GB2073324A - Rotary gas-compressor - Google Patents

Abstract

The compressor is of the type having a pair of co-operating rotors (20, 22) with parallel rotation axes and dissimilar asymmetric lobed configurations. The rotor (22) has a profile (in cross-section) that is greater in area than that of the other rotor (20) and comprises a sector II of radius R, which is somewhat greater than the radius r of the outer edge of a discharge port (44). The gas flow is thereby improved. Balancing plugs (100) may be lodged in the rotors. Strips (70, 72) of a material e.g. a lead-tin alloy, that can be worn-down easily may be attached to the rotor (22) and the inner peripheral wall of the casing, Fig. 5 (not shown), to facilitate "lapping-in" of the rotors. Sealing strips made of said material may also be attached to the casing end-walls, Figs. 9a to 11 (not shown). The compressor may comprise two stages disposed one above the other, Figs. 1 (a)-(c) (not shown). <IMAGE>

Description

SPECIFICATION Rotary compressor This invention relates to the field of rotary compressors. More particularly, this invention relates to the field of rotary compressors wherein a pair of rotors are mounted in adjacent bores in a casing, each rotor having a single projecting lobe or tooth which engages a corresponding recess in the other rotor as the rotors counter-rotate. This invention especially relates to rotary compressors of this type which can be operated without the addition of lubricant and/or coolant fluid in the compressor chamber as is done in oil or water flooded compressors.

Dual rotor rotary compressors of the general type which are the subject of the present invention have been known in the art for many years. Generally speaking, the compressors include a pair of counter-rotating rotors mounted in cylindrical bores (i.e., cylinders) in a casing. Each rotor has a projecting tooth or lobe (and some designs have more than one) which engages a corresponding recess in the other rotor. The rotor tips are intended to seal against the interior walls of the cylinders, and parts of the rotor hubs are intended to seal against each other as they rotate relative to each other whereby a gas may be internally compressed as the rotors rotate. Gas inlet and discharge ports are provided which are sealed from each other during the rotation cycle so that internal compression can occur.

It has been suggested that a machine of this type could be provided with a pair of identical or essentially identical rotors which rotate in chambers or cylinders of equal size and shape. That is, the profile of each of the rotors and its projected tooth or teeth are identical. It is also known to have rotors of dissimilar size or shape, in which event the rotors may be mounted in dissimilar bores. It has also been proposed to mount rotors of dissimilar profiles in cylinders with identical size bores, but some such arrangements have suffered the serious deficiency that continuous fluid communication between the chambers is severely reduced or eliminated, and most or all internal compression must occur in only one of the chambers, thereby reducing the efficiency of the machine.

By way of example, some of the principal prior art known to the inventors may be seen in the following patents: Hupe 1,304,394 (U.K.), Northey (Lister) 900,881 (U.K.), Northey 752,437 (U.K.), Brown U.S.

3,535,060, Brown U.S. 3,472,445 and Northey U.S. 2,097,037. Other known relevant prior art is as follows: Northey 661,749 (U.K.), Rowlands et al 341,324 (U.K.), Weatherston U.S. 4,076,469, Weatherston U.S. 4,033,708, Weatherston U.S.

3,941,521, Weatherston U.S. Re. 29,627, Brown U.S. 3,723,031 and Northey U.S.

2,058,817.

A persistent problem in the design and construction of these rotary compressors is to maximize flow through the machine and maximize the discharge. capacity while avoiding high internal discharge velocities which cause severe losses, especially when axial inlet and discharge ports are used. Another problem involves the dynamic balancing of each rotor, especially when rotors of dissimilar profiles are used. Sealing problems have also been persistent in these machines. The sealing problems have included sealing between the rotors, sealing between the rotor tips and the cylindrical walls or chambers, and sealing between the ends of the rotors and the end covers or plates which enclose the cylinders of bores.

The purpose of the present invention is to overcome the above discussed problems of the prior art and to achieve an improved rotary compressor of the dual rotor type.

According to the invention, there is provided a rotary compressor of the type having shaped rotors rotating in intersecting cylindrical bores, the compressor including a first cylinder, a second cylinder, an inlet rotor in said first cylinder, said inlet rotor having a hub and a lobe rotating on a first shaft, a discharge rotor in said second cylinder, said discharge rotor having a hub and a lobe rotating on a second shaft, inlet means in at least said first cylinder for gas intake for compression, discharge means in said second cylinder for the discharge of gas after compression, said discharge means including a port in at least one end wall of said second cylinder having an outer circumferential arc of maximized radius "r" determined by the profile of a segment of said discharge rotor and an inner circumferential arc of minimized radius slightly larger than the radius of said secorid shaft, said discharge rotor having four quadrants starting from the tip of the lobe of said discharge rotor, and extending in sequence around said discharge rotor, the second of said quadrants being an arc of a circle of radius "R", said radis "R" being slightly greater than said radius "r" of said discharge port, and said radius "R" being selected to maximize the radius "r" of the discharge port consistent with the avoidance of unacceptably high flow velocities in said second cylinder, and the first and third quadrants each having a dimension "RE" at their junctions with the second quadrant end each having contoured profiles of other than the radius R for at least major portions of said first and third quadrants, and said inlet rotor and said discharge rotor being of different sizes and profiles, with said discharge rotor being of larger size-than said inlet rotor.

