U.S. patent number 3,759,636 [Application Number 05/234,123] was granted by the patent office on 1973-09-18 for composite variable oil pressure relief and compressor unload valve assembly.
This patent grant is currently assigned to Dunham-Bush, Inc.. Invention is credited to James R. Morin, Sr., Donald D. Schaefer.
United States Patent |
3,759,636 |
Schaefer , et al. |
September 18, 1973 |
COMPOSITE VARIABLE OIL PRESSURE RELIEF AND COMPRESSOR UNLOAD VALVE
ASSEMBLY
Abstract
An axially shiftable piston is pressure biased by screw
compressor oil pressure to a position closing off communication
between the intake and the discharge sides of an axial screw
compressor but is biased to the unload position by the compressor
discharge and a biasing spring acting on the same. The piston
carries a fixed or variable oil pressure relief valve to vary the
compressor oil pressure with change in compressor thrust load
which, in turn, varies with the compressor discharge pressure.
Inventors: |
Schaefer; Donald D.
(Farmington, CT), Morin, Sr.; James R. (Springfield,
MA) |
Assignee: |
Dunham-Bush, Inc. (West
Hartford, CT)
|
Family
ID: |
22880029 |
Appl.
No.: |
05/234,123 |
Filed: |
March 13, 1972 |
Current U.S.
Class: |
417/281; 417/282;
417/310; 418/203; 417/299; 418/84 |
Current CPC
Class: |
F04C
28/26 (20130101); F01C 21/003 (20130101) |
Current International
Class: |
F01C
21/00 (20060101); F04b 049/02 (); F04b
049/08 () |
Field of
Search: |
;417/281,282,310,299
;418/84,203 |
References Cited
[Referenced By]
U.S. Patent Documents
Primary Examiner: Croyle; Carlton R.
Assistant Examiner: Sher; Richard
Claims
What is claimed is:
1. In an axial screw compressor including: a housing defining a
working chamber carrying a pair of intermeshed screws for
compressing a working fluid from a lower intake pressure to a
higher discharge pressure with thrust forces being created within
said screws during compressor operation, oil pressure thrust
balancing means acting on said screws and pump means responsive to
compressor operation for supplying pressurized oil to said thrust
balancing means, the improvement comprising:
a valve assembly including means responsive to compressor discharge
pressure and compressr oil pressure for automatically unloading
said compressor during compressor start up, automatically unloading
said compressor upon reduction in compressor oil pressure below a
predetermined minimum value after start up, and automatically
preventing the compressor oil pressure from rising above a
predetermined maximum value during compressor operation.
2. The axial screw compressor as claimed in claim 1, wherein said
means is responsive to variation in the pressure difference of the
working fluid between compressor intake and discharge for
automatically varying the compressor oil pressure.
3. The axial screw compressor as claimed in claim 1, wherein said
means is responsive to variation in the pressure difference of the
working fluid between the compressor intake and compressor
discharge, for inversely varying the compressor oil pressure.
4. The axial screw compressor as claimed in claim 1, wherein said
axial flow compressor includes a working fluid bypass passage
connecting the discharge side of the compressor to the intake side,
said valve assembly includes a shiftable valve member, said valve
member being operatively positioned with respect to said bypass
passage and said pressurized oil such that the compressor discharge
pressure tends to move said valve member to compressor unload
position, and said compressor oil pressure tends to move said valve
member to compressor load position, and said valve assembly further
includes spring biasing means tending to move said valve member to
compressor unload position in the absence of compressor oil
pressure.
5. The axial screw compressor as claimed in claim 1, wherein: said
compressor includes a bypass passage fluid connecting said
discharge and intake sides of said compressor, and said valve
assembly comprises: a valve block including an elongated bore, a
cylindrical main piston slidably and sealably positioned within
said bore and having one end in a first position closing off one
end of said bypass passage to load the compressor with the
compressor discharge acting on that end of the piston when in valve
closed position, means subjecting the other end of said piston to
said pressurized oil for maintaining said main piston in said first
position and spring means acting on said main piston in opposition
to said oil pressure to shift said piston to a second position
within said bore to permit said discharge passage to be in fluid
communication with the compressor intake, whereby; upon loss of oil
pressure or during compressor start up, said spring means maintains
said piston at said second compressor unload position.
6. The axial screw compressor as claimed in claim 3, wherein: said
compressor includes a bypass passage fluid connecting said
discharge and intake sides of said compressor, and said valve
assembly comprises: a valve block including an elongated bore, a
cylindrical main piston slidably and sealably positioned within
said bore and having one end in a first position closing off one
end of said bypass passage to load the compressor with the
compressor discharge acting on that end of the piston when in valve
closed position, means subjecting the other end of said piston to
said pressurized oil for maintaining said piston in said first
position and spring means acting on said main piston in opposition
to said oil pressure to shift said piston to a second position
within said bore to permit said discharge passage to be in fluid
communication with the compressor intake, whereby; upon loss of oil
pressure or during compressor start up, said spring means maintains
said piston at said second compressor unload position.
