FIELD OF THE INVENTION The present invention relates generally to a two-stage rotary screw fluid compressor.

More specifically, the present invention relates to a two-stage rotary screw fluid compressor including a valve for relieving compression work performed by the second stage, and intercooling means for cooling the air stream between the stages, and means for reducing leakage into the gearbox.

BACKGROUND OF THE INVENTION Rotary screw compressors are used in numerous industries to provide a supply of compressed air for supporting applications such as automatic machines, tools, material handling devices, and food processing equipment. In comparison to the predominate reciprocating-piston type compressor, rotary screw compressors operate more efficiently and at a lower compressor specific power. Other advantages include reduced space requirements and lower vibration levels. Two types of rotary screw compressors are oil-injected and oil- less.

The oil-injected type rotary screw compressor includes a casing with two intersecting bores having parallel axes, an inlet port adjacent one end wall, and a compressed air outlet port adjacent another end wall. Disposed within the bores are a pair of meshing rotors with each rotor having helical lands and intervening grooves with a wrap angle of less than 360°.

The leading and trailing faces of each land form leading and trailing flanks. Minimal clearances are maintained between the rotors and the end walls and bores of the casing. One rotor is a male rotor type, i. e. , a rotor having at least the major portions of its lands and grooves disposed outside the pitch circle of the rotor. The other rotor is a female rotor type, i. e. , a rotor having at least the major portions of its lands and grooves disposed inside the pitch circle of the rotor. The lands of one rotor follow the envelopes developed by the

grooves of the other rotor to form a continuous sealing line there between. Chambers are formed between the sealing line, land tops, casing end walls and bores. One of the rotors is driven while the other rotor is driven by the first.

In operation, a gaseous fluid is displaced and compressed within the chambers from the inlet port to the outlet port of the compressor. Three phases make up this process: a filling phase, a compression phase, and a discharge phase. During the filling phase each compression chamber communicates with the air inlet port, during the compression phase the chamber undergoes a continued reduction in volume, and during the discharge phase the chamber communicates with the compressed air outlet port.

Because the flanks of the rotors of the above described screw compressor are in meshing contact with one another, oil must be injected into the compressor to prevent excessive contact wear which would ultimately lead to premature failure of the compressor.

Oil also acts as a cooling fluid to remove heat generated through the compression of gasses.

The essential difference between oil-injected and oil-less air compressors is that the male and female rotors of oil-less systems are timed so they do not come into contact with each other during operation. In other words, total backlash between the rotors is proportioned, not necessarily equally, between the male rotor leading flank and female rotor trailing flank, and the male rotor trailing flank and female rotor leading flank. Rotor timing is typically provided by either helical or spur gears having pitch circles matching the pitch circles of their respective rotors.

Compressors often supply compressed air through non-continuous operation. For example, a system of compressors is often employed wherein the compressors are loaded or unloaded so as to provide a constant stream of compressed gas to a machine with variable gas requirements. When more compressed gas is needed, additional compressors are employed to provide compressed gas. Similarly, one compressor can be used in an application where a non-constant rate of compressed gas flow is required.

Under changing demand conditions, it is inefficient to continue to operate the compressor under load and thereby compressing gas that is not required. Prior art solutions to this problem have been to unload the compressor by opening lift valves such that the gas supply to the compressor is bypassed. This allows the rotors to turn without meeting resistance from compressing gas, thereby reducing the energy required to turn the rotors.

This results in an operational cost savings by only loading the compressor when more compressed gas is required.

There is, however, some residual gas inside the compressor after the lift valves have been opened. The compressor will compress this residual gas forcing the rotors to exert more work and require more energy than if they were allowed to rotate freely. This is particularly noticeable in a two-stage compressor where the first stage remains at relatively low pressure while unloaded, but the pressure builds in the second stage. It would therefore be desirable to reduce the residual gas inside the second stage of the compressor so as to relieve the rotors of this unproductive compression work during unloaded conditions.

The principles of thermodynamics provide that when a gas is compressed the temperature will rise. This temperature rise, in turn, leads to a less efficient compression of the gas. Cooler gas is more easily compressible that hot gas. Therefore, to improve the thermodynamic efficiency of the machine, it is desirable to maintain lower operating temperatures within the gas stream.

