U.S. patent application number 14/513009 was filed with the patent office on 2015-01-29 for bearing housing and assembly of a screw compressor.
The applicant listed for this patent is Trane International Inc.. Invention is credited to Dennis M. Beekman, Daniel R. Crum, Dennis R. Dorman, Timothy Sean Hagen, John R. Sauls.
Application Number | 20150030490 14/513009 |
Document ID | / |
Family ID | 52390674 |
Filed Date | 2015-01-29 |
United States Patent
Application |
20150030490 |
Kind Code |
A1 |
Beekman; Dennis M. ; et
al. |
January 29, 2015 |
Bearing Housing and Assembly of a Screw Compressor
Abstract
An improved bearing housing of a rotary screw compressor is
described. The bearing housing is generally shorter than a
convention bearing housing. The bearing housing can be configured
to enclose and support radial bearings of the screw compressor. The
bearing housing can be configured not to enclose axial bearings of
the screw compressor in an axial direction.
Inventors: |
Beekman; Dennis M.; (La
Crosse, WI) ; Crum; Daniel R.; (La Crosse, WI)
; Hagen; Timothy Sean; (Onalaska, WI) ; Dorman;
Dennis R.; (La Crosse, WI) ; Sauls; John R.;
(La Crosse, WI) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Trane International Inc. |
Piscataway |
NJ |
US |
|
|
Family ID: |
52390674 |
Appl. No.: |
14/513009 |
Filed: |
October 13, 2014 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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12840018 |
Jul 20, 2010 |
|
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|
14513009 |
|
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61890180 |
Oct 12, 2013 |
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Current U.S.
Class: |
418/201.1 |
Current CPC
Class: |
F01C 21/02 20130101;
F04C 28/08 20130101; F01C 21/10 20130101; F25B 49/025 20130101;
Y02B 30/741 20130101; F25B 1/047 20130101; F04C 2240/52 20130101;
F04C 18/16 20130101; Y02B 30/70 20130101; F25B 2600/021
20130101 |
Class at
Publication: |
418/201.1 |
International
Class: |
F04C 18/16 20060101
F04C018/16; F01C 21/02 20060101 F01C021/02; F04C 28/08 20060101
F04C028/08 |
Claims
1. A bearing housing of a screw compressor, comprising: a cavity
defined by the bearing housing, wherein a depth of the cavity is
configured to be no larger than a length of a radial bearing of a
screw compressor.
2. The bearing housing of claim 1, wherein the bearing housing is
configured to support a radial bearing of the screw compressor.
3. The bearing housing of claim 1, wherein the bearing housing is
configured to be positioned to cover an axial end of the screw
compressor.
4. A screw compressor, comprising: a rotor including a shaft; one
or more radial bearings attached to the shaft; one or more axial
bearings attached to the shaft; and a bearing housing, wherein the
bearing housing defines a cavity that is configured to support the
one or more radial bearings but not the one or more axial
bearings.
5. The screw compressor of claim 4, wherein the cavity has a depth
and the one or more radial bearings have a length, and the depth of
the cavity is not larger than the length of the one or more radial
bearings.
6. The screw compressor of claim 4, further comprising: a bearing
cover, the bearing cover configured to be attached to the bearing
housing; wherein the bearing cover is configured to have a
clearance with the one or more axial bearings.
7. The screw compressor of claim 4, further comprising: a bearing
cover, the bearing cover defined a space; wherein the space defined
by the bearing cover is configured to enclose the one or more axial
bearings.
8. The screw compressor of claim 4, wherein the one or more radial
bearings are positioned closer to the rotor than the one or more
axial bearings.
9. A variable capacity screw compressor comprising: a rotor housing
comprising a suction port, a working chamber, a discharge port, and
at least two screw rotors that include a female screw rotor and a
male screw rotor being positioned within the working chamber for
cooperatively compressing a fluid; wherein the suction port, the at
least two screw rotors and the discharge port being configured in
relation to a selected rotational speed; the selected rotational
speed operates at least one screw rotor at an optimum peripheral
velocity that is independent of a peripheral velocity of the at
least one screw rotor at a synchronous motor rotational speed for a
rated screw compressor capacity; a motor operable to drive the at
least one screw rotor at a rotational speed at a full-load capacity
that is substantially greater than the synchronous motor rotational
speed at the rated screw compressor capacity; and a variable speed
drive to receive a command signal from a controller and to generate
a control signal that drives the motor at the rotational speed.
Description
FIELD
[0001] The disclosure herein relates to a rotary type compressor,
such as a rotary screw compressor, which can be used in, for
example, a heating, ventilation, and air-conditioning ("HVAC")
system. More specifically, the disclosure relates to a bearing
housing of a rotary screw compressor to support and enclose
discharge axial bearings. The bearing housing herein can improve
location of the rotors which may result in improved compressor
performance and reliability, and machining capability of the
housing may also be improved.
BACKGROUND
[0002] A screw compressor is a type of positive displacement
compressor that can be used to compress various working fluids,
such as for example refrigerant vapor. Such screw compressors may
be used in refrigeration units, such as for example, water chillers
as part of a HVAC system. The screw compressor typically includes
one or more rotors that rotate relative to bearings such as for
example radial and axial bearings at the discharge end. A bearing
housing and cover are often part of the assembly of the screw
compressor to enclose and support the bearings, e.g. radial and
axial bearings. During operation, the bearing cover can have a
discharge outlet or port so that a compressed working fluid (e.g.
refrigerant vapor) can be discharged from an axial end of the
rotors and out of the bearing housing and cover.
SUMMARY
[0003] According to an embodiment of the present invention, a
variable capacity screw compressor comprises a rotor housing, a
motor, and a variable speed drive. The rotor housing comprises a
suction port, a working chamber, a discharge port, and at least two
screw rotors that comprise a female screw rotor and a male screw
rotor being positioned within the working chamber for cooperatively
compressing a fluid. The suction port, the at least two screw
rotors and the discharge port are configured in relation to a
selected rotational speed. The selected rotational speed operates
at least one screw rotor at an optimum peripheral velocity that is
independent of a peripheral velocity of the at least one screw
rotor at a synchronous motor rotational speed for a rated screw
compressor capacity. A motor is operable to drive the at least one
screw rotor at a rotational speed at a full-load capacity that is
substantially greater than the synchronous motor rotational speed
at the rated screw compressor capacity. A variable speed drive
receives a command signal from a controller and generates a control
signal that drives the motor at the rotational speed.
[0004] In another embodiment, a method for sizing at least two
screw compressors is provided. The target capacity for each screw
compressor is selected. Each screw compressor has a different rated
capacity and further comprises a suction port, a working chamber, a
discharge port, and at least two screw rotors being positioned
within the working chamber for cooperatively compressing a fluid.
The rotational speed is selected to operate at least one screw
rotor in each screw compressor at an approximately constant optimum
peripheral velocity that is independent of the rated capacity of
each screw compressor. The suction port, the at least two screw
rotors and the discharge port are configured together with the
rotational speed for each screw compressor.
