122

, respectively. Outboard of the abutting engagement of surfaces

120

,

122

, radial surfaces

124

,

126

of fixed scroll

56

and flame

60

, respectively, are in sliding engagement. Frame

60

is provided with an axial annular surface

128

and fixed scroll

56

is provided with a stepped axial surface

130

which faces surface

128

of the frame. Frame

60

is also provided with an outer annular lip

132

which extends upwardly from surface

128

but does not extend so far as to abut surface

130

of the fixed scroll. Surfaces

126

,

128

,

130

and the inner surface of lip

132

define a four-sided chamber in which a conventional O-ring seal

118

is disposed. O-ring

118

is made of conventional sealing material such as, for example, EPDM rubber or the like. O-ring

118

is contacted by surfaces

128

and

130

and is squeezed therebetween, i.e., the seal provided by the above-described configuration of fixed scroll and frame surfaces and seal

118

is an axial seal. In the assembly of the fixed scroll

56

to the frame, O-ring

118

is disposed on surface

128

of the frame, held in place by lip

132

, and the fixed scroll is assembled thereto. As surfaces

120

,

122

are abutted, seal

118

is squeezed into its sealing configuration between surfaces

128

and

130

and, hence, the suction and discharge portions of the compressor are sealably separated.

FIG. 18

shows an alternative sealing structure comprising O-ring seal

118

′, which is provided with a plurality of eyelets

134

on its inside diameter and, as shown in

FIG. 19

, seals fixed scroll

56

′ and frame

60

′ together. The eyelets encircle bolts

62

(FIG.

1

), which fasten fixed scroll

56

′ to frame

60

′. In this alternative embodiment, fixed scroll

56

′ is provided with axial surface

120

′ which abuts axial surface

122

′ of frame

60

′. Radial surface

124

′ of frame

60

′ slidingly engages radial surface

126

′ of fixed scroll

56

′. Fixed scroll

56

′ is provided with an annular step which defines axial surface

130

′, and frame

60

′ is provided with an annular step having frustoconical surface

128

′. As fixed scroll

56

′ is assembled to frame

60

′, with eyelets

134

disposed appropriately about the bolt holes in through which bolts

62

extend, O-ring

118

′ is brought into sealing contact with exterior radial surface

136

and annular axial surface

130

′ of frame

56

′, and with frustoconical surface

128

′ of frame

60

′. Hence, it is shown that in the alternative sealing arrangement, the O-ring seal is in both axial and radial sealing engagement with the fixed scroll and frame.

FIGS. 20 through 24

show one embodiment of an Oldham coupling used in compressor

20

. Oldham ring

93

is disposed between fixed scroll

56

and orbiting scroll

58

and comprises two pairs of somewhat elongate tabs,

204

,

206

and

208

,

210

, which respectively extend from opposite axial sides

224

and

226

of the Oldham coupling. Each of tabs

204

,

206

,

208

and

210

have a rectangular cross section and the tabs of each pair are aligned in a common direction. As seen in

FIG. 22

, tabs

204

,

206

of one pair are aligned in a direction that is generally perpendicular to the direction in which tabs

208

,

210

of the other pair are aligned. Referring to

FIG. 26

, Oldham coupling

93

is disposed in recessed portion

202

of fixed scroll

56

. In

FIG. 26

, recessed portion

202

and Oldham coupling

93

are both shown shaded by perpendicularly oriented lines; overlapping portions of recessed portion

202

and Oldham coupling

93

are thus shaded by a checked pattern formed by their respective, superimposed shading lines.

FIGS. 41

,

42

and

91

also show recess

202

of fixed scroll

56

. As also shown in

FIG. 26

, fixed scroll

56

is provided with, on approximately opposite radial sides, elongated recesses or slots

212

and

214

in which Oldham coupling tabs

204

and

206

are slidably disposed. Also as shown in

FIG. 26

, elongate slots

212

and

214

extend in a direction parallel to plane

220

, along which suction tube counterbore

88

is directed. Plane

220

is generally perpendicular to plane

222

, which is the plane in which orbiting scroll

58

tips at its largest tipping moment. As seen in

FIG. 26

, orbiting scroll

58

is provided with a pair of elongated recesses or slots

216

,

218

in which tabs

208

and

210

are slidably received. It can be readily understood that orbiting scroll

58

is keyed to fixed scroll

56

by Oldham coupling

93

such that it does not rotate relative thereto. Rather, orbiting scroll

58

eccentrically orbits relative to fixed scroll

56

, its orbiting motion guided by tabs

204

,

206

,

208

and

210

which slide within recesses

212

,

214

,

216

, and

218

. It will be noted in

FIG. 26

that as tabs

204

and

206

respectively assume a position at one end of their respective slots

212

and

214

(the shown position), the outer circumferential surface of Oldham coupling

93

on the side of plane

222

on which suction port

88

is located (lower right-hand side of FIG.

26

), conforms very closely to the adjacent, radially interior wall

203

of recess

202

. Similarly, as tabs

204

and

206

respectively assume a position at the opposite end of their respective slots

212

and

214

(position not shown), the outer circumferential surface of Oldham coupling

93

on the side of plane

222

opposite that on which suction port

88

is located (upper left-hand side of FIG.

26

), conforms very closely to the adjacent, radially interior wall

203

of recess

202

. Thus, it will be understood by those skilled in the art that recess

202

is closely sized to accommodate the reciprocating movement of Oldham coupling

93

along axis

240

, which lies in plane

220

. The space necessary to accommodate Oldham coupling

93

is thereby minimized.

Referring again to

FIGS. 20 through 24

, it can be seen that each of opposite axial sides

224

and

226

of Oldham ring

93

is provided with pad surfaces

228

through

236

. Pad surfaces

228

a

,

232

a

,

234

a

and

236

a

are disposed on side

224

; on opposite side

226

of Oldham ring

93

, directly below and matching the shapes of the pad surfaces on side

224

, are corresponding surfaces

228

b

,

230

b

,

232

b

,

234

b

and

236

b

. In each of

FIGS. 20 through 25

, the pad surfaces are shown shaded or cross hatched to clarify their general shape and position.

