REFLUX GAS COMPRESSOR

BACKGROUND OF THE INVENTION

1. Field of the Invention.

The present invention is related to mechanical gas compressors and pumps. More particularly, the present invention is related to positive displacement rotary compressors, specifically including those known as Roots blowers.

2. Description of Related Art Including Information Disclosed Under 37CFR 1.97-1.99.

The present invention is related to the rotary gas compressors disclosed and claimed in the applicant's previously issued U.S. Patents No. 4,859,158, 5,090,879, 5,439,358, and 6,312,300, issued Aug. 22, 1989, Feb. 25, 1992, Aug. 8, 1995, and Nov. 6, 2001, respectively.

The class of positive displacement compressors known as Roots blowers has been known to and has served industry continuously since the mid 1850's. Roots blowers typically include two lobed impellers, also called rotors, which are rotated about parallel drive shafts in opposite directions and which are meshed with one another in a phased relationship by means of timing gears attached to each drive shaft. Commercially available Roots blowers typically have impellers with two or three lobes, but have also been designed to incorporate four or more lobes. Roots blowers having two lobes on each impeller have the greatest volumetric capacity per revolution and are the most common, as volumetric capacity is reduced proportionately by adding additional lobes. Roots blowers are particularly useful for moving large volumes of air or other gases from one volumetric space to another, typically against low pressure differentials. They are commonly referred to as blowers, as opposed to compressors, because compression of gas does not take place within the machine itself, as in a typical reciprocating piston-and- cylinder compressor; but rather takes place only when gas is discharged into a volumetric space that is at a higher pressure than the pressure of the intake gas.

Previously known Roots blowers are useful as compressors for compressing gases from atmospheric pressure up to approximately 5 to 7 psig discharge pressure. They are also useful

for evacuation of gas from one volumetric space to another, and may be used as a vacuum pump or a vacuum booster.

Roots blowers offer a number of advantages over other types of gas compressors, including conventional reciprocating piston compressors, helical screw compressors, fan type blowers, and centrifugal compressors. Among the advantages are simplicity, ruggedness, high volumetric capacity, and trouble-free operation. Roots blowers are particularly useful for sweeping large amounts of gas from one space to another in situations where mixture of intake gas and discharge gas must be prevented, as there is little or no backflow or mixture of discharge gas with intake gas, either when the blower is operating or when it is stopped. Further, Roots blowers do not contaminate a gas being processed, as there are no valves or reciprocating, rubbing, or contacting parts in the flow stream, and lubrication of the impellers is not ordinarily necessary. The Roots blower also maintains constant displacement volume from intake through to discharge, a design feature not found in any other type of positive displacement compressor. Roots blowers have not been previously known as being particularly useful for compressing a gas against a substantial pressure differential. This limitation has been due to heating effects that accompany such compression. As a gas is impelled through a conventional Roots blower it is compressed and undergoes an increase in temperature as it is discharged into a volume of higher pressure discharge gas. Such compression is adiabatic, such that the temperature of the gas increases exponentially with increasing pressure ratios. In addition, heat is generated from dynamic flow effects as discharge pressure gas surges into impeller cavities and is then expelled in the opposite direction.

This increase in temperature of the gas being processed through the blower leads to heating of the impellers, the housing and other mechanical parts of the blower. Such heating is not uniform throughout the compressor and cannot be easily controlled. The compressor housing, for example, can be externally cooled by a number of conventional methods, such as the use of water jackets, radiating fins, heat sinks, and the like. The greatest heating problem lies with the impellers because there is no practical way to directly cool them. Overheating of the impellers leads to their distortion, expansion and eventual binding against the housing. At pressure ratios above about two to one (2:1) such effects become a significant problem and essentially limit the sustained operation of the blower. Overheating of the blower can result in lockup or other mechanical failure of the impellers and associated seals and other components,

causing extensive damage and shutdown. Overheating has been a major limitation on the use of Roots blowers for compressing gas against high pressure differentials.