[0005] In another embodiment, a refrigeration chiller, having at
least one refrigeration circuit, comprises a variable capacity
screw compressor, condenser, expansion valve and evaporator. The
variable capacity compressor comprises a rotor housing, a motor
housing and a variable speed drive. The rotor housing further
comprises a suction port, a working chamber, a discharge port, and
at least two screw rotors that comprise a female screw rotor and a
male screw rotor being positioned within the working chamber for
cooperatively compressing a fluid. The suction port, the at least
two screw rotors and the discharge port are configured in relation
to a selected rotational speed. The selected rotational speed
provides at least one screw rotor to operate at an optimum
peripheral velocity that is independent of a peripheral velocity of
the at least one screw rotor at a synchronous motor rotational
speed for a rated screw compressor capacity. The motor housing
further comprises a motor, the motor is operable to drive the at
least one screw rotor at a rotational speed at a full-load capacity
that is substantially greater than the synchronous motor rotational
speed at the rated screw compressor capacity. The variable speed
drive is configured to receive a command signal from a controller
and to generate a control signal that drives the motor at the
rotational speed. A condenser is coupled to the discharge port of
the variable capacity screw compressor. The condenser is configured
to cool and condense fluid received from the discharge port. An
expansion valve is coupled to the condenser. The expansion valve is
configured to evaporate at least a portion of fluid received from
the condenser by lowering pressure of fluid received from the
condenser. An evaporator is coupled to the expansion valve. The
evaporator is configured to evaporate fluid received from the
expansion valve and to provide fluid to the suction port of the
variable capacity screw compressor.
[0006] An improved bearing housing of a rotary screw compressor is
described. A bearing housing is generally configured to suitably
enclose and support discharge radial bearings which are located at
a discharge side of the compressor, for example toward the axial
end of the rotors.
[0007] In previously known designs, the discharge bearing housing
of a screw compressor is constructed to be a relatively long part,
which encloses and/or supports the discharge radial bearings, the
axial bearings, and the bearing retaining assembly for example the
axial bearing retainers.
[0008] In the bearing housing shown and described herein, a shorter
bearing housing suitably encloses and/or supports the discharge
radial bearings, but does not enclose or support the axial bearings
and retaining assembly, e.g. the axial bearing retainers. A bearing
cover is provided which encloses the axial bearings and the axial
bearing retainers. The shorter bearing housing may be simpler to
fabricate and the accuracy of the discharge axial bearing bores may
be improved due to shorter reaches and shorter machine tool boring
bars, compared to a convention bearing housing. Due to the new
bearing housing design, the bearing cover may be fabricated
relatively easier using machine tools that may not be as precise as
those used to fabricate the bearing housing. That is, the design of
the bearing housing herein can improve for example the machining
capability of the housing, such as for example by enabling short
machine cutter tooling and short reaches for the machining center
of the housing. The shorter bearing housing can improve location of
the rotors such as during assembly, such as for example by
improving the accuracy of the discharge bearing bores, which may
result in improved compressor performance and reliability.
[0009] In one embodiment, a bearing assembly may include a bearing
cover and a bearing housing. The bearing housing includes a cavity
that is configured to enclose and/or support a discharge radial
bearing. The cavity has a depth. The discharge radial bearing has a
length. The depth of the cavity may be configured to be no more
than the length of the discharge radial bearing so that the cavity
can be configured to enclose and/or support the discharge radial
bearing, but not the axial bearing.
[0010] Other features and aspects of the embodiments will become
apparent by consideration of the following detailed description and
accompanying drawings.
BRIEF DESCRIPTION OF THE DRAWINGS
[0011] Reference is now made to the drawings in which like
reference numbers represent corresponding parts throughout.
[0012] FIG. 1 illustrates an embodiment that incorporates a screw
compressor arranged as part of a refrigeration chiller system.
[0013] FIG. 2 illustrates a cross sectional view of a screw
compressor according to one embodiment.
[0014] FIG. 3 illustrates an additional cross sectional view of a
screw compressor according to one embodiment.
[0015] FIG. 4 illustrates an embodiment of a refrigeration chiller
and controller system according to one embodiment.
[0016] FIG. 5 illustrates a partial sectional view of a screw
compressor, with which the embodiments as disclosed herein can be
practiced.
[0017] FIG. 6 illustrates an enlarged sectional view of an area A
of the screw compressor in FIG. 5.
[0018] FIG. 7 illustrates a conventional bearing housing.
DETAILED DESCRIPTION
[0019] As a preface to the detailed description, as used in this
specification and the appended claims, the singular forms "a,"
"an," and "the" also include plural referents, unless the context
clearly dictates otherwise. References in this specification to
"one embodiment," "an embodiment," "an example embodiment," etc.,
indicate that the described embodiment may include a particular
feature, structure, or characteristic; however, every embodiment
may not necessarily include the particular feature, structure, or
characteristic. When a particular feature, structure, or
characteristic is described in connection with an embodiment, other
embodiments may incorporate or otherwise implement such feature,
structure, or characteristic whether or not explicitly
described.
[0020] Referring now to FIGS. 1-4, components of a chiller or
chiller system 10 are illustrated. Chiller 10 includes many other
conventional features not depicted for simplicity of the
drawings.
[0021] Chiller system 10 is directed to refrigeration systems.
Chiller 10 is in the range of about 20 to 500 tons or larger,
particularly where the refrigeration system includes a multiple
stage compressor arrangement. Persons of ordinary skill in this art
will readily understand that embodiments and features of this
invention are contemplated to include and apply to, not only single
stage compressors/chillers, but also to (i) multiple stage
compressors/chillers and (ii) single and/or multistage
compressor/chillers operated in parallel.
[0022] As shown, chiller 10 comprises a screw compressor system 12
(also sometimes referred to as a screw compressor 12), a condenser
14, and an evaporator 20, all of which are serially connected to
form a semi- or fully-hermetic, closed-loop refrigeration system.
Chiller 10 may circulate a fluid 80 (such as, for example, a
refrigerant) to control the temperature in a space such as a room,
home, or building. The fluid 80 may be circulated to absorb and
remove heat from the space and may subsequently reject the heat
elsewhere.
[0023] Fluid 80 may be a refrigerant. The refrigerant may be
selected from an azeotrope, a zeotrope or a mixture or blend
thereof in gas, liquid or multiple phases. For example, such
refrigerants may be selected from: R-123, R-134a, R-1234yf, R-410A,
R-22 or R-32. Because embodiments of the present invention are not
restricted to the refrigerant chosen, embodiments of the present
invention are also adaptable to a wide variety of refrigerants that
are emerging, such as low global warming potential (low-GWP)
refrigerants.
[0024] FIG. 1 illustrates the condenser 14. Condenser 14 is shown
as a shell and tube flooded-type. The condenser 14 can be arranged
as a single evaporator or multiple evaporators in series or
parallel, e.g. connecting a separate or multiple evaporators to
each compressor. Condenser 14 may include condenser tubing 16.
Fluid 80 may pass across the condenser tubing 16 through which cool
air or cool liquid flows.
