US3307777A - Screw rotor machine with an elastic working fluid - Google Patents

Screw rotor machine with an elastic working fluid Download PDF

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Publication number
US3307777A
US3307777A US420370A US42037064A US3307777A US 3307777 A US3307777 A US 3307777A US 420370 A US420370 A US 420370A US 42037064 A US42037064 A US 42037064A US 3307777 A US3307777 A US 3307777A
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Prior art keywords
rotor
rotors
compressor
male
lands
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Expired - Lifetime
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US420370A
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English (en)
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Schibbye Lauritz Benedictus
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Svenska Rotor Maskiner AB
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Svenska Rotor Maskiner AB
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C23/00—Combinations of two or more pumps, each being of rotary-piston or oscillating-piston type, specially adapted for elastic fluids; Pumping installations specially adapted for elastic fluids; Multi-stage pumps specially adapted for elastic fluids
    • F04C23/001—Combinations of two or more pumps, each being of rotary-piston or oscillating-piston type, specially adapted for elastic fluids; Pumping installations specially adapted for elastic fluids; Multi-stage pumps specially adapted for elastic fluids of similar working principle
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
    • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C18/082—Details specially related to intermeshing engagement type pumps
    • F04C18/084—Toothed wheels
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
    • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C18/12—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type
    • F04C18/14—Rotary-piston pumps specially adapted for elastic fluids 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 toothed rotary pistons
    • F04C18/16—Rotary-piston pumps specially adapted for elastic fluids 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 toothed rotary pistons with helical teeth, e.g. chevron-shaped, screw type

Definitions

  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • both rotors 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.
  • 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.
  • the advantage of increased capacity can especially be with their axes coinciding with the bore axes.
  • 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.
  • 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.
  • 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.
  • 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.
  • a male rotor 28 and a female rotor 30 located in the working space 12 in 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • a male rotor 90 and a female rotor 92 are journaled in hearings in the casing 62 but for the sake of simplicity the bearings are shown only diagrammatically as bushes.
  • 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.
  • rotors 98 and 100 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.
  • the male rotors 90 and 98 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Applications Or Details Of Rotary Compressors (AREA)
US420370A 1963-12-23 1964-12-22 Screw rotor machine with an elastic working fluid Expired - Lifetime US3307777A (en)

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SE14418/63A SE310751B (ja) 1963-12-23 1963-12-23

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BE (1) BE657477A (ja)
DE (1) DE1428277C3 (ja)
GB (1) GB1077517A (ja)
SE (1) SE310751B (ja)

Cited By (15)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3467300A (en) * 1967-02-06 1969-09-16 Svenska Rotor Maskiner Ab Two-stage compressor
US3574488A (en) * 1968-04-19 1971-04-13 Plenty & Son Ltd Screw pumps
US3796526A (en) * 1972-02-22 1974-03-12 Lennox Ind Inc Screw compressor
US3910731A (en) * 1970-07-09 1975-10-07 Svenska Rotor Maskiner Ab Screw rotor machine with multiple working spaces interconnected via communication channel in common end plate
US3931718A (en) * 1970-04-16 1976-01-13 Hall-Thermotank Products Ltd. Refrigerant screw compression with liquid refrigerant injection
US4076468A (en) * 1970-07-09 1978-02-28 Svenska Rotor Maskiner Aktiebolag Multi-stage screw compressor interconnected via communication channel in common end plate
US4119392A (en) * 1975-11-27 1978-10-10 Demag Ag Screw compressor with axially displaceable motor
EP0149446A2 (de) * 1983-12-14 1985-07-24 Boge Kompressoren Otto Boge GmbH & Co. KG Drehkolbenverdichter
EP0320956A2 (en) * 1987-12-18 1989-06-21 Hitachi, Ltd. Screw type vacuum pump
US5131817A (en) * 1990-03-22 1992-07-21 The Nash Engineering Company Two-stage pumping system
BE1014461A3 (nl) * 2001-11-08 2003-10-07 Atlas Copco Airpower Nv Oliegeinjecteerde schroefcompressor.
US20040086409A1 (en) * 2002-11-01 2004-05-06 Kabushiki Kaisha Kobe Seiko Sho (Kobe Steel, Ltd.) Screw compressor
US20090232691A1 (en) * 2005-08-25 2009-09-17 Gert August Van Leuven Low-pressure screw compressor
WO2022179130A1 (zh) * 2021-02-26 2022-09-01 珠海格力电器股份有限公司 转子组件、压缩机及空调
US11578723B2 (en) 2016-09-21 2023-02-14 Knorr-Bremse Systeme Fuer Nutzfahrzeuge Gmbh Screw compressor for a utility vehicle

Families Citing this family (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB1503488A (en) 1974-03-06 1978-03-08 Svenska Rotor Maskiner Ab Meshing screw rotor fluid maching
US20160208801A1 (en) * 2015-01-20 2016-07-21 Ingersoll-Rand Company High Pressure, Single Stage Rotor
FR3078449B1 (fr) * 2018-02-27 2020-03-06 Cooper Capri Sas Dispositif d'entree de cable pour cable electrique sous conduit
CN111345484A (zh) * 2020-03-11 2020-06-30 四川自立机械有限公司 节能型保压旋转阀
CN112746958B (zh) * 2021-01-04 2022-07-12 西安交通大学 一种燃料电池用双螺杆压缩膨胀一体机

