US3414189A - Screw rotor machines and profiles - Google Patents

Description

Dec. 3, 1968 J. E. PERSSON SCREW ROTOR MACHINES AND PROFILES 6 Sheets-Sheet 1 Filed June 22, 1966 INVJ%ITOR.

J n EJl/arcz Dec. 3, 1968 J. E. PERSSON 3,414,189

SCREW ROTOR MACHINES AND PROFILES Filed June 22, 1966 6 Sheets-Sheet :2'1

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SCREW ROTOR MACHINES AND PROFILES Filed June 22, 1966 6 Sheets-Sheet 4 Dec. 3, 1968 J. E. PERSSON 3,414,189

SCREW ROTOR MACHINES AND PROFILES Filed June 22, 1966 6 Sheets-Sheet 5 Fig. 10 4 8 INVENTOR.

/en ECZl/ard Arssc m Dec. 3, 1968 J. E. PERSSON 3,414,189 7 SCREW ROTOR MACHINES AND PROFILES Filed June 22, 1966 e Sheets-Sheet 6 I ENTOR.

Jcqn 50 1 3/17] QrSSo/ Patented Dec. 3, 1968 3,414,189 SCREW ROTOR MACHINES AND PROFILES Jan Edvard Persson, Ektorp, Sweden, assignor to Atlas Copco Aktiebolag, Nacka, Sweden, a corporation of Sweden Filed June 22, 1966, Ser. No. 564,469 12 Claims. (Cl. 230-143) ABSTRACT OF THE DISCLOSURE A screw rotor machine comprising intermeshing helically grooved rotors in a casing therefor with clearance spaces being provided between the intermeshing rotors and between them and the casing so as to avoid direct contact between the rotors and reduce the leakage of working fluid from the working chambers.

This invention relates to screw rotor machines comprisring a housing with intersecting parallel bores and inlet and outlet passages communicating with said bores through inlet and outlet ports arranged in the housing and having main and gate rotors with at least three helical threads and grooves on the main rotor intermeshing with at least four helical threads and grooves on the gate rotor. Said main and gate rotors cooperate with one another and with the walls of said bores to define pockets for a working fluid moving from the inlet towards the outlet ports of the housing. In the type of screw rotor machines to which the present invention is related said threads and grooves of the main rotor lie substantially outside the pitch circle of the main rotor and said main rotor threads have a generally convex profile on the leading and trailing flanks, and said threads and grooves of the gate rotor being at least one more than the main rotor threads and grooves and lie substantially inside the pitch circle of the gate rotor and said gate rotor threads have generally concavely curved profiles on the leading and trailing flanks. In machines of this type working as compressors the volumes of said pockets decrease towards the outlet port and in machines working as elastic fluid motors said volumes increase towards said outlet port. Screw rotor machines of this type may also operate as vacuum pumps, gas meters and similar devices, or as pumps or motors operating on incompressible fluids such as oil, water or other material.

In U.S. Patent 1,821,523 Montelius suggests a screw rotor machine with a main and a gate rotor with threads of the above type and in which a root portion of the main rotor thread cooperates with a crest portion of the gate rotor thread to provide rolling contact along said portions which are situated close to the pitch circles. Between said portions a band seal is consequently obtained and the root portion is in a manner well known from cycloid gears produced by travelling generation i.e. by points moving along the crest portion of the gate rotor thread. Montelius discloses a single threaded main screw rotor and a double threaded gate screw rotor which for this and other reasons is not suitable for screw rotor machines according to the invention.

In U.S. Patent 2,174,522 Lysholm has suggested a screw rotor machine in which the main rotor has threads with convex profile which are situated outside the pitch circle of the main rotor and the gate rotor has threads with concave profile which lie inside the pitch circle of the gate rotor. The leading flank of the main rotor thread in Lysholms machine has circular profile and the trailing flank of said thread has a profile generated by the outermost edge of the leading flank of the gate rotor. The trailing flank of the gate rotor threads in Lysholms machine has a circular profile corresponding to the circular profile of the leading flank of the main rotor thread. The leading flank of the gate rotor thread in Lysholms machine has profile which is generated by the summit of the main rotor thread. Lysholms profile as above described and illustrated in FIGS. 2-5 in his patent has the advantage that it produces a continuous sealing line between high and low pressure spaces in the machine thus avoiding leakage except through said seal during the operation of the machine.

Lysholms profile, however, has the disadvantage that the edges of the gate rotor thread which generate the trailing flanks and part of the leading flanks of the main rotor profile and during operation are intended to form space packing between the rotors are very difficult to manufacture with enough accuracy to give the sealing characteristics which are needed to obtain an acceptable efficiency of the machine. Also the same edges are extremely exposed to damage during operation and any damage done to said edges spoils a considerable length of the sealing line and thereby reduces the eificiency of the machine.

When the fluid carrying pockets of Lysholms machine during operation reduce their volume in the outlet portion of the machine and the fluid is being displaced through the outlet port the rotor profiles and the end wall of the housing will close the outlet port area for said pocket before it is fully emptied. This volume which is closed from the outlet port has to be emptied in the inlet port which means a certain loss since the energy of said already compressed volume can not be recovered.