[0025] Condenser 14 may be fabricated from carbon steel and/or
other suitable material, including copper alloy heat transfer
tubing. Condenser tubing 16 can be of various diameters and
thicknesses, and comprised typically of copper alloy. In addition,
condenser tubing 16 may be replaceable, mechanically expanded into
tube sheets and externally finned seamless tubing. Other known
types of condenser 14 are contemplated.
[0026] Condenser 14 may be configured to communicate fluid 80 from
a discharge passage 36. Discharge passage 36 may be configured to
receive the fluid 80, or may be coupled to the condenser 14 through
an oil separator 24, as depicted in FIG. 1. Other configurations
are contemplated. The oil separator 24, when employed, separates
oil from the fluid 80 and returns the oil via an oil supply passage
26 to the screw compressor 12 for reuse. The oil may be reused to,
for example, cool the fluid 80, cool screw rotors 42, seal the
interfaces between the screw rotors 42 themselves, seal the
interfaces between the screw rotors 42 and the walls of a working
chamber 44, and/or lubricate bearings 46, 48.
[0027] Condenser 14 may transform the fluid 80 from a superheated
vapor to a saturated liquid. As a result of the cool air or cool
liquid passing across the condenser tubing 16, fluid 80 may reject
or otherwise deliver heat from the chiller 10 to another fluid,
like air or liquid, in a heat transfer relation, which in turn
carries the heat out of the system.
[0028] An expansion valve 18 may employed, as shown in FIG. 1.
Expansion valve 18 may be configured to receive fluid 80 from
condenser 14. Fluid 80 received from condenser 14 typically is in a
thermodynamic state known as a saturated liquid. The expansion
valve 18 may abruptly reduce the pressure of the fluid 80. The
abrupt pressure reduction may cause adiabatic flash evaporation of
at least a portion of the fluid 80. In particular, the adiabatic
flash evaporation may result in a liquid and vapor mixture of the
fluid 80 that has a temperature that is colder than the temperature
of the space to be cooled.
[0029] Evaporator 20 is shown in FIG. 1 as a shell and tube
flooded-type. The evaporator 20 can be arranged as a single
evaporator or multiple evaporators in series or parallel, e.g.
connecting a separate or multiple evaporators to each compressor.
Evaporator 20 may include evaporator tubing 22. Fluid 80 may pass
across the evaporator tubing 22 through which cool air or cool
liquid flows.
[0030] Evaporator 20 may be fabricated from carbon steel and/or
other suitable material, including copper alloy heat transfer
tubing. Evaporator tubing 22 can be of various diameters and
thicknesses, and comprised typically of copper alloy. In addition,
evaporator tubing 22 may be replaceable, mechanically expanded into
tube sheets and externally finned seamless tubing. Other known
types of evaporator 20 are contemplated.
[0031] Evaporator 20 is configured, as illustrated in FIG. 1, to
receive fluid 80 communicated from the expansion valve 18. Fluid 80
received by the evaporator 20 in the refrigeration loop may be
relatively colder than it was when discharged from the screw
compressor 12. The oil return apparatus 28, when employed,
separates oil from the fluid 80 and returns the oil via an oil
return passage 30 to the screw compressor 12 for reuse. The oil may
be reused to, for example, cool the fluid 80, cool screw rotors 42,
seal the interfaces between the screw rotors 42 themselves, seal
the interfaces between the screw rotors 42 and the walls of a
working chamber 44, and/or lubricate the bearings 46, 48.
[0032] The evaporator 20 may absorb and remove heat from the space
to be cooled, and the condenser 14 may subsequently reject the
absorbed heat to air or liquid that carries the heat away from the
space to be cooled. In operation, warm air or liquid may be
circulated from the space to be cooled across the evaporator tubing
22. The warm air or liquid passing across the evaporator tubing 22
may cause a liquid portion of the cold fluid 80 to evaporate. At
the same time, the warm air or liquid passed across the evaporator
tubing 22 may be cooled by the fluid 80. It should be understood
that any configuration of the condenser 14 and/or evaporator 20 may
be employed that accomplishes the necessary phase changes of fluid
80.
[0033] The chilled or heated water is pumped from the evaporator 20
to an air handling unit (not shown). Air from the space that is
being temperature conditioned is drawn across coils in the air
handling unit that contains, in the case of air conditioning,
chilled water. The drawn-in air is cooled. The cool air is then
forced through the air conditioned space, which cools the
space.
[0034] Additionally, though not shown, an economizer 32 may be
incorporated to include an economizer cycle. Economizer 32 or a
subcooling cycle (not shown), or both, may be employed in the
refrigeration cycle and return the fluid 80 to the screw compressor
12 via suction passage 34 or other passage (not shown) depending on
the configuration required the application conditions.
[0035] Referring to FIGS. 2 and 3, screw compressor 12 typically
comprises a rotor housing 40 and an electric motor housing 50.
Screw compressor 12 may be formed, all or in part, of gray cast
iron, for example. Other materials may be used to form the screw
compressor 12. Screw compressor 12, according to embodiments of the
present invention, facilitates highly efficient operation at
full-load and part-load conditions over a preselected screw
capacity range.
[0036] Motor housing 50 houses a motor 52 in an embodiment of the
present invention. Electric motor 52 may coupled to a variable
frequency drive 38. The electric motor 52 drives meshed screw
rotors 42. Motor housing 50 may be integral to the rotor housing
40.
[0037] The rotor housing 40 may have a low pressure end and a high
pressure end that each contain a suction port 76 and discharge port
78, respectively. Suction port 76 and discharge port 78 are in
open-flow communication with the working chamber 44. The suction
port 76 and the discharge port 78 may each be an axial, a radial or
a mixed (a combination of a radial and an axial) port.
[0038] The suction port 76 may receive the fluid 80 at a suction
pressure and a suction temperature. The suction port 76 may receive
fluid 80 from suction passage 34 in thermodynamic states known as a
saturated vapor or a superheated vapor. The screw compressor 12 may
compress the fluid 80 as the screw compressor 12 communicates the
fluid 80 from the suction port 76 to the discharge port 78. Fluid
80 passing through the discharge port 78 discharges into discharge
passage 36.
[0039] Compressing the fluid 80 may also result in the fluid 80
being discharged at a discharge temperature that is higher than the
suction temperature. The fluid 80 discharged from the discharge
port 78 may be in a thermodynamic state known as a superheated
vapor. Accordingly, fluid 80 discharged from the screw compressor
12 may be at a temperature and a pressure at which the fluid 80 may
be readily condensed with a cooling air or a cooling liquid.
[0040] Suction port 76 and discharge port 78 are configured to
minimize flow losses, when at least one of the rotors 42 is
operated at an approximately constant peripheral velocity. The
suction port 76 may be located where fluid 80 exits the suction
area of screw compressor 12 and is drawn into the working chamber
44. The suction port 76 may be sized to be as large as possible to
minimize, at least, the approach velocity of the fluid 80. The
location of the suction port 76 in the rotor housing 40 also may be
configured to minimize turbulence of fluid 80 prior to entry into
the rotors 42.
