US3535060A - Rotary displacement machines - Google Patents

Rotary displacement machines Download PDF

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US3535060A
US3535060A US3535060DA US3535060A US 3535060 A US3535060 A US 3535060A US 3535060D A US3535060D A US 3535060DA US 3535060 A US3535060 A US 3535060A
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rotor
tooth
hub
profile
rotors
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Arthur E Brown
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Ingersoll Rand Co
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Arthur E Brown
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    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01CROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
    • F01C1/00Rotary-piston machines or engines
    • F01C1/08Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing
    • F01C1/12Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing of other than internal-axis type
    • F01C1/123Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing of other than internal-axis type with tooth-like elements, extending generally radially from the rotor body cooperating with recesses in the other rotor, e.g. one tooth

Description

A. E. BROWN ROTARY DISPLACEMENT MACHINES Oct. 20, 1970 6 Sheets-Sheet 1 Filed March 21. 1969 Oct. 20, 1970 A. E. BROWN 3,535,060
ROTARY DISPLACEMENT MACHINES Filed March 21,, 1969 FIG. IV 27 6 Sheets-Sheet '3 Oct. 20, 1970 A. E. BROWN ROTARY DISPLACEMENT MACHINES 7 6 Shets-Sheet 4 Filed March Zl, 1969 Oct. 20, 1910 A, 5, BROWN 3,535,060
ROTARY DISPLACEMENT MACHINES I 6 Sheets-Sheet 5 Filed March El, 1969 FIG. |x
United States Patent 3,535,060 ROTARY DISPLACEMENT MACHINES Arthur E. Brown, 117 E. th St., Corning, N.Y. 14830 Continuation-impart of application Ser. No. 719,360, Apr. 8, 1968. This application Mar. 21, 1969, Ser. No. 809,309
Int. Cl. F04c 17/04, 17/10, 27/00 US. Cl. 418114 23 Claims ABSTRACT OF THE DISCLOSURE A rotary compressor, vacuum pump, or expansion engine. Two interengaging rotors rotate within a casing structure. Each rotor has a hub and a tooth. One of the rotors has a disk located at each end of the rotor. The disks control the flow of the working fluid through ports in the end walls of the casing. The ports have outer radii equal to that of one rotor tooth and therefore these ports have large areas (high capacity) for flow of the Working fluid. The rotors and the outlet ports have configurations such that the clearance volume can be zero (or near zero) when operated as a compressor, so that nearly all of the compressed gas is delivered to the outlet ports. The single tooth rotors are internally balanced.
This application is a continuation-in-part of my 00- pending application, Ser. No. 719,360, filed Apr. 8, 1968, now Pat. No. 3,472,445.
DISCUSSION OF PRIOR ART U.S. Pats. 2,058,817 and 2,580,006 and Great Britain 661,749 show machines wherein disks are located at the ends of rotors so as to control ports in the end walls. In said patents, the rotors and ports are shaped in such a way that during each cycle there is a substantial volume of gas which is compressed but not delivered to the outlet port. Instead, the non delivered gas is throttled back to inlet pressure resulting in a loss in capacity and overall efficiency. In said prior patents, the ports do not extend to the outer radius of the rotor and therefore maximum port area (with minimum flow loss) is not secured.
A problem with screw type compressors such as in US. Pat. 2,287,716 is that at the discharge end of each thread space, there is a pocket of gas which becomes trapped and would undergo high compression, if not for bleed-oft schemes.
OBJECTS AND ADVANTAGES OF THIS INVENTION (1) The primary object is to secure more flow area through the end ports 25 and 61; and thereby permit higher r.p.m. and/or longer rotors compared with the machines shown in said copending application. The rotors in this present application are (by necessity) more complex than those in the copending application.
The machines disclosed in the earlier application are suited for a lower capacity range (because of their simpler construction) whereas the machines disclosed in this present application are suited for a higher capacity range.
(2) Clearance volume in a rotary compressor is defined as a volumetric space into which gas is compressed but not delivered to the discharge line. In a rotory compressor of the class described, it is desirable to secure a minimum (or zero) clearance volume because the gas in such a clearance volume is throttled back to inlet pressure (instead of being reexpanded). Such a throttling process causes a loss in bothoverall efiiciency and capacity.
Therefore, an object is to provide novel rotor profiles and a novel discharge port configuration such that the clearance volume can be reduced to zero (or near zero). This is particularly desirable in a high pressure ratio machine where the weight of the compressed gas contained in-a given clearance volume becomes a larger percentage of the delivered weight.
(3) A specific object is to provide simple novel profiles such that the rotors are economical to manufacture and at the same time secure the advantages of zero or low clearance volume, no trapped pockets, and low leakage.
(4) Another object is to reduce certain throttling losses through ports 25 which occur during the last portion of each delivery stroke. This is accomplished by providing the wider angle A and the notches 32.
(5) Another object is to internally balance the single tooth rotors.
(6) Another advantage is that there is no axial thrust on the rotors (due to gas pressure) such as is associated with most screw type compressors. This eliminates the need, cost, space requirement, and power consumption of thrust bearings.
(7) Another object is to provide rotors (as shown in FIGS. XI and XII) which .are suited for a higher pressure ratio, yet which also retain certain low loss characterisics of the rotors shown in FIGS. VII and VIII.
(8) Another advantage is reduced cost compared with a screw type machine because;
(a) The rotors have simple profiles which are easy to fabricate and there is generally one tooth per rotor.
(b) No large thrust bearings are required.
(c) No speed up gear box and no high speed shaft coupling are required. The rotors turn at electric motors speeds.
BRIEF DESCRIPTION OF THE DRAWINGS FIG. I is a section view taken along the lines I-I in FIG. III.
FIG. II is a section view taken along the lines IIII in FIG. III. One rotor disk 9 and its mounting bolts 10 and dowel pins 11 have been removed in FIG. II. The rotors in FIG. H have been rotated about 42 degrees from their positions shown in FIG. I.
FIG. III is a section view taken along the line IIL-III in FIG. 1.
FIG. IV is a section view taken along the lines -IVIV in FIG. I.
FIGS. V and VI are isometric views of the rotors (shafts removed) used in FIG. I.
FIGS. VII and VIII are larger scale views which illustrate in more detail the rotors used in FIG. I.
FIG. IX is an elevation view of casing end plate 6 when no notches 32 are used.
FIG. X is a view of casing end plate 6 when notches 32 are used.
FIG. XI and XII are reduced scale section view of rotors for use in a higher pressure ratio machine. These rotors are operable in the FIG. I casing structure.
