US3307777A

US3307777A - Screw rotor machine with an elastic working fluid

Info

Publication number: US3307777A
Application number: US420370A
Authority: US
Inventors: Schibbye Lauritz Benedictus
Original assignee: Svenska Rotor Maskiner AB
Current assignee: Svenska Rotor Maskiner AB
Priority date: 1963-12-23
Filing date: 1964-12-22
Publication date: 1967-03-07
Anticipated expiration: 1984-03-07
Also published as: BE657477A; DE1428277A1; GB1077517A; DE1428277B2; DE1428277C3; SE310751B

Description

March 7, 1967 L. B. SCHIBBYE SCREW ROTOR MACHINE WITH AN ELASTIC WORKING FLUID Filed Dec. 22, 1964 '7 Sheets-Sheet 1 3+ Fig.1

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I IV VEN TOR MM M ATTORNEY March 7, 1967 l... B. sea-mama 3 SCREW ROTOR MACHINE WITH AN ELASTIC WORKING FLUID Filed Dec. 22, 1964 '7 Sheets-Sheet 3 Fig.3

I NVEN TOR C] ATTORNEY arch 7, 1967 L. B. SCHIBBYE 3739793777 SCREW ROTOR MACHINE WITH AN ELASTIC WORKING FLUID Filed Dec. 22, 1964 7 Sheets-Sheet 4 r1 tot.

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I N VEN TOR March 7, 1967 L. B. SCHIBBYE 3,307,777

SCREW ROTOR MACHINE WITH AN ELASTIC WORKING FLUID Filed Dec. 22, 1964 '7 Sheets-Sheet 5 Fig.6

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'Filed Dec. 22, 1964 7 Sheets Sheet 6 March 7, 1967 B. SCHIBBYE 3,307,777

SCREW ROTOR MACHINE WITH AN ELASTIC WORKING FLUID FiledDec. 22, 1964 7 7 Sheets-Sheet "7 Fig.8

I NVENTOR '4..- ATTORNEY United States Patent f 3,307,777 SCREW ROTOR MACHINE WITH AN ELASTIC WGRKING FLUID Lauritz Benedictus Schihbye, Saltsjo-Duvnas, Sweden, as-

signor to Svenska Rotor Maskiner Aktiebolag, Nacka, Sweden, a corporation of Sweden Filed Dec. 22, 1964, Ser. No. 420,370 Claims priority, application Sweden, Dec. 23, 1963, 14,418/63 14 Claims. (Cl. 230143) A screw rotor machine with an elastic Working fluid is a machine which principally comprises at least two coplanar cooperating rotors with helical lands having a wrap angle of less than 360 and with intervening grooves, and a casing enclosing the rotors. The rotors are of male and female rotor type, which means that the lands of the male rotor have at least their main portion located outside the pitch circle of the rotor and have substantially convex flanks, and that the lands of the female rotor have at least their main portions located inside the pitch circle of the rotor and have substantially concave flanks. The casing enclosing the rotors is further provided with a working space having low pressure and high pressure ports which casing substantially is composed of coplanar intersecting bores each sealingly enclosing a rotor.

When the rotors cooperate a land of one rotor goes into mesh with a groove of the other rotor whereby a chevron shaped chamber is formed between the rotors and the walls of the working space comprising a portion of a male rotor groove and a portion of a female rotor groove communicating therewith. Each such chamber has its base end located in a stationary plane transverse to the axes of the rotors which plane is common for all the chevron shaped chambers and situated at the high pressure port, and its apex end located at the intermesh between the cooperating lands of the different rotors. When the rotors revolve the apex end of the chevron shaped chamber moves axially in relation to the stationary plane, whereby the volume of the chamber varies.