[0041] Discharge port 78 may be sized larger than theoretically
necessary to provide a thermodynamic optimum size and thereby,
reduce the velocity at which the fluid 80 exits the working chamber
44. The discharge port 78 may be generally located where fluid 80
exits the working chamber 44 of screw compressor 12. The discharge
port 78 location in the rotor housing 40 may be configured such
that the maximum discharge pressure can be attained in the rotors
42 prior to being delivered into the discharge passage 36. In
addition, screw compressor 12 may incorporate a muffler 58 or other
apparatus suitable for noise reduction.
[0042] Referring again to FIG. 3, rotors 42 are mounted for
rotation in a working chamber 44. The working chamber 44 comprises
a volume that is shaped as a pair of parallel, intersecting
flat-ended cylinders, and is closely toleranced to the exterior
dimensions and geometry of the intermeshed screw rotors 42. The
plurality of meshed screw rotors 42a, 42b may define one or more
compression pockets between the screw rotors 42a, 42b and the
interior chamber walls of the rotor housing 40. The rotor housing
40 has little separation from the rotors 42. Milling, machine
grinding or molding can be employed to achieve high accuracy and
tight tolerances between rotors 42 flutes and lobes and the rotor
housing 40.
[0043] First screw rotor 42a and second screw rotor 42b are
disposed in a counter-rotating, intermeshed relationship and
cooperate to compress a fluid. At least one of rotors 42 is
cooperatively configured with motor 52 to be operable at a
rotational speed for a screw compressor capacity within a
preselected screw compressor capacity range. The selected
rotational speed at full-load capacity is substantially greater
than a synchronous motor rotational speed at a rated capacity (also
referred to herein as rated screw compressor capacity) for screw
compressor 12.
[0044] In the embodiment illustrated, rotor 42a may be called a
female screw rotor and comprise a female lobed/fluted body or
working portion (typically a helical or spiral extending land and
groove). Rotor 42b may be called a male screw rotor and comprise a
male lobed/fluted body or working portion (typically a helical or
spiral extending land and groove).
[0045] Rotors 42 include shaft portions, which are, in turn,
mounted to the housing of screw compressor 12 by, for example, one
or more bearings 46, 48. The exemplary bearings 46, 48 will also be
configured with tight clearances in relation to at least rotors 42
and rotor housing 40.
[0046] Compression of the fluid 80 in screw compressor 12 produces
axial and radial forces. The configurations of embodiments of the
present invention may also mitigate time varying and non-uniform
rotor movements and forces against chamber walls, bearings, and end
surfaces of the screw compressor 12 caused by the interaction of
the screw rotors 42a, 42b, the axial forces, and the radial
forces.
[0047] As mentioned, a lubricating fluid, typically oil, may be
delivered from oil supply passage 26 or oil return passage 30 to
the screw compressor 12. The lubricating fluid provides cushioning
films for the walls of the working chamber 44, rotors 42a, 42b, and
bearings 46, 48 of the screw compressor 12, but does little to
prevent the transmission of the time varying and non-uniform axial
and radial forces. The screw compressor 12 may also utilize an
expander (not shown), which may also be integral to screw
compressor 12, to recover energy available from the refrigeration
cycle as the high pressure liquid expands through the expander to a
lower pressure.
[0048] The electric motor 52 in one exemplary embodiment may drive
at least one of the rotors 42 in response to command signals 62
received from the controller 60. The horsepower of preferred motor
52 can vary in the range of about 125 horsepower to about 2500
horsepower. Torque supplied by the electric motor 52 may directly
rotate at least one of the screw rotors 42. Employing motor 52 and
variable speed drive 38, screw compressor 12 of embodiments of the
present invention may have a rated screw compressor capacity within
the range of about 35-tons to about 150-tons or more and have a
full-load speed range within about 4,000 revolutions per minute to
about 15,000 revolutions per minute, when the fluid is an R-134a
refrigerant.
[0049] While conventional types of motors, like induction motors,
can be used with and will provide a benefit when employed with
embodiments of the present invention, a preferred motor 52
comprises a direct drive, variable speed, hermetic, permanent
magnet motor. Permanent magnet motor 52 can increase system
efficiencies over other motor types. The choice of motor 52 may be
affected by cost and performance considerations.
[0050] Referring to FIGS. 2 and 3, the permanent magnet motor 52
comprises a motor stator 54 and a motor rotor 56. Stator 54
consists of wire coils formed around laminated steel poles, which
convert variable speed drive 38 applied currents into a rotating
magnetic field. The stator 54 is mounted in a fixed position in the
screw compressor 12 and surrounds the motor rotor 56, enveloping
the rotor 56 with the rotating magnetic field. Motor rotor 56 is
the rotating component of the motor 52 and may consist of a steel
structure with permanent magnets, which provides a magnetic field
that interacts with the rotating stator magnetic field to produce
rotor torque. In addition, permanent magnet motor 52 may be
configured to receive variable frequency control signals and to
drive the at least two screw rotors per the received variable
frequency control signals.
[0051] The motor rotor 56 may have a plurality of magnets and may
comprise magnets buried within the rotor steel structure or be
mounted at the rotor steel structure surface. Motor rotor 56
surface mount magnets are secured with a low loss filament, metal
retaining sleeve or by other means to the rotor steel support.
Further manufacturing, performance, and operating advantages and
disadvantages can be realized with the number and placement of
permanent magnets in the motor rotor 56. For example, surface
mounted magnets can be used to realize greater motor efficiencies
due to the absence of magnetic losses in intervening material, ease
of manufacture in the creation of precise magnetic fields, and
effective use of rotor fields to produce responsive rotor torque.
Likewise, buried magnets can be used to realize a simpler
manufactured assembly and to control the starting and operating
rotor torque reactions to load variations.
[0052] The performance and size of the permanent magnet motor 52 is
due in part to the use of high energy density permanent magnets.
Permanent magnets produced using high energy density magnetic
materials, typically at least 20 MGOe (Mega Gauss Oersted), produce
a strong, more intense magnetic field than conventional materials.
With a motor rotor 56 that has a stronger magnetic field, greater
torques can be produced, and the resulting motor 52 can produce a
greater horsepower output per unit volume than a conventional
motor, including induction motors. By way of comparison, the torque
per unit volume of permanent magnet motor 52 can be at least about
75 percent higher than the torque per unit volume of induction
motors used in refrigeration chillers of comparable refrigeration
capacity. The result is a smaller sized motor to meet the required
horsepower for a specific compressor assembly.
[0053] The permanent magnet motor 52 of an embodiment of the
present invention is compact, efficient, reliable, and relatively
quieter than conventional motors. As the physical size of the screw
compressor 12 is reduced, motor 52 used can be scaled in size to
fully realize the benefits of improved fluid flow paths and
compressor element shape and size. Motor 52 is reduced in volume by
approximately 30 percent or more, when compared to conventional
existing designs for compressor assemblies that employ induction
motors and have refrigeration capacities in excess of 35-tons. The
resulting size reduction of embodiments of the present invention
provides a greater opportunity for efficiency, reliability, and
quiet operation through use of less material and smaller dimensions
than has been achieved through more conventional practices.