7. The axial screw compressor as claimed in claim 5, wherein said
main piston is centrally bored to fluid connect said bypass passage
with the pressurized oil acting on the end of said piston opposite
that subjected to compressor discharge, and said valve assembly
further includes an oil pressure relief valve carried by said
piston for closing off said bore, whereby the oil pressure is
maintained below a certain predetermined maximum value by bleeding
off compressor oil into the compressor discharge.
8. The axial screw compressor as claimed in claim 5, wherein; said
main piston is centrally bored and counterbored, to effect fluid
communication between respective ends of said piston, a ball of a
diameter less than the counterbore but greater than the bore is
carried within said counterbore, said counterbore is axially
threaded, a hollow screw is threadably received by the threaded
counterbore, and a coil spring is axially compressed between the
ball and the screw to permit setting of said valve to a
predetermined maximum desired oil pressure.
9. The axial screw compressor as claimed in claim 5, wherein; said
main piston is hollow for placing the ends of said piston in fluid
communication, an oil relief valve plunger is slidably and sealably
positioned within hollow main piston, a floating piston is slidably
positioned within said main piston, a calibration compression
spring is positioned coaxially between said floating piston and
said plunger to bias said plunger and floating piston in opposite
directions, a power spring is positioned between said floating
piston and the main piston end wall remote from said plunger for
biasing said plunger and said floating piston in a common direction
to close off fluid communication between said main piston ends, and
said valve assembly includes means permitting the working fluid
pressure differential between the compressor inlet and the
compressor discharge to act on the floating piston and to thereby
control the extent of compression of said calibration spring,
whereby the maximum oil pressure permitted by said oil relief
plunger varies inversely with the working fluid pressure
differential across the compressor.
10. The axial screw compressor as claimed in claim 9, wherein said
main piston includes means defining an annular chamber at the end
of said piston remote from said plunger, said floating piston
includes a tubular wall portion slidably received within said
annular chamber, and said main piston includes means permitting
compressor working fluid discharge pressure to act on one end of
said floating piston tubular wall portion to compress said power
spring, and said main power piston further includes means for fluid
communicating the compressor working fluid intake with said annular
chamber to permit the compressor inlet pressure to act on the other
end of said tubular wall portion of said floating piston and in
conjunction with said power spring.
11. The axial rotary screw compressor as claimed in claim 10,
wherein said power spring comprises a coil spring carried by said
annular chamber having one end in contact with the end of said main
piston, and the opposite end in contact with said floating piston
tubular wall portion.
12. The axial screw compressor as claimed in claim 1, wherein; said
compressor housing section carrying said intermeshed screws is
bored axially the length of the same to define a bypass passage, a
valve block is fixed to the inlet end of said axial screw
compressor housing section and includes means defining a compressor
working fluid intake passage, said block includes a bore coaxial
with the bypass passage, acts as an extension of the same and
intersects said compressor inlet, a main piston is slidably and
sealably carried by said bore within said valve block and includes
means for shutting off the end of said bypass passage in a first
position but permitting fluid communication between said bypass
passage and said compressor intake passage when in a second
position, spring means coupled to said main piston acts to move
said main piston to compressor unload position, means directs
compressor oil to the end of said main piston remote from said
bypass passage for producing a force in opposition to said springs
and the compressor discharge acting on the opposite end of said
main piston, said main piston is bored and counterbored to define a
fluid passage connecting the ends of the same, means define an
annular cavity within said main piston, a floating piston including
a tubular portion is slidably and sealably received within said
annular cavity, means permit compressor discharge to act on one end
of said tubular portion of said floating piston, means fluid
communicate the compressor intake to said annular chamber to permit
the working fluid at the compressor inlet pressure to act on the
other end of said floating piston, a power compression coil spring
is coaxially carried within said annular cavity to bias said
floating piston away from said fluid bypass passage end of said
main piston passage, a plunger is slidably carried by said floating
piston and adapted to seat within said bore and acts as a pressure
relief valve, and a calibrating compression coil spring is
concentrically carried by said plunger with one end in contact with
the same and the other end abutting said floating piston, whereby
at high compressor working fluid pressure differential, said
floating piston compresses the power the to lower the biasing force
acting on the plunger and thereby relieve the oil pressure at a
relatively low pressure value, while at low working fluid pressure
differential between the compressor intake and compressor
discharge, the power spring forces the floating piston to compress
the calibrating spring and increase pressure at which the oil
pressure relief valve operates and to thereby effectively control
oil pressure thrust compensation in terms of the axial thrust
forces on the compressor screw resulting from operation of the
compressor.