In machines of this type which have been most used up to now the different rotors have been connected by a synchronizing gear which has the double purpose partly to fix the rotors connected therethrough in the certain relative angular position, partly to transmit the torque acting upon the rotor having no stub shaft passing out from the machine. In consideration of the fact that the adjustment of the angular position must be very accurate in order to avoid direct contact between the rotors it is in such machines of high importance that the wear of the wheels in the synchronizing gear is kept as low as possible. The power transmitting shaft extending outside the machine has for this reason been connected to the rotor having the greatest capacity for power transmission to the working fluid which owing to the profiles of the rotors always has been the male rotor. A further reason to connect the power transmitting shaft to the rotor having the largest power transmitting capacity is to limit the distortion of the rotors and the synchronizing gear as far as possible and thus the corresponding angular movement between the rotors. This feature is of special importance as the machines owing to designing points of view normally has been shaped so that the synchronizing gears has been located at the end of the rotors opposite to that at which the external shaft is located.

Recently, however, also a modified type of screW rotor machines has been put on the market which type is a compressor without any synchronizing gear connecting the rotors and with oil injection which .among others 3,307,777 Patented Mar. 7, 1 967 has the purpose to form an oil film on the flanks of the rotor lands whereby torque transmission by a direct contact between the flanks of the rotors has been possible. In order to limit the torque transmitted by the flank contact as much as possible the external shaft has in such machines been connect-ed to the male rotor. Owing to the fact that the male rotor always obtains a larger axial force than the female rotor from the working fluid and owing to the direction of the screw pitch which is necessary for the function of machine the axial forces deriving from the working fluid and the torque transmission have been added to each other on the male rotor while they have been subtracted from each other on the female rotor so that the difference between the thrusts acting on the axial bearings of the different rotors has increased further. This is serious as the male rotor normally has a smaller pitch circle and thus the higher speed of rotation for which reason there has been difficulties to obtain an acceptable service life for the thrust bearing of the male rotor. In some cases it has thus been necessary to use a thrust bearing for the male rotor with adiameter which is longer than the one which owing to the fixed centre distance between the rotors has been desirable from practical point of view as the thrust bearing of the female rotor has had to be designed with a smaller diameter so that thrust bearings of two types have been used which is improper with respect to manufacture and spare parts. Another possibility which also has been used in certain cases is to provide the male rotor with an oil-actuated balancing piston which of course is undesirable owing to the more complicated design of the machine deriving therefrom.

Another disadvantage of screw rotor machines with liquid injection is that the tip speed of the male rotor necessary for optimum efficiency is higher than the speed normally obtainable by direct drive connection between the prime mover and the male rot-or. For instance for air compressors with a built-in pressure ratio of 7:1 the tip speed of the male rotor necessary for maximum efliciency is thus about 20 m./s.

The efficiency varies 'with the tip speed in such a way that it decreases rapidly for speeds below the optimum one While the decrease for speeds above the optimum one is considerably more gradual. This can be illustrated by the fact that approximately the same efliciency is obtained for tip speeds of 15 m./s. and of 40 m./s. The determining fact-or is the characteristic of the input torque which increases rapidly for decreasing speed of rotation of one and the same machine.

It is essential for machines of the type mentioned above to 'be able to vary the capacity of the machine by variation of the speed of the prime mover. The prime movers normally used for driving of such compressors, viz. diesel engines, have, however, such a torque characteristic in dependence on the speed of rotation that the torque decreases for low speeds where the torque demand of the compressor increases. For a ceitain size of engine which is determined with regard to the necessary power for maximum capacity of the compressor the possibility to decrease the speed of the engine and thus the capacity of the compressor is restricted by the relation of the torques of the engine and of the compressor so that a tip speed of the male rotor falling below about 15 m./ s. only can be obtained at throttling of the compressor inlet. This means increased losses and an unacceptable decrease of the efliciency of the compressor plant. In order to obtain the desired possibility to vary the capacity of the compressor down to about 50% of maximum capacity by adjustment of the speed it is thus necessary to design the plant so that the tip speed of the male rotor for maximum capacity is about 30 m./s. Owing to the flatness of the efliciency curve above the maximum point for a tip speed of about 20 m./s. this means, however, no considerable decrease of the efficiency for maximum capacity and this slight decrease is compensated by high efficiency at part loads which gives an optimum efliciency for the total working cycle comprising part loads as well as maximum load.