One method currently employed to reduce the temperature between the first and second stages is to employ a heat exchanger, for example a shell and tube heat exchanger, to reduce the temperature of the gas stream. However, such an operation also causes a pressure drop across the heat exchanger thereby negating the effects of the cooling. It would therefore be desirable to provide intercooling of the compressed gas stream between the first and second stages of the compressor without a corresponding pressure drop. This would allow the second stage to begin with an inlet pressure approximately equal to the first stage outlet pressure, but at a lower temperature, and therefore more thennodynamically efficient compression.

The general configuration of prior art two-stage compressor includes a gas flow path that begins through an inlet into the first stage. The inlet is located near the gearbox at the front end of the compressor. The gas is compressed as it is moved toward the rear of the compressor and the end of the first stage. The gas then passes from the first stage to the second stage near the rear of the compressor. The second stage further compresses the gas as it moves the gas forward toward the front end, or gearbox end of the compressor. The second stage gas outlet is then located somewhere near the gearbox at the end of the second stage.

This configuration creates a low-pressure area in the location of the gearbox and shaft connection of the first stage rotor, and a high-pressure area in the location of the gearbox and shaft connection of the second stage rotor. The shaft connections and gearboxes are not effectively sealed to entirely prevent pressure leakage from within the compressor, thus the proximate location of the low-pressure first stage inlet and high-pressure second stage outlet create leakage problems through the gearbox.

Any leakage along the air flow path from the first stage inlet to the second stage outlet increases the required horsepower necessary to run the compressor and reduces the amount of compressed air being delivered by the compressor. Having a high-pressure area in close proximity to a low-pressure area increases the likelihood of leakage from the high-pressure area to the low-pressure area. It would therefore be desirable to design the compressor so as to eliminate any configuration that provides close proximity between the highest and lowest pressure regions of the machine.

SUMMARY OF THE INVENTION The present invention provides an apparatus and method for constructing and operating a two stage rotary screw compressor so as to minimize the unloaded power requirements, provide enhanced intercooling of the air stream between the two stages, and minimize pressure leakage between the stages and the gearbox.

In one aspect of the present invention, a compressor is provided comprising a first stage comprising a first stage bypass cavity and a first stage outlet, an intercooling zone in communication with said first stage outlet, a second stage comprising a second stage inlet and a second stage outlet in communication with said intercooling zone, a gearbox disposed proximate to the first stage outlet and the second stage inlet, wherein both stages communicate with the gearbox through their respective rotor shafts, a valve, wherein the valve provides means for communications between the first stage bypass cavity and the second stage during unloaded conditions, and intercooling means operable for injecting cooling fluid into the intercooling zone.

In another embodiment of the present invention, the valve is a double acting lift valve or a plurality of valves disposed along the length of the second stage. When the valve allows communication between the first stage bypass cavity and the second stage, the overall power consumption of the compressor comprises less than 20 percent of the maximum loaded power.

In a further embodiment of the present invention, the valve is located within the latter half of the second stage as defined by the rotor length. In a still further embodiment of the present invention the valve is located within the final 20 percent of the length of the second stage as defined by the rotor length.

In another embodiment of the present invention, the intercooling means are operable for injecting the cooling fluid into the intercooling zone at an angle of between 90 degrees

and 270 degrees, and preferably about 180 degrees relative to the air stream entering the intercooling zone.

In one embodiment of the present invention, the intercooling means comprise a plurality of orifices for injecting cooling fluid into the first stage discharge air stream. The plurality of orifices further comprises between 6 and 12 orifices.

In a further embodiment of the present invention, the compressor is designed such that the first stage discharge gas pressure is approximately equal to the second stage inlet pressure. Further, the pressure differential between the first stage outlet and the second stage inlet is preferably the smallest pressure differential of any two points along the first stage and second stage.

In a still further embodiment of the present invention, the first stage discharge gas pressure is located at an end of the first stage near the gearbox and opposite a first stage inlet pressure, and the second stage inlet pressure is located at an end of the second stage near the gearbox and opposite a second stage outlet pressure, such that the first stage inlet pressure, the lowest pressure within the compressor and the second stage outlet pressure, the highest pressure within the compressor are located at opposite ends of the air stream within the compressor.