13. The axial screw compressor as claimed in claim 12, wherein said
main piston comprises first and second cylindrical members
including spaced concentric tubular wall portions coupled at the
end closest to said bypass passage to define said annular cavity,
and said floating piston comprises spaced inner and outer tubular
wall portions coupled together by a base portion and said main
piston includes a bore and two counterbores in that order and
wherein the shoulder between the two counterbores defines a stop
determining the maximum compression of said calibrating spring,
while the end of inner tubular wall portion of said main piston
contacts the base portion member coupling the inner and outer
tubular wall portions of said floating piston to limit the extent
of expansion of said calibrating spring compressively positioned
between the floating piston and the plunger.
Description
BACKGROUND OF THE INVENTION
1. Field of the Invention
This invention relates to axial screw compressors and, more
particularly, to such compressors wherein the compressor system oil
pressure is employed to create a counterthrust to balance the axial
thrust created by the compressor screws acting on the fluid to be
compressed. The invention is particularly directed to axial screw
compressors for compressing refrigerants.
2. Description of the Prior Art
In axial screw compressors, the proper management of axial forces,
generated during the compression process, is mandatory if a
successful design is to be achieved. This axial force is the sum
of:
A. the resultant gas force due to the pressure difference across
the machine, and
B. the contact forces which are the result of one rotor driving the
other.
The resultant gas force is the sum of the increased gas pressure in
the rotors interlobe spaces, and the gas pressure acting on the end
faces of the rotor blades and the supporting and driving shaft
ends.
In compressors currently being manufactured, these axial forces are
balanced by:
1. THE USE OF MATCHED BALL BEARING SETS ON EACH ROTOR "TO ABSORB"
THE THRUST;
2. THE USE OF BALANCE PISTONS, WHICH ARE GROOVED OVERSIZED SHAFT
ENDS, WHICH HAVE THE COMPRESSOR FLUID DISCHARGE PLUS OIL PRESSURE
ACTING ON ONE SIDE AND THE SUCTION PRESSURE ACTING ON THE OTHER,
WHEREIN THE PRESSURE DIFFERENTIAL ACROSS THE BALANCE PISTONS
GENERATES THE NECESSARY COUNTERTHRUST; OR
3. THE USE OF A COVER OVERLYING THE OUTLET SHAFT ENDS TO PERMIT THE
GAS PRESSURE TO BE KEPT AT SUCTION PRESSURE AND THUS LESSEN THE
THRUST; OR
A COMBINATION OF ALL THREE.
In the current modes of balancing or compensating for the thrust
forces, there are unfavorable features to each of these three
approaches. For instance, where ball bearings are employed to
absorb the thrust, a fixed high side rotor end clearance is
provided so that the rotors do not come in contact with the
housing. This clearance results in fluid leakage from the high
pressure side to the low pressure side of the compressor. This, in
turn, causes a loss in compressor capacity, etc. Further,
anti-friction bearings are noisy and comparatively expensive.
Where balance pistons are employed, they require a large amount of
oil because of their poor sealing ability and, if too much oil is
"consumed" by the pistons, the suction gas will be displaced in the
compressor by oil, thus the compressor will be flooded with oil and
will be unable to pump an efficient amount of refrigerant, if the
working fluid constitutes a refrigerant, greatly lowering the
efficiency of the compressor. While covers which overlie the shaft
ends may be suitable for machines which are being driven from the
suction side, this approach at overcoming the thrust created by the
operation of the screws is complicated when the screws are driven
from the compressor discharge side, since the covers must employ
dynamic seals which are expensive and somewhat unreliable.
SUMMARY OF THE INVENTION
The present invention is directed to an axial screw compressor
which employs oil pressure acting on balance pistons to provide the
counter-thrust necessary to balance out the thrust forces created
by the pressure differential across the compressor screws, and by
one of the screws being driven by the other. The invention itself
is directed to a composite valve assembly which performs the
functions of:
1. insuring that the compressor starts unloaded;
2. if for any reason the oil pressure is insufficient, unloading
the compressor; and
3. providing an oil pressure relief valve which may be selectively
preset or which may vary with the compressor thrust load.
Specifically, the invention is directed to a screw compressor which
includes a housing defining a working chamber which carries a pair
of intermeshed screws for compressing fluids entering the chamber
from an inlet at low pressure, to a higher pressure for discharge
at the outlet, and wherein an axial thrust force is developed
within the screws during fluid compression. The invention is
further directed to such a screw compressor in which oil pressure
thrust balancing means counteract this thrust and wherein an oil
pump responsive to compressor operation provides the thrust
balancing means with pressurized compressor oil.