FIG. XIII is an elevation view of a modified disk for use in the FIG. I machine (after modifying port 25).
FIG. XIV is an end view of a modified second rotor to 'be used when the FIG. XIII disk is employed.
FIGS. XV and XVIII illustrate modified rotors operable in the FIG. I casing structure.
FIG. XIX is a section view of a second embodiment or species of the invention.
Like numbers indicate similar parts throughout the drawings.
3 DETAILED DESCRIPTION OF FIGS. 1 TO VIII A first rotor 1 and a second rotor 2 are mounted for rotation (on shafts and hearings) in the intersecting bores 3 and 4 in the casing structure 5. A bolted on end plate 6 is part of the casing structure. The first rotor 1 includes a hub 7, a tooth '8, and two disks 9 bolted 10 and dowelled 11 to each axial end of the hub. A groove 12 passes in an axial direction through the hub and disks. FIGS. VI and VIII illustrate the second rotor which has a major cross section profile taken perpendicular to the rotor axis and midway (or part way) along its length. The second rotor also has a minor cross section profile taken perpendicular to the rotor axis and near or at each end of the rotor. The number 14 refers to the rotor tooth as seen in said major cross section profile. The number 16 refers to the rotor tooth as seen in the minor cross section profiles. Teeth 14 and 16 coincide (have the same profile) outside the pitch circle H. Inside the pitch circle, the profile of each minor tooth 16 contains the concave face J-I. The minor hubs are smaller in diameter than the major hub 13. A groove 17 passes through the major hub.
When operating as a compressor, a source of power is applied to the drive shaft 18 turning rotor 2 in the direction indicated. Timing gears 19 and 20 drive rotor 1 in timed relation. Gas (or the working fluid) enters the inlet port 21 and fills main working chambers 22 and 23. Internal compression begins when tip E reaches the bore juncture 24. Two ports 25 (located in each end wall of the casing) are covered over by the disks 9 during the compression and dwell portions of the rotor cycle so as to prevent backfiow. After the gas in chambers 22 and 23 has been partially compressed, the disks 9 begin to uncover ports 25. The compressed gas is delivered through ports 25, and is then conducted through passages 26 to a manifold 27 which provides a single outlet. The manifold 27 is omitted in FIGS. I and II.
The disks 9 thus serve as rotary valves for controlling the flow of gas through the end ports 25. The ports 25 extend radially outward to the housing bore and are therefore larger in fiow area so as to permit higher r.p.m. and/or longer rotors (compared to the machines in the copending application).
The passages 28 are 'for liquid coolant.
When operating as an expansion engine, the cycle is reversed with higher pressure gas entering at 27 and exhausting at 21 after expanding internally within the machine.
CLEARANCE VOLUME IS R EDU CED TO ZERO (OR NEAR ZERO) AND THERE ARE NO BROKEN OR INTERRUPTED SEAL LINES At the rotor positions shown in FIG. II, there is a certain chamber 29 partially bounded by the concave face 30 of tooth 8, one side wall 31 of groove 12, and the other tooth 14. It is desirable that the gas in chamber 29 be delivered to the discharge line (when operating as a compressor). Otherwise, chamber 29 would constitute a clearance volume and the compressed gas therein would be throttled back to inlet pressure with a resulting loss in both capacity and overall efficiency. The concave face 30 of tooth 8 and one side wall 31 of groove 12 have profiles such that the outer end E of teeth 14 and 16 sweep in sealing proximity across the concave face 30 and wall 31 as the rotors rotate. Also, the convex face J-C-D of teeth 14 and 1 6 and the remaining portion O-P-K of groove 12 have profiles such that these parts move in sealing proximity to each other so as to prevent loss of the compressed gas in chamber 29. The disks 9 and the minor profiles of rotor 2 are also shaped in such a way as to prevent clearance volume or loss of seal line.
The rotor profiles for accomplishing these objectives The ports 25 (in the two end walls of the first rotor bore) are shaped and located in such a way that all the compressed gas (including that in chamber 29 less leakage) is delivered through the end ports 25. The specific shape and location of the ports 25 are subsequently discussed in detail.
The outer edge E of the second rotor sweeps in sealing proximity across the concave face 30 of the tooth on the first rotor and this portion of the rotor cycle is defined as phase P. The preceding definition of phase P applies to one or more claims. For purposes of clarification, phase P is further identified as follows: When operating as a compressor, the beginning of phase P would be at the rotor positions shrown in FIG. II and would end when edge E next intersects ,the hub radius (or pitch circle) of the first rotor 1. With the proportions shown in FIG. II, phase P would occupy about 20 degrees of rotor rotation. The ports 25 (in the end walls) are in open communication with chamber 29 throughout the defined phase P portion of the rotor cycle. Thus the compressed gas is free to flow from chamber 29 through the ports 25 (or reverse flow in case of an expansion engine) throughout the defined phase P.
In the description and claims a single 360 degree rotation of a rotor is defined as being one complete rotor cycle.
The machine shown in FIG. I can be designed to have a substantially zero clearance volume. If the optional notches 32 are employed, then there will be a small clearance volume (in the notches) as will subsequently be explained.
DETAILED DESCRIPTION OF ROTORS- FIGS. V TO VIII The rotors 1 and 2 are designated first rotor and second rotor respectively, for purposes of identification in the description and claims. Arcs T-N, O-P, P-K, RA, J-I, C-D and D-E are constant radius. Rotors 1 and 2 rotate with equal r.p.m. and therefore the diameter of each pitch circle S and H is equal to the distance between rotor axes. The diameters of hubs 7 and 13 are (in this case) equal to the diameters of picth circles S and H (less running clearance). Profile N-O is generated by tip E. Profiles C] and K-R are generated by edges K and J respectively. Profile surfaces T-U and EL-M are wide clearance surfaces, have no sealing function, and therefore it is not necessary to form them precisely. The edge B should be accurate, however.
In order to prevent the formation of trapped pockets, non delivered clearance volume, and/ or a broken proximity seal line, the distance from point I to point G (FIG. VIII) should be substantially equal to or greater than the distance from point G to point D. Points J, G, and D lie in a common plane perpendicular to the rotor axis. Point G is the intersection of radial line B-D and pitch circle H. Arc C-D as shown is tangent at D to arc D-E. In the particular tooth profile shown, the center of circular arc CD is at G.