Bearing in mind that all rotary compressor designs are limited in the amount of speed and capacity that any one size is capable of achieving efficiently, a principal feature of the present invention is to maximize the exposed area of the discharge port (in a configuration where the rotor shape is used to time the opening of the discharge port). This maximization of the exposed area of the discharge port is achieved by a compressor design in which the discharge rotor (i.e., the rotor which cov ers and uncovers the axial discharge port) is of larger size than and has a different profile than the inlet rotor (i.e., the rotor which covers and uncovers axial and/or radial inlet ports).The discharge rotor profile has a section with a minimum radial dimension (R) which determines the outer limit of the discharge port and which is made as large as possible consistent with (1) the avoidance of unacceptably high internal flow velocities and discharge velocities and (2) the need to maintain a suficient size of the inlet or suction rotor to be able to balance it effectively. This maximization of discharge port area consistent with the two considerations itemized above is achieved by the use of rotors of dissimilar size and profile which, nonetheless, rotate in cylinders of equal size. The outer profile of the discharge rotor is made up of a series of segments, some of which are circular arcs and other segments of which are noncircular curves or profiles.A major circular arc segment of the discharge rotor having a radius R is selected to maximize the radial dimension of the axial discharge port, since the maximum radial dimension of the discharge port must be slightly less than the radial dimension R of the discharge rotor. Another circular segment of the discharge rotor is selected to be approximately equal to the radius of the cylinder in which the rotor is to rotate so that sealing can be effected between the rotor and the cylinder wall. Other profile segments of the discharge rotor are either determined by design considerations such as balancing of the rotor or merging or blending of profile segments or by being generated from mating surfaces of the inlet rotor.Similarly, the profile of the inlet rotor is determined to some extent by similar design considerations and by considerations of sealing with the discharge rotor and the cylinder wall or by being generated from mating parts of the discharge rotor.

Another principal feature of the present invention relates to rotor sealing. Predetermined patterns of crushable and/or abradable material are loated on the faces of the housings at each end of the cylinder and/or the end faces of the rotor for purposes of sealing the leakage space between them. Sealing between the tips of the rotors and the interior cylindrical wall is also achieved by the use of a similar crushable and/or abradable material on the tips of the rotors or the walls of the cyliner to both form a wear-in surface and also form a labyrinth seal. A solid coating or thin strips of crushable and/or abradable material may also be positioned on one or both rotors wherever the rotors are required to be in mating contact for sealing purposes.Since the rotors are of dissimilar size and shape, the dissimilar profiles result in a relative velocity between the rotors at their closest approach.

This relative velocity results in the coating material being abraded away so that optimum sealing surfaces are worn into the sealing strips on the rotor profiles. This sealing mechanism cannot be employed with rotors of identical profile, since the necessary relative, velocity between the rotors at the closest approach point does not exist for the majority of the profile.

Other features and advantages of the present invention will be apparent to and understood by those skilled in the art from the following detailed description and drawings, wherein: Figure 1(a) is a schematic representation of rotary compressor system, Figures 1(b) and 1 (c) show side and rear elevation views of rotary compressor casing and staging in accordance with the present invention, Figure 2 is a view taken along line 2-2 of Fig. 1 (a) showing the rotary compressor of the present invention, Figure 3 is a detailed view of the discharge rotor of the rotory compressor of Fig. 2, Figure 4 is a detaled view of the intake rotor of the rotary compressor of Fig. 2, Figures 5-8 show the rotary compressor with the rotors at different locations during a cycle of operation, Figure 9(a) shows details of a preferred seal structure on the face of the end housings, Figure 9(b) shows a variation on the seal structure of Fig. 9(a) and Figures 10 and 11 show details of the seal structure.

Referring first to Figs. 1 (a) and 1 (c) an overall system schematic and a staging and drive arrangement for a two stage rotary compressor system is shown. The system shown schematically in Fig. 1 (a) has a first stage rotary compressor 10 and a second stage rotary compressor 1 2. Rotary compressors 10 and 1 2 are conceptually identical, although they may differ in specific details of rotor size, profile or rotor axial length. Accordingly, only the essential elements of rotary compressor 10 will be described, and it will be understood that similar elements are present in rotory compressor 12, unless otherwise indicated. Rotary compressor 10 has an external casing 14 in which are located a pair of generally cylindrical dnd preferably identical and intersecting chambers or cylindrs 1 6 and 18. An inlet or suction rotor 20 is rotatably mounting in cylinder 16, and a discharge rotor 22 is rotatably mounted in a cylinder 1 8. End housings 24 and 26 close off the opposite ends of the cylinders. Rotors 20 and 22 are mounted, respectively, on shafts 28 and 30, which are, in. turn, driven to rotate at equal rotational speeds by gears 32 and 34.

Gear 32 drives gear 34, and gear 32 is driven by a bull gear 36 which is, in turn, driven by a motor 38.

Inlet air is introduced into compressor 10 in both an axial direction (i.e., in a direction generally parallel to the axes of rotation of the rotors), and in a radial direction (i.e., generally radial and perpendicular to the axes of rotation of the rotors 20 and 22). The axial and radial air inlets are shown schematically at 40 and 42 in Fig. 2 and are shown in more detail in Figs. 1(b), .1(c) and 2. The air taken into rotary compressor 10 is internally compressed and is then discharged at a first stage discharge pressure through two axial discharge ports 44 in end housings 24 and 26.The discharged first stage compressed air is delivered to a pulsation dampener 46 (which essentially functions to smooth out discharge pulses) and thence to interstage cooler 48 (where the temperature of the discharge air is reduced) and then to a water separator 50 (where normally present excess moisture may be removed). The first stage compressed air is then delivered via an axial inlet 40(a) (see Fig.

1 (c)) to second stage compressor 1 2 where it undergoes a second stage of internal compression and is then discharged to a second array of pulse dampener 46', cooler 48' and water separator 50' and is then delivered for its eventual intended end use. As indicated by the relative axial lengths of compressors 10 and 12, the axial length of the rotors in second stage compressor 1 2 is less than in compressor 10, because of the reduced volume of the compressed air delivered to compressor 12.

With regard to the previous description of Fig. 1, it will be understood that the elements of the two stage compressor system are shown only in schematic configuration for purposes of illustration, with no attention being paid to details such as rotor configurations and dimensional clearances. Also, the compressors will have rotor bearings, cooling jackets and other conventional structure, the details of which are not shown. It will also be understood that components such as the gears, motors, pulsation dampeners, coolers, and water separators may be conventional components known in the art. In actuality the elements of the compressor are arranged and staged as shown in Figs. 1 (b) and 1(c).

