US20150030490A1

US20150030490A1 - Bearing Housing and Assembly of a Screw Compressor

Info

Publication number: US20150030490A1
Application number: US14/513,009
Authority: US
Inventors: Dennis M. Beekman; Daniel R. Crum; Timothy Sean Hagen; Dennis R. Dorman; John R. Sauls
Original assignee: Trane International Inc
Current assignee: Trane International Inc
Priority date: 2010-07-20
Filing date: 2014-10-13
Publication date: 2015-01-29

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.

Description

FIELD

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

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

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.
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.
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.
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.
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.
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.
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.
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

Reference is now made to the drawings in which like reference numbers represent corresponding parts throughout.

FIG. 1 illustrates an embodiment that incorporates a screw compressor arranged as part of a refrigeration chiller system.

FIG. 2 illustrates a cross sectional view of a screw compressor according to one embodiment.

FIG. 3 illustrates an additional cross sectional view of a screw compressor according to one embodiment.

FIG. 4 illustrates an embodiment of a refrigeration chiller and controller system according to one embodiment.

FIG. 5 illustrates a partial sectional view of a screw compressor, with which the embodiments as disclosed herein can be practiced.

FIG. 6 illustrates an enlarged sectional view of an area A of the screw compressor in FIG. 5.

FIG. 7 illustrates a conventional bearing housing.

DETAILED DESCRIPTION

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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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 42 a, 42 b may define one or more compression pockets between the screw rotors 42 a, 42 b 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.
First screw rotor 42 a and second screw rotor 42 b 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.
In the embodiment illustrated, rotor 42 a 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 42 b 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).
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.
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 42 a, 42 b, the axial forces, and the radial forces.
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 42 a, 42 b, 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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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 42 b. 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.
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.
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.
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.
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.
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 42 b and the female rotor 42 a. 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.
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 42 b has a wrap angle of about 347 degrees and the female rotor 42 a has a wrap angle that is 6/7ths of the male rotor 42 b. The wrap angle of the female rotor 42 a varies with the ratio of number of lobes. The female rotor 42 a has a radius of about 2.5 inches (6.35 centimeters) and 7 lobes and the male rotor 42 b 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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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.
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 140 a and 140 b and/or axial bearings 150 a and 150 b (sometimes referred to as “thrust bearings”) respectively. The discharge radial bearings 140 a and 140 b and the axial bearings 150 a and 150 b can be attached to the first and second shafts 112 and 122. The axial bearing 150 a and 150 b, as well as the discharge radial bearings 140 a, 140 b can be retained in the axial direction by axial bearing retainers 142.
The bearings 140 a, 140 b, 150 a and 150 b 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 140 a and 140 b and the axial bearing 150 a and 150 b can help withstand forces that may be produced during operation.
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 140 a and 140 b. The discharge radial bearings 140 a and 140 b are typically positioned closer to the rotor housing 130 than the axial bearings 150 a and 150 b in the axial direction.
The bearing housing 160 includes a first cavity 162 and a second cavity 164 configured to enclose and/or support the discharge radial bearings 140 a and 140 b respectively. The bearing housing 160 can be a slab-like structure, which may help provide a relatively uniform thermol expansion. The discharge radial bearings 140 a and 140 b 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 140 a and 140 b during operation, such as for example, help the discharge radial bearings 140 a and 140 b withstand forces that may be produced during operation.
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 150 a and 150 b, portions of the first and second shafts 112 and 122, and the axial bearing retainers 142 can be enclosed inside the space 180.
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 140 a and the second discharge radial bearing 140 b respectively.
The first discharge radial bearing 140 a has a first length L1 and the second discharge radial bearing 140 b 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 140 a and the second discharge radial bearing 140 b 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 150 a and 150 b.
In some embodiments, the first discharge radial bearing 140 a can be fully supported by the first cavity 162. In some embodiments, the second discharge radial bearing 140 b 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 140 a or the second discharge radial bearing 140 b respectively will stop increasing as the first depth D1 or the second depth D2 increases.
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.
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.
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 140 a and/or the second radial bearing 140 b. 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 140 a and/or 140 b 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 140 a and/or the second radial bearing 140 b. 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 140 a and 140 b is not enclosed and/or unsupported by the first and/or the second cavities 162 and 164.
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 150 a and 150 b can be enclosed in the space 180. Referring to FIG. 6, the first axial bearing 150 a (and/or the second axial bearing 150 b, which is not illustrated in FIG. 5) can have a clearance 182 between the first axial bearing 150 a (and/or the second axial bearing 150 b) and the bearing cover 170. The first axial bearing 150 a and/or the second axial bearing 150 b does not generally need a support from the bearing housing 160 and/or the bearing cover 170 because the first axial bearing 150 a and/or the second axial bearing 150 b typically are configured to withstand forces in the axial direction.
It is to be appreciated that because the bearing cover 170 can have the clearance 182 with the first axial bearing 150 a (and/or the second axial bearing 150 b), therefore the bearing cover 170 does not have to be precision machined.
In the illustrated embodiment, the first axial bearing 150 a and/or the second axial bearing 150 b are arranged next to the first discharge radial bearing 140 a and/or the second discharge radial bearing 140 b in the axial direction. The first axial bearing 150 a and the second axial bearing 150 b are positioned further away from the first rotor 110 and/or the second rotor 120 relative to the first and second discharge radial bearings 140 a and 140 b respectively in the axial direction. Because the axial bearing 150 a and/or the second axial bearing 150 b 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 150 a and/or the second axial bearing 150 b 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 150 a, 150 b. The first axial bearing 150 a, the second axial bearing 150 b, and the axial bearing retainer 142 can be enclosed by the bearing housing 170.
It is noted that the arrangement of the first and second radial bearings 140 a and 140 b, as well as the first and second axial bearings 150 a and 150 b are exemplary. The arrangement of bearings can be varied. The number of discharge radial bearings and/or axial bearings can also vary.
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 340 a, 340 b, a first and a second axial bearings 350 a, 350 b, as well as axial bearing retainers 342. The bearing housing 360 is covered by a bearing cover 370.
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 140 a, 140 b, but does not enclose the first and second axial bearings 150 a, 150 b 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 140 a, 140 b, 150 a and 150 b, as well as the axial bearing retainers 342.
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 140 a, 140 b), 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 140 a and 140 b more accurately compared to the first and second radial bearings 150 a and 150 b, 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.
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.
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
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.

Claims

What is claimed is:

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.