A significant advance in the art was the development of recirculation cycles to effect a moderate reduction in the heating of Roots compressors. In a recirculating Roots compressor, a portion of the discharge gas, which is compressed to a higher pressure than the intake gas, is recirculated back into the compressor so as to effectively increase the pressure of gas being processed through the compressor. In some recirculating compressors a portion of the discharge gas is cooled prior to being recirculated back into the compressor. In both cases the operating temperature of the compressor is effectively reduced, thereby mitigating the heating problems noted above. By this means, a capability for sustained operation has been obtained in some cases up to pressure differentials of approximately 2.7:1.

U.S. Patent No. 2,489,887 to Houghton, for example, discloses the general concept of cooling a Roots compressor by introducing recirculated gas of a lower temperature into the intake gas to reduce heating of the compressor passages which allow a portion of the high pressure discharge gas to be recirculated back into the pump.

U.S. Patent No. 3,351,227 to Weatherston discloses a multi-lobed Roots type compressor having feed-back passages which allow a portion of the high-pressure discharge gas to be recirculated back into the pump housing. Weatherston, however, discloses only the use of quite small feedback passages, the size of which are not related to the sizes of the intake and discharge ducts. This results in uneven flow velocities and pressures. As a' result, the Weatherston compressor does not significantly mitigate overheating of the process gas.

German Patent No. 2,027,272 to Kruger discloses the concept of cooling and recirculating discharge gas in a two-lobed Roots compressor. The compressor of Kruger, due to its two-lobed configuration, has no provision for preventing communication and backflow from the discharge port into the recirculation ports.

French Patent No. 778,361 to Bucher discloses four-lobed Roots compressors having recirculation ports. The recirculation ports are however small, with the intended purpose of using small, nozzle-like ports to allow the recirculated gas to _. cool upon entry into the compressor housing.

U.S. Patent No. 4,390,331 to Zimmerly discloses a lobed-impeller, positive displacement rotary pump that is designed primarily for pumping liquids, with no provision for recirculation.

U.S. Patent No. 4,390,331 to Nachtrieb discloses a rotary compressor having four-lobed impellers, but likewise with no provision for recirculation.

U. S. Patent No. 2,906,448 to Lorenz discloses a positive displacement rotary compressor having two-lobed impeller, with a double-walled housing construction for cooling purposes.

British Patent No. 282,752 to Kozousek discloses a rotary pump that is characterized by rotor lobes that are particularly shaped so as to provide the maximum possible working space or displacement volume per revolution, and thereby maximize the volumetric capacity of the pump. The pump disclosed in Kozousek discloses small recirculation ports which are for the purpose of obtaining even delivery of the gas.

In some prior art recirculating Roots compressors, such as the compressor described in Houghton, the flow of recirculating gas is periodically interrupted each time a rotor lobe passes the recirculation entry port, or is halted or possibly reversed as a displacement cavity is simultaneously opened to a recirculation entry port and the discharge port. This results in a loss of momentum and flow of the recirculation fluid, creating heat and reducing the efficiency of the recirculation fluid in cooling the compressor flow. This problem, which is inherent in many previously known recirculating Roots compressors, is overcome in the present invention and overcome in the previously issued U.S. Patents No. 5,439,358 and 6,312,300 to Weinbrecht, as will be apparent from the description set forth below.

In the applicant's previously issued U.S. Patents cited above, certain aspects are disclosed which achieve a substantial change in the thermodynamic nature of the compression cycle, such that the resulting compression process is significantly more isothermal than adiabatic, and such that heat generated in the process is reduced.

Accordingly, it is an object and purpose of the present invention to provide an improved positive displacement, transverse flow, rotary gas compressor.

It is also an object and purpose of the present invention to provide a positive displacement, transverse flow, rotary gas compressor having an improved gas recirculation or refluxing means for reducing heating of the compressor.

It is a further object and purpose of the present invention to provide a positive displacement, transverse flow, rotary gas compressor that is characterized by having an internal peripheral counter flow of refluxing gas which flows back from discharge into advancing positive displacement cavities.

It is also an object and purpose of the present invention to provide a rotary, positive displacement, transverse flow gas compressor that produces significantly less heat inside the compressor, and is thus capable of operating at higher sustained pressure ratios than have previously been attainable.