Many attempts have been made to improve the Lysholm profile with respect to the above mentioned disadvantages and also many suggestions have been made to improve the efficiency by means of reducing the internal leakage, by reducing the length of the sealing line and/ or by changing the character of the space packing from its original edge to surface seal to a more eflective surface to surface seal. Lysholm himself partly realized the problem with the sharp edges and in his above mentioned patent, FIG. 6, he suggests for practical reasons to provide a small radial addendum portion on the gate rotor threads and a corresponding radial dedendum portion in the main rotor grooves lying outside the pitch circle of the gate rotor and inside the pitch circle of the main rotor, respectively. This embodiment is illustrated in FIG. 3 on the accompanying drawings which shows a cross section contour of a main rotor thread and meshing gate rotor groove and threads. The addendum portion on Lysholms profile forms rounded leading and trailing edges on the gate rotor thread crests cooperating with correspondingly shaped root portions of the main rotor thread. M and C are main and gate rotor axes. R and R are main and gate rotor pitch circle radii and R and R are bore radii of main and gate rotor bores a b is the main root flank generated by a on the gate rotor thread. a b is the gate rotor flank generated by the summit b of the main rotor thread. r is the radius of the circular main and gate rotor flank portions b c and b c The portion c -di of the main rotor flank is generated by 0 on the gate rotor thread. Addendum f fits in dedendum f For the sake of better clearness the addendum beyond the pitch circle of the gate rotor and the dedendum inside the pitch circle of the main rotor are shown very much larger in proportion to the radii of the rotors than would be desirable in practice. This is what Lysholm says about the size of this modification. The arrows WM and w indicate main and gate rotor directions of rotation.

Lysholm machine is designed to operate with small clearances everywhere between the main and gate rotors and between the housing and the rotors to form the above mentioned space packing alongthe sealing line.

An improved screw rotor machine of similar general design as Lysholms has been suggested by Nilsson in U.S. Patent 2,622,787 which has the advantages that the above mentioned closed pockets in the outlet area are avoided and that the sharp outer edges of the Lysholm gate rotor profile is only generating root portions of the main rotor profile and the critical importance of said edges is reduced. Also a rounded addendum bigger than suggested by Lysholm is added outside the generating point of the gate rotor flank, i.e. outside the pitch circle of the gate rotor in order to keep this still important point less exposed to damage.

FIG. 4 is a cross section contour similar to FIG. 3 of a main rotor thread and meshing gate rotor groove end threads of one embodiment of Nilssons machine. In this embodiment the profiles of both rotors have symmetrically shaped leading and trailing flanks. The gate rotor profile is shaped as a circular arc a -c d with the radius 1' inside the pitch circle and has the above mentioned addendum portion f outside the pitch circle. The main rotor profile complements the gate rotor profile in so far as the main potrion of the flanks adjacent the crest is a circular arc b c d mating the one on the gate rotor. The root portions ti -b and dr-1 of the main rotor flanks are generated by the points a and d on the gate rotor pitch circle and the dedendum portion f inside the pitch circle fits the addendum portion of the gate rotor profile. In FIG. 4 M and G again indicate main and gate rotor axes. R and R are again main and gate rotor pitch circle radii and R and R bore radii of main and gate rotor bores.

Nilssons machine is also designed to operate with small clearances to form a space packing everywhere between the main and gate rotors as well as between the rotors and the housing.

As above mentioned the original Lysholm profile formed a continuous sealing line between the trapped volumes having different pressures. This is not the case with the Nilsson profile. Certain blow holes form passages between these volumes allowing a continuous flow of working medium from high to low pressure spaces. This means a considerable loss and the total efficiency of machines with Nilssons profile is hardly better than that of Lysholms.

The above mentioned screw rotor machine profiles are described and discussed in the Russian book, Screw Compressors, by I. A. Sakun, Moscow, 1960, from which FIGS. 3 and 4 are copied and enlarged. Sakun also describes another modification of the original Lysholm machine which is illustrated in FIG. 5 and which was carried out on a compressor built for the Leningrad Shipbuilding Institution (LSI) in 1949. In FIG. 5 M and G again indicate main and gate rotor axes and R and R the pitch circle radii. R and R are again the bore radii. (l -b is a trailing root portion of the main rotor thread which is generated by a straight line portion (1 -12 of the gate rotor thread. Similarly e f is a leading root portion of the main rotor thread generated by a straight line portion d e of the gate rotor thread. b -c on the main rotor thread is generated by the point b of the gate rotor thread and b c on the gate rotor thread by the summit c of the main rotor thread. c -d on the main rotor thread and c -d on the gate rotor thread are circles with radius r struck on the point P. d e on the main rotor thread is generated by the point d of the gate rotor thread.