FIG. 25

shows alternative Oldham ring

93

′ which is substantially identical to Oldham ring

93

except that it is prepared by a sintered powder metal process rather than a metal machining process. It can be seen the primary distinction of Oldham ring

93

′ is that the material area surrounding each of the tabs is slightly enlarged.

As shown in

FIG. 1

, it can be seen that Oldham ring

93

,

93

′ is disposed between fixed scroll member

56

and orbiting scroll member

58

. Also, surface

74

of orbiting scroll member

58

has an outlying, peripheral surface portion

205

, which lies outside of its scroll wrap

76

, and which faces lower side

226

of Oldham ring

93

,

93

′. Similarly, recessed area

202

of fixed scroll

56

has downwardly facing surface

238

(

FIG. 91

) which faces upper side

224

of Oldham ring

93

,

93

′. Pads 228 through

236

on opposite sides of Oldham ring

93

,

93

′ slidingly contact surfaces

205

and

238

. Referring to

FIGS. 22 and 25

, pad surfaces

228

a

and

228

b

have portions which lie on opposite sides of plane

220

.

FIGS. 22

,

24

and

25

show axis

240

which extends centrally through the thickness of Oldham coupling

93

,

93

′, and which lies in plane

220

. During compressor operation, orbiting scroll member

58

tends to tip in plane

222

, about an axis in plane

220

which is parallel with axis

240

. As orbiting scroll

58

tips in plane

222

, outlying portion

205

of surface

74

will be alternatingly urged into contact with pad surface portions on side

226

of Oldham ring

93

,

93

′ on only opposite sides of plane

220

. Referring to

FIGS. 1

,

22

,

24

and

25

, as orbiting scroll member

58

tips in plane

222

in a clockwise direction as viewed in

FIG. 24

about an axis generally parallel to axis

240

and proximal plane

220

, a portion of surface portion

205

is swung upward and into abutting contact with Oldham ring

93

,

93

′ abutting pads

234

b

and

236

b

and a portion of

228

b

. This action urges opposite side pad surfaces

234

a

and

236

a

and a portion of

228

a

(all on the left hand side of plane

220

in

FIGS. 22

,

25

) into abutting contact with the adjacent portion axial surface

238

in fixed scroll recessed area

202

. Conversely, as orbiting scroll member

58

tips in plane

222

, in a counterclockwise direction as viewed in

FIG. 24

about an axis generally parallel to axis

240

and proximal plane

220

, the radially opposite portion of surface portion

205

is swung upward and into abutting contact with the Oldham coupling, abutting pads

230

b

,

232

b

and a portion of

228

b

. This action urges opposite side pad surfaces

230

a

and

232

a

and a portion of

228

a

(all on the right hand side of plane

220

in

FIGS. 22

,

25

) into abutting contact with the adjacent portion axial surface

238

in fixed scroll recess

202

. The tipping of orbiting scroll

58

in plane

222

oscillates between the above-described clockwise and counterclockwise motions during compressor operation. Thus it can be seen that the travel of Oldham coupling

93

,

93

′ is aligned to support surface

205

of the orbiting scroll member and prevent its tipping. As will be understood with reference to FIG.

26

, surface

205

of the orbiting scroll member is supported by the Oldham coupling at locations which oppose the maximum values of the oscillating tipping moments on the orbiting scroll, thereby preventing wobbling of the orbiting scroll member.

Upon compressor shutdown, orbiting scroll member

58

is no longer orbitally driven by motor

32

and crankshaft

34

and is free to move in response to gas pressures acting thereon, including the pressure differential between discharge port

100

and suction port

88

. Further, upon compressor shut-down, a pressure differential which exists between the fluid contained in the discharge chamber and the fluid contained in the scroll set, which is at a pressure lower than that contained in the discharge chamber. As the two volumes seek pressure equilibrium, a reverse flow of fluid refrigerant from the discharge chamber back into the scroll set. Unimpeded, this pressure differential acts upon orbiting scroll member

58

so as to cause it to orbit in a reverse manner with respect to fixed scroll member

56

. Such reverse orbiting results in refrigerant flowing into discharge port

100

in a reverse direction and exiting through suction port

88

into the refrigerant system. This problem of reverse scroll rotation during compressor shutdown has long been associated with scroll compressors. Valve assembly

102

is provided to alleviate this problem by using the fluid flowing from the discharge chamber into the scroll set to act on the discharge check valve so as to quickly move the check valve to a closed position covering the discharge port. In this manner, reverse orbiting is prevented and more gradual equilibrium may be achieved.

Shown in FIGS.

1

and

27

-

45

are various components and embodiments of discharge check valve assemblies

102

,

102

′ which may be used with compressor

20

. Each of these embodiments comprises a lightweight plastic or metallic pivoting valve that is positioned adjacent to and directly over discharge port

100

provided in fixed scroll member

56

and is held in place by valve retaining member

310

or

324

. Alternative valve members

302

,

302

′ and

302

″ are shown in

FIGS. 27

,

28

;

35

-

37

;

38

-

40

, respectively. The valve member may be provided with either of pivot ears

309

or a bore

322

for receiving a roll spring pin

320

, on which are provided bushings

318

. Ears

309

or bushings

318

are received in bushing recesses

318

,

318

′ in the valve retaining member.

With the compressor in operation, refrigerant fluid at suction pressure is introduced through suction tube

86

, which is sealingly received into counterbore

88

provided in fixed scroll member

56

and is communicated into suction pressure chamber

98

which is defined by fixed scroll member

56

and frame member

60

. The suction pressure refrigerant is compressed by scroll mechanism

38

. As orbiting scroll member

58

is caused to orbit with respect to fixed scroll member

56

, refrigerant fluid within suction chamber

98

is compressed between fixed wrap

68

and orbiting wrap

76

and conveyed radially inwards towards discharge port

100

in pockets of progressively decreasing volume, thereby causing an increase in refrigerant pressure.