In a compressor of the most common size, i.e. a male rrotor diameter of about 200 mm. this means that the normal speed of the compressor is about 2900 r./ min. A normal speed for a diesel engine having the actual power of about 100 HP. is, however, only about 1800 r./ min. In order to obtain an optimum total efliciency it is thus necessary to provide a gear box having a ratio of 1.5: l- 1.75:1 between the engine and the compressor which means complications and increase in the price of the compressor plant and also losses in the gear box which influence the efficiency in an unfavourable way.

The present invention relates to a machine which does not have any of the disadvantages mentioned above, viz. the unequal load of the thrust bearings and the need of a gear box connected with the exterior shaft of the machine. This is obtained by connecting the exterior shaft to the female rotor in contradiction to the principles followed up to now and in spite of the larger torque which then must be transmitted between the rotor flanks and the considerably greater contact forces therebetween. The contact surfaces between the rotors have owing to the large length of the rotors such a large area that it is possible by strengthening the edges of the female rotor lands with addendums located outside the pitch circle of the rotor and by providing an oil film on the rotor flanks by oil injection to transmit the torque between the rotors without any provable wear of the flanks of the lands arises. By designing both rotors with about the same outer diameter which is advantageous especially from manufacturing point of view and wit-h regard to similar deflection of both rotors the pitch circle of the female rotor will have a considerably larger diameter than the pitch circle of the male rotor as the female rotor lands are located substantially inside and the male rotor lands are located substantiailly outside the pitch circle of the respective rotor. The speed of the female rotor is thus lower than that of the male rotor.

It has shown to be suitable that the male rotor is provided with three or four lands and that the female rotor is provided with a number of lands which exceeds the number of the lands of the male rotor with at least two and which does not exceed the double number of the lands of the male rotor but preferably lies below this double number. These combinations of lands give the lands of the female rotor a sufficient thickness to transmit the necessary torque between the rotors without any deflection of the lands. These combinations of lands give further the male rotor a speed which is 50% higher with a land combination of 4+6, 67% higher with a land combination of 3+5 and 75% higher with a land combination of 4+7 than the speed of the exterior shaft without use of a separate gear box and the power losses following therefrom, so that the speed of the male rotor practically without losses is the one necessary for an optimum efficiency. In comparison with a machine with the exterior shaft connected to the male rotor and directly connected with the engine the advantage is further obtained that the capacitor of the machine with the same dimensions of the male rotor increases with at least 50%, 67% and 75%, respectively, which means decreased dimensions not only in comparison with direct male rotor drive but also in comparison With a male rotor driven compressor with a gear box. At the same time considerably lower manufacturing costs are obtained as these are about proportional to the weight of the compressor. A decrease of the dimensions is further especially important for an air compressor of the type mentioned above as they often are mounted on a transportable car the dimension of which is depending on the bulk of the compressor plant.

The advantage of increased capacity can especially be with their axes coinciding with the bore axes.

used as a link of combination in a two stage tandem compressor, i.e. a compressor with a common driving shaft on which the driving rotors are fixed and with separate driven rotors in the two stages. Owing to the fact that the driving and driven rotors are not angularly fixed in relation to each other through a synchronizing gear the two driven rotors of the different stages can without any disadvantages being obtained rotate with different speeds as in such a two stage compressor it is a problem to obtain a volume ratio between the two stages of enough size so that the best relation between the pressure ratios of the stages from efliciency point of view can be obtained. In earlier machines the relation of the volume has made it necessary to use a first stage having a large axial length in relation to the diameter of the male rotor, and a second stage having small axial length in relation to the diameter of the male rotor so that the relation length to diameter for both stages has differed from the one giving an optimum efficiency. By connecting the female rotor of the first stage and the male rotor of the second stage to the driving shaft the volume capacities of the two stages can be made so different that desirable distribution of the pressure rise between the two stages giving optimum total efiiciency can be obtained at such a relation between length and diameter of the rotor in each stage that a good eificiency is obtained also in each stage. The advantages obtained in this way cause that the total efficiency of the machine can be increased above the one up to now obtainable by machines of similar type.