In order to prevent the formation of trapped pockets, non delivered clearance volume, and/or a broken proximity seal line, the distance from point K to point 34 should be substantially equal to or greater than the distance from point 34 to point A. Points K, 34, and A lie in a common plane perpendicular to the rotor axis. Point 34 is the intersection of a radial line to point A and pitch circle S. Arc RA as shown is tangent at A to the outer periphery of disk 9. In the particular disk profile shown, the center of circular arc RA is at point 34.
The disks 9 have profile surfaces N-O-P-K which coincide with hub 7 and tooth 8 and therefore these surfaces may be finish profiled with the disks fastened in place; and this leads to more accuracy and minimizes cost. The disks 9 have the same outer radius as the tooth 8 and therefore the disks fit directly into bore 3 and no further machining (or counterboring) of bore 3 is necessary.
The disks 9 should be thick enough to prevent excessive deflection due to gas back pressure. Typical dimensions are: rotor diameter 12 inches, hub length 11.25 inches, and thickness of each disk 0.375 inch. In some cases it may be desirable to make the disks 9 of thicker proportions in order to increase the Wr of the first rotor. To obtain smoother running, the Wr of each rotor system (including drive motor) should be equal as discussed in said copending application.
The second rotor 2 can be made from a single piece of material by first shaping the major profile of the rotor the full length of the rotor. Then, the two minor profiles can be completed by end milling are H and end milling hub 15.
The angle A (FIG. VI'II) denotes the total included angle occupied by tooth 14 or tooth 16. The reason for making such a large angle will next be explained. Referring to FIG. II the preferred leading edge of tooth 14 (or 16) is shown by the solid line D-J. An alternate leading edge is shown by dotted line D'-J'. By using the preferred profile DI, the volume of gas remaining in chamber 29 is substantially less than if alternate profile D'J' were employed. This results in less throttling loss of gas through end ports 25 during the final critical portion of the delivery stroke (as a compressor). The A as drawn in FIGS. II and VIII occupies an included angle of 65 degrees. The optimum value for angle A depends on the distance between rotor axes, r.p.m. and rotor width. If angle A were reduced to 55 degrees (for example), the results would still be a substantial improvement over line D'-J (FIG. II). p
The wide angle A provides several additional advantagcs which are: (a) The are DE is longer hence this leak path has more resistance; (b) The coacting arc O-P has a wider angle, hence there is more uncovered flow area through ports 29 during the whole delivery stroke; and (c) The inlet port 21 can be made wider (in the circumferential direction) with little loss in total displacement. A large inlet port reduces flow losses through same.
A honeycomb material or a ductile coating may be applied to the disks 9 for the purpose of reducing running clearance, yet prevent seizure in case of contact. Labyrinth seal grooves are shown at 35.
The removable adjustable wear strip 36 has several special functions which are: (a) The tip E moves with substantial relative velocity (non rolling) across the generated face N-O and therefore, if contact between solid hard rotor materials should occur, seizure would likely result. The wear strip 36 (of a soft material such as brass) will prevent such seizure. (b) The wear strip 36 is independently adjustable relative to its rotor by means of shims and screws 37. This will permit independent adjustment between profile N-O and edge E so as to reduce leakage between rotors. The optional wear strip is not shown in FIG. VI.
DESCRIPTION OF FIG. IX
The location and geometry of the end wall ports (in combination with the particular rotors) permits the machine to have no trapped pockets, zero clearance volume, and no built-in geometrical leak path or interrupted seal line. FIG. IX illustrates the end wall port configuration to be used when no notches 32 are employed. A similar opposite hand port is located in the opposite end wall of bore 3. The lines VWXYZ-V define the outline of a port 25. Are V-W is the outer edge of the port and its radius 38 is equal to the outer radius of the housing bore 3. The circular arc W-X is termed the trailing edge of the port and has its center at B, the axis of the second rotor. The arc W-X has a radius 39 equal to the outer radius of the second rotor. The arc W-X as shown is the optimum shape for the trailing edge of the port. That is, it provides maximum port area and there are no leak paths other than running clearances between moving parts. An alternate less preferred trailing edge is shown by dotted line W'- which will provide slightly less performance. Both edges W-X and WX' are considered to substantially conform to the circular path described by the outer radius of the second rotor. The claims are intended to include these variations.
The circular arc X-Y has a radius equal to that of arc O-P. The shape of the leading edge YZV is not critical. For optimum performance, the discharge port should open as rapidly as possible. Therefore, the leading edge YZV has two circular arcs as shown so as to coincide with profile PKRA on the first rotor (FIG. VII).
The leading edges YZV of port 25 should be located such that the ports begin to open prior to when the desired discharge pressure is reached internally within the machine. This will result in some back fiow but the larger port area will result in a reduction in flow losses during the main portion of the delivery phase of each rotor cycle.
The end plate 6 is recessed at 21 in FIG. IX so as to provide axial inlet porting in addititon to the radial inlet porting shown at 21 in FIG. I.
ROTOR BALANCING The first rotor 1 may be completely balanced by means of three holes 40 (sized and located as shown) which pass clear through the hub section 7. The ends of the holes 40 are covered over with the disks 9. One disk is shown broken at 41 so as to show one balancing hole. In some cases, balancing holes may also be applied to the disks 9, but the disks are generally thin in relation to hub length so that it is generally preferable to do at least the major portion of the balancing by means of holes 40 in the hub 7 as shown.
The second rotor 2 is completely balanced by means of three holes 42 which pass through hub section 13. Separate end caps 43 seal the ends. The outer end of tooth 14 is cast with a hollow interior at 44 (FIGS. VI and VIII) for balancing purposes. The hollow interior opens directly to the trailing face of tooth 14 so as to permit support and withdrawal of foundry cores. It is not necessary to close this opening.
DESCRIPTION OF OPTIONAL NOTCHES 32 As the rotors approach the end of their delivery stroke (FIG. II) the uncovered areas of ports 25 become less. Under some operating conditions (high r.p.m. and/or wide rotors) the pressure drop through the discharge ports 25 would become too high near the end of each delivery stroke. To overcome this problem, the notches 32 provide more flow area through the ports 25. The notches overlap the added areas 454647Y (FIG. X).
DESCRIPTION OF FIGS. XI AND XII These are reduced scale section views of modified rotors (suited for higher pressure ratio) operable in the FIG. I casing structure. Each section view is taken midway along the length of the rotor. A disk 48 is located at each axial end of hub 7 and tooth 8. The FIG. XI rotor is similar to the FIG. VII rotor except that the leading edge 49 (of the groove in the disk 48) does not coincide with the groove in the hub 7, but instead lags same by about 14 degrees. The FIG. XII rotor is similar to the FIG. VIII rotor except that the leading edge 50 of the minor profile tooth lags the leading edge D-] of the major profile tooth by about 14 degrees. The purpose of these modifications is to obtain a later opening of the end wall ports 25 (when operating as a compressor) and thereby obtain a higher built-in pressure ratito. Referring to FIG. IX, the leading edge Y-ZV of end wall port 25 should (in the case of a higher pressure ratio machine) be relocated in a clockwise direction. To obtain optimum port area-timing, the angle occupied by port 25 (about the axis of the first rotor) should be approximately equal to the angle occupied by the groove in disk 9 or disk 48.