Refrigerant fluid at discharge pressure is discharged upwardly through discharge port

100

and exerts an opening force against rear face

306

of valve member

302

,

302

′,

302

″, causing it to move to or remain in an open position. The refrigerant is expelled into discharge plenum or chamber

104

as defined by discharge gas flow diverting mechanism

106

and top surface

108

of fixed scroll member

56

. From the discharge gas flow diverting mechanism the compressed refrigerant is introduced into housing chamber

110

where it exits through discharge tube

112

into a refrigeration system in which compressor

20

is incorporated.

Discharge check valve assembly

102

,

102

′ prevents the reverse flow of refrigerant upon compressor shutdown, thereby preventing the reverse orbiting of scroll mechanism

38

. Referring to

FIGS. 42-45

, check valve assembly

102

comprises rectangular valve member

302

having front face

304

, rear face

306

, and pivot portion

308

, valve member retaining member

324

, bushings

318

, and spring pin

320

. Rear face

306

faces and preferably has an area greater than discharge port

100

. Pin

320

extends through hole

322

in pivot portion

308

and is fitted with bushings

318

on opposite sides of valve member

302

, with the radial flanges of bushings

318

adjacent the valve member. Bushings

318

are rotatably disposed in two opposite-side bushing recesses

316

of member

324

. During compressor operation, refrigerant acts upon front and rear faces

304

and

306

, thereby causing valve member

302

to pivot relative to member

324

, which is fixed relative to fixed scroll member

56

. Valve retaining member

324

mounts over and around the valve member and includes two mounting extensions

312

, which may be secured to the fixed scroll member such as by bolts. In assembly, spring pin

320

is received in bore

322

of valve member

302

and bushings

318

are attached at the ends of the pin. Valve retaining member is positioned over the valve member with the two bushings being received in the two recesses and the two mounting extensions positioned adjacent mounting bores provided in the upper surface of fixed scroll member

56

. The valve assembly is then secured to the fixed scroll by two mounting bolts or the like. Valve members

302

′ (

FIGS. 35-37

) and

302

″ (

FIGS. 38-40

) have integral bushings or ears

309

and no spring pin; each may be used with retaining member

310

or

324

as described above.

Valve

320

is urged against valve stop

314

,

314

′ by the force of discharge refrigerant acting on rear face

306

. Notably, valve

320

is not bistable, and would tend to return, under the influence of gravity, to its closed position if the discharge refrigerant force acting on rear face

306

were removed. During compressor shutdown, refrigerant in the discharge pressure housing chamber

110

of the compressor moves towards the suction pressure chamber

98

through discharge port

100

. With relief hole

326

provided in valve stop

314

, refrigerant travels through stop

314

and acts against the large surface area of front face

304

of valve member

302

, causing it to quickly pivot towards the discharge port and engage the surrounding surface

108

of fixed scroll member

56

such that front face

304

covers and substantially seals the opening of discharge port

100

. Relief hole

326

also prevents “stiction”, which tends to cause the valve member to stick to the stop, which may occur during compressor operation. In this manner refrigerant is prevented from flowing in a reverse direction from discharge pressure housing chamber

110

to suction chamber

98

and through suction passage

96

. A discharge check valve employing valve retainer member

310

functions in a similar manner, which stop

314

′ providing a large area of valve front face

304

exposed to reversely-flowing discharge gases on compressor shut-down. The fuller interface of face

304

with stop

314

vis-a-vis stop

314

′ is expected to provide better valve wear.

With housing chamber

110

effectively sealed off from suction chamber

98

the pressure differential is effectively eliminated thereby preventing reverse orbiting of orbit scroll member

58

. The pressurized refrigerant contained within scroll compression chambers between the interleaved scroll wraps acts upon scroll mechanism

38

to cause the wraps of orbiting scroll member

58

to radially separate from the wraps of fixed scroll member

56

. With scroll members

56

and

58

no longer sealed with one another, the refrigerant contained therein is permitted to leak through scroll member wraps

68

and

76

and the pressure within scroll mechanism

38

reaches equilibrium.

During normal scroll compressor operation, discharge pressure refrigerant is discharged through the discharge port causing the discharge check valve to move to an open position. A biasing spring (not shown) may be provided to prevent cycling of the discharge check valve and resulting chatter due to pressure pulsations which occur during compressor operation.

As shown in

FIG. 1

, discharge gas flow diverting mechanism

106

is attached to fixed scroll member

56

and surrounds annular protuberance

402

of the fixed scroll member.

FIGS. 46

,

47

, and

48

illustrate a first embodiment of the discharge gas flow diverting mechanism.

FIGS. 49

,

50

, and

51

illustrate a second embodiment of the gas flow diverting mechanism.

FIGS. 52

,

53

, and

54

illustrate a third embodiment of the gas flow diverting mechanism. The gas flow diverting mechanism may be attached to the fixed scroll member as by crimping the whole or portions of lower circumference

404

into an annular recess provided in annular protuberance

402

. In the alternative, a series of notches may be formed in the annular protuberance to permit a series of crimps along the lower circumference of the gas flow diverting mechanism. Other means, such as interference fit, locking protuberances, etc., may be employed to secure the gas flow diverting mechanism to the fixed scroll member. Also, as shown in third embodiment gas flow diverting mechanism

106

″ (FIG.

53

), the gas diverting mechanisms may be provided with a plurality of holes

414

which are aligned above a plurality of tapped holes

416

provided in fixed scroll member surface

108

(FIG.

5

), the gas diverting mechanism attached to the fixed scroll member with threaded fasteners (not shown).

During compressor operation, compressed refrigerant fluid is forced from discharge port

100

through discharge check valve

102

and into discharge chamber

104

, which is defined by the inner surface of the gas flow diverting mechanism and upper surface

108

of the fixed scroll member. Gas flow diverting mechanism

106

may be positioned so that discharge gas exiting chamber

104

through outlet

406

is directed downward through gap

408

(

FIGS. 1

,

2

) formed between housing

22

, fixed scroll member

56

and frame

60

, and is further directed into housing chamber

110

along path

411

to optimally flow over and about the motor overload protector

41

which is attached to stator windings

410

. Hence, the gas diverting mechanism provides an additional measure of motor protection by ensuring that hot discharge gases are immediately directed towards the overload protector.