Referring to Figs. 1 (b) and 1 (c) it can be seen that the two compressor stages are arranged one above the other. That is, the first stage compressor 10 is positioned above the second stage compressor 1 2, as distinguished from conventional multi-stage systems where the compressor stages are located side by side. A single bull gear 36 drives both of the drive gears 32 for each compressor stage, and all of the gears are housed in common drive gear casing 37 so that the gears are all at one end of the assembly and the two compressor stages extend one above the other on one side of the gear case 37.The dampeners 46, 46', coolers 48, 48' and separators 50, and 50' are omitted from Figs. 1 (b) and 1(c) for purposes of clarity, but it will be understood that they would be incorporated as shown in Fig. 1(a).

The staging arrangement shown in Figs.

1(b) and 1(c) is important in that it achieves a very compact arrangement requiring less floor space and permitting a smaller overall size for a compressor enclosure than would otherwise be the case for a conventional side by side arrangement for compressors of similar capacity or rating. The width of the overall machine (i.e., the dimensions from ieft to right) can be reduced about 25% by the arrangement of stages one over the other rather than side by side. Additional savings can also be realized in the overall enclosure housing (i.e., the generally square or rectangular enclosure which would ordinarily house the entire compressor), since the space in front of the bull gear case can be used to locate discharge silencers or other accessory components.

Still referring to Figs. (b) and 1(c) it is to be noted that the first stage inlet ports 40 and 42 are located in a large plenum chamber 43 which may be a part of the inlet siliencer, and the second stage inlet 40(a) is part of a large plenum chamber 45. The location of the various inlets in the large plenum chambers 43 and 45 provides large inlet volumes right up to the actual inlet passages to minimize areas of high velocity air flow and thus reduce flow losses.

Referring now to Fig. 2, a cross sectional view taken along line 2-2 of Fig. 1 (a) is shown. As shown in Fig. 2, inlet or suction rotor'20 rotates in a clockwise direction, and discharge rotor 22 rotates counterclockwise.

In the position of the rotors shown in Fig. 2, both the inlet port and the discharge port 44 are shown partially uncovered, with the covered portions of each port being shown in dotted lines. It will be understood that discharge port 44 is covered and uncovered (i.e., closed and opened) by the action of discharge rotor 22 as it rotates, while suction port 40 is covered and uncovered by the combined action of inlet rotor 20 and discharge rotor 22.

Also, the communication of radial inlet port 42 with the interior of casing 14 is intermittently interrupted by suction rotor 20 as it slips past inlet portion 42 to begin each compression cycle.

Rotor 20 may be considered as being comprised of a hub portion 52 and a projecting lobe or finger 54. Similarly, rotor 22 may also be considered to be comprised of a hub portion 56 and a projecting lobe or finger 58.

The particular shapes and profiles of rotors 20 and 22 are of special importance in the present invention. The shape and profile of discharge rotor 22 is dictated, in part, by the desired objective of maximizing the area of discharge port 44. This maximization of the area of the discharge port is accomplished primarily by a design which achieves the maximum radial dimension "r" for the outer dimension of the discharge port. The dimension "r" is, in turn, determined by determining the maximum dimension "R" which can be accommodated in a hub portion of the discharge rotor, consistent with avoidance of excessive flow velocities and balancing requirements. In discussing the design considerations for the discharge rotor 22, it will be helpful to keep in mind that rotors 20 and 22 are designed to rotate in timed relation to perform the functions of air intake, compression and discharge.The rotors rotate at equal rotational speeds, and the tips of each of the fingers 54 and 58 meet twice in each revolution at the cusps formed by the junctions of cylindrical chamber sections 1 6 and 1 8.

In determining the contour of profile of discharge rotor 22, it will be advantageous to view the rotor in a series of 90 quadrants.

Quadrant I covers the 90 arc in the clockwise direction beginning at the tip of lobe 58.

Quadrant II covers the next 90 of arc, quadrant Ill the third 90 or arc, and quadrant IV the fourth 90' of arc. To avoid confusion in the drawing of Fig. 2, only quadrants I and II of discharge rotor 22 are labeled. All four of these quadrants are labeled in detail in Fig. 3 which shows an enlarged and detailed view of the profile of discharge rotor 22. Similarly, as shown in Fig. 4, intake rotor 20 may also be viewed as being formed of four quadrants starting from the tip of finger 54 and proceeding in the counterclockwise direction and being coordinated with the correspondingly numbered quadrants of the discharge rotor.

The relationship between the coordinated like numbered quadrants of the two rotors will be explained hereinafter.

Quadrant II of discharge rotor 22 will be discussed first (with particular reference to Fig. 3) because its dimension is critical to the design of the rotor and controls or influences several other dimensions. The portion of discharge rotor 22 included in quadrant II is the part of the discharge rotor which covers the radial extremity of discharge port 44 at the beginning of the compression part of a cycle of operation of the rotor (when the pointed tips of lobes 54 and 58 are at the upper cusp of casing 14 as shown in Fig. 5 and are moving away from each other). Thus, consistent with the objective of maximizing the radial dimension of the discharge port, the profile of quadrant II of rotor 22 is an arc of a circle.The radial dimension R which defines the circular arc of quadrant II is selected to be a maximum consistent with the need to avoid narrow passages with high flow losses between the rotor and cylinder and the need to have a suction rotor shape capable of being balanced. Another constraint on the size of radius R is the need to have adequate suction volume remaining in the cylinder bores. The radial dimension "r" of the discharge port is slightly less than dimension R of the rotor. It is to be noted that when the present invention speaks about increasing and maximizing the dimensions r and R, the frame of reference is the values which those dimensions would have in a compressor of the same casing size if the inlet and discharge rotors were essentially the same size and shape.

The surface of quandrant Ill of rotor 22 in the disclosed embodiment is then made in the form of an elipse, but other contors may be used. Then the formula for the elipse being determined by the volume of material desired in that part of the rotor hub to provide rotational dynamic balance and also by determining the desired spacing between the largest dimension (at point p,) of that segment of the rotor and the cylinder wall, which spacing determines flow velocity in the compressor.