The total included angle A occupied by the major profile tooth still remains the same as in FIG. VIII (about 65 degrees) and this is a desirable feature as it reduces the flow rate (and pressure drop) through the ports 25 near the end of each delivery stroke as previously explained. To summarize: The FIGS. XI and XII rotor configurations accomplish two objectives which are:
(l) The desired Wide angle A of the major tooth is re- 7 tained and (2) Optimum porting in a high pressure ratio machine is obtained.
DESCRIPTION OF FIGS. XIII AND XIV The disk 51 (a modification of disk 9) fastens to the same rotor hub 7. The radius 52 is less than the outer radius of tooth 8. A small tooth 53 is attached directly to the disk 51 and has the same thickness as the disk.
The coacting minor hub 54 has a larger radius 55 than its counterpart 15. Outer radius 38 (FIG. IX) must be reduced accordingly.
DESCRIPTION OF FIGS. XV TO XVIII These figures (drawn smaller scale) illustrate a modified set of rotors operable in the FIG. I casing structure. FIGS. XVII and XVIII are end views. FIGS. XV and XVI are section views taken as indicated. The profiles of the rotors (perpendicular to the rotor axis) are similar to the rotors shown in FIG. I. The transverse sections are part toroidal (FIG. XV) and nearly spherical (FIG. XVI).
DESCRIPTION OF FIG. XIX
This figure illustrates a second embodiment of the inven tion. The rotors have hubs 7 and 13, teeth 8 and 14, and grooves 12 and 17 which are similar in cross section profile to the same numbered parts shown in FIG. I. A disk 56 is fastened to each end of the hub 7. The disks rotate within counterbores 57 located in each end wall.
Each disk has a cut out section at 58. A portion of the cut out section is filled with thin metal circular tubes 59 brazed together and to the disk. The length of each tube 59 is equal to the thickness of the main body portion 56. The complete disk assembly including tubes may be finish machined after the tubes are brazed in place. Such a fabricating procedure will assure a disk of uniform thickness at all locations including the tube section. The axial length of the second rotor 60 is such that it can rotate with close running clearance between the flat faces of the two disks 56-. The end ports 61 are identical to the ports 25 except for a straight leading edge 62. The two disks 56 serve as rotary valves for controlling the flow of gas through the ports 61. The disks contain a small clearance volume (about 2% The function of the tubes 59 will next be explained. For purposes of explanation, imagine for a moment that the tubes 59 were removed, leaving a single large opening in the disk. Then when the FIG. XIX rotors reach the rotative position wherein the tip E begins to sweep across the concave face of tooth 8, pressurized gas would be free to flow in a transverse direction around the axial ends of tooth 14. The function of tubes 59 therefore is to permit gas to flow in an axial direction (through the tubes and between the tubes) but restrict gas flow in a transverse direction. It is not necessary to provide the tubes 59 throughout the complete opening 58 as the tubes are needed only in the last portion of the delivery stroke. The novel brazed tube construction 59 (instead of drilled holes in a metal plate) permits more flow area. The circular tubes may, of course, be replaced by tubes of hexagonal or other cross section. In some cases, it is desirable to place tubes of small diameter next to the concave face of tooth 8 and use larger diameter tubes at locations further away from said concave face. The smaller tubes are required near the face so as to prevent excessive spanning of narrow tip E.
In the case of a higher pressure ratio machine, the leading edge of the groove 58 in the disk may be relocated so as to lag the leading edge of groove 12 and thereby obtain a later port opening.
MODIFICATIONS NOT ILLUSTRATED A variation is to move the center of arc CD radially inward along the line DB and make radius F longer. Then profile P-K should be modified so as to conform in proximity sealing relationship with modified arc DCJ. These variations are described and illustrated in more detail in FIGS. XVI and XVII of said co-pending application. Arcs K-R-A and JI may be modified in the same manner as was just described for arcs JCD and P-K.
An alternate fabrication method would be to make the hub 7 and disks 9 from. a single piece of material and the tooth 8, a separate removable piece.
The diameter of hub 7 may be larger than that of hub 13 (or vice versa). If hub 7 is larger, then the hub 13 pro file will have a reverse curvature at the pitch line just as tooth 16 in FIG. VIII contains a reverse curvature at I.
It is not practical (in a 1 to 1 speed ratio) to make the outer radius of rotor 1 substantially larger than the outer radius of rotor 2 as this will result in tooth interference. The outer radius of tooth 8 can be made smaller than the outer radius of tooth 14, and no such interference will result. It would be possible to operate if the outer radius of tooth 8 were only slightly larger than the outer radius of tooth 14 in which case it would be necessary to round off either or both edges E and N slightly to prevent interference.
The disks and hubs could have deviations from being true circular arcs.
It is possible to let each rotor hub vary in diameter along the axial length of the hub as shown in FIGS. XV and XVI.
It is possible to employ two diametrically opposite teeth on each rotor. Such an arrangement provides 20% more displacement than the single tooth rotors but the disadvantages are more expensive rotors and reduced end port area.
Another modification is to provide a first rotor with two teeth (and a valve disk at each end of the rotor) and let it rotate at one half the rpm. of a coacting single tooth rotor. In such a case, the pitch circle diameter of the first rotor would be twice that of the second rotor.
The rotors as illustrated are nonhelical in form. Unlike screw type compressors, it is not necessary to make the rotors helical in order to secure internal compression. The rotors are generally simpler to construct if made with straight (nonhelical) teeth and grooves. There are some conditions, however, where it is desirable to make the rotors helical.
The claims include the above described modifications.
STATEMENTS PERTAINING TO THE CLAIMS The machines described in this application contain certain parts which have similar names but different specific construction and different specific function. For example, the machine shown in FIG. I contains two rotor teeth and three hub diameters. In order to make the claims more clear and to eliminate specific identification of certain parts more than once, reference letters and numbers are used in the claims to identify certain parts and a mode of operation. If such a procedure were not employed, some of the claims would become unwieldy, verbose, and difficult to follow. Examples of such reference letters and numbers are: first rotor, second rotor, tooth M, tooth X, phase P, line C, point D, etc.