As shown in the embodiment of

FIGS. 49 through 51

, gas flow diverting mechanism outlet

406

′ may be provided with a downwardly turned hood

412

to further direct the outwardly flowing discharge gas downward toward gap

408

.

Notably, discharge check valve assembly

102

is oriented toward gas diverting mechanism outlet such that, when the valve is open, front face

304

is exposed to the reverse inrush of discharge pressure gas from chamber

110

to chamber

104

through outlet

406

upon compressor shutdown, thereby facilitating quick closing of the valve.

The scroll compressor of

FIG. 1

is provided with an intermediate pressure chamber

81

into which is introduced refrigerant gas at an intermediate pressure which urges orbiting scroll member

58

into axial compliance with fixed scroll member

56

. Intermediate pressure chamber

81

is defined by surfaces of the orbiting scroll member

58

and the main bearing or frame

60

which lie between a pair of annular seals

114

,

116

respectively disposed in grooves

502

,

504

provided in downwardly-facing axial surfaces

72

,

506

of orbiting scroll member

58

and which are in sliding contact with interfacing surfaces of frame

60

. Referring to

FIGS. 1

,

10

and

14

, it can be seen that intermediate pressure chamber

81

is generally defined as the annular volume between a step provided in the frame

60

and the downwardly depending hub portion

516

of the orbiting scroll

58

. Seals

114

and

116

respectively seal the intermediate pressure from the suction pressure region and the discharge oil pressure region.

Referring to

FIG. 12

, it can be seen that downwardly depending hub portion

516

of the orbiting scroll member

58

has outer radial surface

508

which adjoins planar surface

72

. Surface

508

extends from surface

72

to bottommost axial surface

506

of the hub portion

516

. Radial surface

508

is provided with wide annular groove

510

having upper annular surface

512

. Aperture

85

extends from surface

512

to surface

74

, at which it opens into an intermediate pressure region between the scroll wraps of the orbiting and fixed scroll members. As seen in

FIG. 12

, aperture

85

may be a single straight passageway which extends at an angle from surface

512

to surface

74

. Alternatively, aperture

85

may comprise a first axial bore (not shown) extending from surface

74

in parallel with surface

508

into a portion of hub

516

radially inboard of groove

510

, and a radial crossbore (not shown) extending from the first bore to the radial surface of groove

510

. For ease of manufacturing, it is preferable to provide a single, angled aperture as shown in FIG.

12

.

Referring now to

FIG. 17

, it can be seen that seal

116

is provided in groove

504

and is in sliding contact with surface

514

of frame

60

which interfaces surface

506

of hub portion

516

. The portion of surface

506

radially inboard of groove

504

, i.e., to the right as shown in

FIG. 17

, is at discharge pressure and is ordinarily filled with oil. As seen in

FIG. 17

, seal

116

is generally C-shaped having outer portion

518

and inner portion

520

disposed within the annular channel provided in outer portion

518

, the channel facing radially inboard. Outer seal portion

518

may be a polytetrafluoroethylene (PTFE) material, or other suitable low-friction material, which provides low friction sliding contact with surface

514

. The interior of inner seal portion

520

is exposed to discharge pressure oil, which causes seal

116

to expand axially and radially outward in groove

504

, thereby ensuring sealing contact between the sealing surfaces of seal

116

and the uppermost and outermost surfaces of groove

504

and surface

514

of the frame.

Referring now to

FIGS. 14 and 16

, it can be seen that planar surface

72

of orbiting scroll member

58

is provided with annular groove

502

in which is disposed seal

114

. Seal

114

includes outer portion

522

having a c-shaped channel which is open radially inwardly, and an inner portion

524

disposed within the c-channel. The C-channel of portion

522

opens radially inwardly so as to be exposed to intermediate pressure fluid within intermediate pressure chamber

81

, which urges seal

114

radially outward in groove

502

and axially outward against the opposing axial surfaces of groove

502

and surface

78

of frame

60

on which seal

114

slidingly engages. Outer seal portion

522

may be made of PTFE material, or other suitable low-friction material, thereby allowing low friction sliding engagement with surface

78

. Inner seal portion

114

may be Parker Part No. FS16029, having a tubular cross section. Grooves

504

and

502

may be provided with seals

114

and

116

of a common cross-sectional design, which may be as illustrated in either

FIG. 16

or FIG.

17

. That is, the cross-sectional design of seal

114

may be adapted for use in groove

504

. Conversely, cross-sectional design of seal

116

may be adapted for use in groove

502

. The pressure within intermediate pressure chamber

81

may be regulated by means of a valve as disclosed in U.S. Pat. No. 6,086,342 (Utter), issued Jul. 11, 2000, which is expressly incorporated herein by reference.

Referring to

FIG. 1

, main bearing or frame

60

is provided with downwardly depending main bearing portion

602

which is provided with bearing

59

in which journal

606

of crankshaft

34

is radially supported. Crankshaft journal portion

606

is provided with radial crossbore

608

(

FIGS. 55

,

56

) which extends from the outer surface of crankshaft journal portion

606

to upper oil passageway

54

within the crankshaft. A portion of the oil conveyed through passageway

54

is provided through crossbore

608

to lubricate bearing

59

. Oil flowing from crossbore

608

through bearing

59

may flow downward along the outside of crankshaft journal portion

606

where it may be radially distributed by a rotating counterweight

614

, after which it is returned to sump

46

. From crossbore

608

, oil may also flow upwards along bearing

59

and along the outside of journal portion

606

and into annular oil gallery

610

, which is in communication with housing chamber

110

and sump

46

through passageway

612

in frame

60

. Passageway

612

is oriented in frame

60

such that the rotating counterweight

614

will pick up and sling the oil coming through passageway

612

to disperse the oil in the radial side of the compressor opposite the inlet of discharge tube

112

. The terminal end opening

732

of oil passageway

54

is sealed with plug

616

which is flush with or somewhat below the terminal end surface of crankpin

61

.