[0054] Any bearings employed with motor 52 may be rolling element
bearings (REB) or hydrodynamic journal bearings. Such bearings may
be oil lubricated. Oil-free bearing systems may be employed. A
special class of bearing which is refrigerant lubricated is a foil
bearing and another bearing type uses REB with ceramic balls. Each
bearing type has advantages and disadvantages that should be
apparent to those of skill in the art. Bearings should be selected
to facilitate highly efficient operation of the screw compressor 12
at reduced speeds for capacity modulation and to minimize rotor
dynamics and vibration associated with reduced speeds. Any bearing
type may be employed that is suitable of sustaining rotational
speeds in the range of about 2,000 RPM to about 20,000 RPM.
[0055] The motor rotor 56 and motor stator 54 end turn losses for
the permanent magnet motor 52 are very low compared to some
conventional motors, including induction motors. The motor 52,
therefore, may be cooled by means of fluid 80 (typically,
refrigerant). When fluid 80 is employed for cooling motor 52, fluid
80 may only need to contact the outside diameter of the stator 54.
Cooling the motor 52 in this way allows for the elimination of the
motor cooling feed ring that is typically used in induction motor
stators. Alternatively, refrigerant may be metered to the outside
surface of the stator 54 and to the end turns of the stator 54 to
cool the motor 52.
[0056] In addition, the torque that is needed from motor 52 comes
essentially from the internal pressure distribution in the rotors
42, which is a function of rotors 42 geometry and the operating
conditions. That internal pressure distribution within the rotors
42 provides the load against which the motor 52 has to work.
Employing embodiments of this invention without a mechanical
unloader results in a theoretical torque that may be essentially
constant over a full range of operating conditions, and for a given
operating condition, a ratio of theoretical to actual torque on the
motor 52 that may be approximately constant, despite decay in the
actual torque during operation due to changing losses and leakage,
for example. In contrast, for a given operating condition,
conventional screw compressors invoking a mechanical unloader will
have significant torque fluctuations or variations over time.
[0057] As illustrated in FIG. 4, a variable speed drive 38 may
drive the motor 52 and in turn, screw compressor 12. The speed of
the motor 52 can be controlled by varying, for example, the
frequency of the electric power that is supplied to the motor 52.
Use of a permanent magnet motor 52 and variable speed drive 38
moves some conventional motor losses outside of the refrigerant
loop. The efficiency of the variable speed drive 38, line input to
motor shaft output, preferably can achieve a minimum of about 95
percent over the system operating range.
[0058] The variable speed drive 38 drives the screw compressor 12
at the optimum, or near optimum, rotational speed at each capacity
over the preselected screw compressor capacity range for a screw
compressor 12 of a given rated capacity. The variable speed drive
38 may be refrigerant cooled, water cooled or air cooled. As
mentioned, similar to cooling of motor 52, the variable speed drive
38, or portions thereof, may be by using a refrigerant circulated
within the chiller system 10 or by other conventional cooling
means. How the motor 52 and/or variable speed drive 38 are cooled
should be understood as dependent on the operational and
environmental conditions in which the motor 52 and/or variable
speed drive 38 reside in operation.
[0059] The variable speed drive 38 typically will comprise an
electrical power converter comprising a line rectifier and line
electrical current harmonic reducer, power circuits and control
circuits (such circuits further comprising all communication and
control logic, including electronic power switching circuits).
Conditions in which the screw compressor 12 is employed may justify
employing more than one variable speed drive 38 for chiller 10.
[0060] The variable speed drive 38 can be configured to receive
command signals 62 from a controller 60 and to generate a control
signal 64. The variable speed drive 38 will respond, for example,
to signals 62 received from a microprocessor (also not shown)
associated with controller system 60 to increase or decrease the
speed of the motor 52 by changing the frequency of the current
supplied to motor 52. Controller 60 may be configured to receive
status signals 82 indicative of an operating point of the screw
compressor, and to generate command signals that requests the
electric motor system to drive the screw compressor per a
preselected operating parameter. Status signals 82 may deliver
similar or different status information depending, for example, on
the intended purpose of the sensor selected. Controller 60 may
generate command signals 62 per a preselected operating parameter,
like a torque profile for screw compressor 12. Control signal 64
can drive the high energy density motor 52 at a rotational speed
substantially greater than a synchronous motor rotational speed for
the rated screw compressor capacity and drive the motor 52, and in
turn at least one screw rotor 42, at an optimum peripheral velocity
independent of the rated screw compressor capacity.
[0061] The motor 52 and the variable speed drive 38 have power
electronics for low voltage (less than about 600 volts), 50 Hz and
60 Hz applications. Typically, an AC power source (not shown) will
supply multiphase voltage and frequency to the variable speed drive
38. The AC voltage or line voltage delivered to the variable speed
drive 38 will typically have nominal values of 200V, 230V, 380V,
415V, 480V, or 600V at a line frequency of 50 Hz or 60 Hz depending
on the AC power source.
[0062] By the use of motor 52 and variable speed drive 38, the
speed of motor 52 can be varied to match varying system
requirements. Speed matching results in approximately 30 percent
more efficient system operation compared to a compressor without a
variable speed drive 38. By running compressor 12 at lower speeds
when the load on the chiller is not high or at its maximum,
sufficient refrigeration effect can be provided to cool the reduced
heat load in a manner which saves energy, makes the chiller 10 more
economical from a cost-to-run standpoint, and facilitates highly
efficient chiller 10 operation as compared to chillers which are
incapable of such load matching at the rotational speeds possible
via embodiments of the present invention. For example, a rated
screw compressor capacity of about 100-tons configured according to
embodiments of the present invention could be efficiently operable
over a preselected screw capacity range of about 75-tons to about
125-tons.
[0063] Screw compressor 12 can be operated at rotational speeds
substantially higher than synchronous motor rotational speeds for a
given rated capacity of the screw compressor 12. The specific
optimum speed for the rated screw compressor capacity range is a
function of screw compressor capacity and head pressure.
Embodiments of the present invention dramatically improve the
discharge porting of fluid 80 and in turn, allow for screw
compressor 12 to be operated at a significantly increased
rotational speed over the rotational speed that gives the best
performance for conventionally sized rotors and ports. For example,
the selected rotational speed for a rated screw compressor capacity
of about 100-tons, according to embodiments of the present
invention, is about 5800 revolutions per minute, when the fluid is
an R-134a refrigerant. In contrast, a conventional screw compressor
with a rated capacity of about 100-tons has a synchronous motor
rotational speed is about 3400 revolutions per minute, when the
fluid is an R-134a refrigerant.