The difference between Lysholms original profile and this LSI profile is that in the LS1 profile the point b on the gate rotor profile which is generating the trailing flank of the main rotor profile is moved inside the pitch circle. The portion a b of'the gate rotor profile between the point 12 and the pitch circle, i.e. in this case the outer diameter, is a straight radial line. The corresponding part (1 -12 of the main rotor profile is generated by the line a b i.e. by a point continuously moving along said line. The same modification was carried out on the opposite flanks where the point d on the gate rotor has generated the part d e of the main rotor profile and the straight line d e on the gate rotor has generated the part 61-f of the main rotor. This modification means that the edges on the outer diameter of the gate rotor profile are not relied upon for producing the main part of the sealing line, and the generating points are less exposed to damage. This is the main point in this modification and the differences as compared to the Lysholm profile are so small that as far as efficiency and other operating characteristics are concerned no measurable influence is recognized.

In order to obtain machines according to the above designs which have so good efiiciency that they are competitive with other compressor and gas motor designs and particularly with the piston compressor or motor screw rotor machines of the above described and similar designs have to be carried out with extremely close tolerances and small clearances. Close tolerances are always expensive from the manufacturing point of view and small clearances make the machines sensitive for termal distortion or distortion due to varying high pressures, bearing wear, and impurities in the working fluid, which is usually gas or air. Due to the high operating speed even a slight contact between the parts will usually be disastrous for such machines.

One object of the present invention is to obtain a screw rotor machine in which seizing of the rotors is practically avoided by providing a profile wherein:

(1) Certain portions near the pitch circles of the respective rotors where the relative speed between the rotors is low have a shape providing good gear action which means that the mating surfaces have smallest possible differences between their radii of curvatures and that any contact area between the two surfaces moves continuously relative to both surfaces.

(2) The clearances between the rotors varies along the profile in accordance with the principle that with the rotors timed in a position with contact between the gear portions the clearances between other portions of the profiles are large enough to ensure that contact will not occur at said other portions during any running conditions.

(3) The gate rotor does not need to be driven externally but is continuously driven in the rotation direction by the internal fluid pressure and this internal driving torque does not exceed what the gear portions of the profiles may safely carry. This also means that in oilflooded compressors or other machines handling a lubricating medium synchronizing gears may be dispensed with without any risk of wear of the rotor profiles.

Another object of the invention is to increase the eificiency of the screw rotor machine by providing a profile wherein:

(l) The length of the sealing line compared to the effective volume of the machine is shorter than for earlier known profiles.

(2) The main parts of the space packings along the sealing lines are formed by surfaces with little difference in radii of curvature which means a more efficient seal than with space packing formed by a sharp edge cooperating with a surface.

(3) The clearances and consequently the leakage areas are kept at a minimum between the trailing flank of the main rotor and the leading flank of the gate rotor over a part of the profile which has a long sealing line and also the still remaining part with space packing of the less efficient type sharp edge to surface is reduced. This is obtained by the qualities of the profile mentioned in the previous paragraph under the points 1 and 3. At the first assembly of the machine it is possible to put the above mentioned mating gear portions on the sides in question of the rotor profiles into slight contact, i.e. to let the mentioned gear portions carry the load normally taken by the synchronizing gear. If the rotors of a dry running machine are covered by a wearaway coating of suitable thickness this coating will be exposed to a certain wear under a running-in period until the synchronizing gear takes over the load. The result will be practically zero-clearance adjusted for manufacturing inaccuracies as well as for any distortion due to the working conditions.

(4) The blow hole area between volumes having different pressures, above mentioned as a disadvantage particularly for the profile suggested by Nilsson, is kept small.

(5) The closed pockets in the outlet area above mentioned as a disadvantage particularly for the Lysholm profile but also for the LSI profile are reduced.

In the accompanying drawings FIG. 1 is a longitudinal vertical section and FIG. 2 a plan view of a typical screw compressor, the rotors of which can be carried out with the above mentioned known profiles as well as with a profile according to the present invention.

FIGS. 3-5 illustrate as above described the Lysholm, Nilsson and LSI profiles, respectively, in order to give substantially the present state of the art.

FIGS. 6a and b show diagrammatically and with a heavy line the sealing lines on the main and gate rotors of the LS1 profile shown in FIG. 5. FIG. 60 is a transverse section through the rotor of FIG. 6a. FIG. 6d is a transverse section through the rotor of FIG. 6b.

FIG. 7 illutrates a cross section contour similarly as in FIGS. 3-5 of a main rotor thread and meshing gate rotor groove and threads of what might preferably be called the basic profile according to the present invention.

FIGS. 8a and b are views of the sealing lines or areas on the main and gate rotors with the profile shown in FIG. 7. FIG. 80 is a transverse section through the rotor of FIG. 8a. FIG. 8d is a transverse section through the rotor of FIG. 8b.

FIG. 9 shows on an exaggerated scale the gate rotor profile of the invention together with a theoretical profile with constant clearance in order to indicate the variations of the clearances according to the invention. FIGS. 10a;f show diagrammatically the meshing main and gate rotors in different relative rotor positions.