Quadrant I of discharge rotor 22 is made of two parts, the first being a circular arc 12" duration extending from the tip of lobe 58 in the clockwise direction, and the remaining 78 of arc is a modified elipse, with the elipse formula being modified to get equal slopes at the junction between the eliptical portion and 12" of circular arc. the 12" of circular arc have a radius R1 which is essentially equal to the internal radius of the cylindrical portion of casing 14. The 12" of that arc are selected to be qual to the arc length of radial inlet port 42.Also, the relationship between the modified elipse and the circular segment having the radius R at the junction of quadrants I and II is that the elipse be at a distance R from the rotor center and be perpendicular to a radius R at the junction with the curve of radius R of quadrant II to provide equal slopes and a smooth profile at the junction of quadrants I and Il.

Quadrant IV of discharge rotor 22 is generated, i.e., determined, by surfaces on intake rotor 20 and will be explained in more detail after discussion of the determination of the shape of intake rotor 20.

Referring now to inlet rotor 20 as shown in Figs. 2 and 4, it will be noted that inlet rotor 20 is smaller than discharge rotor 22 as a direct result of the fact that discharge rotor was increased in size to maximize the area of the discharge port as previously described.

Starting with quadrant Il of the inlet rotor, which coordinates with quadrant II of the discharge rotor, the profile of quadrant II of the inlet rotor is a circular arc of radius R2 equal to the difference between the separation "S" between the axes of the rotors 20 and 22 and the dimension R of rotor 22. The surfaces of quadrants Il of each rotor are in sealing contact with each other during each cycle of the compressor, and these surfaces have sliding motion relative to each other (with no rolling component). Quadrant Ill of intake rotor 20 is a surface which is generated, i.e., fully determined, by the eliptical shape of the surface of the quadrant Ill of discharge rotor 22.The surfaces of quadrant Ill of each rotor are also in'sealing cpntact with each other during each cycle of the compressor, and these surfaces have sliding motion relative to each other. (Sealing contact refers to a condition of actual contact or a very small clearance on the order of .001-.002 inches which is obtained after wear-in of abradable seals, as will be explained in more detail hereinafter).

With regard to quadrant I of inlet rotor 20, the first portion, i.e., the part extending coun te?clockwise from the beginning of the quadrant for about 52 to the point marked P2 is arbitrary so long as it does not interfere with the corresponding 52 of arc of quadrant I of the discharge rotor. This section of the inlet rotor can be somewhat arbitrary, because it does not'have to effect any sealing contact with the coordinate part of the discharge rotor. Similarly, the remaining 38 portion of the arc of quadrant I of the inlet rotor is a surface generated by the modified elipse which forms the remaining coordinate 38 portion of quadrant I of the discharge rotor.

Referring now to quadrant IV of the inlet rotor 20, the portion starting at the pointed tip of lobe 54 and extending in a clockwise direction is a 12" arc of a circle of radius R i.e., a radius equal to the internal radius of the cylindrical section 1 4 in which the rotor is located (and equal to R1 of quadrant I of rotor 22). A 12" arc length is selected to match the 12" arc length of the opening of radial inlet 42.The remainder of this fourth quadrant of the intake rotor proceeding in the clockwise direction is defined by two circular arcs; one circular arc having its center at a point P3 on the radius R, at the 12" location and having a radius R3 and extending for about 98 to P4 on the rotor surface; and the other circular arc having a radius R4 which is centered on the Y ordinate which defines the end of qadrant IV.

It is, of course, important that the circular arcs defined by R1, R3 and R4 have equal slopes where the arcs join to preserve profile continuity.

Quadrant IV of the discharge rotor 22 is a fully generated surface determined entirely by the shape of the quadrant IV of inlet rotor 20.

The first 78 of the surface profile of quadrant IV of the discharge rotor is generated directly by the portions having arc lengths determined by the radii R3 and R4 of the surface profile of quadrant IV of the inlet rotor. Similarly, a last 12" of quadrant IV of the discharge rotor is a circular arc of appropriate radius to slide on the 12" circular arc of the first section of quadrant IV of the inlet rotor (i.e., the arc length of the tip of lobe 54 at the end of quadrant IV of the inlet rotor determined by the 12" arc of R1).

The entire inner tooth or lobe recess profile of the discharge rotor, i.e., the recessed portion extending from the tip of tooth 58 to point P5 at the end of the quadrant IV is a generated surface determined by the path of the tip of tooth 54 as it passes along that surface of the discharge rotor during rotation.

That similar recessed inner tooth or lobe profile of intake rotor 20 extending from the tip of tooth 54 to point P5 at the beginning of quadrant I is a completely noncritical profile from the standpoint of rotational mating.

Other than for the fact that this recess profile on the intake rotor must be shaped to avoid any interference with the profile of quadrant I of the discharge rotor as it sweeps past this recess profile on the intake rotor, this recess profile on the intake rotor is arbitrary and may be determined by balance, weight and strength considerations.

As previously indicated, the avoidance of excessive flow velocities and attendant high losses is a limiting factor in maximizing radius dimension R of quadrant II of the discharge rotor which, in turn, limits the radius r of the discharge port. Radius R of quadrant II of the discharge rotor determines the start of the eliptical shape of the profile of quadrant Ill which is desirable for balancing purposes.

That eliptical shape is perpendicular to a radius R at the end of quadrant li and is limited in its outward projection from the center of the rotor by the radial gap which exists between point P1 on the surface of the discharge rotor in the quadrant Ill and the inner surface of the casing of chamber 1 8 during the compression and discharge parts of the cycle of rotor operation (point P1 defining a point on the surface of the discharge rotor in quadrant Ill which has the largest radial separation from the center of rotation of the rotor). The radial gap betwee point P, and the inner surface of the casing of chamber 1 8 defines a throat which determines the velocity of at least some of the gas in the compressed state as it moves from space between the rotors toward the discharge port during compression and/or discharge. The smaller the radial gap between point P, and the chamber wall is permitted to become, the greater will be the gas velocity, and the greater will be the flow losses encountered. Thus, a compromise must be effected between the desired enlargement of the discharge port and the increasing flow losses which are encountered as the discharge port, and hence the required size of the-discharge rotor, are increased.