Radial oil passage

622

in roller

82

and radial oil passage

624

in crankpin

61

are maintained in mutual communication (FIG.

61

C), although roller

82

may pivot slightly about crankpin

61

, its pivoting motion is limited by the sides of bore

618

engaging the sides of limiting pin

83

. The remaining oil which flows through oil passageway

54

in the crankshaft, which flows beyond crossbore

608

, flows through communicating oil passages

622

and

624

to lubricate bearing

57

. Because oil passage

54

is oriented at an angle relative to the axis of rotation of shaft

34

, oil passage

54

forms a type of centrifugal oil pump which may be used in conjunction with pump assembly

48

disposed in oil sump

46

and described further hereinbelow. The pressure of the oil which reaches radial oil passages

608

and

624

is thus greater than the pressure of the oil in sump

46

, which is substantially discharge pressure. Oil flowing through bearing

57

may flow upwards into oil receiving space or gallery

55

(

FIGS. 15

,

63

B) which is in fluid communication with an intermediate pressure region between the scroll wraps through oil passage

626

. The oil in oil gallery

55

is at discharge pressure, and flows through passageway

626

by means of the pressure differential between gallery

55

and the intermediate pressure region between the scrolls. The oil received between the scrolls through passageway

626

serves to cool, seal and lubricate the scroll wraps. The remaining oil which flows along bearing

57

flows downward into annular oil gallery

632

, which is in communication with annular oil gallery

610

(FIG.

1

).

As best shown in

FIG. 64

, axial bore

84

of roller

82

is not quite cylindrical, and forms, along one radial side thereof, clearance

633

between that side of the bore and the adjacent cylindrical side of the crankpin

61

, which extends therethrough. Clearance

633

provides part of a vent passageway which, during conditions when intermediate pressure between the scroll wraps is greater than discharge pressure, would prevent a backflow gas flow condition through roller bearing

57

. With reference now to the flowpath represented by arrows

635

of

FIG. 63A

, if intermediate pressure is greater than discharge, such as during startup operation of a compressor, refrigerant may be vented through passageway

626

, into oil gallery

55

, and through clearance

633

between bore

84

and the outer surface of crankpin

61

into a region defined by countersink

628

provided in the lower axial surface of the roller

82

about bore

84

and crankpin

61

. This region is in communication with a radial slot

630

provided in the lower axial surface of roller

82

. This vented refrigerant may flow into annular oil gallery

632

and back to housing chamber

110

of the compressor through passageway

612

in frame

60

. In this manner, venting of refrigerant during startup operation assures that oil gallery

55

does not pressurize to the point of restricting oil flow to bearing

57

or, as indicated above, flush the oil from bearing

57

with the venting refrigerant during compressor startup.

As seen in

FIGS. 14

,

15

and

63

, downwardly-facing surface

636

of the orbiting scroll member inside the central cavity of hub portion

516

is provided with a short cylindrical protuberance or “button”

634

which projects downwardly approximately 2-3 mm from surface

636

. Button

634

is, in one embodiment, approximately 10-15 mm in diameter and its axial surface abuts portions of the interfacing uppermost axial surfaces of crankpin

61

and/or roller

82

, which are generally flush with one another. Button

634

provides the function of locally loading crankpin

61

and/or roller

82

so as to minimize frictional contact over the entire upper axial roller and crankpin surfaces and thus serves as a type of thrust bearing. The interface of button

634

and crankpin

61

and/or roller

82

is near the centerlines of hub portion

516

and roller

82

, where the relative velocity between the button and the crankpin and roller assembly is lowest, thereby mitigating wear therebetween.

Positive displacement type oil pump

48

is provided at the lower end of crankshaft

34

and extends into oil sump

46

defined by compressor housing

22

. A first embodiment of the oil pump is disclosed in

FIGS. 65 through 79

and an alternative second embodiment is disclosed in

FIGS. 80 and 81

. In the first embodiment, as shown in the fragmentary sectional side views of

FIGS. 65 and 66

, positive displacement pump

48

is disposed about lower end

702

of crankshaft

34

and is supported by outboard bearing

36

.

The pump is comprised of oil pump body

704

, vane or wiper

706

, which may be made injection molded of a material such as Nylatron™ GS, for example, circular reversing port plate or disc

708

, the planar upper, axial surface of which is in sliding contact with the lower surface of vane

706

, retention pin

710

, wave washer

713

, circular retainer plate

715

and snap ring

712

. The pump components are arranged within pump body

704

in the order shown in

FIG. 68

, and wave washer

713

urges the pump components into compressive engagement with each other. An annular groove is provided in the lower end of the pump body to receive snap ring

712

. Slot

714

, as shown in

FIGS. 55-57

, is provided in lower end

702

of shaft

34

and receives rotary vane

706

, which is longer than the diameter of lower shaft end

702

, and which is caused to rotate by the rotation of the crankshaft. The vane slides from side to side within the slot and contacts the surface of pump cylinder

716

formed in pump body

704

. As best shown in

FIGS. 65 and 73

, pump cylinder

716

is larger in diameter than, and is eccentric relative to, portion

709

of bearing

36

. Further, the centerline of pump cylinder

716

is offset with respect to the center line of crankshaft

34

and lower axial oil passage

52

.

The diameter of portion

709

of bearing

36

is somewhat larger in diameter than lower shaft end

702

, thereby providing a small clearance therebetween, through which oil may leak from pump

48

, as will be described further hereinbelow, to lubricate the lower journal portion

719

of shaft

34

, which is radially supported by journal portion

717

, and axially supported by surface

726

, of bearing

36

.