In the following portion of the specification the invention will be explained more in detail with reference to some preferred embodiments of screw rotor machines shown in the accompanying drawings in which FIG. 1 shows a vertical section through a screw compressor taken on line 11 in FIG. 2.

FIG. 2 shows a transverse section through the compressor of FIG. 1 taken on line 2-2 in FIG. 1.

FIG. 3 shows a horizontal section of the compressor of FIG. 1 taken on line 33 in FIG. 1.

FIG. 4 shows the overall efl'lciency of the compressor of FIGS. 1-3 as a function of the peripheral speed of the male rotor.

FIG. 5 shows the torque characteristics of the compressors of FIGS. 1-3 and the torque characteristic of a diesel engine connected to the female rotor of the compressor, both as a function of the peripheral speed of the male rotor.

FIG. 6 shows a transverse section corresponding to FIG. 2 through a modified type of a screw compressor.

FIG. 7 shows a transverse section corresponding to FIG. 2 through a further modified type of a screw compressor and FIG. 8 a horizontal section through a two stage compressor designed according to the invention.

The screw compressor shown in FIGS. 1-3 comprises a casing 10 forming a working space 12 substantially in the form of two intersecting cylindrical bores having parallel axes. The casing 10 is further provided with a low pressure channel 14 and a high pressure channel 16 for the working fluid which channels communicate with the working space 12 through a low pressure port 18 and a high pressure port 20, respectively.

In the compressor shown the low pressure port 18 is located in its entirety in the low pressure end wall 22 of the working space 12 and extends mainly on one side of the plane containing the axes of the bores. The high pressure port 20 of the compressor shown is located partly in the high pressure end wall 24 of the working space 12 and partly in its barrel wall 26 and it is in its entirety located on the side of the plane through the bore axes opposite to the low pressure port.

In the working space 12 are provided two cooperating rotors, viz, a male rotor 28 and a female rotor 30, located These rotors are journaled in' the casing 10 in cylindrical roller bearings 32 in the low pressure end wall and in pairs of ball bearing-s 3'4 with shoulders in the .high pressure end wall 24. The female rotor 30 is further provided with a stub shaft 36 projecting outside the c-asing 10.

The male rotor 28 has four helical lands 38 and in tervening grooves 40 having a wrap angle of about 300. The female rotor 30 has six helical lands 42 and intervening grooves 44 having a wrap angle of about 200. The female rotor lands 42 are provided with addendums 48 located radially outside the pitch circle 46 of the female rotor 30 and the male rotor grooves 40 are provided with corresponding dedendums 52 located radially inside the pitch circle 50 of the male rotor 28.

In the barrel wall 26 of the working space 12 are provided a plurality of oil injection channels 54 opening at the inter-section line 56 between the two bores forming the working space 12. These channels 54 form communications between anoil supply chamber -8 and'the working space 12. Oil is supplied to this chamber 58 from a pressure oil source not shown through a supply opening 60 under a pressure higher than the pressure prevailing in the working space 12 at the openings of the channels 54.