Another compromise feature which must be effected is dictated by balancing considera tions. As the radius R of quadrant II of the discharge rotor is increased, the radius R2 of the coordinate mating quadrant II of the intake rotor must be correspondingly decreased, and the base or root of lobe 54 in the vicinity where R3 and R4 merge in quadrant IV of the intake rotor must also be decreased. As the radius R2 and the root of lobe 54 of the intake rotor decrease in size, it becomes increasingly difficult and eventually impractial to balance the intake rotor and the strength of the intake rotor across the base or root of the lobe becomes jeopardized. Thus, enlargement of the discharge rotor is also limited by the attendant considerations of maintaining the intake rotor large enough easily balance intake rotor and by strength considerations.

Turning now to Figs. 5-8, various positions.

of the rotors and various modes of a cycle of the compressor are shown. In Fig. 5, the rotors have just sealed off the inlet ports from the chamber volumes 16 and 1 8 between the outer peripheries of the rotors and the inner walls of the chambers. In the position shown in Fig. 5, the chambers 1 6 and 1 8 are full of air or gas at inlet pressure, and the compression cycle is about to begin. Referring now to Fig. 6, as the rotors continue to rotate in the indicated directions, the volume between the advancing surfaces of the rotors is being constantly reduced, whereby the volume of gas in that space is compressed. The space behind the rotors is, of course, expanding and filling with gas at inlet pressure for the next cycle.

The discharge port is sealed by parts of quadrants lil and IV of rotor 22. The circular arcs at the tips of each of the teeth of the rotors 54 and 58 seal against the inner walls of the chambers to seal the compressing volume between the advancing surfaces of the rotors from the inlet pressure behind the rotors, and the mating surfaces between the rotors where they abut similarly seal the area of reducing volume and increasing pressure from the inlet pressure.

In the position shown in Fig. 6, the compression cycle is almost, but not quite, completed and discharge port 44 is about to open as discharge rotor 22 continues to rotate in the counterclockwise direction. When the discharge port is first uncovered by the trailing edge of the rotor in quadrant IV, the gas has not yet been fully compressed to discharge pressure. Thus, gas flows back into the compressor from the pulse dampener momentarily as compression continues in the compressor.

As rotor 22 continues to rotate counterclockwise, internal compression of the gas continues and more of the discharge port is uncovered. When full discharge pressure is reached, gas then flows out of the discharge port to the pulse dampener. This tactic of opening the discharge port early (i.e., before compression is complete) makes it possible to achieve a larger exposed discharge area at the beginning of the-discharge process with a resultant reduction in overall discharge losses.

Referring now to Fig. 7, the two rotors have rotated to the position where they are at the cusp formed at the lower part of the casing at the intersection of chambers 1 6 and 1 8. In the position shown in Fig. 7 the rotors are in sealing contact along portions of the contours in segments IV of each rotor quadrant to seal the inlet from the discharge, and the pressurized gas or air in the space between the rotors is being discharged through discharge port 44 as the closed space between the rotors continues to be reduced in volume as rotation of the' rotors continues.

Fig. 8 shows the rotors further advanced in rotation, and Fig. 8 illustrates that the rotors continue to be in sealing engagement along surfaces of quadrants IV of each rotor as the trapped space between the rotors continues to be reduced in volume and all of the pressurized air is forced out of the discharge port. As the rotors continue to rotate, all of the trapped and compressed air in the diminishing volume between the rotors will be delivered through the diminished open profile of the discharge port, and there will be essentially no communication between the inlet port and the discharge port. As the rotors continue to rotate beyond the position shown in Fig. 8, the discharge port will be fully closed by quadrants I and II of rotor 22.As the rotors continue to rotate back to the position shown in Fig. 5, the intake or suction port remains open while the discharge port has been closed off. When the rotors return to the position shown in Fig. 5, the suction port will again be closed off from the compression space, and another cycle of compression and discharge will be initiated.

It is important to note that during the compression cycle the circular arc segments at the tip of each of the lobes 54 and 58 seal against the inner walls of their respective casings, and the corresponding parts of the rotor profiles in corresponding quadrant sections will be in sealing engagement between the axes of rotation of the two rotors, whereby the inlets will always be sealed from the discharge port to prevent any backflow. This sealing between the corresponding points of rotor profiles occurs at or near the point where the rotor profiles intersect an imaginary line between the axes of the rotors as the rotors travel through their paths of rotation.

In order to optimize the efficiency of the compressor, it is very important to have an effective sealing system in the compressor. To that end, the present invention incorporates a sealing system which includes seals between the counter-rotating rotors, seals between the rotors and the cylinder walls, seals for the axial shafts of the rotors, and seals between the end faces of the rotors and the faces of the end housings.

Figs. 3 and 5 show the sealing system between the two rotors and between the rotors and the cylinder walls. Sealing between the two rotors is accomplished by strips of abradable material 70 on the surface of discharge rotor 22. These strips of abradable material, which could alternatively be on suction rotor 20, are spaced apart on the outer surface of the rotor profile and extend from the beginning of quadrant II through all of quadrant II and quadrant Ill and through about 60 of quadrant IV. These strips of abradable material extend the full axial depth of the rotor profile (i.e.,, for the full length of the rotor between end housings). These strips of abradable material are preferably solder (such as a 50 lead 50 tin solder), or they may be teflon or a nickel graphite coating or some similar materials.The important feature is that these strips of material are relatively soft and abradable so that they can be easily worn by the sliding action between rotor 22 and rotor 20 at the closest point between the two rotors (i.e., where the rotors meet for sealing on or near the line between the axes of the rotors).