It is also an objective and purpose of the present invention to provide a rotary, positive displacement, transverse flow gas compressor which establishes a compression cycle having a thermodynamic nature that is significantly closer to isothermal than to adiabatic, and that does not require internal cooling for operation at pressure ratios of up to approximately five to one (5:1).

It is also an objective and purpose of the present invention to provide a rotary, positive displacement, transverse flow gas compressor which achieves improved efficiency through a substantially isothermal thermodynamic compression cycle.

It is yet another object and purpose of the present invention to provide a rotary, positive displacement, transverse flow gas compressor which is characterized by having a means of refluxing internal to the rotor housing, thereby significantly reducing the amount of fabrication effort and material required for compressor housing production.

SUMMARY OF THE INVENTION

The compressor of the present invention includes a housing having opposing end walls and mutually opposing interior sidewalls. The compressor includes a pair of intermeshed, involutely lobed impellers, also referred to as rotors, which are rotatably journalled in the housing. The impellers are driven to rotate in opposite directions so as to sweep a gas from intake through the housing from an intake port to a discharge port. The impellers may have from four to nine lobes. The volumetric spaces defined by adjacent lobes of the impellers, the opposing end walls of the housing, and the interior sidewalls of the housing are referred to herein as displacement cavities, in which parcels of gas are transported from the intake port of the compressor to the discharge port.

Upstream sidewall portions of the interior housing sidewalls are cylindrically curved, with a radius of curvature as close to the radii of the rotating impeller lobes as can be achieved within normal machining tolerances while avoiding sliding contact between the impeller lobes and said upstream sidewall portions, so as to form a substantially gas-tight seal between the tips of the lobes and the upstream sidewall portions in the manner of a conventional Roots blower.

The upstream sidewall portions of each sidewall extend from the intake port over an angular sector at least equal to the angular sector between adjacent lobes of the impeller, which for example in the case of a six-lobed impeller is 60 degrees, so as to impede or restrict backflow of gas into the intake port. In the preferred embodiment the cylindrically curved upstream sidewall portions of the interior sidewalls each extend through an angular sector equal to approximately twice the angular sector between adjacent lobes of the impellers.

The interior housing sidewalls further include downstream sidewall portions, which are spaced radially from the outer ends of the impeller lobes so as to allow limited peripheral counterflow, or reflux, of the compressed discharge gas back into the displacement cavities formed between adjacent lobes of the impellers. The downstream sidewall portions extend over an angular sector, extending upstream from the discharge port, of at least the angle between adjacent lobes of the impellers, which in the case of a six-lobed impeller is 60 degrees.

In the preferred embodiment the upstream sidewall portions of the sidewalls each extend over an angular sector that begins at the intake port and is approximately equal in magnitude to the angular sector over which any two adjacent lobes of the impeller extend, or 120 degrees in the case of a six-lobed impeller. The downstream sidewall portions extend upstream from the discharge port over an angular sector also approximately equal to the angular sector represented by any two adjacent lobes of the impeller, or 120 degrees in the case of the six-lobed impeller.

The reflux discharge gas that is admitted into the displacement cavities by peripheral backflow along the downstream wall portions of the sidewalls serves to raise the pressure of intake gas within the displacement cavities, so that the gas pressure within each displacement cavity is nearly equal to that of the discharge pressure as the contained gas is swept into the discharge port.

In the preferred embodiment the impellers each have six lobes, and the upstream and downstream portions of the interior housing sidewalls each extend over a combined angular sector of approximately 300 degrees, with the upstream wall portions each extending through an angular sector of approximately 120 degrees. This embodiment is preferred because it results in slippage or backfill flow between the tips of the impeller lobes and the interior housing sidewalls being collected in a following cavity and carried forward into discharge, and is thereby characterized by improved volumetric efficiency.

The compressor of the present invention is useful in applications requiring the continuous compression of large volumes of gas or vapor. The transverse flow arrangement and the rugged design permit in-line multiple staging driven by a single power source, so that very high compression system pressure ratios can be achieved. The near-isothermal thermodynamic nature of the compression process provides an inherent energy efficiency advantage of from 8% to 14% when compared to any prior art method of compression.

These and other aspects of the present invention will become more apparent upon consideration of the detailed description of the invention set forth below and in the accompanying drawing.