FIG. ll shows diagrammatically a cross section and FIG. 12 a longitudinal section of the rotors of the invention with the sealing lines between volumes of different pressures indicated in rotational projection with a view to illustrate the surfaces on the gate rotor profile where two opposite flanks of the engaging threads are exposed to different pressures which cause driving or braking torques on the rotors.

The present invention is described hereinbelow in connection with an embodiment of a screw rotor machine in the form of a screw rotor compressor. It should be understood, however, that this embodiment is described as an example only, since the screw rotor compressor is the most common screw rotor machine, but screw rotor machines according to the invention may also be carried out as screw rotor motors, screw rotor pumps, or other screw rotor machines with variable working chamber volumes Within the scope of the claims. Screw rotor compressors according to the invention may have one or more main rotors and one or more gate rotors cooperating therewith and with the housing to provide working pockets which change volume as they move along the screw rotors from an inlet port to an outlet port. Screw rotor machines according to the invention may, furthermore, be singlestage, two-stage or multiple-stage machines, and the number of compression stages, for instance, does not influence the character of the screw rotors according to the invention. The illustrated machine is a single-stage screw rotor compressor with one main rotor and one gate rotor mounted for rotation in intersecting bores with parallel axes in a housing and having intermeshing helical threads and grooves.

The screw compressor illustrated in FIGS. 1 and 2 consists of a housing which comprises a lower portion 31, an upper portion 32, a two-piece inlet end portion 33, 34, and a synchronizing end portion 35. Air or other gaseous fluid is supplied to the screw compressor housing 31, 32 through the inlet connection 36 and compressed air or other gaseous fluid handled by the compressor leaves the compressor through the outlet connection 37. In the compressor housing 31-35 two screw type rotors 38 and 39 are mounted for rotation in the intersecting bores 10 and 11, the rotor shafts being carried in bearings 12 and 13, which may be plain hearings or hell or roller bearings. The rotors are fixed in position endwise by thrust bearings 14, which may be plain cam hearings or ball or roller bearings capable of taking axial thrust. Conventional sealing rings 15 and 16 are provided around the shafts at the rotor ends to prevent leakage into or from the compression chambers. The compressor is provided with synchronizing gears 17, 18. 19 and 20 are oil sealing rings which prevent oil leakage from the bearing housings. From the inlet connection 36 an inlet passage 21 which is carried out with two branches around the rotor shafts leads to an inlet port 22 communicating with the bores 10, 11. An outlet port indicated at 23 in the lower end portion and the bottom of the housing 31 communicates with an outlet passage 24- and the outlet connection 37 of the compressor. The compressor may be driven by any suitable source of power, such as a diesel engine or an electric motor, over a shaft connection 25. The inlet and outlet ports may be shaped in various different ways and, for instance, as illustrated in my application for Screw Rotor Machines filed concurrently with this application.

The main and gate rotors 38 and 39 cooperate in conventional manner to define with one another and with the walls of the bores 10 and 11 pockets for a working fluid moving from the inlet port 22 towards the outlet port 23 of the housing and having a volume which changes during said movement. The shape of the main and gate rotors 38, 39 which is characteristic of the present invention is disclosed in FIGS. 712. FIG. 7 is a cross section of a portion of the housing 31, 32 and intermeshing threads and grooves of the rotors 38 and 39 and shows the general shape of the profiles without consideration to the clearances and seals which are more particularly illustrated in connection with FIGS. 810. The same reference letters have been used in FIG. 7 as in the cross section contours of known screw rotor machines illustrated in FIGS. 35 and, consequently, the letters M and G in FIG. 7 indicate the main and gate rotor axes. R and R are the main and gate rotor pitch circle radii and R and R are the bore radii of the main and gate rotor bores. The bore radius R of the gate rotor is in the illustrated embodiment the same as the pitch circle radius R of the gate rotor. a b -c is the main rotor trailing flank and a b c the leading flank of the gate rotor. a b is a band seal portion formed as a straight line and is a portion of a radius of the gate rotor. b2C2 is without consideration to the clearances generated by the summit c of the main rotor thread. (l -b is the trailing root portion of the main rotor thread and is without consideration to the clearances generated by the straight line portion a b of the gate rotor thread so that a band seal is obtrined between said portions. The point a is generated by a and the point b by b also without consideration of clearances. The portion b c of the main rotor trailing flank is generated by the point 11 or the inner end portion of the band seal portion a b on the gate rotor also without consideration to the clearances.

The portion of the main rotor thread from c d is circular with a radius r struck on the point P which is the point of intersection between the pitch circles. The portion ai -e of the leading main rotor thread is without consideration to the clearances generated by a portion d e which is a straight radial portion of the trailing gate rotor thread. The portion c d of the trailing gate rotor thread is a circular arc with substantially radius r which with suitable clearance conforms with the arc c d FIGS. 8a and 1) illustrate the sealing between a pair of screw rotors such as disclosed in FIG. 7 and it should be noted that an uninterrupted sealing line is obtained substantially all the way to the line of intersection 26 between the intersecting bores 10 and 11. The blow hole obtained in connection with the above mentioned known compressor is here as in the above mentioned LSI-compressor reduced to a minimum of no importance since the sealing lines between the rotors meet substantially with the line of intersection 26. The relative quality of the seal along the sealing lines of the profiles according to the invention are indicated by the number of parallel dot and dash lines in FIGS. 8a and b and are referred to later on for comparison with FIGS. 11 and 12.