[0064] The allowable range of rotational speed for a particular
rated capacity of a screw compressor 12 is selected to achieve an
optimum peripheral velocity of at least one of the screw rotors
independent of the rated capacity of screw compressor 12 that
results in a relatively uniform high efficiency across the screw
compressor product family (e.g. 60-tons, 80-tons, 100-tons and
150-tons.) The optimum peripheral velocity is a constant product of
the rotational speed and the radius of at least one of the rotors
42, typically, the male rotor 42b. The approximately constant
optimum peripheral velocity is, for example, in the range between
about 131 feet per second (about 40 meters per second) to about 164
feet per second (about 50 meters per second). In one embodiment,
the approximately constant optimum peripheral velocity is between
about 42 meters per second (about 137 feet per second) to about 45
meters per second (about 147 feet per second) in high pressure
applications, when R-134a refrigerant is the fluid 80. Persons of
skill in the art would understand that, for a low pressure
application or for a different primary fluid 80, or both, the
optimum peripheral velocity may be different.
[0065] The rotational speed of the motor 52 may be selected in
combination with configuring rotors 42, suction port 76 and
discharge port 78 for each target capacity to achieve an
approximately constant optimum peripheral velocity of at least one
of the screw rotors 42 regardless of the rated capacity of the
screw compressor 12. That is, specific combinations of screw rotors
42, inlet port 76, discharge port 78 and the operational rotational
speed are selected such that each specific combination enables each
screw compressor 12 to run at approximately the same optimum
peripheral velocity for each different rated capacity and, in turn,
to produce relatively the same high efficiency between or among
each different rated capacity of screw compressor 12.
[0066] Embodiments of the present invention include a method of
sizing of at least two screw compressors 12 with different rated
capacities that achieve approximately constant efficiency across
the screw compressor product family (e.g. 60-tons, 80-tons,
100-tons and 150-tons.). By employing embodiments of this
invention, the isoentropic efficiency versus capacity (in tons) of
screw compressor 12 is significantly increased, on the order of 15
percent, over a conventional screw compressor. In addition, because
screw compressor 12 is operated at relatively higher speed, the
screw compressor 12 can slowed down on the order of 20-30 percent
of the speed for the operating capacity and still have an
approximately constant peak efficiency or efficiency plateau as
compared to the efficiency at the rated screw compressor
capacity.
[0067] The target capacity for each screw compressor 12, each
having a different rated capacity, is selected. The rotational
speed is also selected based on the target capacity of each screw
compressor 12 to operate at least one screw rotor 42 in each screw
compressor 12 at an approximately constant optimum peripheral
velocity that is independent of the rated capacity of each screw
compressor 12. The suction port 76, the at least two screw rotors
42 and the discharge port 78 are configured together with the
rotational speed selected for each screw compressor 12.
[0068] Specifically, driving screw compressor 12 at an optimum
peripheral velocity allows for each rotor 42 to have a geometry and
a profile that may remain the same for a wide range of preselected
screw compressor capacities for the rated screw compressor
capacity. Each of the rotors 42, though, may have a different
geometry and a profile for each different rated screw compressor
capacity that will enable at least one screw rotor to be operated
at a selected rotational speed that produces an approximately
constant optimum peripheral velocity between or among each rated
capacity of each screw compressor 12. The volumetric ratio of the
screw compressor 12 is selected as a function of the loading
conditions in which the screw compressor 12 will be used. By way of
example, in embodiments of the present invention, more than two
volumetric ratios, potentially four, five or more, are contemplated
over a range of rated screw compressor capacities. The volumetric
ratio may also be such that the system compression ratio and the
internal compression ratio closely match. The rotor 42 profile may
be a balance of the length of the sealing line, flow cross
sectional area and blow-hole area size.
[0069] The geometry and profile are generally defined, in part, by
the number of lobes in each rotor, the wrap angle, the length of
the rotors and the diameter of the rotors, for example. Screw rotor
42 has a profile taken in a plane transverse to the parallel axes
of the male rotor 42b and the female rotor 42a. The profile of
rotors 42 can be symmetric or asymmetric, and circular, elliptical,
parabolic, hyperbolic, for example. Rack generation of rotors 42
profile may be employed. Selecting a profile of rotors 42 is a
balance of the internal leakage path of fluid 80 during operation
of screw compressor 12 and the porting configuration of suction
port 76 and discharge port 78, such that screw compressor 12 has an
approximately constant optimum peripheral velocity.
[0070] More specifically, for example, at an about 44 m/s optimum
peripheral velocity for at least one rotor 42 of a 100-ton screw
compressor, the resulting male rotor 42b has a wrap angle of about
347 degrees and the female rotor 42a has a wrap angle that is
6/7ths of the male rotor 42b. The wrap angle of the female rotor
42a varies with the ratio of number of lobes. The female rotor 42a
has a radius of about 2.5 inches (6.35 centimeters) and 7 lobes and
the male rotor 42b has a radius of about 3 inches (7.62
centimeters) and 6 lobes. The length of rotors 42 is significantly
smaller, on the order of about 20-30 percent smaller, than a
conventionally sized screw compressor at the rated screw compressor
capacity. A person of skill in the art will appreciate that
analytical techniques can be employed for other combinations of
rotor 42 profiles for a given rated screw compressor capacity
within the scope of the present invention.
[0071] Employing a geometry/profile of rotors 42 for a screw
compressor 12 having a preselected screw compressor range and
operable at an approximately constant optimum peripheral velocity,
allows for operation of the screw compressor 12 at 25 or more
percent less than the rated screw compressor capacity without
significant adverse rotor dynamic effects. Screw compressor 12 has
an improved rotor profile that maximizes internal flow area,
internal friction due to relative motion of the rotor 42 surfaces
is minimized, and leakage paths are reduced. This reduced leakage
and higher flow tend to increase the screw compressor 12 efficiency
and reduce power wasted, which increases overall efficiency.
[0072] Referring now to FIG. 4, further details regarding an
embodiment of the chiller 10 are presented. In particular, chiller
10 may include a controller or controller system 60. Controller 60
may be arranged to communicate with the variable frequency drive
38, screw compressor 12, condenser 14 and evaporator 20. Chiller 10
may further include one or more sensors. Sensors 66, 68, 70, 72 and
74, for example, may be employed to sense and/or communicate
torque, suction pressure and/or temperature, discharge pressure
and/or temperature, and/or other measurable parameter. Other
sensors could be employed depending on the application in which
screw compressor 12 is used. Signals 82 may be communicated via
wiring, fiber optics, wireless and/or a combination of wiring,
fiber optics and wireless. The sensors 66, 68, 70, 72 and 74
communicate status signals 82 to controller 60 with data that are
indicative of the operation of various components of the chiller
10.
[0073] The controller 60 may include processors, microcontrollers,
analog circuitry, digital circuitry, firmware, and/or software (not
shown) that cooperate to ultimately control operation of the screw
compressor 12. The memory may comprise non-volatile memory devices
such as flash memory devices, read only memory (ROM) devices,
electrically erasable/programmable ROM devices, and/or battery
backed random access memory (RAM) devices to store an array of
performance related characteristics for the screw compressor 12.
The memory may further include instructions which the controller 60
may execute in order to control the operation of the screw
compressor 12.