In another aspect of the present invention, a method for operating a compressor is provided comprising, providing a first stage comprising a first stage inlet cavity and a first stage outlet, providing an intercooling zone in communication with said first stage outlet, providing a second stage comprising a second stage inlet and a second stage outlet in communication with said intercooling zone, providing a gearbox disposed proximate to the first stage outlet and the second stage inlet, providing a valve, wherein the valve provides means for communications between the first stage bypass cavity and the second stage during unloaded conditions, providing intercooling means operable for injecting cooling fluid into the intercooling zone, and wherein the valve is opened during unloaded conditions to allow compressed gas from the second stage into the first stage bypass cavity, thereby reducing the compression work of the second stage during unloaded conditions, wherein the intercooling means are operable for injecting cooling fluid into the intercooling zone at an angle of about 180 degrees relative to the direction of the gas entering the intercooling zone, thereby inducing mixing of the coolant and compressed gas and cooling the compressed gas stream before entering the second stage of the compressor.

In one aspect of the present invention, the valve is located within the final 20 percent of the length of the second stage as defined by the rotor length. In a further embodiment of

the present invention, the first stage discharge gas pressure is minimized and approximately equal to the second stage inlet pressure, thereby substantially reducing pressure leakage into the area of the gearbox.

Features of a compressor of the present invention may be accomplished singularly, or in combination, in one or more of the embodiments of the present invention. As will be appreciated by those of ordinary skill in the art, the present invention has wide utility in a number of applications as illustrated by the variety of features and advantages discussed below.

A compressor of the present invention provides numerous advantages over prior compressor designs. For example, the present invention advantageously provides a means for reducing the second stage compression work during unloaded conditions so as to minimize the power consumption during such unloaded periods.

Another advantage is that the present invention provides a means for intercooling the compressed air stream between the first and second stages of the compressor without experiencing a significant pressure drop.

A still further advantage is that the present invention provides a compressor design that minimizes the pressure differential between the two stages in the area of the gearbox, so as to reduce or eliminate pressure leakage between the stages and the gearbox cavity.

As will be realized by those of skill in the art, many different embodiments of a compressor according to the present invention are possible. Additional uses, objects, advantages, and novel features of the invention are set forth in the detailed description that follows and will become more apparent to those skilled in the art upon examination of the following or by practice of the invention.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a cross sectional view from the front of a two-stage rotary screw fluid compressor according to one embodiment of the present invention, highlighting the double acting lift valve capacity bypass for the second stage and the bypass cavity.

FIG. 2 is a view of the right side of a two-stage rotary screw compressor according to one embodiment of the present invention showing the location of the first stage capacity control valves and the second stage capacity bypass valve.

FIG. 3 is a cross sectional view from the right side of a two-stage rotary screw compressor according to one embodiment of the present invention, highlighting the general

arrangement of each stage, the means for managing the gas flow path so as to minimize leakage and the intercooling means.

FIG. 4 is a cross sectional view from the rear of a two-stage rotary screw compressor according to one embodiment of the present invention, highlighting the intercooling means and arrangement of the coolant injection nozzles.

DETAILED DESCRIPTION OF THE INVENTION The present invention provides various apparatus and methods relating to the construction and operation of two stage rotary screw compressors. These compressors are used in various applications providing high-pressure gas. Most frequently, the gas is atmospheric air, although the compressor designs described herein can be used to compress most other compressible fluids. For exemplary purposes, the description of the preferred embodiments of the present invention will refer to the compression of air. However, as one skilled in the art will recognize, the apparatus and methods of the present invention may be employed in conjunction with compressors used to compress a variety of compressible fluids.

Referring to the Figures where like numbers represent like components, in a first aspect of the present invention, a compressor 1 is provided comprising a capacity control valve 18 or valves to regulate airflow through the compressor 1. The compressor 1 is a two- stage rotary screw compressor comprising a first stage 10 with a first stage inlet 14 and first stage outlet 16, and a second stage 20 with a second stage inlet 24 and a second stage outlet 26.