The reference letters and numbers used in the claims correspond with letters or numbers used in the drawings. The claims are not to be avoided merely by labeling a part by a different reference letter or number or by omitting a reference letter or number from a part.
In claims 20 and 23, the points of novelty are underlined as an aid.
While the preferred embodiments of the invention have been described, it will be understood that the invention is not limited thereto since it may be otherwise embodied within the scope of the following claims.
I claim: 1. In a rotary displacement machine adapted to handle a working fluid, the combination of: a casing structure having a pair of intersecting bores; a first rotor mounted for rotation in one said bore; a second rotor mounted for rotation in the other said bore; timing gear means constraining said two rotors to rotate in timed relation;
said first rotor having a first cross section (taken perpendicular to the rotor axis and part way along the length of the rotor) which is designated profile M; said profile M having a hub M and a tooth M; said tooth M projecting radially outward from said hub M and attached thereto; said hub M having a groove therein; said casing structure having an inlet port for the entrance of the working fluid into said bores and an outlet port for the exit of the working fluid from said bores; at least one of said ports being located in an end wall of the bore containing said first rotor; the radially outward portion of said port in an end wall being located (from the axis of the first rotor) at a radial distance larger than the outer radius of said hub M; said first rotor having a second cross section (ta-ken perpendicular to the rotor axis and at an end of the rotor) which is designated profile E; said profile E having a land segment and an opening E interrupting said land segment; said land segment having an outer radius larger than the outer radius of said hub M; the radially outward periphery of said land segment occupying a total angle (about the first rotor axis) substantially larger than the angle occupied by said tooth M; said land segment serving to cover said port in an end wall during a portion of each rotor cycle; said opening E serving to uncover said port in an end wall during another portion of each rotor cycle; said land segment thus serving to control the flow of the working fluid through said port in an end wall; said second rotor having a hub X and a tooth X; said tooth X projecting radially outward from said hub X and attached thereto; said hub X having a groove therein; each said tooth being adapted to enter into and recede from the groove in the opposite rotor hub; said hubs M and X being proportioned so as to rotate in sealing proximity to each other during a substantial portion of each rotor cycle; said two rotors cooperating with each other and the casing structure to provide main working chambers which vary in volume as the rotors rotate; said tooth M having a concave face the profile of which is generated by the outer end of said tooth X: the outer end of said tooth X being adapted to sweep in sealing proximity across said concave face and this small portion of each rotor cycle is designated phase P; said concave face, said groove in hub M, and said tooth X forming the partial boundary of a certain small chamber during said phase P; the maximum volume of said small chamber being substantially less than the maximum volume of said main working chambers; said small chamber being of variable volume as said phase P progresses; said opening E serving to uncover said port in an end wall also throughout said phase P; and said port in an end wall being shaped and located such that it is in direct open communication with said small chamber throughout said phase P so as to permit flow of the working fluid through said port in an end wall during said phase P. 2. A rotary displacement machine as defined in claim 1 wherein said port in an end wall has an edge which conforms substantially to a portion of the circular path described by the outer radius of said second rotor.
3. A rotary displacement machine as defined in claim 1 wherein: said tooth X has a convex face surface which substantially begins at the pitch circle and extends radially outward to the outer radius of the tooth; the profile of said convex face surface is designated line I CD; said line JCD intersects the pitch circle of the second rotor at a general point of intersection which is designated point J; an imaginary circle passes through the outer radial end of said tooth X and has its center at the axis of the second rotor; the outer radial end of said line I CD intersects said imaginary circle at a point of general intersection which is designated point D; an imaginary straight radial line interconnects said point D and the axis of the second rotor and this line is designated line DGB; said line DGB intersects the pitch circle of the second rotor at point G; said points I, D and G are all located in a common plane which is perpendicular to the axis of the second rotor; the minimum distance from said point G to said point I is substantially equal to the distance from said point G to said point D;
the groove in said hub M has a profile such that it moves in sealing proximity relationship with said convex face surface (of tooth X) during part of each rotor cycle;
and a purpose of shaping said tooth X and said groove in hub M (as specified in this claim) is to minimize clearance volume and also to substantially maintain a proximity sealing relationship between the two rotors during that portion of each rotor cycle when said tooth X engages said groove in hub M.
4. In a rotary displacement machine adapted to handle a working fluid, the combination of: a casing structure having a pair of intersecting bores; a first rotor mounted for rotation in one said bore; a second rotor mounted for rotation in the other said bore; timing gear means constraining said two rotors to rotate in timed relation;
said first rotor having a hub and a tooth projecting radially outward from the hub; said first rotor hub having a groove therein; said casing structure having an inlet port for the entrance of the working fluid into said bores and an outlet port for the exit of the working fluid from said bores; at least one of said ports being located in an end wall of the bore containing said first rotor; said port in an end wall having an outer edge which is located (from the axis of the first rotor) at a radial distance larger than the outer radius of said first rotor hub;
said first rotor having a disk attached to an end of its hub and rotatable therewith; said disk having an outer radius larger than the outer radius of said first rotor hub; said disk having an opening passing in an axial direction through the disk; said disk serving to cover said port in an end wall during a portion of each rotor cycle; said opening in the disk serving to uncover said port in an end wall during another portion of each rotor cycle; said disk thus serving to control the flow of the working fluid through said port in an end wall;
said second rotor having a hub and a tooth projecting radially outward from the hub; said second rotor hub having a groove therein;
each said tooth being adapted to enter into and recede from the groove in the opposite rotor hub; said two hubs being proportioned so as to rotate in sealing proximity to each other during a substantial portion of each rotor cycle; said two rotors cooperating with each other and said casing structure to provide main working chambers which vary in volume as the rotors rotate;
said first rotor tooth having a concave face the profile of which is generated by the outer end of said second rotor tooth; the outer end of said second rotor tooth being adapted to sweep in sealing proximity across said concave face and this small portion of each rotor cycle is designated phase P; said concave face, said groove in the first rotor hub, and said second rotor tooth forming the partial boundary of a certain small chamber during said phase P; the maximum volume of said small chamber being substantially less than the maximum volume of said main working chambers; said small chamber being of variable volume as said phase P progresses;
said opening in the disk serving to uncover said port in an end wall also throughout said phase P; and said port in an end wall being shaped and located such that it is in direct open communication with said small chamber throughout said phase P so as to permit flow of the working fluid through said port in an end wall during said phase P.