[0074] The controller 60 may receive status signals from one or
more sensors 66, 68, 70, 72 and 74 that provide information
regarding operation of the screw compressor 12. Based upon the
status signals, the controller 60 may determine an operating mode
and/or operating point of the screw compressor 12 and may generate,
based upon the determined operating mode and/or operating point,
one or more command signals 62 to adjust the operation of the screw
compressor 12. The controller 60 may then generate command signals
62 that request the motor 52 to operate according to a preselected
operating parameter(s) (e.g. a torque profile). For example, the
controller 60 may enable operation at an optimal torque and speed
of screw compressor 12 to minimize losses, mechanical wear and
losses. Further disclosure of a controller system 60 suitable for
use with embodiments of the present invention may be found in
co-pending application U.S. patent Ser. No. 12/544,582, assigned to
the assignee of the instant application, which is hereby
incorporated by reference.
[0075] It should be apparent that variations on the control system
60 described above will be apparent to those skilled in the art.
The control system 60 may be implemented with electronic digital,
analog, or a combination of digital/analog control elements and
low-voltage wiring. Other conventional pneumatic tubing,
transmitters, controllers, and relays are contemplated.
[0076] In addition, it also will be readily apparent to one of
ordinary skill in the art that the compressor system disclosed can
be readily implemented in other contexts at varying scales. Use of
various motor types, drive mechanisms, and configurations with
embodiments of this invention should be readily apparent to those
of ordinary skill in the art.
[0077] Employing embodiments of the present invention, as compared
to conventional approaches, increase full-load efficiency, yield
higher part-load efficiency and have a practically constant
efficiency over a given capacity range, controlled independently of
power supply frequency or voltage changes. Also, an advantage of
embodiments of the present invention is that screw compressors 12
of different rated capacity can each have a variable capacity and
still have the approximately same the level of efficiency and
without mechanical unloading.
[0078] Additional advantages include a reduction in the physical
size of the screw compressor and chiller system arrangement,
improved scalability of the screw compressors throughout the
operating range and a reduction in total sound levels. Employing
embodiments of screw compressor 12 can also effectively reduce
costs for the manufacturer, because it allows for one screw
compressor at a rated screw compressor capacity (e.g. 100-tons) to
serve as an efficient screw compressor at a range of preselected
screw compressor capacity range (e.g. 80 tons and 125 tons) without
the need for multiple other screw compressors to be manufactured at
each additional target capacity within the preselected screw
compressor rated capacity range. Practically, embodiments of the
present invention also allow for lower physical part count and
inventory for a product family with no loss in capacity or
performance due to power supply because, for a given rated capacity
of screw compressor (e.g. 100-tons), the screw compressor 12 at 50
Hertz and 60 Hertz are nearly identical.
[0079] Embodiments of a bearing housing of a screw compressor are
further described. A screw compressor of a HVAC system may include
one or more rotors. The one or more rotors of the screw compressor
can be typically supported by bearings, such as for example axial
and/or radial bearings. In some screw compressors, the bearings
supporting the rotors can be enclosed and/or supported by a bearing
housing.
[0080] FIG. 5 illustrates a partial side sectional view of a screw
compressor 100 that includes a first rotor 110 and a second rotor
120. The relative motions of the first rotor 110 and the second
rotor 120 can be used to compress a working fluid, such as for
example refrigerant vapor.
[0081] The first and second rotors 110 and 120 are positioned
inside a rotor housing 130. A bearing assembly including a bearing
cover 170 and a bearing housing 160 that is positioned at an axial
end of the rotor housing 130 along an axial direction defined by an
axis C of the first rotor 110. The bearing assembly generally
covers the rotor housing 130 at the axial end. The compressed
working fluid is typically discharged through the bearing
assembly.
[0082] The first rotor 110 and the second rotor 120 include an
axially extended first shaft 112 and a second shaft 122
respectively along the axial direction defined by the axis C of the
first rotor 110. The first shaft 112 and the second shaft 122 can
be supported by one or more bearings, such as for example,
discharge radial bearings 140a and 140b and/or axial bearings 150a
and 150b (sometimes referred to as "thrust bearings") respectively.
The discharge radial bearings 140a and 140b and the axial bearings
150a and 150b can be attached to the first and second shafts 112
and 122. The axial bearing 150a and 150b, as well as the discharge
radial bearings 140a, 140b can be retained in the axial direction
by axial bearing retainers 142.
[0083] The bearings 140a, 140b, 150a and 150b can help support the
first rotor 110 and the second rotor 120 during operation, and
facilitate the rotation of the first and second rotors 110 and 120
around the axis C. For example, the discharge radial bearings 140a
and 140b and the axial bearing 150a and 150b can help withstand
forces that may be produced during operation.
[0084] The bearing housing 160 is positioned at the axial end of
the rotor housing 130. The bearing housing 160 can be configured to
enclose and/or support the discharge radial bearings 140a and 140b.
The discharge radial bearings 140a and 140b are typically
positioned closer to the rotor housing 130 than the axial bearings
150a and 150b in the axial direction.
[0085] The bearing housing 160 includes a first cavity 162 and a
second cavity 164 configured to enclose and/or support the
discharge radial bearings 140a and 140b respectively. The bearing
housing 160 can be a slab-like structure, which may help provide a
relatively uniform thermol expansion. The discharge radial bearings
140a and 140b can be supported by walls of the first cavity 162 and
the second cavity 164 respectively, and may form interference fit
with the walls of the cavities 162 and 164 respectively so that the
bearing housing 160 can support the discharge radial bearings 140a
and 140b during operation, such as for example, help the discharge
radial bearings 140a and 140b withstand forces that may be produced
during operation.
[0086] The screw compressor 100 also includes the bearing cover
170, which can be attached to the bearing housing 160 to form an
enclosed space 180. As illustrated in FIG. 5, the axial bearings
150a and 150b, portions of the first and second shafts 112 and 122,
and the axial bearing retainers 142 can be enclosed inside the
space 180.
[0087] FIG. 6 illustrates an enlarged view of an area A of FIG. 5.
The slab-like bearing housing 160 includes the first cavity 162 and
the second cavity 164 to enclose the first discharge radial bearing
140a and the second discharge radial bearing 140b respectively.
[0088] The first discharge radial bearing 140a has a first length
L1 and the second discharge radial bearing 140b has a second length
L2 in the axial direction. The first cavity 162 has a first depth
D1 and the second cavity 164 has a second depth D2 in the axial
direction. The first depth D1 can be configured to be about the
same as the first length L1. In some embodiments, the first depth
D1 can be configured to be no more than the first length L1. The
second depth D2 can be configured to be about the same as the
second length L2. In some embodiments, the second depth D2 can be
configured to be no more than the second length L2. The first depth
D1 and the second depth D2 generally are configured so that the
first cavity 162 and the second cavity 164 can enclose and/or
support the first discharge radial bearing 140a and the second
discharge radial bearing 140b respectively. In some embodiments,
the first cavity 162 and the second cavity 164 typically are
configured not to enclose and/or support the first and second axial
bearings 150a and 150b.