The strips of abradable material extend over quadrants II and lil and about 60 of quadrant IV, because those are the areas of engagement between rotor 22 and the corresponding parts of rotor 20 where differential pressure between the compressed gas and suction make it necessary to effect sealing between the two rotors during a cycle of operation. Because of the differences in size and profile between rotors 20 and 22, the rotors always have sliding action relative to each other rather than rolling action at the point of contact, and this relative sliding motion is necessary and extremely effective in promoting the "wearing in" of the abradable strips of effect sealing between the rotors.

Referring now to Fig. 5, the mechanism for sealing between the ends of the rotor tips and the cylinder walls is shown. Seal strips 72 of abradable material like strips 70 are spaced about the inner wall of each cylinder from about the six o'clock position (as shown in Fig. 5) to the cusp between the intersecting cylinders near inlet 43. Strips 72 also run the full axial length or depth of the cylinders between the end housings. As rotors 20 and 22 rotate, the ends of lobes 54 and 58 adjacent to the cylinder walls rub against the abradable strips 72 to wear in the strips and form sealing surfaces between the rotor tips and the sealing strips. This "wearing in" occurs by rubbing action between the 12" circular arc at the end of each rotor tip and cylinder walls. The wearing in is accomplished by the rotor tips, in effect, cutting into the strips 72.This cutting in is accomplished very effectively by the pointed tip of lobe 58 because its direction of rotation drives the tip into the strips 72. However, the desired cutting action will not be accomplished as effec tively at the tip of lobe 54, because the pointed edge of lobe 54 is dragged across the strips 72 with the pointed tip trailing the direction of rotation. Therefore, a cutting notch 73 (see also Fig. 4) is formed in profile surface of rotor 20 at the end of the 12" circular arc. This cutting notch extends the full axial length of the rotor profile, and it serves as a cutting edge to cut sealing surfaces in the strips 72 as rotor 20 moves past the strips 72.

Alternatively, seal strips could be located on the 1 2 circular arc at the end of each rotor tip, and these. strips could seal against the cylinder walls (which would not have seal strips).

Referring now to Figs. '9(a) and 10, the seal mechanism between the end faces of the rotors and the inner faces of end housings 24 and 26 are shown. The sealing mechanism between each of the opposed rotor faces and its adjacent end housing is the same, so only one will be described, and it will be under stood that the same structure applies in both cases. Fig. 9 shows end housing 24. The end housing has a pair of shaft holes 76 and 78 to receive shafts 28 and 30. Surrounding each shaft opening is an annular recess in the end housing. A seal element 80 is positioned in each of the recesses, the seal elements 80 each being secured in place by four counter sunk screws 81. The upper surface of seal element 80 has a pattern of concentric rings of seal material 82.The concentric rings 82 are abradable material such as solder, etc. as with seal strips 70 and 72. The concentric rings 82 define recesses 84 between the concentric rings 82, so that there are, in effect, a series a labyrinth steps or seals around the shaft openings to trap any leaked air and minimize leakage. A pattern of sealing strips 85, also of the same type of abradable material, radiates outwardly from each of the concentric ring structures around the shaft openings, and the discharge port is also sur rounded by seal strip structure. The seal strip structure around discharge port 44 includes a first seal strip 86 which extends from the concentric ring structure in the vicinity of the shaft seal arrangement along the side con tours of the discharge port and the circumfer ential arc of the outermost radial extent of the discharge port. The discharge port seal struc ture also includes a second seal strip 88 spaced slightly outwardly from the first strip 86 and following the pattern of the first strip 86, and the second strip 88 has a series of sawtooth or V-shaped seal components 90 radiating outwardly from the span of strip 88 around the outer circumferential contour of discharge port 44. The end faces of the rotors 20 and 22 bear against all of the seal strip structures shown in Fig. 9 to wear in the sealing surfaces and establish effective sealing between the end faces of the rotor and the sealing strips on the end housings. Seal strips 85, 86, 88 and 90 are all applied to the surface of end housing 24.

Referring to Fig. 9(b) a modification of the seal structure of Fig. 9(a) is shown 'wherein the concentric rings 82 are interrupted by radial bars 83 to form a series of separated pockets or dams to further aid in trapping any leaked air. The securing screws 81 have been omitted from Fig. 9(b).

Referring to Fig. 10, an important detail of the concentric circle seal structure around the shaft openings 76 and 78 is shown. The concentric circular seal strips 82 are formed in the shape of trapezoids with inclined slides sloping outwardly at an angle of about 15 (i.e., forming an angle of about 15 (i.e., forming an angle of 75 with the face of the end housing). Mating concentric grooves 94 are formed in each of the rotor faces. These grooves 94 are concentric circles around the rotor shafts, and these concentric grooves 94 are positioned to engage and mate with the raised strips 82 on the end housing.The concentric grooves 94 are formed with inclined sides sloping outwardly at an angle of about 5 . Thus, when a strip 82 mates with a groove 94, the strips will penetrate only part way into the groves and form a sealing engagement between the differently angled inclined sides of the strips and the grooves. The concentric groove structure around the shaft on the rotor faces wears into the seal rings 82 and/or bars 83 on the end housing to form an effective sealing arrangement which constitutes, in effect, a series of labyrinth seals. If the rotor is displaced away from the end housing, the leakage gap opened between the inclined faces of the seal is always less than the amount of rotor axial displacement, thereby maintaining good sealing over a wide range of operative conditidns.

In all cases, the seal structure is in the form of relatively soft abradable or crushable material such as solder, etc. discussed above. That material is readily 'worn by the action of the rotors to form effective seals shaped in accordance with the shapes of the various surfaces of the rotor elements, including any irregularities. After wear in, the height of the seal strips may be in the range of 0,025-0,05 mm. The sealing system permits the various elements of the compressor, such as the rotors, cylinders, and end housings, to be made to a looser tolerance than might otherwise be possible and operate over a wider temperature range without encountering either mechanically harmful rotor contacts or excessive leakage which causes both performance deterioration and a dangerous build up of heat and expansion capable of causing a machine seizure.The sealing strips 70 on rotor 22 form an effective seal between the moving rotors while avoiding direct contact between the rotors.