5. A rotary displacement machine as defined in claim 4 for use in compressing said working fluid wherein: said port in an end wall is the outlet port, a portion of said inlet port is radially directed and interrupts the concave cylindrical wall of the bore containing said second rotor, another portion of said inlet port is axially directed and interrupts the end wall of the bore containing said second rotor, said axially directed portion of the inlet port is located adjacent said radially directed portion, and said axially directed portion of the inlet port is located outside the circle described by the periphery of said disk.
6. A rotary displacement machine as defined in claim 4 wherein: said first rotor is provided with a notch which interrupts the end surface of said disk, said notch also interrupts said opening in the disk at a radially inward location of the opening, said port in an end wall extends radially inward so as to cyclically register with said notch, and the purpose of said notch and modified port being to provide said port in an end wall with additional area for flow of the working fluid.
7. A rotary displacement machine as defined in claim 4 wherein the outer radius of said first rotor hub is a substantially circular arc and the outer radius of said second rotor hub is a substantially circular arc.
8. A rotary displacement machine as defined in claim 4 wherein said outer end of said second rotor tooth is provided with a strip of wearable material such that said strip is capable of wearing away in case contact should occur when said strip sweeps across said concave face of the first rotor tooth during said phase P, and wherein said strip is made of a softer material than the main body portions of said two rotors.
9. A rotary displacement machine as defined in claim 4 wherein said port in an end wall has an edge which conforms substantially to a portion of the circular path described by the outer radius of said second rotor.
10. A rotary displacement machine as defined in claim 4 wherein: said second rotor tooth has a convex face surface which substantially begins at the pitch circle and extends radially outward to the outer radius of the tooth; the profile of said convex face surface is designated line JCD; said line JCD intersects the pitch circle of the second rotor at a general point of intersection which is designated point I; an imaginary circle passes through said outer radial end of the second rotor tooth and has its center at the axis of the second rotor; the outer radial end of said line JCD intersects said imaginary circle at a general point of intersection which is designated point D; an imaginary straight radial line interconnects said point D and the axis of the second rotor and this line is designated line DGB; said line DGB intersects the pitch circle at point G, said points I, D, and G are all located in a common plane which is perpendicular to the axis of the second rotor; the minimum distance from said point G to said point I is substantially equal to the distance from said point G to said point D;
said groove in the first rotor hub has a profile such that it moves in sealing proximity relationship with said convex face surface during part of each rotor cycle;
and the purpose of shaping said second rotor tooth and said groove in the first rotor hub (as specified in this claim) is to minimize clearance volume and also to substantially maintain a proximity sealing relationship between the two rotors during that portion of each rotor cycle when said second rotor tooth engages with said groove in the first rotor hub.
11. A rotary displacement machine as defined in claim 10 wherein said port in an end wall has an edge which conforms substantially to a portion of the circular path described by the outer radius of said second rotor.
12. In a rotary displacement machine adapted to handle a working fluid, the combination of: a casing structure having a pair of intersecting bores, a first rotor mounted for rotation in one said bore, a second rotor mounted for rotation in the other said bore, timing gear means constraining said two rotors to rotate in timed relation,
said first rotor having a cross section (taken perpendicular to the rotor axis and part way along the length of the rotor) which is designated and the first profile, said first profile having a hub and a tooth projecting radially outwardly from the hub, said first profile hub having a groove therein,
said casing structure having an inlet port for the entrance of the working fluid into said bores and an outlet port for the exit of the working fluid from said bores, at least one of said ports being located in an end wall of the bore containing said first rotor, the radially outward portion of said port in an end wall being located (from the axis of the first rotor) at a radial distance larger than the outer radius of said first profile hub,
said first rotor having a second cross section (taken perpendicular to the rotor axis and at an end of the rotor) which is designated the second profile, said second profile having a land segment, said land segment having a groove interrupting its outer peri hery, said land segment having an outer radius larger than the outer radius of said first profile hub, said land segment occupying a total angle (about the first rotor axis) substantially larger than the angle occupied by said first profile tooth,
said land segment serving to cover said port in an end wall during a portion of each rotor cycle, said groove (in the land segment) serving to uncover said port in an end wall during another portion of each rotor cycle, said land segment thus serving to control the flow of the working fluid through said port in an end Wall,
said second rotor having a cross section (taken perpendicular to the rotor axis and part way along the length of the rotor) which is designated the major profile, said major profile having a hub and a tooth projecting radially outward from the hub, said major profile hub having a groove therein,
said second rotor having a second cross section (taken perpendicular to the rotor axis and at an end of the rotor) which is designated the minor profile, said minor profile having a hub and a tooth projecting radially outward from the hub, the outer radius of said minor profile hub being smaller than the outer radius of said major profile hub,
said first profile tooth being adapted to enter into and recede from said groove in the major profile hub, said major profile tooth being adapted to enter into and recede from said groove in the first profile hub, said minor profile tooth being adapted to enter into and recede from said groove in the land segment, said first profile hub and said major profile hub being proportioned so as to rotate in sealing proximity to each other during a substantial portion of each rotor cycle, said minor profile hub and said land segment 13 being proportioned so as to rotate in sealing proximity to each other during a substantial portion of each rotor cycle, and said two rotors cooperating with each other and said casing structure to provide working chambers which vary in volume as the rotors rotate.
13. In a rotary displacement machine adapted to handle a working fluid, the combination of: a casing structure having a pair of intersecting bores, a first rotor mounted for rotation in one said bore, a second rotor mounted for rotation in the other said bore, timing gear means constraining said two rotors to rotate in timed relation,
said first rotor having a hub and a tooth projecting radially outward from the hub, said first rotor hub having a groove therein,
said casing structure having an inlet port for the entrance of the working fluid into said bores and an outlet port for the exit of the working fluid from said bores, at least one of said ports being located in an end wall of the bore containing said first rotor, said port in an end wall having an outer edge which is located (from the axis of the first rotor) at a radial distance larger than the outer radius of said first rotor hub,
said first rotor having a disk attached to an end of its hub and rotatable therewith, said disk having an outer radius larger than the outer radius of said first rotor hub, said disk having a groove passing in an axial direction through the disk and interrupting the outer periphery of the disk, said disk serving to cover said port in an end wall during a portion of each rotor cycle, said groove in the disk serving to uncover said port in an end wall during another portion of each rotor cycle, said disk thus serving to control the flow of the working fluid through said port in an end wall,
said second rotor having a cross section (taken perpendicular to the rotor axis and part way along the length of the rotor) which is designated the major profile, said major profile having a hub and a tooth projecting radially outward from the hub, said major profile hub having a groove therein,
said second rotor having a second cross section (taken perpendicular to the rotor axis and at an end of the rotor) which is designated the minor profile, said minor profile having a hub and a tooth projecting radially outward from the hub, said minor profile hub having an outer radius smaller than that of said major profile hub,
said first rotor tooth being adapted to enter into and recede from said groove in the major profile hub, said major profile tooth being adapted to enter into and recede from said groove in the first rotor hub, said minor profile tooth being adapted to enter into and recede from said groove in the disk,
said first rotor hub and said major profile hub being proportioned so as to rotate in sealing proximity to each other during a substantial portion of each rotor cycle, said minor profile hub and said disk being proportioned so as to rotate in sealing proximity to each other during a substantial portion of each rotor cycle, and said two rotors cooperating with each other and said casing structure to provide working chambers which vary in volume as the rotors rotate.
