EP0149446A2 - Compresseur à piston rotatif - Google Patents

Compresseur à piston rotatif Download PDF

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Publication number
EP0149446A2
EP0149446A2 EP84890238A EP84890238A EP0149446A2 EP 0149446 A2 EP0149446 A2 EP 0149446A2 EP 84890238 A EP84890238 A EP 84890238A EP 84890238 A EP84890238 A EP 84890238A EP 0149446 A2 EP0149446 A2 EP 0149446A2
Authority
EP
European Patent Office
Prior art keywords
rotor
main rotor
tooth
secondary rotor
screw
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Withdrawn
Application number
EP84890238A
Other languages
German (de)
English (en)
Other versions
EP0149446A3 (fr
Inventor
Gerold Dipl.-Ing. Dr. Riegler
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Boge Kompressoren Otto Boge GmbH and Co KG
Original Assignee
Boge Kompressoren Otto Boge GmbH and Co KG
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Boge Kompressoren Otto Boge GmbH and Co KG filed Critical Boge Kompressoren Otto Boge GmbH and Co KG
Publication of EP0149446A2 publication Critical patent/EP0149446A2/fr
Publication of EP0149446A3 publication Critical patent/EP0149446A3/fr
Withdrawn legal-status Critical Current

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Classifications

    • 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

Definitions

  • the invention relates to a parallel and external-axis rotary compressor with a housing provided with inlet and outlet openings and at least two rotors as the main and secondary rotor, which are equipped with helically wound tooth gaps and are arranged axially parallel and interlocking, the tooth profiles of the main rotor being essentially convex and located outside the pitch circle and the tooth flanks of the secondary rotor are essentially kcnkav and located within the pitch circle.
  • screw compressors have been built for around thirty years, they have only become widely used as series machines in the past fifteen years. In terms of quantities, it is primarily the smaller injection oil-cooled air or refrigerant compressors that make up the majority of the world production of screw compressors today.
  • screw compressors are twin-shaft rotary or rotary piston compressors. They work similarly to the well-known piston compressors according to the principle of displacement.
  • the working spaces in screw compressors are the gaps in the teeth of two intermeshing helical toothed rotors that run in a housing that closely encloses the rotors.
  • tooth flank profile is known for example from DE OS 26 39 870 and DE OS 27 35 670.
  • tooth flank profiles described for screw rotors which are composed of a large number of flank sections, such as arcs, ellipse curves, involutes, cycloids and hyperbolic curves, in the face cut. Because of the complexity of these flank curves, extremely complex manufacturing processes or very expensive tools are usually required, which often do not allow economical manufacture of the rotors, particularly in the case of small screw compressor units.
  • DE OS 31 40 107 proposes a rotor profile in which the tooth flanks of the main rotor are not composed of curve segments, but are formed from one head point of the main rotor to the next by a steady, uniform, analytically definable curve shape.
  • the version with pointed secondary rotor teeth known from DE-OS 31 40 107 can also expect high wear and heating during operation, which in turn increases the leakage gas quantities and the compression efficiency is deteriorated.
  • the invention has for its object to avoid the disadvantages described above and to create a toothing for a screw compressor system, which is characterized in the same way for main and secondary rotor by simple manufacture, robust design and small leakage gaps. Furthermore, an inexpensive manufacture is said to be simple procedures to be carried out, such as hobbing possible and low-wear operation can be ensured.
  • both flanks of the main rotor teeth are formed continuously from the tooth base to the tooth head from a screw turret with a continuous curve.
  • a screw gate is created as the envelope of a screwed plane, provided that it is at an angle to the screw axis. Due to the unwinding of the screw gates, very simple production by hobbing or by planing or butting is possible.
  • the toothing according to the invention results in an involute which, according to the known toothing setters, also involves an involute and thus a screw gate on the secondary rotor, which, like the main rotor, can be produced in a simple manner.
  • the toothing according to the invention results in a curve which is continuous from the tooth base to the tooth tip. This avoids inequalities and cracks in the tooth flanks and peaks in the power transmission during operation and thus prevents unnecessary sliding and wear.
  • the main rotor teeth are symmetrical in frontal section with respect to a partial plane containing the screw axis, the root circle radius of the main rotor tooth preferably being equal to the pitch circle radius of the main rotor, while the fillet circle radius of the screw pusher is preferably smaller than the pitch circle radius.
  • a symmetrical profile shape has the advantage that the production of both tooth flanks on the main and secondary rotor is identical.