BRIEF DESCRIPTION OF THE DRAWING

The accompanying drawing is incorporated into and a part of this specification and, when taken in combination with the detailed description below, illustrates the operation and construction of the best mode of the invention known to the inventor.

Figure 1 is an end view in cross section of the preferred embodiment of the rotary compressor of the present invention.

DESCRIPTION OF THE PREFERRED EMBODIMENT

Referring to Figure 1, there is illustrated a preferred embodiment of the positive displacement, transverse flow, rotary compressor 10 of the present invention. The compressor 10 includes two involutely lobed impellers 12 and 14, each having six lobes, which are rotatably journalled within a hollow housing 16 and which are driven in opposite rotational directions as indicated by the directional arrows in Figure 1. Impellers 12 and 14 are shaped and intermeshed with one another so as to form a substantially gas-tight seal that prevents gas from passing between them at all stages of their rotation. In operation, gas is drawn into the compressor 10 through an intake port 18 and is discharged from a discharge port 20 at the opposite side of the compressor 10. For reasons that will be apparent below, it is noted that the individual lobes of the six-lobed impellers 12 and 14 are spaced at 60 degree angular intervals from one another.

The housing 16 has interior surfaces which include two opposing, parallel, planar end walls (only one end wall 22 of which is shown), each of which are orthogonal to the axes of rotation of the impellers 12 and 14.

Housing 16 further includes upper and lower opposing interior sidewalls 24 and 26, respectively, which each extend from the intake port 18 to the discharge port 20 across the upper

and lower halves of the housing 16, respectively. The volumetric spaces defined by adjacent lobes of the impellers 12 and 14, the opposing end walls of the housing, and the interior sidewalls 24 and 26 of the housing 16 are referred to herein as displacement cavities, in which parcels of gas are transported from the intake port 18 of the compressor to the discharge port 20.

As described in more detail below, the sidewalls 24 and 26 include upstream and downstream sidewall portions of slightly different sizes and shapes, which function to permit a limited amount of reflux backflow of high pressure discharge gas into the compressor 10 during transport of gas through the compressor 10, while nevertheless preventing backflow of discharge gas into the intake port 18.

Specifically, the upper interior sidewall 24 includes an upstream sidewall portion 24a and a downstream sidewall portion 24b, which are separated by a short transition sidewall portion 24c.

The upstream sidewall portion 24a is cylindrically curved and has a radii of curvature that is as close to the maximum radii of the lobes of impeller 12 as can be achieved within normal machining tolerances, while avoiding frictional contact between the tips of the lobes of impeller 12 and the upstream sidewall portion 24a. The impeller 12 and the upstream sidewall portion 24a, taken alone, thus function in the manner of a conventional Roots compressor to sweep parcels of gas from the intake port 18 into the compressor 10, while preventing backflow of gas into the intake 18. The upstream sidewall portion 24a extends over an angular sector, as measured from the upper edge of the intake port 18, of approximately 120 degrees, or the angular sector defined by any two lobes of the six-lobed impeller 12.

The downstream sidewall portion 24b is at all points at a greater distance from the axis of rotation of impeller 12 than is the upstream sidewall portion 24a. More specifically, in the preferred embodiment the downstream sidewall portion 24b has a noncylindrical curvature that is characterized by a slightly but progressively increasing distance from the axis of rotation of impeller 12, as measured moving from the upper Hp of discharge port 20 toward the transition sidewall portion 24c; such that the backflow reflux of discharge gas past the tips of the lobes of impeller 12 diminishes at greater distances upstream from the discharge port 20.