14. A rotary displacement machine as defined in claim 13 wherein: said first rotor tooth has a concave face surface with a profile such that the outer end of said major profile tooth sweeps in sealing proximity across said concave face surface, said groove in the first rotor hub has a contour such that it moves in sealing proximity relationship with said major profile tooth during a portion of each rotor cycle, and an edge of said port in an end wall conforms substantially to a portion of the circular path described by the outer radius of said second rotor.
15. A rotary displacement machine as defined in claim 13 wherein: an edge of said port in an end wall conforms substantially to a portion of the circular path described by the outer radius of said second rotor;
said first rotor tooth has a concave face with a profile such that the outer end of said major profile tooth sweeps in sealing proximity across said concave face;
said major profile tooth has a convex face surface which substantially begins at the pitch circle and extends radially outward to the outer radius of the tooth; the profile of said convex face surface is designated line JCD; said line JCD intersects the pitch circle of the second rotor at a general point of intersection which is designated point I; an imaginary circle passes through the outer radial end of said major profile tooth and has its center at the axis of the second rotor; the outer radial end of said line JCD intersects said imaginary circle at a general point of intersection which is designated point D; an imaginary straight radial line interconnects said point D and the axis of the second rotor and this line is designated line DGB; said line DGB intersects the pitch circle at point G; said points I, D, and G are all located in a common plane which is perpendicular to the axis of the second rotor; the minimum distance from said point G to said point I is substantially equal to the distance from said point G to said point D;
said groove in the first rotor hub has a profile such that it moves in sealing proximity relationship with said convex surface during part of each rotor cycle;
and the purpose of shaping said major profile tooth and said groove in the first rotor hub (as specified in this claim) is to minimize clearance volume and also to substantially maintain a proximity sealing relationship between the two rotors during that portion of each rotor cycle when said major profile tooth en gages with said groove in the first rotor hub.
16. A rotary displacement machine as defined in claim 13 wherein: said major profile tooth has a convex face outside the pitch circle, said minor profile tooth has a convex face which coincides with that of said major profile tooth,
said groove in the first rotor hub coincides with said groove in the disk at locations radially inside the outer radius of said first rotor hub,
said first rotor tooth has a concave face with a contour such that the outer end of said major profile tooth sweeps in sealing proximity across said concave face, and a portion of said groove in the disk coincides with said concave face of the first rotor tooth.
17. A rotary displacement machine as defined in claim 13 wherein: one side of said groove in the disk has a convex profile outside the pitch circle of the first rotor, said minor profile tooth has a concave face surface inside the pitch circle of the second rotor, and said convex profile of the groove in the disk moves in sealing proximity to said concave face surface during a portion of each rotor cycle.
18. A rotary displacement machine as defined in claim 13 wherein: said machine is suited to operate both as a compressor machine and as an expansion engine, said major profile tooth occupies a total angle (about the axis of the second rotor) equal to a minimum of 50 degrees,
the leading edge of said groove in the disk lags behind the leading edge of said groove in the first rotor hub when operating as a compressor machine, the trailing edge of said groove in the first rotor hub lags behind the trailing edge of said groove in the disk when operating as an expansion engine, the purpose of placing the two groove edges out of phase with each other (as specified in this claim) is to delay the time of opening of said port in an end wall when operating as a compressor machine and to advance the time of closing of said port in an end wall when operating as an expansion engine, yet, retain said 50 degree minimum angle for the major profile tooth as specified.
19. A rotary displacement machine as defined in claim 13 wherein said first rotor hub is circular over the main portion of its periphery and wherein said major profile hub is circular over the main portion of its periphery.
20. In a rotary displacement machine adapted to handle a working fluid, the combination of: a casing structure having a pair of intersecting bores, a first rotor mounted for rotation in one said bore, a second rotor mounted for rotation in the other said bore, timing gear means constraining said two rotors to rotate in timed relation,
said first rotor having a hub and a tooth projecting radially outward from the hub, said first rotor hub having a groove therein,
said casing structure having an inlet port for the entrance of the working fluid into said bores and an outlet port for the exit of the working fluid from said bores, at least one of said ports being located in an end wall of the bore containing said first rotor, the radially outward portion of said port in an end wall being located (from the axis of the first rotor) at a radial distance substantially egual to the outer radius of said first rotor tooth,
said first rotor having a disk attached to an end of its hub and rotatable therewith, the outer radius of said disk being substantially egual to the outer radius of said first rotor tooth, E dis); having a groove passing in an axial direction through the disk, fig
groove interrupting the outer periphery of said @5, said disk serving to close said port in an end wall during a portion of each rotor cycle, said groove in the disk serving to uncover said port in an end wall during another portion of each rotor cycle, said disk thus serving to control the flow of the working fluid through said port in an end wall,
said second rotor having a hub and a tooth projecting radially outward from the hub, said second rotor hub having a groove therein,
said first rotor tooth being adapted to enter into and recede from the groove in said second rotor hub as the rotors rotate, said second rotor tooth being adapted to enter into and recede from the groove in said first rotor hub as the rotors rotate, said first rotor hub and said second rotor hub being proportioned so as to rotate in sealing proximity to each other during a substantial portion of each rotor cycle, and said two rotors cooperating with each other and said casing structure to provide main working chambers which vary in volume as the rotors rotate.
21. A rotary displacement machine as defined in claim 4 wherein: said disk contains a plurality of conduits passing in an axial direction through the disk, said conduits are for the passage of the working fluid, and the purpose of employing a plurality of conduits (instead of a single large opening) is to restrict leakage flow of the Working fluid across the axial end of said second rotor tooth (during said phase P).
22. A rotary displacement machine as defined in claim 21 wherein said conduits are in the form of hollow tubes fastened to each other and fastened to the main body portion of said disk.