The compressor operates in the following way. The female rotor 30 is driven through the stub shaft 36 projecting from the casing by a prime mover not shown and thereby the male rotor 28 is rotated through the direct contact between the flanks of the lands 42 of the female rotor and the lands 38 of the male rotor 28. This flank contact takes place substantially between portions of said lands 42, 38 which lie adjacent to and preferably somewhat outside the pitch circles 46 and 50 of the rotors, that is between flank portions which have practically the same peripheral speed so thatthe inale .land 42 enters the male rotor groove and then also a male rotor land 38 enters the female rotor groove 44. In this way two 'intercommunicating grooves form a chevron shaped chamber sealed against the low'pressure port 18 as well as against the high pressure port 20 and the apex of this chamber is located at the point of engagement of said lands in said grooves and its base is located at the high pressure end wall 24. The volume of this chamber is then continually decreased during the further rotation ofwthe rotors as the apex of the chamber is moved in the directions towards the stationary high pressure end wall 24. At this decrease of the volume the pressure in the chamber increases. During this compression thus taking place oil is injected in the chevron shaped chamber through the injection channels 54 which injection due to the fact that the channels 54 open at the intersection line, 56 will occur at the apex of the chamberwere the compression is initiated as the lands while entering the corresponding grooves function .as axially movable pistons. The oil serves as a cooling fluid which cools the working fluid and exactly that portion thereof in which the heat generation due to the compression is maximum. However, as the rotors rotate the oil is thrown by the centrifugal force against the lands of the rotors and particularly against the tips thereof where it produces an oil film which seals the leak-age clearances between the rotors and the casing; As the rotors within the chevron shaped chamber. rotate in a direction towards each other oil is also thrown into the region of engagement between the cooperatingrotors. This has the double function of forming an oil seal for the working fluid and of-forming an oil film on the torquetransmitting land flanks so that the flanks are prevented from coming in direct mechanical contact with each other whereby the torque transmission between the rotors takes place without wear of the land flanks. After a certain rotation of the rotors determined by the size and form of the high pressure port the chevron shaped chamber is openedtowards the high pressure port 20 and the discharge of the working fluid to the high pressure channel commences and this discharge then continues until the exhaust of the chevron shaped chamber through the high pressure port is finished.

As shown in FIG. 4 which relates to an air compressor having a built-in pressure ratio of about 7:1 the overall efliciency 1; of the compressor is dependent on the peripheral speed of the male rotor 28 and has its maximum at a peripheral speed of about 20 m./s. The shape of the efficiency curve is however such that at peripheral speeds lower than the speed which corresponds to maximum efliciency the variation of the efficiency is considerably larger in relation to the variation of the peripheral speed than the variation of the efficiency in relation to the variation of the peripheral speed at peripheral speeds higher than the speed which corresponds to maximum efficiency so that the efiicien-cies are about the same at 15 m./s. and at 40 m./s. This shape of the efficiency curve 17 is dependent substantially on the torque demand ,of the compressor at diiferent peripheral speeds which is shown by the curve T in FIG. 5. In FIG. 5 there is also shown a curve T which illustrates the torque produced by a diesel engine or normal type the size of this engine being adapted such that its power corresponds to the powerdemand of the compressor at full load at a peripheral speed of the male rotor of about 28.5 m./s. In a compressor having a male rotor with a diameter of about 200 mm. the power demand is about 100 HR A normal diesel engine of this order of magnitude has a normal operating speed of about 1800 r./min. which at direct drive of the female rotor of this compressor ,having four male rotor lands and six female rotor-lands results in a peripheral speed of the male rotor of 28.5 m./s. As shown in FIG. 5 the two torque curves T and T intersect each other at a peripheral speed of by reduction of the engine speed. The difference between this last mentioned value and the theoretical one basedsolely on the speed ratio depends on the fact that the. volumetric efficiency of the compressor decreases somewhat at decreasing peripheral speed of the male rot-or.

The compressor. shown in FIG. 6 differs constructively from that shown in FIGS. 1-3 only in that the female rotor 30 is provided with seven lands 42 and intervening grooves 44 instead of six and in that the outer diameter of the female rotor 30 has at the same time been correspondingly increased. As to its function it differs from the. compressor shown in FIGS. 1-3 in that the gear ratio between the male and female rotors has been increased from 1.5 :1 to 1.75:1 which renders it possible at the same male rotor diameter further to increase the control range downto about 40% of the full load capacity or to maintainithe the same control range in compressors in which the diameter of the male. rotor is only about of the diameter of a corresponding compressor of the type shown in FIGS. 13.