Generally, the greater radial distance of the downstream sidewall portion 24b from the axis of rotation of the impeller 12, as compared with that of the upstream sidewall portion 24a which effectively forms a gas-tight seal with the tips of the impeller lobes, allows a controlled

and limited amount of high pressure discharge gas to flow back into the displacement cavities that are bounded by downstream sidewall portion 24b, before they open into the discharge port 20. More specifically, the progressively increasing distance between the surface of the downstream sidewall portion 24b and the impeller axis of rotation, as measured moving toward the discharge port 20, allows for a greater amounts of high pressure discharge gas to flow into the displacement cavity nearest the discharge port 20, while allowing a lesser amount of discharge gas to flow into the immediately preceding displacement cavity, and an even lesser amount to flow into the next preceding displacement cavity. As shown in Figure 1, limited and progressively decreasing amounts of high pressure discharge gas are allow to backflow into the three displacement cavities located upstream from the discharge port 20, while at all times the displacement cavity or cavities nearest the intake port 18, that is, any displacement cavity bounded by the upstream sidewall portion 24a, is effectively sealed and thus does not permit backflow of high pressure discharge gas into such cavity or into the intake port 18. As will be further seen from Figure 1, as any lobe of the impeller 12 passes the transition sidewall portion 24c, which merely represents a transition in the machined interior sidewall 24 of the housing 16, higher pressure discharge gas is progressively and increasingly admitted past the tip of the lobe so as to increase the pressure in the preceding displacement cavity, such that by the time the displacement cavity is opened to the discharge port 20 the pressure in the displacement cavity is substantially increased, thereby reducing the increase in temperature occasioned by opening of the displacement cavity into the discharge port 20.

The upper sidewall portions 24a, 24b and 24c of the illustrated preferred embodiment, combined, extend over an angular sector of somewhat less than 270 degrees, as measured from the upper edge of the intake port 18 and extending across the upper side of the housing 16 to the upper edge of the discharge port 20.

The lower interior sidewall 26 includes an upstream sidewall portion 26a, a downstream sidewall portion 26b, and a short transition sidewall 26c, all of which function in the same manner as the corresponding portions of upper sidewall 24, to admit limited amounts of high pressure discharge gas to pass by the tips of the lobes of impeller 14 and thereby increase the pressure in the displacement cavities before they open into the discharge port 20, yet without allowing backflow of high pressure discharge gas into the intake port 18.

The lower sidewall portions 26a, 26b and 26c likewise extend together over an angular sector of somewhat less than 300 degrees, as measured from the lower edge of the intake port 18 to the lower edge of the discharge port 20. In this regard, it will be noted that in the preferred embodiment the size of the intake port 18 is larger than the size of the discharge port 20, which is a consequence of the gas being discharged from the discharge port 20 being at a higher pressure and lower volume than the gas drawn into the intake port 18.

From the intake port 18, the upper and lower cylindrically curved upstream sidewall portions 24a and 26a each extend, in the illustrated preferred embodiment, over an angular sector of approximately 128 degrees, which angular sector is slightly greater than the angle spanning two displacement cavities between any two successive pairs of lobes of the six-lobe rotors 12 and 14. Over this sector the sidewall portions 24a and 26a have a substantially cylindrical curvature, with a preferable tolerance of not more than two one thousandths of an inch between the outside lobe tips of the impellers 12 and 14 and the cylindrical surfaces of the sidewall portions 24a and 26a.

In contrast, the surfaces of the upper and lower downstream sidewall portions 24b and 26b of the housing 16 are at a greater distance from the axes of the impellers 12 and 14 than are the surfaces of the sidewall portions 24a and 26a, so as to provide a controlled clearance between the tips of the impeller lobes and the surfaces of sidewall portions 24b and 26b, in order to allow controlled amounts of internal reflux counterflow of high pressure discharge gas back into the displacement cavities between the lobes of the impellers 12 and 14.

It should also be recognized that, in accordance with the invention, the upstream sidewall portions 24a and 26a need only span an angular sector of at least 60 degrees in order to avoid any backflow of compressed discharge gas back into the intake port 18, while still allowing controlled reflux counterflow of compressed discharge gas into the displacement cavities formed between adjacent lobes of each rotor 12 and 14. Conversely, the upper and lower downstream sidewall portions 24b and 26b need only span an angular sector of at least 60 degrees from the upper and lower lips of the discharge port 20, respectively, in order to allow controlled reflux counterflow of compressed discharge gas back into at least one displacement cavity before it opens into the discharge port 20.

In the illustrated preferred embodiment, the transition sidewall portions 24c and 26c are centered at approximately the midpoint between the lips of the intake and discharge ports 18 and

20, or approximately 128 degrees from each of the upper and lowers lips of the ports 18 and 20, such that the angular sectors of the upstream sidewall portions 24a and 26a and the angular sectors of downstream sidewall portions 24b and 26b are approximately the same, i.e. approximately 128 degrees.