The improvement resides in a valve assembly responsive to
compressor start up and/or low oil pressure for automatically
unloading the compressor during compressor start up and at any time
the oil pressure drops below a preset value. An unload or bypass
passage connects the compressor outlet to the inlet and a shiftable
main valve piston is operatively positioned within the passage and
is subjected to oil pressure acting on one end of the piston to
maintain the unload passage closed, while spring biasing means acts
in conjunction with the pressure of the compressor discharge on the
other end to open the valve.
The main valve piston carries an oil pressure relief valve which in
one form is adjustably set to prevent the oil pressure from rising
above a preset maximum level, the pressure relief valve opening to
permit the oil to enter the unload passage when the main piston is
in valve closed, compressor load position, whereupon the oil mixes
with the compressor working fluid discharge. Alternatively, the
main piston which itself is shiftable to permit or shut off fluid
communication between the discharge side of the compressor and the
intake side may be bored and counterbored to receive a floating
piston which is spring biased toward the oil pressure end of the
main piston by a main power spring. A plunger which closes the bore
and is subjected to compressor oil pressure, reciprocates within a
central bore of the floating piston, being spring biased by means
of a calibration spring towards valve closed position. The oil
pressure acting on the seated area of the plunger is opposed by the
compressive force of the calibration spring and the discharge
pressure of the compressor, while the same discharge pressure acts
on the floating piston carrying the plunger and is opposed by the
bias of the compression force of the main power spring and the
intake suction pressure of the compressor acting on the opposite
end of the floating piston. This arrangement permits the control of
oil pressure inversely to working fluid pressure differential
across the compressor.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a sectional view of a portion of an axial screw
compressor embodying the improved, combined compressor unload and
oil pressure relief valve assembly of the present invention, in one
form, with the main piston in valve closed, compressor load
position.
FIG. 2 is a sectional view of the combined compressor unload and
oil pressure relief valve assembly of FIG. 1, with the main piston
in valve open, compressor unload position.
FIG. 3 is a sectional view of the combined compressor unload and
oil pressure relief valve assembly of FIGS. 1 and 2, with the main
piston in valve closed, compressor load position, but with the oil
pressure relief valve in open position.
FIG. 4 is a sectional view of a portion of a rotary screw
compressor incorporating a second embodiment of the combined
compressor unload and oil pressure relief valve assembly of the
present invention in compressor load position and at relatively low
compressor working fluid pressure differential.
FIG. 5 is a sectional view similar to that of FIG. 4 of the valve
assembly, with the floating piston shifted and the plunger of the
oil pressure relief valve in valve open position under high
compressor working fluid pressure differential conditions.
DESCRIPTION OF THE PREFERRED EMBODIMENTS
Turning to FIG. 1 of the drawings, a first embodiment of the
present invention is illustrated as being applied to a rotary screw
compressor of which only a portion is shown, indicated generally at
10. Compressor 10 comprises a sealed outer housing including a
generally cylindrical casing or housing section 14 carrying at the
center thereof a configured working space or working chamber 16
supporting a pair of intermeshed rotors screws such as the female
screw 18 which, in this case, is directly driven by means of an
electric motor (not shown) fitting space 20 whose rotor is directly
coupled to screw 18 by means of an integral connecting shaft 22.
The intake and discharge ends of the compressor housing section 14
are sealed by means of end plates 24 and 26 respectively, housing
12 further including an additional plate or section 28 at the
discharge side which forms, in this case, an end plate portion of
the electric drive motor 20. At the intake or inlet side of the
screw compressor 10, there is further provided a relatively thick
valve block 30 which is bolted or otherwise fixed to the exposed
face of compressor end plate 24, the components 30, 24, 14, 26 and
28 of the compressor carrying appropriate recesses which receive
O-rings 33 to insure fluid sealing between these sections of the
compressor housing 12. The means for coupling the sections of the
compressor housing together are not illustrated, but may in fact
constitute bolts or the like which extend through multiple
sections. Valve body 30 in addition to carrying within bore 32, the
main valve body or piston 34 of a combined compressor unload and
oil pressure relief valve assembly 31 which reciprocates therein,
is provided with a relatively large intake or inlet passage 35,
which constitutes the compressor inlet permitting, in the
illustrated embodiment, a refrigerant gas which forms the working
fluid for the compressor to pass through openings 36 within end
plate 24 and enter compressor working chamber 16 for compression
between the female screw 18 and the intermeshed driven male screw
(not shown) which is laterally offset from screw 18. The compressed
working fluid is discharged from the screw compressor through
discharge opening 38 of end plate 26 at the right hand end of the