The compressor shown in FIG. 7 differs constructively from that shown in FIGS. 1-3 only in that the male rotor 28 is provided with three lands 38 and intervening grooves 40 instead of four and in that the female rotor 30 is provided with five lands 42 and intervening grooves 44 instead of six and in that the outer diameter of the female rotor 30 at the same time has been correspondingly increased. As to its function it differs from the compressor shown in FIGS. l-3 in that the gear ratio between the male and the female rotors has been increased from 1.5:1 to 1.67:1 which renders it possible at the same male rotor diameter to further increase the control range down to about 45% of the full load capacity or to maintain the same control range in compressors in which the male rotor diameter is only about 90% of that of the corresponding compressor of the type shown in FIGS. 1-3.

The two stage tandem compressor shown in FIG. 8 comprises a casing 62 forming two working

spaces

64 and 66 each substantially consisting of two intersecting .cylindrical bores having parallel axes, these two working spaces being separated by a partition 68. The bores forming the working

spaces

64, 66 are coaxial in pairs and all formed with one and the same diameter. The casing 62 is further provided with a low pressure channel 70 for the working fluid communicating with the low pressure working space 64 through a low pressure port 72, an intermediate pressure discharge port 74 leading from the low pressure working space to an intermediate pressure discharge channel not shown, an intermediate pressure inlet channel 76 communicating with the high pressure working space 66 through an intermediate pressure inlet port 78, and a high pressure port 80 forming the outlet port from the high pressure working space 66 and communicating with a high pressure channel not shown. The intermediate pressure discharge port 74 is further directly connected with the intermediate inlet channel 76 through channels or conduits not shown. In the compressor shown the low pressure port 72 is located in its entirety in the low pressure end wall 82 of the low pressure working space 64. The intermediate pressure discharge port 74 is in a manner not shown located partly in the partition 68 and partly in the barrel wall 84 of the low pressure working space 64. The intermediate pressure inlet port 78 is located entirely in the low pressure end wall 86 of the high presure working space 66 while the high pressure port 80 in a manner not shown is located partly in the partition 68 and partly in the barrel wall 88 of the high pressure working space.

In the low pressure working space 64 are provided two cooperating rotors, viz. a male rotor 90 and a female rotor 92, with their axes coinciding with the bore axes. These rotors are journaled in hearings in the casing 62 but for the sake of simplicity the bearings are shown only diagrammatically as bushes. However, the invention is not limited to slidebearings for the rotors. The female rotor 92 is provided with a stub shaft 94 projecting outside the casing 62 through the low pressure end wall 82 and with a torsion shaft 96 passing through the partition 68 and through the high pressure working space 66.

In the high pressure working space 66 there are likewise two cooperating rotors, viz. a male rotor 98 and a female rotor 100, the male rotor 98 being coaxial with the female rotor 92 in the low pressure working space 64 and nonrotatably connected to the latter by means of the torsion shaft 96 while the female rotor 100 is coaxial with but totally free from the male rotor 90 in the low pressure working space 64. The

rotors

98 and 100 are journaled in bearings in the casing 62 in the same manner as the

rotors

90 and 92 the bearings being shown as bushes without thereby limiting the invention. When seen in a transverse plane the

male rotors

90 and 98 have the same profile as the male rotor 28 in FIG. 2 (not illustrated) and in the same manner, when seen in a transverse plane, the