As shaft

34

rotates, vane

706

reciprocates in shaft slot

714

, its opposite ends

744

,

746

(

FIGS. 74

,

75

) sliding on the cylindrical wall of pump cylinder

716

. Having opposite ends

744

,

746

facilitates multi-direction operation of vane

706

. The vane may alternatively be formed with a spring (not shown) in the middle or may be of a two-piece design with two vane end portions connected by a separate, intermediate spring (not shown). The intermediate spring urges the vane ends outward toward the inner surface of the pump body for a tighter more efficient pumping operation. Such alternative configurations would better seal vane ends

744

,

746

to the cylindrical wall of pump cylinder

716

, thereby reducing pump leakage. The pump relies on some amount of leakage, however, to provide lubrication of lower bearing

36

. Oil leakage past vane

706

as it is rotated in pump cylinder

716

travels upward through the small clearance between lower shaft portion

702

and portion

709

of bearing

36

, providing a source of lubricant to the journal and thrust bearings above. Hence, lower bearing

36

of compressor

20

is lubricated by leakage from pump

48

rather than by oil pumped thereby through lower shaft passageway

52

.

As shown in FIG.

66

and

74

-

79

, oil from sump

46

enters the pump via inlet

50

and is acted upon by a side surface of rotating vane or wiper

706

. The vane forces oil into anchor-shaped inlet

718

provided in the planar, upper axial surface of reversing port plate

708

, where, due to the decreasing volume, the oil is forced to travel into the central reversing port outlet

720

and upwards into axial oil passage inlet

722

, past scallops

750

,

752

in the sides of vane

706

. The anchor shape of the reversing port plate permits effective pumping operation regardless of the direction of rotation of the crankshaft, for oil will be allowed to enter inlet

718

at or near either of its two anchor “points”. Hence, oil will be provided to the compressor's lubrication points even during reverse rotation of the compressor upon shutdown, should that occur. Circumferential retention pin channel

711

is provided in the planar, lower axial surface of reversing port plate

708

to slidably receive retention pin

710

. Pin

710

is fixed relative to the pump body, retained within notch

754

provided in the cylindrical wall of pump cylinder

716

(

FIGS. 68

,

73

) below pump inlet

50

. This permits rotational repositioning of the reversing port plate to properly accommodate multi-direction operation, opposite end surfaces of channel

711

brought into abutment with pin

710

as shaft

34

changes rotational direction. Port plate

708

thus having rotatably opposite first and second positions. Referring to

FIG. 94

, it can be seen that when the shaft, and thus vane

706

, rotates in the direction of arrow

758

, reversing port plate

708

is urged into and assumes its first position as shown. Referring to

FIG. 95

, it can be seen that when the shaft, and thus vane

706

, rotates in the opposite direction, as indicated by arrow

760

, port plate

708

is urged into and assumes its second position, as shown. Plate

708

is urged into its first or second position through frictional engagement with the slidably abutting surface of vane

706

.

As mentioned above, pump cylinder

716

is eccentric relative to the centerline of crankshaft

34

, and the crankshaft centerline is located on the radial side of centerline

762

of pump cylinder

716

which is opposite pump inlet

50

. Oil received from inlet

50

is directed, by one lateral side of vane

706

, to anchor shaped inlet

718

in port plate

708

. This oil is then conveyed, through the channel extending between inlet

718

and outlet

720

of port plate

708

. Oil forced from port plate outlet

720

flows past scallops

750

,

752

in the sides of vane

706

and into inlet

722

of crankshaft oil passageway

52

.

Lower bearing thrust washer

724

rests on lower bearing thrust surface or shoulder

726

to provide a thrust bearing surface for crankshaft

34

. Oil leakage from pump mechanism

48

travels upward from vane

706

through the interface between lower shaft end

702

and lower bearing portion

709

, as described above, to provide lubricating oil to the interface between crankshaft thrust surface

726

and thrust washer

724

, and crankshaft journal portion

719

and bearing journal portion

717

. Provided in bearing portion

709

is recess

756

(

FIGS. 69

,

71

and

96

), which better facilitates the conveyance of oil from the clearance between bearing portion

709

and lower shaft end

702

, to the interface between journal portion

717

of bearing

36

and portion

719

of crankshaft

34

. Grooves (not shown) are formed in thrust washer

724

to assist in the delivery of lubricating oil to thrust surface

726

. In addition, slots (not shown) may be provided in the pump body to assist oil leakage from the pump mechanism to the thrust surface. Also, slot, flat or other relief

728

(

FIGS. 55

,

56

) may be provided in the crankshaft journal portion

719

to provide further rotational lubrication to the interfacing surfaces of the lower journal bearing. In this manner, leakage from the pump, rather than the primary pump flow traveling along the crankshaft axial oil passageway, provides both rotational and thrust lubrication to the lower bearing surfaces. This concentrates the delivery of primary pump oil flow to destinations further up the crankshaft. The pump thus provides a means of lubricating the lower bearing of the compressor which allows relatively loose tolerances of the interfacing surfaces of the pump body and shaft and simple machining of the crankshaft.

As shown in

FIG. 1

, oil from pump

48

travels upwards along lower axial oil passageway

52

and offset upper oil passageway

54

. The offset configuration of the upper oil passageway

54

provides an added centrifugal pumping effect on the primary oil flow of the pump. The upper opening

732

of passageway

54

is provided with plug

616

. Part of the oil flow through passageway

54

is discharged through radial passageway

608

in shaft journal portion

606

(

FIGS. 55

,

56

) and is delivered to bearing

59

. The remainder of the oil flow through passageway

54

is discharged through radial passageway

624

in crankpin

61

and communicating radial passageway

622

in roller

82

, and is delivered to bearing

57

(FIG.

63

B). Oil flows upwards along bearing

57

and into oil gallery

55

, which is defined by the upper surfaces of crankpin

61

and eccentric roller

82

, and the surface

636

of orbiting scroll member

58

. Oil is delivered to the scroll set via axial passage

626

provided in the orbiting scroll member.

Oil pump

48

′ of the second embodiment, as shown in the exploded view of FIG.