  • the choice of the pitch circle as the root circle results in the advantage that when the main and secondary rotor are rolled off, there is no relative speed between the main rotor tooth root and the secondary rotor tooth head, and therefore no wear occurs, which leads to higher leakage losses and thus poorer compressor efficiency.
  • the outer diameter of the secondary rotor is equal to the pitch circle of the secondary rotor.
  • the main rotor consists of at least 3 teeth
  • a particularly advantageous embodiment of the profile according to the invention in accordance with the characterizing part of the claim also consists in the fact that the main rotor teeth are asymmetrical with respect to a partial plane containing the screw axis in the end section, the throat radius of the suction-side screw gate of the main rotor tooth in engagement with the secondary rotor being smaller than the throat radius of the pressure-side screw gate of the main rotor tooth which is in engagement with the secondary rotor.
  • Conductor proves to be advantageous if, according to the invention, the circular radius of the pressure-side screw gate is equal to the rolling radius of the main rotor and the two unequal screw gates of the asymmetrical profile come to rest in such a way that the two involutes intersect at one point on the outer diameter of the main rotor.
  • the leakage gap currents through this blow hole are negligibly small and thus improve the compressor efficiency compared to known embodiments.
  • the manufacturing accuracy can be increased and the rotor manufacturing can be made cheaper.
  • a screw compressor system according to the invention is significantly cheaper both in operation and in purchase compared to the systems currently available on the market.
  • the profile according to the invention that, by means of a center distance correction, the main rotor turrets in the end section in the zero position rest in two points on the secondary rotor turrets, as a result of which three-point contact of the toothing occurs in the zero position.
  • the zero position of the toothing is defined by the rotor position at the end of the compression or extension. If there is no contact of the main rotor and secondary rotor flanks in three points in this position, the teeth are not tight in this position and there is an additional leakage gap opening, which leads to gas mass losses and thus a lower efficiency.
  • the housing inner wall of the secondary rotor can at the same time be designed as a bearing shell for the secondary rotor. This eliminates the need for an expensive bearing construction for the secondary runner.
  • This storage according to the invention has the further advantage that this storage is very robust and practically does not allow deflection of the secondary runner. On the one hand, this prevents tarnishing of the secondary rotor teeth on the housing bore, on the other hand there is no enlargement of the leakage gap areas during operation due to deflection, so that the internal compressor efficiency is improved. Lubrication of the bearing points is ensured at all times by injection of Ule during compression.
  • the rotary piston compressor is expediently driven via the secondary rotor.
  • the constant pressure angle in involute toothing means that the power transmission angle can be freely selected when designing the toothing, as a result of which correspondingly favorable transmission angles can be achieved.
  • the advantage of driving the secondary rotor is that the circumferential speed of the rotor pair is greater and thus the leakage gap quantities per compressed gas quantity unit are smaller. This is shown positively in the specific power requirement and the internal compression efficiency.
  • the main rotor and secondary rotor pitch of the screw connection can be freely selected. Depending on the slope selected, this results in very short, robust and cheap runners, which are characterized by a slight deflection compared to known runners.
  • the toothing according to the invention also has advantages in the case of intermittent operation of screw compressors, since, due to the favorable, freely selectable power transmission angle, the blows are easily absorbed each time the compressor starts up and increased wear can be avoided.
  • FIGS. 1-8 Further details of the invention are explained in more detail with reference to the exemplary embodiments shown schematically in FIGS. 1-8
  • Fig. 1 the end section of a screw compressor system 1 according to the invention with a main rotor 2 and a secondary rotor 3 is shown schematically in a compressor housing 4.
  • the rotors each standing in meshing engagement with each other is 2.3 in the usual way seen from tiksei- g ti en end of the compressor unit 1 of the zero position, that is shown on Ausschubende.
  • the gas to be compressed is sucked into the screw-type compressor system 1 on the suction side 5 axially opposite the front section in FIG. 1 and, after the compression chambers 6 and 7 have been closed, passes through the rotation of the main rotor 2 in the direction of the arrow. 8 and the associated rotation of the secondary rotor 3 in the direction of arrow 9 to the pressure side 10, wherein it is compressed by reducing the compression chamber 11 and, after reaching the compression end pressure, is connected to the pressure line through an outlet opening (not shown) in the housing 4.
  • the main rotor 2 has four teeth 12, while the secondary rotor 3 has six teeth 13.