female rotors

92 and 100 have the same profie as the female rotor 30 in FIG. 2. However, within the scope of the invention it is fully possible to use also other forms of profiles, for instance those shown in FIGS. 6 and 7, and it is not either necessary to use the same forms of profiles in the different stages. Preferably, however, the profiles are made such that the distance between the rotor shafts is the same in both stages. The compressor is further provided with oil injection openings in the same way as shown in FIG. 1. The length to diameter ratio of the male rotors is about 1.75:1 and 1.05:1 in the low pressure stage and high pressure stage, respectively, instead of 2.0:1 and 0.811 in a corresponding compressor of standard type in which the two male rotors are non-rotatably connected with each other. The ratios between rotor length and rotor diameter just mentioned are not limiting the present invention but only serve to illustrate the technical effect attained by the invention in relation to earlier types of compressors. Said length to diameter ratios result, in fact, in the same ratio between the volume capacities in the different stages as in the older compressor type due to the speed ratio between the male rotors in the different stages as a result of the female rotor drive in the low pressure stage and male rotor drive in the high pressure stage, said speed ratio amounting to 1.5 :1 in the compressor shown. However, this speed ratio is wholly dependent on the selected rotor profiles and may be varied as shown in FIGS. 6 and 7. The ratio between the volume capacities of the stages is not either limiting the invention as the optimum ratio may vary in dependence on the properties of different working fluids, different over all pressure ratios and pressure differentials and on the quantity of the oil injected and its distribution between the stages.

What is claimed is:

1. A two stage compressor of the screw rotor type having low and high pressure stages arranged in tandem and with each stage comprising male and female rotors with coplanar axes and intermeshing helical lands and grooves enclosed in ported casing structure providing for flow of working fluid serially through said stages from a low pressure inlet to a high pressure outlet, the male rotor of one of said stages being connected to the female rotor of the other of said stages to cause the connected rotors to rotate at the same speed, and an external driving connection for driving one of said connected rotors.

2. A compressor as defined in claim 1, in which the connected rotors comprise the female rotor of the low pressure stage and the male rotor of the high pressure stage.

3. A compressor as defined in claim 2, in which said external driving connection is to the female rotor of the low pressure stage.

4. A compressor as defined in claim 1, in which the distance between the axes of rotation of the intermeshing rotors is the same in both of said stages.

5. A compressor as defined in claim 1, in which the outer diameter of the male and female rotors of one stage are the same, respectively, as the outer diameters of the female and male rotors of the remaining stage.

6. A compressor as defined in claim 5, in which the profiles of both male rotors are the same and the profiles of both female rotors are the same.

7. A compressor as defined in claim 1, in which in at least the low pressure stage the male rotor has at least three lands and the female rotor at number of lands exceeding the number of male lands by at least two, but not in excess of double the number of male lands.

8. A compressor as defined in claim 7, in which the number of male lands is three and the number of female lands is five.

9. A compressor as defined in claim 7, in which the number of male lands is four and the number of female lands is six.

11). A compressor as defined in claim 7, in which the 9 number of male lands is four and the number of female lands is seven.

11. A compressor as defined in claim 1, in which means is provided for introducing liquid into the working spaces of the compressor and in Which the connected rotor of the low pressure stage drives the remaining rotor of that stage through the action of the intermeshing lands of the rotors.

12. A compressor as defined in claim 11, in which the connected rotors of both stages drive the remaining rotors of the respective stages through the action of the intermeshing lands of the pairs of rotors of the respective stages.

13. A compressor as defined in clai 1, in which the connected rotors are connected by a torsion shaft fixed to the respective rotors at axially spaced apart places.

14. A compressor as defined in claim 1, in which all of said rotors are separately carried by bearings each wholly independent of the other bearings.

' 1 0 References Cited by the Examiner UNITED STATES PATENTS 2,287,716 6/1942 Whitfield 230-143 2,481,527 9/1949 Nilsson 230--143 2,486,770 11/1949 Whitfield 103128 2,622,787 12/1952 Nilsson 230143 2,659,239 11/1953 Nilsson et a1 103-128 2,975,963 3/1961 Nilsson 230-143 3,073,513 1/1963 Bailey 230143 3,074,624 1/1963 Nilsson et al 230-143 3,138,320 6/1964 Schibbye 230143 3,179,330 4/1965 MacColl 230-443 FOREIGN PATENTS 464,475 4/ 1937 Great Britain.

DONLEY J. STOCKING, Primary Examiner.

WILBUR J. GOODLIN, MARK NEWMAN, Examiners.