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GB464475A (en) * 1934-10-16 1937-04-16 Milo Ab Improvements in rotary engines
US2287716A (en) * 1941-04-22 1942-06-23 Joseph E Whitfield Fluid device
US2481527A (en) * 1944-06-29 1949-09-13 Jarvis C Marble Rotary multiple helical rotor machine
US2486770A (en) * 1946-08-21 1949-11-01 Joseph E Whitfield Arc generated thread form for helical rotary members
US2622787A (en) * 1947-07-16 1952-12-23 Jarvis C Marble Helical rotary engine
US2659239A (en) * 1949-10-07 1953-11-17 Jarvis C Marble Independent synchronization
US2975963A (en) * 1958-02-27 1961-03-21 Svenska Rotor Maskiner Ab Rotor device
US3073513A (en) * 1960-04-26 1963-01-15 Svenska Rotor Maskiner Ab Rotary compressor
US3074624A (en) * 1960-03-11 1963-01-22 Svenska Rotor Maskiner Ab Rotary machine
US3138320A (en) * 1959-01-15 1964-06-23 Svenska Roytor Maskiner Aktieb Fluid seal for compressor
US3179330A (en) * 1960-08-30 1965-04-20 James Howden And Company Ltd Rotary engines and compressors

Patent Citations (11)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB464475A (en) * 1934-10-16 1937-04-16 Milo Ab Improvements in rotary engines
US2287716A (en) * 1941-04-22 1942-06-23 Joseph E Whitfield Fluid device
US2481527A (en) * 1944-06-29 1949-09-13 Jarvis C Marble Rotary multiple helical rotor machine
US2486770A (en) * 1946-08-21 1949-11-01 Joseph E Whitfield Arc generated thread form for helical rotary members
US2622787A (en) * 1947-07-16 1952-12-23 Jarvis C Marble Helical rotary engine
US2659239A (en) * 1949-10-07 1953-11-17 Jarvis C Marble Independent synchronization
US2975963A (en) * 1958-02-27 1961-03-21 Svenska Rotor Maskiner Ab Rotor device
US3138320A (en) * 1959-01-15 1964-06-23 Svenska Roytor Maskiner Aktieb Fluid seal for compressor
US3074624A (en) * 1960-03-11 1963-01-22 Svenska Rotor Maskiner Ab Rotary machine
US3073513A (en) * 1960-04-26 1963-01-15 Svenska Rotor Maskiner Ab Rotary compressor
US3179330A (en) * 1960-08-30 1965-04-20 James Howden And Company Ltd Rotary engines and compressors

Cited By (21)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3467300A (en) * 1967-02-06 1969-09-16 Svenska Rotor Maskiner Ab Two-stage compressor
US3574488A (en) * 1968-04-19 1971-04-13 Plenty & Son Ltd Screw pumps
US3931718A (en) * 1970-04-16 1976-01-13 Hall-Thermotank Products Ltd. Refrigerant screw compression with liquid refrigerant injection
US3910731A (en) * 1970-07-09 1975-10-07 Svenska Rotor Maskiner Ab Screw rotor machine with multiple working spaces interconnected via communication channel in common end plate
US4076468A (en) * 1970-07-09 1978-02-28 Svenska Rotor Maskiner Aktiebolag Multi-stage screw compressor interconnected via communication channel in common end plate
US3796526A (en) * 1972-02-22 1974-03-12 Lennox Ind Inc Screw compressor
US4119392A (en) * 1975-11-27 1978-10-10 Demag Ag Screw compressor with axially displaceable motor
EP0149446A2 (de) * 1983-12-14 1985-07-24 Boge Kompressoren Otto Boge GmbH & Co. KG Drehkolbenverdichter
EP0149446A3 (de) * 1983-12-14 1985-08-07 Boge Kompressoren Otto Boge GmbH & Co. KG Drehkolbenverdichter
EP0320956A3 (en) * 1987-12-18 1990-02-21 Hitachi, Ltd. Screw type vacuum pump
EP0320956A2 (en) * 1987-12-18 1989-06-21 Hitachi, Ltd. Screw type vacuum pump
US4984974A (en) * 1987-12-18 1991-01-15 Hitachi, Ltd. Screw type vacuum pump with introduced inert gas
US5131817A (en) * 1990-03-22 1992-07-21 The Nash Engineering Company Two-stage pumping system
BE1014461A3 (nl) * 2001-11-08 2003-10-07 Atlas Copco Airpower Nv Oliegeinjecteerde schroefcompressor.
US20040086409A1 (en) * 2002-11-01 2004-05-06 Kabushiki Kaisha Kobe Seiko Sho (Kobe Steel, Ltd.) Screw compressor
US7104772B2 (en) * 2002-11-01 2006-09-12 Kobe Steel, Ltd. Screw compressor
CN100394029C (zh) * 2002-11-01 2008-06-11 株式会社神户制钢所 螺杆式压缩机
US20090232691A1 (en) * 2005-08-25 2009-09-17 Gert August Van Leuven Low-pressure screw compressor
US7828536B2 (en) * 2005-08-25 2010-11-09 Atlas Copco Airpower, Naamloze Vennootschap Low-pressure screw compressor
US11578723B2 (en) 2016-09-21 2023-02-14 Knorr-Bremse Systeme Fuer Nutzfahrzeuge Gmbh Screw compressor for a utility vehicle
WO2022179130A1 (zh) * 2021-02-26 2022-09-01 珠海格力电器股份有限公司 转子组件、压缩机及空调

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BE657477A (ja) 1965-04-16
DE1428277A1 (de) 1970-01-08
GB1077517A (en) 1967-08-02
DE1428277B2 (de) 1980-01-17
DE1428277C3 (de) 1980-09-11
SE310751B (ja) 1969-05-12

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