As can be best seen in FIG. 3, an air flow path is defined by the first stage inlet 14 in communication with the first stage 10 of the compressor, further in communication with the first stage outlet 16. The first stage outlet 16 is in communication with the second stage inlet 24 which is in communication with the second stage 20 of the compressor, further in communication with the second stage outlet 26.

As depicted in FIGS. 1 and 2, one embodiment of the compressor 1 of the present invention comprises a plurality of first stage capacity control valves 18 each comprising an open and closed position. The valves may be opened or closed to regulate the quantity of air passing by the first stage of the compressor 10 depending on the demand requirements on the compressor 1. The capacity control valves 18 allow air into a bypass cavity 12, as seen in FIG. 1, which is in communication with the first stage inlet 14.

In a further embodiment of the present invention, the compressor 1 further comprises a second stage capacity bypass valve 22. The second stage capacity bypass valve 22 provides

means for reducing the unloaded power consumption by providing communication between the second stage and the first stage bypass cavity 12. In effect, this connects the first stage 10 and second stage 20 in a manner so as to relive pressure in the second stage 20.

When the compressor 1 is unloaded, the inlet valves 18 open to bypass air from the first stage compression process through the bypass cavity 12 into the first stage inlet 14. This greatly reduces the amount of air being compresses in the first stage 10 and thus reduces the amount of power required to turn the first stage rotors.

Similarly, the second stage capacity bypass valve 22 connects the second stage 22 with the first stage inlet 14 via the bypass cavity 12. Any residual air in the first stage after the inlet capacity control valves 18 are open is moved to the second stage. Without the capacity bypass valve 22, this air would be compressed in the second stage causing the second stage to perform unnecessary compression work on the air in the second stage.

By opening the capacity bypass valve 22, the air in the second stage is allowed to enter the first stage bypass cavity 12 thereby reducing the second stage compression work and correspondingly reducing the power consumption of the second stage. By removing the compression work in the second stage 20, the compressor 1 requires even less power to operate than unloading only the first stage 10.

In one embodiment of the present invention, the second stage capacity bypass valve 22 is located at least at the halfway point along the length of the second stage, as defined by the rotor length. By positioning the bypass valve 22 further down the second stage, the higher pressure located toward the end of the second stage 20 will be reduced. In a preferred embodiment of the present invention, the second stage bypass valve 22 is located within the final 20 percent of the second stage, as defined by the rotor length.

In a further embodiment of the present invention, the second stage capacity bypass valve 22 comprises a plurality of capacity bypass valves positioned along the second stage 20. For example, one capacity bypass valve located at the halfway point along the second stage 20 and one located toward the end of the second stage 20. Employing a plurality of capacity bypass valves 22 allows for enhanced communication between the second stage 20 and the first stage bypass cavity 12 thereby further reducing the buildup of pressure in the second stage 20.

The capacity bypass valve 22 of the present invention comprises any commonly known compressor valve. Acceptable valves include, but are not limited to, lift valves, turn valves, spiral valves, and the like. In a preferred embodiment of the present invention the second stage capacity bypass valve 22 comprises a lift valve.

In one embodiment of the present invention, the power consumption of a two stage compressor is reduced to less than 20 percent of the maximum loaded power consumption when the second stage capacity bypass valve 22 is open, in conjunction with the first stage capacity control valves 18, to allow communication with the first stage bypass cavity 12.

Thus, by opening the capacity bypass valve 22, a significant energy savings is realized.

In a further embodiment of the present invention, the capacity bypass valve 22 provides direct communication with the first stage 10. This embodiment eliminates the first stage bypass cavity 12 and directly connects the two stages. The same results are achieved as described above with regard to energy savings in an unloaded state.

In a still further embodiment of the present invention, the capacity bypass valve 22 provides communication between the second stage 20 and the atmosphere. In this embodiment any pressure in the second stage is allowed to vent outside the compressor. The position of the capacity bypass valve or valves 22 along the second stage will allow for a compression work reduction along the second stage so as to minimize the work being done by the compressor 1.

In another aspect of the present invention, an apparatus and method for intercooling a compressed air stream between a forst stage and a second stage is provided. As shown in FIGS. 3 and 4, the first stage outlet 16 and second stage inlet 24 are in communication through an intercooling zone 30. The intercooling zone 30 further comprises at least one injection orifice 32 for injecting cooling fluid into the compressed air stream.