Although the eal atrip may have a height little as 0,025-0,05 mm after wear in, they serve not only to effect a seal between the rotors at the closest point between the rotors, but also they reduce the leakage gap between the rotors on either side of the closest approach point to provide an additional sealing effect.

Because of the presence of the elaborate seal structure, the compressor can, during break in periods of operation, the run under conditions which cause heating of the rotors beyond that which would be expected in normal operation. This will result in crushing .

and abrading of the seals during break in; and will lead to running with nominal clearances during normal operation. During normal operation, the performance of the compressor will be more predictable because the seals have been formed to their final shape during the break in period. Also, no seal material will enter the fluid stream during normal operg- tion. As indicated, the heights of the seal strips may be as little as 0,025-0,05 mm when worn in. The overall height and width of the seal elements may vary from position to position within the machine, and the dimensions of the seal elements should be controlled to values such that the number of elements being crushed or abraded during break in does not impose harmful mechanical or thermal loads on the rotors, bearings or other components of the compressor.

Referring again to Figs. 2, 3 and 4, structure for balancing of the rotors is shown.

Holes are drilled and tapped through each rotor (parallel to the axis of each rotor). These holes are then capped with threaded plugs 100 which are secured by staking or other means. The balancing holes are drilled completely through the rotor in the axial direction, and the plugs are inserted in each opposed end face of the rotor, and the outer surface of each plug is made to be flush with the face of the rotor with spaces between the opposed plugs. Three of the plugs 100 are shown positioned in one face of rotor 22, and four of the plugs 100 are shown positioned in one face of rotor 20.

As shown in Figs. 2, 3 and 4, the plugs may be of different sizes, and they are located about the axis of each of the rotors in what- ever manner may be required for general balancing of each particular rotor. Preferably, two plugs on each face of each rotor are on mutually perpendicular axes (see plugs 100(a) and 100(b) on rotor 20 and plugs 100(c) and 100(d) on rotor 22). Balancing of the rotors may then be fine tuned by changing the weight of either or both of the mutually perpendicular plugs, either by replacing a plug with another plug to add weight or by removing material from the inside of a plug. In this way, a positive or negative change in centrifgual force can be created as may be required to fine tune the balance of each rotor. Loca tion of the fine tuning plugs on mutually perpendicular axes simplifies the calculations required for determining plug weights to achieve fine balance. By having fine tuning plugs on mutually perpendicular axes a change in one plug has no effect on the other axis. Thus fine tuning can be effected merely by adjustment of the plugs on the mutually perpendicular axes.

Claims (34)

1. A rotary compressor of the type having shaped rotors rotating in intersecting cylindrical bores, the compressor including a first cylinder, a second cylinder, an inlet rotor in said first cylinder, said inlet rotor having a hub and a lobe rotating on a first shaft, a discharge rotor in said second cylinder, said discharge rotor having a hub and a iobe rotating on a second shaft, inlet means in at least said first cylinder for gas intake for compression, discharge means in said second cylinder for the discharge of gas after com- pression, said discharge means including a port in at least one end wall of said second cylinder having an outer circumferential arc of maximized radius "r" determined by the profile of a segment of said discharge rotor and an inner circumferential arc of minimized radius slightly larger than the radius of said second shaft, said discharge rotor having four quadrants starting from the tip of the lobe of said discharge rotor and extending in sequence around said discharge rotor, the second of said quadrants being an arc of a circle of radius "R", said radius ''R'' being sightly greater than said radius "p" of said discharge port, and said radius "R" being selected to maximize the radius "r'' of the discharge port consistent with the avoidance of unacceptably high flow velocities in said second cylinder, and the first and third quadrants each having a dimension "RE" at their junctions with the second quadrant and each having contoured profiles of other than the radius "R'' for at least major portions of said first and third quadrants, and said inlet rotor and said discharge rotor being of different sizes and profiles, with said discharge rotor being of larger size than said inlet rotor.

2. A rotary compressor as claimed in claim 1, wherein said first and second cylinders are of equal dimensions.

3. A rotary compressor as claimed in claim 1 or 2, wherein said first quadrant of said discharge rotor extends from the tip of said lobe of said discharge rotor and has a first portion in the form of an arc of a circle having a nominal radius R, about equal to the radius of the second cylinder and a second portion in the form of an elipse, said elipse being modified if necessary to blend the slope of said circular arc portion with said elipse, and said elipse having the dimension "R" at the end of said first quadrant and being perpendicular-to a radius "R" at the end of said first quadrant, the third quadrant of said discharge rotor is an elipse, which is determined at least in part by factors of dynamic balance and flow velocities, and the fourth quadrant of said discharge rotor is generated by mating surfaces on said inlet rotor.

4. A rotor compressor as claimed in any of claims 1 to 3, wherein said inlet rotor has four quadrants coordinated with the four quadrants of the discharge rotor, the first quadrant of the inlet rotor has a first part which is arbitrary as long as it does not interfere with the corresponding portion of the first quadrant of the discharge rotor and a second part which is generated by part of the elipse of the second portion of the discharge rotor, the second quadrant of the inlet rotor is an arc of a circle of radius "R2" equal to the difference between distance between the axes of said first and second rotors and the dimension "R" of said second quadrant of said discharge rotor, the third quadrant of the inlet rotor is generated by the elipse of the third quadrant of the discharge rotor, and the fourth quadrant of the inlet rotor is made up of three circular arcs of radius "R4", "R3,, and "R,", the arc of radius "R1" extending from the tip of the lobe of the inlet rotor and being about equal to the radius of the first cylinder, the arc of the radius "R3,, having its center on a radius of the arc of radius ''IR," and the arc of the radius "R4" centered on a line which defines the junction between the third and fourth quadrants.