80

and the sectional view of

FIG. 81

, functions essentially as described above but is different structurally as it is designed for use in compressors having no lower bearing. Oil pump

48

′ includes anti-rotational spring

738

, which is attached to compressor housing

22

or some other fixed support. Spring

738

supports oil pump body

704

′ axially within housing

22

, and against rotation with shaft extension

740

, which includes axial inner oil passage

742

and is attached to the lower end of a crankshaft (not shown). Slot

714

′, similar to slot

714

of shaft

34

, is provided in shaft extension

740

; vane

706

′ is slidably disposed in the slot for reciprocation therein, the vane rotatably driven by the slot as described above. Instead of wave washer

713

, retainer plate

715

and snap ring

712

, pump assembly

48

′ may alternatively comprise split spring washer

712

′ to urge the pump components into compressive engagement with each other. Pump assembly

48

may be similarly modified. Vane

706

′, reversing port plate

708

′ and retention pin

710

′ are substantially identical to their counterparts of the first embodiment pump assembly, and pump assembly

48

′ functions as described above.

Those skilled in the art will appreciate that pump assemblies

48

,

48

′, although described above as being adapted to a scroll compressor, may also be adapted to other types of applications, such as, for example, rotary or reciprocating piston compressors.

Compressor assembly

20

may be provided with an offset between fixed scroll centerline

802

and crankshaft centerline S. This offset affects the crank arm and radial compliance angle so as to flatten cyclic variations in crankshaft torque and flank sealing force between the scroll wraps. The compressor may incorporate either a slider block radial compliance mechanism or, as shown in the above-described embodiments, a swing link radial compliance mechanism. The following nomenclature is used in the following discussion:

e

orbiting radius (eccentricity);

b

distance from crankpin 61 centerline P to orbiting scroll center

of mass O;

d

distance from crankpin 61 centerline P to eccentric swing link

center of mass R;

r

distance from crankpin 61 centerline P to crankshaft 34

centerline S;

D

offset distance from fixed scroll wrap centerline to crankshaft

centerline

F

force;

M

mass;

O

orbiting scroll center line and center of mass;

P

crankpin 61 center line;

R

swing link center of mass;

S

crankshaft 34 centerline and rotation axis;

RPM

revolutions per minute;

Sub-

scripts

b

swing link

§

flank sealing

ib

swing link inertia

P

drive pin

s

orbiting scroll

tg

tangential, gas

rg

radial, gas

tp

tangential, eccentric pin

rp

radial, eccentric pin

Greek

symbols

θ

radial compliance (phase) angle

α

swing link center of mass angular offset

ξ

Crankshaft angle

There are three characteristics which distinguish the scroll compressors from other gas compression machines, respectively the quiet operation, the ability to pump liquid, and high energy efficiency. The scroll compressor has an advantage over reciprocating or rotary compressors in that it does not suffer mechanical damage during liquid ingestion. This is because the scrolls are provided with a radial compliance mechanism that allows the scrolls to disengage in the event of liquid compression. In such a case, the compressor turns merely into a pump. Typical radial compliance mechanisms also split the driving force into a tangential force meant to balance the friction and compression forces and a radial component to ensure the flank contact between wraps and thus the sealing between compression pockets.

Another advantage is the smoother variation of the crankshaft torque as the compressing gas is distributed in multiple pockets with only two openings each crankshaft cycle. The crankshaft torque is directly proportional to the compression force and the torque arm, respectively the distance between the compression force vector and crankshaft rotation axis. A means of further leveling the crankshaft torque variation is to provide varying distance to the vector, with a minimum value of this distance coinciding with the maximum compression force. However, a corresponding increasing variation in flank sealing force may result. The swing link radial compliance mechanism can level this variation as well.

A radial compliance mechanism often used in scroll compressors is a slider block. The ability of the slider block version to reduce the torque variation in scroll compressors is presented in Equation 1, below. The slider block allows the orbiting scroll to move the center of mass during crankshaft rotation. A side effect of the center of this movement is that the centrifugal force and thus the radial flank sealing force varies with crankshaft angle.

The radial compliance mechanism considered in the present study is a swinglink as described above as with respect to the illustrated embodiments. The force diagram for this swing link is presented in FIG.

82

.

The force balance in X and Y directions as well as the moments about orbiting scroll centerline O (

FIG. 82

) are presented in Equations 1-3:

Σ

F

x

=0=

F

is

−F

fs

−F

fg

−F

rp

+F

ib

*Cos(α) (1)

Σ

F

y

=0=

F

tg

−F

tp

−T

rg

+F

ib

*Sin(α) (2)

where:

F

is

=M

*(2*π*

RPM/

60)

2

*e

and

F

ib

=M

b

*(2*π*

RPM/

60)

2

*{square root over (e

2

+L +((

d−b

+L )*Cos(π−δ))

2

+L )}

Σ

M

o

=0=

F

rp

*b*

Cos(θ)−

F

tp

−F

rg

*b*

Sin(θ)+

F

ib

*e*

Sin(α) (3)

The fixed scroll may be physically translated by an offset defining a locus shown in FIG.

82

. Consequently the orbiting radius (eccentricity) will vary with the crankshaft angle.

With reference to

FIGS. 89

,

90

, as proven in Equation 1, fixed scroll centerline

802

to crankshaft center S offset D causes flank contact force variation only because of the variation in centrifugal force. The swing link brings an additional effect. The centrifugal force changes in same manner the flank sealing force, respectively a positive offset increases the distance between the orbiting scroll center of mass O and crankshaft rotation axis S, thus the flank contact force is increased. However, the positive fixed scroll to crankshaft center offset D causes an increase of the radial compliance angle θ. The increased radial compliance angle decreases the flank contact force due to the radial component of the drive force. Thus, the swing link mechanism has an inherent compensating effect.

The fixed scroll to crankshaft center offset (assumed along line e of

FIG. 82

) causes a change of the radial compliance angle. Table I shows the relation between offset values and the radial compliance angle.

TABLE I

Offset, inches

−0.10

−0.08

−0.06

−0.04

−0.02

0.00

0.02

0.04

0.06

0.08

0.10

Compliance angle, degree

−14.1

−10.2

−6.8

−3.8

−1.1

1.4

3.7

5.9

8.0

10.0

12.0

FIG. 83

is a graph in which the values of the flank contact force versus orbiting radius variation due to the offset for different instantaneous values of the tangential gas force obtained by solving the system of Equations 1-3 are plotted.