housing 14. Opening 38 is aligned with an opening 40 within the
motor end plate 28, allowing the compressed gas to pass over the
rotor and stator of the drive motor 20 to cool the same in
conventional fashion. This general arrangement of elements is
conventional and forms no part of the present invention. Further,
while the bearings supporting the screws are not illustrated, screw
18 rotates about its axis and centerline 42, being positively
driven by a direct connection with the electric rotor shaft 22
which is in axial alignment therewith. Valve block 30 is further
provided at the left hand end with an end plate or cover 44 which
overlies bore 32 and is sealably affixed to the outer side wall of
block 30 by means of plurality of mounting screws 46. Block 30 is
further provided with a recess 48 which carries oil under pressure
from a compressor driven oil pump (not shown) which may, for
instance, be directly connected to the electric motor rotor shaft
22. Recess 48 extends to bore 32 so that pressurized oil fills bore
32 to the extent permitted by a slidable and sealable cylindrical
main valve body or main piston 34. Multiple ring seals 39 fill a
peripheral recess within the main piston 34. Main piston 34
constitutes a composite cylindrical valve body including a first
section 49 having a diameter slightly less than that of bore 32, a
second section 50 of slightly less diameter, and a third section 52
in the form of a relatively thin plate whose diameter is in excess
of both sections 49 and 50. Sections 50 and 52 are coupled to the
longer section 49 by several screws 54, the section 52 constituting
the headed end of the main piston and being provided with an
annular O-ring 56 constituting a peripheral seal for contact with
the main piston valve seat defined by left hand end wall 58 of the
compressor housing section 14 surrounding bore 60. Bore 60 within
housing section 14 forms a compressor unload or bypass passage 62
and acts in conjunction with openings 64 and 66 within right hand
end plate 26 and the electric motor housing end wall 28,
respectively, to permit fluid connection between the intake and
discharge sides of the screw compressor. With the main piston 34 in
valve closed position as illustrated in FIG. 1, the working fluid
for the compressor such as a refrigerant gas, may enter compressor
inlet passage 35 within block 30 and pass through the opening 36
within the compressor end plate 24 for compression by the screws
and discharge at a higher pressure downstream of the screws via
opening 38 within end plate 26. The unload passage 62 is cut off by
the main piston 34 since its right hand end abuts the left hand
wall 58 of the compressor housing surrounding bore 60. The main
left hand section 49 of the main piston 34 has coupled thereto, a
pair of tension coil springs 68 by means of mounting screws 70,
while the springs at their opposite ends are fixed to a spring
retainer or plate 72 which extends transversely across bore 32. The
coil springs 68 thus tend to bias the main piston 34 into valve
open or compressor unload position as illustrated in FIG. 2.
Section 49 of the main piston 34 is centrally bored at 74 and is
counterbored at 76, while sections 50 and 52 carry, respectively,
axially aligned bores 77 and 78, thus forming a fluid passageway
through the center of the main piston 34. A portion of the
counterbore 76 is threaded as at 80 and receives a threaded,
cylindrical adjusting screw 82 which itself is centrally bored at
84. Further, a ball 86 of a diameter in excess of bore 74 but less
than counterbore 76 is provided within counterbore 76, and is
biased to valve closed position against an inclined seat 87 defined
by the wall connecting bore 74 with counterbore 76 within section
49 of the main piston 34. Ball 86 is biased by means of a
compression coil spring 88, one end of which abuts hollow adjusting
screw 82. Thus, the main piston 34 carries a compressor oil
pressure relief valve 85 which may be set to open at a desirable
oil pressure by axial shifting of the threaded adjusting screw 86
in the embodiment of FIGS. 1-3 inclusive.
As mentioned previously, the axial screw compressor which employs
the combined unload and oil pressure relief valve assembly of the
present invention, is the type in which oil pressure is employed to
counter balance the axial thrust created by operations of the
compressor screw compressing the working fluid. As illustrated in
FIG. 1, this is the normal condition of the composite valve 31
during sustained operation of the compressor. In such case, the
working fluid entering inlet 35 passes through opening 36 within
end plate 24 and at intake or suction pressure is compressed
conventionally by rotation of the intermeshed screw, as it passes
axially from left to right for discharge through aligned openings
38 and 40 into the interior of the electric motor section of the
screw compressor assembly downstream and to the right of the
compressor itself. The high pressure refrigerant gas, assuming
refrigerant is the working fluid, enters the unload passage 62
through aligned openings 64 and 66 and acts directly on the right
hand end of the main piston 34 tending to unseat the main piston 34
and permit return of the pressurized working fluid to the intake
side of the compressor. The force of the compressed gas acts
jointly with the coil springs 68 tending to return the piston 34 to
an extreme left hand position as defined by abutment 90 which stops
the radially enlarged or headed end of the main piston 34. In
contrast, the oil pressure acting on the left hand end of the main
piston 34 tends to maintain piston 34 in the extreme right hand
position or compressor load position.