FIG. 9 shows in somewhat exaggerated scale an example of a gate rotor profile according to the invention and a constant clearance profile shown in dotted lines and illustrates the intended variations in clearances for a gate rotor as compared with a conventional thread when cooperating with a main rotor profile. The same clearance variations may of course be obtained by suitable modification of the main rotor profile instead or even by different combinations of modifications of both rotor profiles and all such combinations which give the cooperating characteristics according to this invention are intended to be included under the scope of this patent.

FIG. 10 shows cooperating parts of rotor profiles according to the invention in different angular positions and illustrates the varying clearances. The clearances in various portions of the profiles are marked Cl in FIGS. 10af with a number and in a suitable embodiment Cl 1 is smaller than Cl 7, Cl 5 and Cl 4; Cl 2 is smaller than Cl 4 and Cl 8 is smaller than Cl 5 and Cl 6.

This means that the above described cross sectional contour of the rotors according to the invention results in minimum clearance or no clearance between the generated root portion tl b and the band seal portion a b of the gate rotor. In this area very little relative motion occurs between the gate rotor and the main rotor during operation of the machine. It has therefore been found that contact can be permitted without lubrication in this area of a screw compressor running at high speed. If contact occurs in screw compressors of this design provided with synchronizing gears such as 17, 18 such contact only results in a polishing effect on the contacting surface portions. Due to the small relative motion between the gate rotor and the main rotor in this area very little heat through friction or rubbing is produced and consequently no deformation of the rotors occurs from this reason. In the area between b and c the main rotor summit c moves with considerable speed relative to the surface of the gate rotor flank and in this area the clearance Cl 5, Cl 6 according to the invention has therefore been made so large that under all conditions contact will not occur between the summit c and any point of the surface 17 In this way applicant provides a screw rotor machine in which dangerous or disastrous contacts between the screw rotors is avoided and seizing of the rotors from this reason can therefore not occur. Similarly, any rotor deformation due to compression heat or fluid pressure on the rotors can only result in contact between the band seal portions a b and the root portions a b which is not injurious to the rotors.

The clearance Cl 4 between the circular arc portions c d and c d is according to the invention made large enough to prevent seizing in this area, and the clearance Cl 2 is smaller than the clearance Cl 4 so that any contact due to disturbances in the synchronization, for example, occurs between the root portion d e and the portion d 2 of the gate rotor in areas where the relative movement between said portions is small. Furthermore, the clearance Cl 1 is made less than Cl 7 and consequently any contact due to bearing wear or distortion of the rotors transversely to the rotor axes results in nondangerous rolling contact between the gate rotor crests and the cylindrical bottom of the main rotor grooves.

The size of the above described different clearances depends upon several factors such as:

(1) the rotor diameters (2) working conditions for the compressor (3) cooling arrangements at the compressor housing and the rotors (4) manufacturing tolerances for the rotors, bearings,

synchronizing gears and compressor housing.

As an example may be mentioned that for a rotor diameter of 200 mm. for a compressor intended to operate at a maximum pressure ratio of 4.5: l, a maximum inlet temperature of 50 C., with oilcooled housing and rotors and with the manufacturing tolerances suited for normal production in series the size of the variations of the clearances in the different areas as compared to conventional constant clearance machines may amount to about :30%. This means that the clearance between a "e and the bottom of the main rotor grooves, between a "-b and a b on the main rotor trailing flank, and between d e and al -e on the main rotor leading flank is about of the clearance in conventional constant clearance machines. Similarly the clearance between b "a and b c on the trailing main rotor flank is about of the clearance in said constant clearance machines, and the same applies to the clearance between 0 and b "c or at least in the area close to b FIGS. 11 and 12 illustrate in axial view and rotational projection the sealing lines between the intermeshing portions of the rotors 38, 39 and show how the compressed gas acts on the gate rotor lobes in the rotation direction. For this purpose all points on the sealing lines are shown at their correct radial distances from the gate rotor centerline as indicated by the arrows x, y, z. This means that the areas A enclosed by a lobe of the sealing line or the areas B enclosed between the sealing line and the periphery C of the gate rotor show the peripherally projected gate rotor flank areas exposed to pressure in opposite rotational directions and represent the correct size of the areas on which said pressures act to give a driving or braking torque on the rotor. The point 1 in FIGS. 11 and 12 corresponds to the point P (2 in FIG. 7, 2 to d 3 to c 4 to b 5 to I2 and 6 to P (a The same numbers as in FIGS. 11 and 12 are used in FIG. 8b to facilitate the comparison between the sealing lines in these views. The cross sections to the left in FIGS. 8a and I) serve to show how the projections on the side views are taken looking on the respective rotor. It would be obvious that with the rotational direction Wm, Wg indicated in FIG. 7 the pressure difference within the area A will cause a driving action and the pressure difference within the area B will cause a braking action on the gate rotor. By changing the length of the profile parts a b and d -e in FIG. 7 the areas A and B in FIG. 2 may be influenced. According to the invention the profile part a b may preferably be chosen 550% of the depth of the gate rotor thread and d e may be chosen 525% of Said depth. To make the profile portion a b short means that the blow hole area is small and the sealing line long. On the contrary a long portion a -b means a bigger blow hole and a shorter sealing line of better quality. With increasing rotor dimensions the blow hole area will increase in proportion to the square of the diameter whereas the length of the sealing tline increases in proportion to the diameter. This means that the optimum length of the profile part a b depends upon the rotor dimensions. Also the manufacturing tolerances which will influence the leakage through the space packing along the sealing line more than the leakage through the blow hole has to be taken into consideration when choosing the length of the part a b As an example it may be mentioned that for a rotor diameter of 200 mm. with the manufacturing tolerances suited for normal serial production the best length of the profile part a b has been found to be about 10% of the depth of the gate rotor thread. The length of the profile portion d -e may thereafter be chosen to give the desired size of the area A so that a suitable driving torque is obtained on the gate rotor. Since the torque required depends upon the rotor bearing friction the use of antifn'ction bearings will be advantageous for the efflciency of the compressor also by indirectly reducing the internal leakage by allowing the profile portion 12-62 to be increased, which also results in a somewhat shorter sealing line. For the abovemcntioned compressor equipped with antifriction bearings the length of the profile part d e may preferably be about 25% of the gate rotor thread depth.