5. A rotary compressor as claimed in claim 4, wherein said arcs of said fourth quadrant of said inlet rotor have equal slopes at the junctions between arcs.

6. A rotary compressor as claimed in any one of claims I to 5, including sealing means between saidl rotors and sealing means between said rotors and said cylinders.

7. ' A rotary compressor as claimed in claim 6, wherein said sealing means between rotors includes strips of abradable material on at least part of the profiles of each rotor, said strips of abradable material being worn-in during operation to effect sealing between the rotors.

8. A rotary compressor as claimed in claim 6, wherein said sealing means between said rotors and said cylinders includes strips of abradable material on the walls of said cylinders, said strips of abradable material being worn in by the tips of said rotors to effect sealing between said rotors and said cylinders.

9. A rotary compressor as claimed in claim 8, including a cutting notch in one of said lobes to engage said strips as said one lobe rotates.

10. A rotary compressor as claimed in claim 6, wherein said sealing means between said rotors and said cylinders includes pat terns of abradable seal strips on opposed end plates of said cylinders said end plates being immediately adjacent to end faces of said rotors, said pattern of abradable material on each end plate including an annular array of a plurality of spaced apart seal strips around the shaft opening in each end plate, and said pattern of abradable material includes patterns of strips of said material radiating outwardly from each annular array.

11. A rotary compressor as claimed in claim 10, including a sawtooth pattern of abradable seal material around at least part of said discharge port.

1 2. A rotary compressor as claimed in claim 10, including a plurality of grooves on said rotor faces for mating engagement with the strips of the annular array of strips around the shaft openings, said strips having.inclined sides of a first slope and said grooves having inclined sides of a second slope.

1 3. A rotary compressor as claimed in any one of claims 1 to 12, including a plurality of balancing holes in opposed face surfaces of said rotor, and balancing plugs in said balancing holes.

14. A rotary compressor as claimed in claim 13, wherein at least two of said balancing holes are on mutually perpendicular axes, and the plugs for said two balancing holes are selectively variable to effect fine balancing of said rotor.

1 5. A rotary compressor as claimed in claim 14 including at least one balancing hole with plugs therein in addition to said two balancing holes.

1 6. A rotary compressor as claimed in claim 14, wherein one of said mutually perpendicular axes runs from the tip of said lobe through the axis of rotation of said rotor.

1 7. A compressor rotor as claimed in claim 16, wherein said two balancing holes are in said hub of said rotor.

1 8. A rotary compressor of the type having counter-rotating rotors in intersecting cylinders the rotors each having a hub and at least one lobe, comprising a seal structure including strips of abradable material on at least part of the profile of at least one rotor, said strips of abradable material being worn-in during operation to effect sealing between the rotors, and means for sealing between said rotors and said cylinders.

1 9. A rotary compressor as claimed in claim 18, wherein said sealing means between said rotors and said cylinders includes strips of abradable material on the walls of said cylinders, said strips of abradable material being worn in by the tips of said rotors to effect sealing between said rotors and said cylinders.

20. A rotary compressor as claimed in claim 19, including a cutting notch in one of said lobes to engage said strips of abradable material on the walls of the cylinders as said one lobe rotates.

21. A rotary compressor as claimed in claim 18, wheren said sealing means between said rotors and said cylinders includes patterns of abradable seal strips on opposed end plates of said cylinders, said end plates being immediately adjacent to end faces of said rotors, said pattern of abradable material'on each end plate including an annular array of a plurality of spaced apart seal strips around a shaft opening in each end plate, and said pattern of abradable material including patterns of strips of said material radiating outwardly from each annular array.

22. A rotary compressor as claimed in claim 21, wherein said rotary compressor has a discharge port in an end plate of a cylinder, the seal structure including a sawtooth pattern of abradable seal material around at least part of said discharge port.

23. A rotary compressor as claimed in claim 21, including a plurality of grooves on opposed rotor faces for mating engagement with the strips of the annular array of strips around the shaft openings, said strips having inclined sides of a first slope and said grooves having inclined sides of-a second slope.

24. A rotary compressor as claimed in claim 18, wherein said sealing means between said rotors and said cylinders includes strips of abradable material on the lobe adjacent the tip of the lobe.

25. A compressor rotor including a hub adapted to be mounted on a shaft for rotation, a lobe on said hub, said hub and lobe having opposed face surfaces, a plurality of balancing holes in opposed face surfaces of said rotor, and balancing plugs in said balancing holes.

26. A compressor rotor as claimed in claim 25, wherein at least two of said balancing holes are on mutually perpendicular axes, and the plugs for said two balancing holes are selectively variable to effect fine balancing of said rotor.

27. A compressor rotor as claimed in claim 26, including at least one balancing hole with plugs therein in addition to said two balancing holes.

28. A compressor rotor as claimed in claim 26, wherein one of said mutually perpendicular axes runs from the top of said lobe through the axis. of rotation of said rotor.

29. A compressor rotor as claimed in claim 28, wherein said two balancing holes are in said hub of said rotor.

30. A multi-stage rotary compressor including a first stage rotary compressor, a second stage rotary compressor, said first and second stage rotary compressors being arranged with one stage positioned above the other stage, and common drive means, connected to each of said first and second stage rotary compressors to drive said compressors.

31. A multi-stage rotary compressor as claimed in claim 30, wherein said first stage compressor has first and second rotary elements driven by first and second interengaged timing gears, said second stage compressor has first and second rotary elements driven by first and second interengaged timing gears, and said common drive means is a drive gear connected to one of each of said timing gears in each stage.

32. A multi-stage rotary compressor as claimed in claim 31, wherein said first and second gears of one stage are above corresponding first and second timing gears of the other stage, and said drive gear is positioned to one side of said timing gears.

33. A rotary compressor substantially as hereinbefore described, and as illustrated in the accompanying drawings.

34. A multi-stage rotary compressor, substantially as hereinbefore described, and as illustrated in the accompanying drawings.