Prior to reaching the condition illustrated in FIG. 1, at start up,
the suction, discharge and oil pressure are all equal, since the
axial thrust in the compressor is in part due to the difference in
gas pressure across the compressor. The low intake pressure at the
inlet may vary from zero to 100 psi and the high discharge pressure
at the outlet of the compressor may range from 200 to 400 psi. It
is seen, therefore, that it is undesirable to let this difference
in pressure occur until the coil pressure has built up to a
sufficient level to counteract the thrust load via the balance
pistons (not shown) on the rotor ends.
The pressurized oil within recess 48 and bore 32 to the left of the
main piston 34 is supplied by means of an oil pump (not shown)
connected directly to the electric motor 20 and oil pressure is
established almost instantaneously. However, the present invention
prevents the gas and contact thrust from adding prior to oil
pressure build-up to the point where the oil pressure establishes a
counterthrust in unusual instances where the oil pressure is slow
in building, where the compressor is started after long shut downs
or where there is high percent of refrigerant in the oil. Since the
gas suction pressure is equal to the gas discharge pressure and oil
pressure at start-up, the only force acting on the main piston 34
is that provided by coil springs 68. The coil springs therefore
shift the piston 34 from the extreme right hand position where the
headed end abuts end wall 58 of screw compressor housing 14
adjacent the left hand end of bore 60 to a position in which the
headed end contacts abutment 90 unloading the compressor, with the
compressor discharge recirculated to the compressor inlet, as seen
in FIG. 2. However, almost immediately the oil pressure builds up
sufficiently to force the piston to move from the position shown in
FIG. 2 to the position shown in FIG. 1 with the oil pressure relief
85 in closed position in the absence of excessive oil pressure.
It is to be noted, that the motor horse power required to start and
run the compressor is proportional to the pressure difference
across the compressor at the time of starting and thus by
eliminating the pressure difference at start up, this reduces the
necessary horse power requirement at a time in which the power
requirements are normally the highest. Thus, the compressor can
start easier, since it is unloaded at start with less amperage
requirement. Momentarily, the lack of oil pressure permits start up
with the discharge gas being bypassed to the suction side of the
compressor, but almost immediately thereafter, oil pressure rises
sufficiently to shift the main power piston 34 from left to right
to seal the bypass or unload passage 62 from the compressor intake
passage 35.
The combined valve assembly 31 of the present invention permits the
compressor to be unloaded automatically at any time during
operation when the oil pressure falls below a preset value. For
instance, if the oil pressure is normally in excess of 40 psi above
compressor discharge pressure, but if it should fall below
approximately 10 psi, the combination of the force created by the
coil springs 68 and the compressor discharge gas acting on the
right hand end of the main piston 34 causes the main piston 34 to
move from right to left, again unloading the compressor and
permitting recirculation of the compressed gas to the discharge
side of the screws. Oil pressure may drop as a result of loss of
oil in the sump, failure of the oil pump (not shown), blockage of
the oil pump inlet, etc. As the bearing loads are partially a
result of the pressure difference across the compressor, it is
desirable to eliminate this loading if the relative oil pressure to
the bearings falls below a safe level, for instance, 25 psi. In
contrast to commercially available oil pressure safety devices of
the type such as the externally mounted bellows operated electrical
contacts, which have a 30 to 90 second delay built into them to
eliminate nuisance stoppage of the compressor, the mechanism of the
present invention immediately lessens the bearing loads in the
unprotected first zero to 90 seconds before the external safety
device shuts the compressor down. If there is an oil failure of
short duration, that is from zero to 90 seconds, the combined valve
assembly 31 of the present invention will partially protect the
compressor without shutting it off by simply unloading the
compressor but permitting the compressor to continue to operate in
unloaded condition.
Turning to FIG. 3, the contrast in conditions for the valve
assembly 31 and those shown in FIG. 1, illustrates the operation of
the oil pressure relief valve 85 which forms a portion of combined
valve assembly 31. Rotary screw compressors, even where the
pressurized oil system does not function to absorb or counteract
the axial thrust set up in the rotary screws, employs a pressure
relief valve to maintain the oil pressure at a preset level
depending upon the design requirements of the compressor. In air
conditioning and refrigeration use, rotary screw compresor oil
pumps are usually oversized by 50 to 100 percent so sufficient oil
and oil pressure will be available upon bearing wear, and where oil
requirements go up, and also to insure adequate lubrication when
the oil becomes less viscous under certain conditions that are
outside of the design limits of the machine such as when the oil is
extremely hot or where there is an excessive amount of refrigerant
within the oil. The present invention is advantageously directed to
the incorporation of oil relief valve 85 as a part of the combined
valve assembly 31, and further, to a valve whose relief pressure is
adjusted by increasing or decreasing the compression of the coil
spring 88 which abuts the ball 86. Assuming that the maximum
relative oil pressure desired for all oil functions such as
lubrication, and in particular, oil pressure balancing of the axial
thrust of the screws is in the neighborhood of 80 psi above
compressor discharge. The screws 82 can be shifted axially to the
left to an extent where the ball 86 will unseat from the end of the
bore 74 at oil pressure of 100 psi for instance. During this time,
the oil pressure acitng on the left hand end of the main piston 34
maintains the valve in closed or compressor load position, that is,
to its extreme right hand position, as illustrated in FIG. 1. If
the oil pressure exceeds 100 psi, the ball 86 unseats to permit
bleeding of the oil into the compressor discharge via unload
passage 62 to the extent necessary to maintain the oil pressure at
100 psi, as evidenced in FIG. 3.