To be sure of correct gear function and smooth running of the compressor under all operating conditions the number of threads of a compressor according to the invention must on the main rotor be at least three and the rotor length and lead such that the main rotor root portion (l -b is in constant mesh with the band seal portion a b of the gate rotor at least in three cross sections along the rotors. The gate rotor must have at least one more thread than the main rotor.

The present invention has given the surprising result that the efiiciency of a screw compressor built with normal manufacturing tolerances for serial production is increased by several percents so that a screw compressor may be built according to the invention which is almost competitive with a reciprocating piston compressor with regard to efliciency. It has also been found that in the case of oil flooded compressors in which cooling and lubricating and sealing oil is injected for taking care of the compression heat and for providing a seal between the rotors and the housing synchronizing gears may be dispensed with without causing any wear of the rotor portions a b (l -b or d e d e which can destroy the important clearances.

In the rotors according to the invention further improvements of the efficiency may be obtained by providing paint or flock or other wear away material on the gate rotor threads or the main rotor threads or both substantially over the areas between the points b c and -11 and also between points [1 -0 and c d The screw compressor above described and illustrated on the drawings should only be considered as an example and may be modified in various ways within the scope of the following claims. The invention also includes separate screw rotors for screw rotor machines according to the invention.

What I claim is:

1. A screw rotor machine comprising (a) a housing with intersecting parallel bores and inlet and outlet passages communicating with said bores through inlet and outlet ports arranged in the housing (b) main and gate rotors mounted for rotation in said housing in said bores and having at least three intermeshing helical threads and grooves, sai-d main and gate rotors cooperating to define with one another and with the walls of said bores pockets for a working fluid moving from the inlet towards the outlet ports of the housing (c) said threads and grooves of the main rotor lying substantially outside the pitch circle of the main rotor and said main rotor threads having a generally convex profile on the leading and trailing flanks, and said threads and grooves of the gate rotor being at least one more than the main rotor threads and grooves and lying substantially inside the pitch circle of the gate rotor and said gate rotor threads having generally concavely curved profiles on the leading and trailing flanks, and

(d) a first band seal portion on the leading flank of said gate rotor thread extending from the crest of the gate rotor thread a short distance towards the gate rot-or axis, a cooperating first portion of the trailing flank of the main rotor thread extending from the root of the main rotor thread to form a first root portion generated with a first clearance by said first band seal portion on the leading gate rotor flank, a second portion between said first root portion and the summit of the main rotor thread substantially generated by an inner end portion of the first band seal portion with substantially larger second clearance than said first clearance, a second band seal portion on the trailing flank of the gate rotor thread extending from the crest of the gate rotor thread a short distance towards the gate rotor axis, a cooperating first portion of the leading flank of the main rotor thread extending from the root of the main rotor thread to form a first portion of the leading main rotor flank generated with a third clearance by said second band seal portion, a second portion between said first portion of the leading main rotor flank and the summit of the main rotor thread substantially formed as a circular arc and cooperating with a portion of the trailing gate rotor flank formed as a circular arc conforming with said second portion on the leading main rotor thread and cooperating therewith with a fourth clearance substantially larger than said third clearance, and a second portion of the leading gate rotor flank between said first band seal portion and the bottom of the gate rotor groove generated by the summit of the main rotor thread with substantially larger fifth clearance than said first clearance.

2. A screw rotor machine according to claim 1, in which said second, third, fourth and fifth clearances are about twice said first clearance.