23. In a rotary displacement machine adapted to handle a working fluid, the combination of: a casing structure having a pair of intersecting bores, a first rotor mounted for rotation in one said bore, a second rotor mounted for rotation in the other said bore, timing gear means constraining said two rotors to rotate in timed relation,
each said rotor having a hub and a single tooth projecting radially outward from each hub, each said hub having a groove therein, each said tooth being adapted to enter into and recede from the groove in the opposite rotor hub, said two hubs being proportioned so as to rotate in sealing proximity to each other during a substantial portion of each rotor cycle, said two rotors cooperating with each other and said casing structure to provide working chambers, said casing structure having an inlet port for the entrance of the working fluid into said bores and an outlet port for the exit of the working fluid from said bores, at least one of said ports being located in an end wall of the bore containing said first rotor,
said first rotor having an end surface which rotates in sealing proximity to said end wall and which controls the flow of the working fluid through said port in an end wall,
the outer periphery of said first rotor hub occupying a total angle (about the axis of the first rotor) exceeding degrees, Q13 total angle occupied Q1 a ll portions o f said second rotor tooth being equal t o a minimum 92 degrees. tl total angle occupied b v all portions gt topfl being at least E degrees l es than that occupied by said second rotor tooth, w purpose 5L" making said second rotor tooth with such a large included angle E specifled in its claim) being t2 reduce LI flo \v losses through Ed p o r t in a n Ll Wall, 2 1n d a purpose if making is; rotor tooth yV i th a smaller included angle 1% specified i2 claim) being t o maximize Lhe total displacement f said machine.
References Cited UNITED STATES PATENTS 2,058,817 10/1936 Northey 230141 2,097,037 10/1937 Northey 230141 2,287,716 6/ 1942 Whitfield 230143 2,578,196 12/1951 Montelius 230143 2,580,006 12/1951 Densham 230--141 FOREIGN PATENTS 661,749 11/ 1951 Great Britain.
WILLIAM L. FREEH, Primary Examiner W. J. GOODLIN, Assistant Examiner U.S. Cl. X.R. 418-179, 191
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Cited By (13)

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US3620655A (en) * 1969-06-18 1971-11-16 Atlas Copco Ab Balanced rotor machine
US3723031A (en) * 1970-11-23 1973-03-27 A Brown Rotary displacement machines
US3748069A (en) * 1969-06-18 1973-07-24 Atlas Copco Ab Toothed rotor piston machine
US3790315A (en) * 1970-10-01 1974-02-05 Atlas Copco Ab Rotary piston compressors with liquid injection
EP0009915A1 (en) * 1978-09-28 1980-04-16 Arthur E. Brown Rotary positive displacement machines
EP0009916A1 (en) * 1978-09-27 1980-04-16 Ingersoll-Rand Company Rotary positive displacement machines
FR2478223A1 (en) * 1980-03-17 1981-09-18 Worthington Compressors Inc ROTARY COMPRESSOR
US4324538A (en) * 1978-09-27 1982-04-13 Ingersoll-Rand Company Rotary positive displacement machine with specific lobed rotor profiles
EP0133629B1 (en) * 1981-01-02 1988-04-13 Ingersoll-Rand Company A rotary positive displacement machine
US20040005235A1 (en) * 2000-10-19 2004-01-08 Didin Alexandr Vladimirovich Volumetric rotary machine
CN1904365B (en) * 2005-07-29 2010-06-16 良峰塑胶机械股份有限公司 Designing method of claw type rotor
WO2012126137A1 (en) * 2011-03-21 2012-09-27 淄博特士德真空设备科技有限公司 Claw type vacuum pump
US20130209306A1 (en) * 2010-10-22 2013-08-15 Peter South Rotary Positive Displacement Machine

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US2058817A (en) * 1931-12-22 1936-10-27 Northey Rotary Engines Ltd Rotary internal combustion engine
US2097037A (en) * 1933-08-25 1937-10-26 Northey Rotary Engines Ltd Rotary compressor or vacuum pump
US2287716A (en) * 1941-04-22 1942-06-23 Joseph E Whitfield Fluid device
GB661749A (en) * 1949-01-17 1951-11-28 Arthur John Northey Improvements in or relating to rotary air or gas compressors and motors
US2578196A (en) * 1946-11-30 1951-12-11 Imo Industri Ab Screw compressor
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US2058817A (en) * 1931-12-22 1936-10-27 Northey Rotary Engines Ltd Rotary internal combustion engine
US2097037A (en) * 1933-08-25 1937-10-26 Northey Rotary Engines Ltd Rotary compressor or vacuum pump
US2287716A (en) * 1941-04-22 1942-06-23 Joseph E Whitfield Fluid device
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Cited By (15)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3748069A (en) * 1969-06-18 1973-07-24 Atlas Copco Ab Toothed rotor piston machine
US3620655A (en) * 1969-06-18 1971-11-16 Atlas Copco Ab Balanced rotor machine
US3790315A (en) * 1970-10-01 1974-02-05 Atlas Copco Ab Rotary piston compressors with liquid injection
US3723031A (en) * 1970-11-23 1973-03-27 A Brown Rotary displacement machines
EP0009916A1 (en) * 1978-09-27 1980-04-16 Ingersoll-Rand Company Rotary positive displacement machines
US4324538A (en) * 1978-09-27 1982-04-13 Ingersoll-Rand Company Rotary positive displacement machine with specific lobed rotor profiles
EP0009915A1 (en) * 1978-09-28 1980-04-16 Arthur E. Brown Rotary positive displacement machines
FR2478223A1 (en) * 1980-03-17 1981-09-18 Worthington Compressors Inc ROTARY COMPRESSOR
EP0133629B1 (en) * 1981-01-02 1988-04-13 Ingersoll-Rand Company A rotary positive displacement machine
US20040005235A1 (en) * 2000-10-19 2004-01-08 Didin Alexandr Vladimirovich Volumetric rotary machine
US7080976B2 (en) 2000-10-19 2006-07-25 Ilya Yakovlevich Yanovsky Volumetric rotary machine
CN1904365B (en) * 2005-07-29 2010-06-16 良峰塑胶机械股份有限公司 Designing method of claw type rotor
US20130209306A1 (en) * 2010-10-22 2013-08-15 Peter South Rotary Positive Displacement Machine
US9435203B2 (en) * 2010-10-22 2016-09-06 Peter South Rotary positive displacement machine
WO2012126137A1 (en) * 2011-03-21 2012-09-27 淄博特士德真空设备科技有限公司 Claw type vacuum pump

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