One of the characteristics of the axial screw compressor is that
the amount of thrust varies as the working fluid pressure
difference across the compressor changes, thus it is possible to
set the oil pressure relief valve to provide the proper oil
pressure to the balance pistons at a given pressure differential
across the compressor to offset the thrust load. It is most
desirable therefore to provide a relief valve that varies in its
setting according to the change in pressure difference across the
compressor and further to provide a pressure relief valve that
provides the correct pressure to balance the thrust load
effectively, regardless of the pressure difference across the
compressor. It was determined that where a 330 psi pressure
differential exists across the compressor, the oil pressure
permitting the proper thrust balancing via the balance pistons
would necessarily be of the order of 40 psi above compressor
discharge pressure, while for a pressure differential across the
compressor on the order of 110 psi, a relative oil pressure of 80
psi is required to permit the balance pistons to properly balance
out the compressor thrust. Thus, the alternate form of the
composite valve 31' of the present invention, illustrated in FIGS.
4 and 5, not only automatically varies the oil pressure to the
needs of the hydraulic balance pistons, but because the oil
pressure requirements are such that an increase in compressor
pressure difference requires less relative oil pressure, the oil
relief valve 85' in the alternate embodiment of the present
invention operates inversely in terms of the pressure difference
across the compressor.
Turning to FIGS. 4 and 5, it is noted that in terms of this
embodiment, the casing or housing section of the axial screw
compressor is identical to that of the prior embodiment. Like
elements carry like primed numerical designations. Thus, compressor
housing section 14' carries a bore 60' defining a compressor bypass
or unload passage 62' which is sealed off at its left hand end by
the headed end 52' of the main piston 34'. In this case, the main
piston 34' is a composite body, formed of steel or like metal
including joined sections 49' and 52', which is carried within bore
32' of a valve block 30'. Portion 48' of bore 32' receives the
pressurized oil which creates a force acting on the left hand end
of the main piston 34' in opposition to the compressor discharge
which acts on the right hand end of piston 34'. Again, the main
piston 34' reciprocates between a closed position in which its
right hand end abuts a portion of the end face 58' of housing
section 14' surrounding bore 60' to shut off the bypass or unload
passage 62' from the compressor inlet passage 35', but upon loss of
oil pressure, the coil springs 68' acting in conjunction with the
compressor discharge move the main piston 34' to a position in
which the shoulder formed by the headed end 52' of the valve body
contacts the abutment 90' to limit further movement to the left and
permit the compressor discharge to re-enter the intake side of the
compressor.
Thus, insofar as unloading the compressor at start or at any time
that the oil pressure decreases below a preset minimum figure, the
embodiment illustrated in FIGS. 4 and 5 operates identically to
that illustrated in FIGS. 1-3 inclusive. As mentioned previously,
the main piston 34' consists of a relatively long section 49' to
which is coupled the enlarged diameter right hand end section 52'
by means of a plurality of screws 100 whose headed ends are
recessed within section 52'. A hollow threaded plug 101 fixes a
thin sealing disc 103 to the end of piston section 52'. The
variable oil pressure relief valve 85' which is carried centrally
of and within the main piston 34', consists essentially of four
parts in addition to the main piston 34'. A plunger 102 is axially
shiftable with respect to a floating piston 104 which supports the
same, the floating piston in turn being seated within counterbore
106 of section 49' of the main piston 34', the plunger being biased
to the left and to valve closed position by a calibration coil
spring 108. In turn, the two part main piston 34' defines an
annular cavity 110, which lies between an elongated inner tubular
wall portion 112 of section 52' and counterbore 106 of section 49'.