[0089] In some embodiments, the first discharge radial bearing 140a
can be fully supported by the first cavity 162. In some
embodiments, the second discharge radial bearing 140b can be fully
supported by the second cavity 164. The term "fully supported" is
referred to a situation where the support (e.g. interference fit)
provided by the walls of the first cavity 162 or the second cavity
164 to the first discharge radial bearing 140a or the second
discharge radial bearing 140b respectively will stop increasing as
the first depth D1 or the second depth D2 increases.
[0090] The bearing housing 160 has a depth W1 in the axial
direction. The depth W1 can have a minimal depth, which may be
configured for example to withstand a compression pressure produced
during operation.
[0091] In some embodiments, it can be beneficial to limit or
minimize the depth W1, so that a mass of the bearing housing 160
can be limited or minimized. This can help limit or minimize, for
example, thermal expansion of the bearing housing 160, which may
help maintain a clearance between the bearing housing 160 and the
first rotor 110 or the second rotor 120.
[0092] The depth W1 can be affected by, for example, the minimal
depth as well as other factors such as the depth D1 and/or the
depth D2. Generally, the larger the depth D1 and/or the depth D2
is, the larger the depth W1. In some embodiments, the depth D1
and/or the depth D2 may be configured so that the first cavity 112
and/or the second cavity 122 can be configured to fully support the
first radial bearing 140a and/or the second radial bearing 140b. In
some embodiments, the depth D1 and/or the depth D2 may be
configured so that the bearing housing 160 does not extend to the
axial bearings 140a and/or 140b in the axial direction. In some
embodiments, the depth D1 and/or the depth D2 may be configured to
be no more than the length L1 and the length L2 respectively. In
some embodiments, the depth D1 and/or the depth D2 may be
configured so that the first cavity 112 and/or the second cavity
122 may be no more than a depth that is required to fully support
the first radial bearing 140a and/or the second radial bearing
140b. In some embodiments, the difference between the depth D1
and/or the depth D2 and the length L1 and/or L2 may be about 3 mm,
so that a maximum of about 3 mm of the first and/or second
discharge radial bearings 140a and 140b is not enclosed and/or
unsupported by the first and/or the second cavities 162 and
164.
[0093] Referring to FIGS. 1 and 2, the bearing cover 170 generally
has a dome shape, which can help define the space 180 with the
bearing housing 160. The first and second axial bearings 150a and
150b can be enclosed in the space 180. Referring to FIG. 6, the
first axial bearing 150a (and/or the second axial bearing 150b,
which is not illustrated in FIG. 5) can have a clearance 182
between the first axial bearing 150a (and/or the second axial
bearing 150b) and the bearing cover 170. The first axial bearing
150a and/or the second axial bearing 150b does not generally need a
support from the bearing housing 160 and/or the bearing cover 170
because the first axial bearing 150a and/or the second axial
bearing 150b typically are configured to withstand forces in the
axial direction.
[0094] It is to be appreciated that because the bearing cover 170
can have the clearance 182 with the first axial bearing 150a
(and/or the second axial bearing 150b), therefore the bearing cover
170 does not have to be precision machined.
[0095] In the illustrated embodiment, the first axial bearing 150a
and/or the second axial bearing 150b are arranged next to the first
discharge radial bearing 140a and/or the second discharge radial
bearing 140b in the axial direction. The first axial bearing 150a
and the second axial bearing 150b are positioned further away from
the first rotor 110 and/or the second rotor 120 relative to the
first and second discharge radial bearings 140a and 140b
respectively in the axial direction. Because the axial bearing 150a
and/or the second axial bearing 150b do not necessarily require a
support provided by the bearing housing 160 or the bearings cover
170, it is not necessary to enclose and/or support the first axial
bearing 150a and/or the second axial bearing 150b in the first
cavity 162 or the second cavity 164. To help minimize or limit, for
example, the mass of the bearing housing 160, the first depth D1
and/or the second depth D2 can be configured so that the first
cavity 162 and/or the second cavity 164 do not extend in the axial
direction to enclose and/or support any portion of the first and/or
second axial bearing 150a, 150b. The first axial bearing 150a, the
second axial bearing 150b, and the axial bearing retainer 142 can
be enclosed by the bearing housing 170.
[0096] It is noted that the arrangement of the first and second
radial bearings 140a and 140b, as well as the first and second
axial bearings 150a and 150b are exemplary. The arrangement of
bearings can be varied. The number of discharge radial bearings
and/or axial bearings can also vary.
[0097] FIG. 7 illustrates a known conventional bearing housing 360
of a screw compressor 300. The bearing housing 360 includes a first
cavity 362 and a second cavity 364. The first cavity 362 and the
second cavity 364 are configured to enclose both a first and a
second discharge radial bearings 340a, 340b, a first and a second
axial bearings 350a, 350b, as well as axial bearing retainers 342.
The bearing housing 360 is covered by a bearing cover 370.
[0098] It can be seen for example that the bearing housing 360 is
relatively longer in an axial direction defined by an axis C3 of a
first rotor 310 relative to the bearing housing 160 of FIG. 6. That
is, the bearing housing 160 in FIG. 6 is relatively shorter in that
the bearing housing 160 encloses the first and second radial
bearings 140a, 140b, but does not enclose the first and second
axial bearings 150a, 150b or the axial bearing retainers 142,
whereas the bearing housing 360 in FIG. 7 extends to enclose both
the first and second discharge radial and axial bearings 140a,
140b, 150a and 150b, as well as the axial bearing retainers
342.
[0099] Screw compressor performance and reliability may be linked
to the precise location of the discharge radial bearings (e.g. the
first and second radial bearings 140a, 140b), which can be
configured to support and locate the rotors (e.g. the first and
second rotors 110 and 120). By reducing the width W1 of the bearing
housing 160 relative to the conventional bearing housing 360, the
first and second cavities 162, 164 can be produced more precisely
and simply than the first and second cavities 362, 364 of the
convention bearing housing 360 due to shorter reaches and shorter
machine boring bars. This can help position the first and second
radial bearings 140a and 140b more accurately compared to the first
and second radial bearings 150a and 150b, which can help both the
performance and reliability of the screw compressor, by for example
more accurately positioning the first and second rotors 110,
120.
[0100] The bearing cover (e.g. the bearing housing 170) is
typically not a critical part, and can be machined relatively
cheaply and generally does not require precision machining tools.
In the embodiment as disclosed herein, the bearing assembly
includes the relatively long bearing cover 170 and the relative
short bearing housing 160 compared to the relatively short bearing
cover 370 and the relatively long bearing housing 360. The overall
cost of making the bearing assembly is reduced compared to a
convention design. It is also simpler to make.
[0101] It is to be understood that it is possible to make the
bearing housing thicker than it is necessary to accommodate the
discharge radial bearing to it enclosed at lead a portion of the
discharge radial bearings
[0102] With regard to the foregoing description, it is to be
understood that changes may be made in detail, without departing
from the scope of the present invention. It is intended that the
specification and depicted embodiments are to be considered
exemplary only, with a true scope and spirit of the invention being
indicated by the broad meaning of the claims.
* * * * *