3. A screw rotor machine according to claim 1, in which the radial length of said first and second band seal portions on the leading and trailing gate rotor flanks is so large that first areas are formed on the gate rotor flanks which are subjected to a driving fluid pressure and second areas are formed on the gate rotor flanks which are subjected to a braking fluid pressure, said first areas being so much larger than said second areas that the frictional resistance of the gate rotor is substantially overcome and the gate rotor is substantially driven by fluid pressure.

4. A screw rotor machine according to claim 1, in which said band seal portions are straight portions of gate rotor radii.

5. A screw rotor machine comprising (a) a housing with intersecting parallel bores and inlet and outlet passages communicating with said bores through inlet and outlet ports arranged in the housing (b) main and gate rotors mounted for rotation in said housing in said bores and having at least three intermeshing helical threads and grooves, said main and gate rotors cooperating to define with one another and with the walls of said bores pockets for a working fluid moving from the inlet towards the outlet ports of the housing (0) said threads and grooves of the main rotor lying substantially outside the pitch circle of the main rotor and said main rotor threads having a generally convex profile on the leading and trailing flanks, and said threads and grooves of the gate rotor being at least one more than the main rotor threads and grooves and lying substantially inside the pitch circle of the gate rotor and said gate rotor threads having generally concavely curved profiles on the leading and trailing flanks, and

(d) a first band seal portion on the leading flank of said gate rotor thread extending from the crest of the gate rotor thread a short distance towards the gate rotor axis, a cooperating first portion of the trailing flank of the main rotor thread extending from the root of the main rotor thread to form a first root portion generated with minimum clearance by said first band seal portion, a second band seal portion on the trailing flank of the gate rotor thread extending from the crest of the gate rotor thread a short distance towards the gate rotor axis, a cooperating first portion of the leading flank of the main rotor thread extending from the root of the main rotor thread to form a first portion of the leading main rotor flank generated with minimum clearance by said second band seal portion, second portions on the trailing and leading main rotor thread flanks extending from said generated root portions to the summit of the main rotor thread and cooperating with said gate rotor with substantially larger clearance than said minimum clearance, and wear away surface coating material on said second portions reducing at least a portion of said larger clearance.

6. A screw rotor machine comprising:

(a) a housing with intersecting parallel bores and inlet and outlet passages communicating with said bores through inlet and outlet ports arranged in the hous- (b) main and gate rotors mounted for rotation in said housing in said bores and having at least three intermeshing helical threads and grooves, said main and gate rotors cooperating to define with one another and with the walls of said bores pockets for a work ing fluid moving from the inlet towards the outlet ports of the housing;

() said threads and grooves of the main rotor lying substantially outside the pitch circle of the main rotor and said main rotor threads having a generally convex profile on the leading and trailing flanks, and said threads and grooves of the gate rotor being at least one more than the main rotor threads and grooves and lying substantially inside the pitch circle of the gate rotor and said gate rotor threads having generally concavely curved profiles on the leading and trailing flanks;

(d) a band seal portion on the leading flank of said gate rotor thread extending from the crest of the gate rotor thread a short distance towards the gate rotor axis, a cooperating first portion of the trailing flank of the main rotor thread extending from the root of the main rotor thread to form a root portion generated by said band seal portion with a certain first clearance, and a second portion on the trailing main rotor flank between said first generated root portion and the summit of the main rotor thread which second portion is substantially generated by a point which lies so much above the surface of the leading gate rotor flank at the inner end of the band seal portion that a second clearance is provided between the band seal portion and said second portion of the trailing main rotor flank between the generated root portion and the main rotor thread summit, said second clearance within the main part of said second portion being at least larger than said first clearance between said band seal portion on the leading gate rotor flank and said first generated root portion of the main rotor thread whereby direct contact between the band seal portion of the leading flank of the gate rotor thread and the main rotor thread is avoided over said second clearance portion;

(e) a band seal portion on the trailing flank of the gate rotor thread extending from the crest of the gate rotor thread a short distance towards the gate rotor axis, a cooperating first portion of the leading flank of the main rotor thread extending from the root of the main rotor thread towards the summit and generated With a third clearance by said band seal portion on the trailing gate rotor thread flank, and a second portion of the leading main rotor flank between said first generated root portion and the summit of the main rotor thread and cooperating with a portion of the trailing gate rotor flank extending from said band seal portion to the bottom of the gate rotor groove and conforming with said second portion on the leading main rotor thread and cooperating therewith with a fourth clearance which is at least 30% larger than said third clearance; and

(f) a second portion of the leading gate rotor thread flank between the band seal portion and the foot of the gate rotor thread generated by the summit of the main rotor thread with a fifth clearance which over the main part of said second portion is at least 30% larger than said first clearance between said first band seal portion and the generated portion at the root of the trailing main rotor thread. 7. A screw rotor machine according to claim 6, in which a band seal portion is provided on the trailing flank of the gate rotor thread extending from the crest of the gate rotor thread a short distance towards the gate rotor axis, a cooperating first portion of the leading flank of the main rotor thread extending from the root of the main rotor thread to form a first root portion generated with minimum clearance by said band seal portion on the trailing gate vrotor flank, and a second portion between said first root portion and the summit of the main rotor thread substantially formed as a circular arc and cooperating with a portion of the trailing gate rotor flank formed as a circular arc conforming with said second portion on the leading main rotor thread and cooperating therewith with a third clearance substantially larger than said minimum clearance.