FIG. 83

shows the flank contact force for a gas tangential force varying from 100 to 1000 lbf. The gas radial force is assumed to be 10% the gas tangential force value. Other numerical values substituted in Equations 1-3 are for a typical four ton scroll compressor. The variable on the X axis represents the fixed scroll offset. A positive offset corresponds to the orbiting scroll center line moving further from the crankshaft centerline. Equations 1-3 show the following changes have opposite effects: (1) in general, an increase of the gas tangential force increases the flank sealing force; and (2) an increase of the orbiting scroll and swing link centrifugal forces increases the flank sealing force.

The curves in

FIG. 83

show also that the fixed scroll to crankshaft center offset effect on flank sealing force depends on the amplitude of the tangential gas force. For gas tangential force less than 400 lbf, the flank contact force increases by increasing the orbiting radius. For gas tangential force greater than 400 lbf, the flank contact force decreases by increasing the orbiting radius. There is negligible change in the value of flank sealing force for a gas tangential force of 400 lbf For a fixed scroll to crankshaft center offset of −0.075 inch, the flank contact force is constant.

The value of the orbiting radius, e, varies with crankshaft angle in a sinusoidal manner. The flank sealing force presented in

FIG. 83

is plotted vs. the crankshaft angle, ξ, in

FIG. 84

for a 0.010 inch fixed scroll to crankshaft center offset D. The orbiting scroll eccentricity is a function of crankshaft angle and it is calculated as follows:

e

(ξ)=

D

*sin(ξ)

where ξ is the crankshaft angle.

FIG. 84

shows the variation of flank sealing force with crankshaft angle for several values of tangential gas force for a radial compliance angle θ of the 0.010 inch offset. The flank sealing force is inversely proportional to the tangential gas force. However, the offset effect changes qualitatively when increasing the tangential gas force. For an optimal choice of the phase angle, the fixed scroll to crankshaft center offset reduces the maximum sealing force and increases the minimum sealing force. This selective effect can be seen for the phase angle case depicted in

FIG. 84

at a crankshaft angle value of about 180 degrees.

For example, the tangential gas force variation versus crankshaft angle as determined for a scroll compressor operating at a highly loaded condition is plotted in FIG.

85

. The radial gas force, F

rg

, for this condition is about 10% the average tangential gas force, F

tg

.

FIG. 86

shows the flank sealing force versus the crankshaft angle for a fixed scroll to crankshaft center offset D of 0.020 inch and a tangential gas force variation as shown in FIG.

85

. Eight different values for the phase between offset and pressure variation are considered. This figure shows the offset effect emphasized in

FIG. 84

for the tangential gas variation illustrated in FIG.

85

. The flank sealing force is inversely proportional to the variation of the gas tangential force. Flank sealing force variation can be reduced for a phase angle about 90 degrees.

FIG. 87

shows the values calculated for torque versus crankshaft angle.

For a better understanding of the fixed scroll to crankshaft center offset effect on torque variation, the peak-to-peak variations are plotted in

FIG. 88

for several offset values versus the phase angle. In

FIG. 88

one can determine for a given offset the phase angle range where a flattening of the crankshaft torque variation can be obtained. Next, from

FIG. 86

the specific phase angle to minimize flank sealing force variation can be obtained.

From the foregoing it has been concluded that the effect of the fixed scroll to crankshaft center offset is more complex in the case of a swing link than in the case of a slider block. It is shown that the centrifugal force has an opposite effect than the radial compliance angle upon the flank sealing force. An appropriate choice of the fixed scroll offset will reduce the torque variation and at the same time reduce the variation of the flank contact force. This implies a reduced value of the maximum flank contact force while the minimum flank contact force still suffices for sealing. The lower value of the maximum sealing force means less friction loading, thus an opportunity for a more efficient compressor as well as a quieter scroll compressor.

While this invention has been described as having certain embodiments, the present invention can be further modified within the spirit and scope of this disclosure. This application is therefore intended to cover any variations, uses, or adaptations of the invention using its general principles.

Number	Name	Date
1934482	Bixler	Nov 1933
2178425	Johnson	Oct 1939
2246276	Davidson	Jun 1941
2583583	Mangan	Jan 1952
2751145	Olcott	Jun 1956
2855139	Weibel, Jr.	Oct 1958
3039677	Nissley	Jun 1962
3082937	Tucker	Mar 1963
3165066	Phelps et al.	Jan 1965
3184157	Galin	May 1965
3343494	Erikson et al.	Sep 1967
3572978	Scheldorf	Mar 1971
4331420	Jones	May 1982
4331421	Jones et al.	May 1982
4406594	Smaby et al.	Sep 1983
4540347	Child	Sep 1985
4623306	Nakamura et al.	Nov 1986
4902205	DaCosta et al.	Feb 1990
4973232	Etou et al.	Nov 1990
5017108	Murayama et al.	May 1991
5176506	Siebel	Jan 1993
5188520	Nakamura et al.	Feb 1993
5306126	Richardson, Jr.	Apr 1994
5370513	Fain	Dec 1994
5375986	Ukai et al.	Dec 1994
5382143	Nakamura et al.	Jan 1995
5409358	Song	Apr 1995
5445507	Nakamura et al.	Aug 1995
5476373	Mantooth et al.	Dec 1995
5494421	Wada et al.	Feb 1996
5505596	Nakamura et al.	Apr 1996
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6086342	Utter	Jul 2000

Number	Date	Country
0 777 051	Jun 1997	EP
1-134088	May 1988	JP
04 292595	Feb 1993	JP
5-240170	Sep 1993	JP
5-272473	Oct 1993	JP
6-272683	Sep 1994	JP

Positive displacement pump rotatable in opposite directions

Information

Patent Number

Date Filed

Date Issued

Inventors

Original Assignees

Examiners

Agents

CPC

US Classifications

Field of Search

US

International Classifications

Abstract

Description

Claims

CROSS-REFERENCE TO RELATED APPLICATION

US Referenced Citations (35)

Foreign Referenced Citations (6)

Provisional Applications (1)