It is within this annular cavity 110 that the floating piston 104
which is irregularly double U-shaped in cross section is permitted
to slide and may be positioned to the left by the main power spring
114 taking the form of a compression coil spring. Section 49' of
the main piston 34' is provided with a small diameter bore 74' and
a first counterbore 116 in addition to the second counterbore 106,
the end of bore 74' being beveled adjacent the headed end of
plunger 102 to define the valve seat 118 for the variable oil
pressure relief valve 85'. The plunger 102 is provided with a
reduced diameter portion 120 about which is concentrically
positioned the forward end of the calibration spring 108. The
floating piston 104 is provided with a base portion 122 which
connects the spaced parallel tubular side walls 124 and 126, the
outer side wall 126 being of extended length and of a thickness
generally on the order of the annular recess 110 receiving the
same, permitting it to be sealable and slidable therein. The end of
the inner tubular wall portion 112 of main piston 34' is positioned
in the path of base 122 to limit the right hand movement of the
floating piston 104 which compresses the main power spring 114
during movement from left to right. Further, face 128 of the base
portion 122 of the floating piston abuts shoulder 130 defined by
counterbores 116 and 106 of section 49' of the main piston 34' to
limit movement of the floating piston to the left. The tubular wall
124 of the floating piston is provided with a flange portion 132
against which the calibration spring 108 abuts, the calibration
spring 108 concentrically surrounding the plunger 102 at its
forward end while the floating piston concentrically surrounds the
calibration spring 108 at its rear to increase the compressive
force acting on the plunger and to maintain it against the valve
seat 118 as the floating piston shifts left from its extreme right
hand position to the point where face 128 abuts the shoulder 130 of
the main piston section 49'. The floating piston is provided with
at least one radial opening 134 permitting the compressor discharge
to act at all times on the left hand end face of the floating
piston and in opposition to the compressive force of the main power
spring 114 and the suction gas pressure acting on the right hand
end face 130 of the outer tubular wall 126 of the floating piston.
Further, section 49' of the main piston is provided with at least
one radial opening 138 which permits gas inlet pressure to reach
the annular cavity 110 carrying the main power spring 114 and the
tubular outer wall 126 of the floating piston. Sufficient clearance
is provided between the plunger 102 and the floating piston 104
which carries the same, such that when the piston is displaced to
the right of valve seat 118, the pressurized oil may escape through
bore 74' and enter the unload passage 62' through the hollowed
interior of the floating piston 104.
FIG. 4 shows the valve assembly 31' of the present invention in its
high relative oil pressure mode, while FIG. 5 shows the composite
valve in its low relative oil pressure mode. Reference to FIG. 4 is
made with the realization that under this mode of operation, the
working fluid pressure difference across the compressor and acting
on floating piston 104, is relatively small, in which case the main
power spring 114 maintains the floating piston 104 at its extreme
forward position, in which case, face 128 of the floating piston
104 abuts the shoulder 130, causing the calibration spring 108 to
be compressed to the point where it is exerting its greatest spring
force and thus the relative oil pressure needed to unseat the
plunger 102 from seat 118 must be at its highest in terms of the
setting for the valve assembly. In the illustrated embodiment, a
spring force of 3.92 pounds maintains the plunger in oil pressure
relief valve seated position and the oil pressure needed to unseat
the plunger would necessarily be 80 psi above the gas discharge
pressure. If it is assumed the working area of the plunger 102 is
0.0491 square inches (80 psi .times. 0.0491 square inches) = to
3.92 pounds, which is equal to that supplied by the calibration
spring 108. In the illustrated version, the mechanism works
linearly, although a non-linear mechanism could be made by using
non-linear springs or two or more linear springs with different
spring rates.
In contrast, turning to FIG. 5, the pressure differential between
the discharge pressure of the compressor and the inlet pressure is
sufficiently high to shift the floating piston 104 from its extreme
left hand position to its extreme right hand position, wherein the
end of the tubular wall 112 abuts the base 122 of the floating
piston. In this case, the main power spring 110 is compressed to
its fullest extent. At the same time, the calibration spring 108
lengthens and its spring force reduces to about 1.96 pounds. The
working area of the plunger is again 0.0491 square inches for the
illustrated embodiment, and oil pressure above 40 psi will unseat
the plunger from its valve seat 108, relieving the oil, since the
force acting against the plunger would be 0.0491 square inches
.times. 40 psi) = 1.96 pounds and the forces on both sides of the
plunger would be equal at that point. It is at this point that the
plunger starts to relieve as shown, and a constant relative oil
pressure is thus maintained at 40 psi, assuming that the pressure
differential across the compressor remains the same and at a
relatively high value. It is noted that the power spring 114 and
the calibration spring have the same travel between extreme limits,
although of course it is not necessarily to maintain this
relationship.
From the above description, it is appreciated that, in particular,
with the embodiment of FIGS. 4 and 5, the combined valve assembly
permits the compressor to start unloaded, where the compressor has
a built-in compression that is not relieved, unloads the compressor
when oil pressure drops below a preset level, maintains the oil
pressure setting in the compressor oil system and varies the oil
pressure with change in compressor thrust load to provide the
necessary varying counterthrust to the changing axial thrust
experienced by the screws during compressor operation.
* * * * *