As noted, the surfaces of upstream sidewall portions 24a and 26a are essentially cylindrical so as to prevent backflow of compressed gas into the intake port 18. However, the surfaces of downstream sidewall portions 24b and 26b may be cylindrical, or may be of progressively increasing diameter from the axes of rotation of the impellers 12 and 14, as in the preferred embodiment. Depending on the level of reflux counterflow of compressed discharge gas desired at various points along the downstream sidewall portions 24b and 26b, the sidewall portions 24b and 26b may be cylindrical along nearly their entire span, or they may be of progressively increasing radius toward the discharge port 20. Further, the transition sidewall portions 24c and 26c may be either abrupt, or gradual as illustrated in Figure 1.

The lobed impellers 12 and 14 are essentially identical to one another, and their function during the operation of the compressor is as described further below. The six lobes of each of the impellers 12 and 14 are substantially identical to one another. In rotation, the lobes of impellers 12 and 14 intermesh in close contact with one another so that there is at all times a high impedance clearance between the impellers, which clearance is small in comparison with the volumetric displacement of the compressor, and which essentially restricts by sonic choking backflow of high pressure discharge gas through to the intake region.

Briefly, the impellers 12 and 14 are driven to rotate in opposite directions about their parallel axes of rotation. The axes of the impellers are also collinear with the central longitudinal axes of the cylindrically curved interior sidewall portions 24a and 26a, respectively. The impellers 12 and 14 are maintained in proper angular relationship to one another, which is at an angular phase relationship of 30 degrees with respect to one another, by their normal intermeshing relationship, and also by means of timing gears (not shown), which are located outside of the primary chamber of the housing 16.

In operation, gas is admitted to the compressor through the intake port 18 that is generally centered between the upper and lower side wall 24 and 26. Individual parcels of gas are swept through the housing 16 by the impellers 12 and 14, with each parcel occupying a displacement cavity which is defined by a pair of adjacent impeller lobes and by the interior walls of the

compressor housing 16. So long as the leading lobe of a displacement cavity is positioned adjacent sidewall portion 24a or 26a, the parcel of gas remains at the intake pressure. As soon as the leading lobe of the displacement cavity reaches sidewall portion 24b or 26b, a limited amount of higher pressure discharge gas begins flowing into the displacement cavity. Depending on the precise shape, sector span, and radii of the downstream sidewall portions 24b and 26b at various points along their surfaces, the rate and amount of reflux counterflow of compressed discharge gas back into the displacement cavity may be vary as the displacement cavity travels through the housing 16. By the time the displacement cavity opens into the discharge port 20, the pressure of the parcel of gas is increased, up to as much as the pressure of the gas in the discharge port 20, and the gas is thus swept into of the discharge port 20 with little or no adiabatic compression and associated heating.

It is believed that compressor of the present invention will find utility in serving a wide variety of applications where high volume, sustained operation is required at single stage pressure ratios of up to five to one (5:1). Inasmuch as Roots type compressors have heretofore only been capable of sustained operation at pressure ratios not exceeding approximately two to one (2:1) due to limitations imposed by overheating of the compressor components, the higher attainable pressure ratio capability of the present invention makes it useful in applications not previously considered feasible.

It will be appreciated that the temperature of the gas being processed is sufficiently reduced by the reflux counterflow of discharge gas that means of heat removal are not ordinarily required, either internal or external, and problems associated with overheating and thermal distortion are reduced. The compressor is characterized by having a more uniform process temperature, so that temperature differences in the transverse flow direction from intake to discharge do not cause thermal distortion difficulties. As a consequence of the substantially isothermal nature of the compression cycle, the reflux compressor has an inherent energy efficiency advantage when compared with other compression processes, an advantage that improves with increasing pressure ratios.

Although the present invention is described herein with reference to a preferred embodiment, it will be understood that various modifications, substitutions and alterations, which may be apparent to one of ordinary skill in the art, may be made without departing from

the essence of the present invention. Accordingly, the present invention is described by the following claims.