8. A screw rotor machine according to claim 6 in which the radial length of the band seal portions of the leading and trailing gate rotor thread flanks is so large, that first areas are formed on the gate rotor thread flanks which are subjected to a driving fluid pressure, and that second areas are formed on the gate rotor thread flanks which are subjected to a braking fluid pressure, said first areas being so much larger than said second areas that the frictional resistance of the gate rotor is substantially overcome and the gate rotor is substantially driven by fluid pressure.

9. A screw rotor machine according to claim 6 in which the band seal portions on the leading and trailing gate rotor thread flanks are formed by straight portions of gate rotor radii.

10. A screw rotor machine according to claim 6 in which a cylindrical portion at the crest of the gate rotor thread and a cylindrical portion at the bottom of the main rotor groove said cylindrical portions are lying on the pitch circles of the gate and main rotors, respectively, so that they can cooperate with rolling contact, a sixth clearance between said cylindrical portions being at least 30% less than a seventh clearance between the summit of the main rotor thread and the bottom of the gate rotor groove, whereby contact is avoided between the main rotor thread summit and the bottom of the gate rotor grooves.

11. A pair of cooperating screw rotors with coplanar axes and with intermeshing helical threads and grooves forming a main rotor and a gate rotor,

(a) said threads and grooves of the main rotor lying substantially outside the pitch circle of the main rotor and said main rotor threads having a generally convex profile on the leading and trailing flanks, and said threads and grooves of the gate rotor being at least one more than the main rotor threads and grooves lying substantially inside the pitch circle of the gate rotor and said gate rotor threads having generally concavely curved profiles on the leading and trailing flanks,

(b) a band seal portion on the leading flank of said gate rotor thread extending from the crest of the gate rotor thread a short distance towards the gate rotor axis, a cooperating first portion of the trailing flank of the main rotor thread extending from the root of the main rotor thread to form a root portion generated by said band seal portion with a certain first clearance and a second portion on the trailing main rotor flank between said first generated root portion and the summit of the main rotor thread which second portion is substantially generated by a point which lies so much above the surface of the leading gate rotor flank at the inner end of the band seal portion that a second clearance is provided between the band seal portion and said second portion of the trailing main rotor flank between the generated root portion and the main rotor thread summit said second clearance Within the main part of said second portion being at least 30% larger than said first clearance between said band seal portion on the leading gate rotor flank and said first generated root portion of the main rotor thread whereby direct contact between the band seal portion of the leading flank of the gate rotor thread and the main rotor thread is avoided over said second clearance portion,

(c) a band seal portion on the trailing flank of the gate rotor thread extending from the crest of the gate rotor thread a short distance towards the gate rotor axis, a cooperating first portion of the leading flank of the main rotor thread extending from the root of the main rotor thread towards the summit and generated with a third clearance by said band seal portion on the trailing gate rotor thread flank, and a second portion of the leading main rotor flank between said first generated root portion and the summit of the main rotor thread and cooperating with a portion of the trailing gate rotor flank extending from said band seal portion to the bottom of the gate rotor groove and conforming with said second portion on the leading main rotor thread and cooperating therewith wit-h a fourth clearance which is at least 30% larger than said third clearance, and

(d) a second portion of the leading gate rotor thread flank between the band seal portion and the foot of the gate rotor thread generated by the summit of the main rotor thread with a fifth clearance which over the main part of said second portion is at least 30% larger than said first clearance between said first band seal portion and the generated portion at the root of the trailing main rotor thread.

12. A pair of cooperating screw rotors according to claim 11 in which a cylindrical portion at the crest of the gate rotor thread and a cylindrical portion at the bottom of the main rotor groove said cylindrical portions are lying on the pitch circles of the gate and main rotors, respectively, so that they can cooperate with rolling contact, a sixth clearance between said cylindrical portions being at least 30% less than a seventh clearance between the summit of the main rotor thread and the bottom of the gate rotor groove, whereby contact is avoided between the main rotor thread summit and the bottom of the gate rotor grooves.

References Cited UNITED STATES PATENTS 2,457,314 12/1948 Lysholm 230143 2,473,234 6/1949 Whitfield 103-128 2,477,002 7/1949 Paget 230-143 XR 2,486,770 11/1949 Whitfield 103-128 2,622,787 12/1952 Nilsson 230143 2,922,377 1/1960 Whitfield 103128 3,245,612 4/1966 Nilsson et a1. 230-143 3,314,598 4/1967 Lysholm 230-143 FOREIGN PATENTS 324,301 10/1957 Switzerland.

FRED C. MATTERN, 111., Primary Examiner.

T. R. HAMPSHIRE, Assistant Examiner.

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