In a preferred embodiment of the present invention, the at least one injection orifice 32 comprises a plurality of injection orifices positioned within the intercooling zone 30. In a most preferred embodiment of the present invention, the intercooling zone comprises 6 to 12 injection orifices.

The injection orifices are positioned so as to inject cooling fluid into the air stream at an angle relative to the direction of the air stream entering the intercooling zone 30. The air stream entering the intercooling zone 30 is ejected from the end of the first stage 10 and is generally directed parallel to the length of the first stage 10. In a preferred embodiment of the present invention, the angle of introduction of cooling fluid is between 90 and 270 degrees relative to the entering air.

In a most preferred embodiment of the present invention, the angle of introduction is about 180 degrees from the entering air stream. By providing a coolant stream injected at a direction 180 degrees from the entering air stream, optimal mixing of the two streams is achieved. As the air enters the intercooling zone 30, it meets the coolant being injected

through the injection orifice 32 and the collision of the two streams provides turbulence and mixing.

As the two streams mix, the coolant absorbs heat from the air stream thereby reducing the temperature of the air. The coolant and lower temperature compressed air stream then depart the intercooling zone 30 and enter the second stage 20 of the compressor for further compression.

In one embodiment of the present invention the plurality of injection orifices are aligned along an axis perpendicular to the length of the first stage 10. In an alternate embodiment of the present invention, the plurality of injection orifices 32 are arrayed throughout the intercooling zone so as to contact the incoming compressed air stream at several different angles so as to provide contact with several aspects of the incoming air stream.

Suitable cooling fluids for use with the present invention comprise those generally known in the art to be suitable cooling fluids for two-stage rotary fluid compressors. In one embodiment of the present invention the cooling fluid also comprises the lubricating fluid and therefore possesses anti-foaming and lubricating properties as well as anti-oxidation properties and the specific heat capacity requirements necessary to effectively lubricate the compressors and cool the compressed air stream.

In yet another aspect of the present invention, a system and method for reducing pressure leakage within a two stage rotary screw compressor is provided. Referring to FIG. 3 the compressor 1 is constructed in such a manner so as to minimize pressure differentials between the first and second stages and the gearbox 40.

The airflow path through the compressor is defined by a first stage inlet 14 located at a distal end from the gearbox in communication with the first stage 10, which in turn is in communication with a first stage outlet 16 located at an end proximate to the gearbox 40.

The first stage 10 is engaged with the gearbox 4Q through the rotor shaft at this end so as to provide power to the first stage rotors from the gearbox 40. The second stage 20 is also engaged with the gearbox 40 through the rotor shaft at this end so as to provide power to the second stage rotors from the gearbox 40. The second stage inlet 24 is in communication with the first stage outlet 16 in the area proximate to the gearbox 40. The second stage inlet 24 is further in communication with the second stage 20 which is in communication with the second stage outlet 26 at a distal end relative to the gearbox 40.

Thus, in the foregoing air flow path, the area proximate to the gearbox 40 comprises the first stage outlet 16 in communication with the second stage inlet 24. The pressure of the

air in the area of the first stage outlet 16 is substantially equal to the pressure at the second stage inlet 24, i. e. the first stage compression is complete and the second stage compression has yet to begin. This provides a minimal pressure differential between the two stages in the area of the gearbox and the pressure inside the gearbox.

In a preferred embodiment of the present invention, the pressure differential between the first stage outlet 16 and second stage inlet 24 is minimized relative to the gearbox pressure. In a most preferred embodiment of the present invention, the pressure of the first stage outlet 16 and second stage inlet 24 is equal and minimized to less than 35 psi. This low pressure differentiation between the two stages and the gearbox 40 reduces leakage between the first stage outlet 16 and second stage inlet 24 into the gearbox 40.

Although the present invention has been described with reference to particular embodiments, it should be recognized that these embodiments are merely illustrative of the principles of the present invention. Those of ordinary skill in the art will appreciate that the apparatus and methods of the present invention may be constructed and implemented in other ways and embodiments. Accordingly, the description herein should not be read as limiting the present invention, as other embodiments also fall within the scope of the present invention.