  • the flanks of each main rotor tooth 12 are each formed from 2 screw gates 14 arranged symmetrically to one another, which result in pointed circular involutes 15 in the end cut.
  • the fillet circle of this pointed circular involute was designated 16, the fillet circle radius being smaller than the radius of the main rotor pitch circle 17.
  • the screw torsion surface 14 ends on the cylinder surface 18, which is determined by the radius of the main rotor rolling circle 17. This results in the end section in FIG. 1 of the screw-type compressor system 1 according to the invention, a short circular arc piece 19 between the sections of the pointed circular involutes 15, whereby the tip 2C of the profile, which cannot be produced, can be avoided.
  • the secondary rotor tooth gap 21 belonging to the main rotor 12 is formed by the point path 22 of the main rotor tooth tip 23 and the envelope path 24 of the main rotor flank 14; the point path 22 represents a convoluted cycloid and the envelope path 24 is an involute in the face cut according to the known toothing laws.
  • the secondary rotor tooth is delimited by a circular arc piece 25 on the outside diameter, the radius of this circular arc piece 25 being equal to the radius of the secondary rotor rolling circle 26.
  • FIG. 2 shows an enlarged section of the pair of runners 2, 3 in meshing engagement with a symmetrical profile according to the invention.
  • the main rotor was again identified with 2, the secondary rotor with 3, the schematic housing with 4, the main rotor flank with 14, the secondary rotor cycloid with 22, the secondary rotor involute with 24 and the secondary rotor arc with 25.
  • the main rotor tip 23 moves in a stationary reference system 30 along the circular arc 31, the main rotor tip 23 being in engagement with the secondary rotor cycloid 22 up to point 32.
  • the pressure-side blow hole which is a leakage area caused by the geometry, arises between the last point of engagement 32 of the main rotor tip 23 with the secondary rotor cycloid 22 and the section of the housing bore 33 is, so that only small amounts of leakage gas can flow through the blow hole.
  • Fi g . 3 shows an end section similar to FIG. 1 of a screw compressor system 1, but with the asymmetrical flank profile according to the invention.
  • Fig. 3 the same machine parts with the same numerals from Figs. 1 and 2 have been designated.
  • each main rotor tooth 12 consists of a screw chest 40 and a screw chest 41, which result in a pointed circular involute 42 and 43 in the face cut.
  • only one branch of screw tors 40 and 41 or of pointed circular involutes 42 and 43 is used.
  • the fillet circle radius is again designated 16, the radius of the fillet circle 16 again being smaller than the pitch circle radius 17 of the main rotor.
  • the screw gate 40 is only used up to the cylinder 18, which is formed by the radius of the pitch circle 17.
  • the screw gate 41 has a fillet radius 44 which is larger than the fillet radius 16 of the screw base 40 and is preferably equal to the radius of the pitch circle 17 of the main rotor.
  • the screw gate 41 is formed in such a way that the two screw gates 40 and 41 converge to a tip 46 at a predetermined outer diameter 45.
  • the secondary rotor cycloid 22 on the pressure side of the pair of teeth in engagement reaches approximately the secondary rotor radius 47, which is preferably equal to the radius of the secondary rotor pitch circle 26.
  • the blow hole section 34 between the last engagement point 32 and the housing intersection 33 can be further reduced, so that the blow hole area and thus the leakage quantities due to this geometry-related gap become negligibly small.
  • FIG. 4 shows that rotary position of the rotor pair 2, 3, in which the main rotor tip 46 touches the secondary rotor cycloid 22 for the first time when the main rotor rotates in the direction of arrow 8, and has thus just exceeded the blow hole section 34.
  • the small blowhole area can also be recognized from this illustration or from the enlargement FIG. 5.
  • FIG. 6 shows an enlarged section of the engagement in the case of toothing of the symmetrical profile corrected for the center distance. This demonstrates the possibility of achieving an exact three-point contact in the zero position by correcting the center distance, which is necessary to reduce the leakage gap areas without impairing the running or rolling properties of the toothing.
  • the intersection between the secondary rotor cycloid 22 and the secondary rotor turret 24 was designated 48 and the flank contact point between the main rotor flank 15 and the secondary rotor turret 24 was designated 49.
  • a secondary rotor bearing according to the invention has also been drawn in in FIG. 3.
  • a bearing shell 51 was inserted into the secondary rotor housing bore 50 by cladding with a UTB Brcnz electrode, the Brinell hardness of the bronze being adjustable in a known manner by the current strength during cladding.
  • Good storage and running conditions can be achieved by using a naturally hard secondary rotor 3, which meshes with a hardened main rotor 2, and using a naturally hard housing into which the hardened bronze bearing 51 has been introduced.
  • FIG. 7 shows a schematic longitudinal section through a screw compressor system 1 according to the invention.
  • the main rotor 2 is rotatably mounted in the housing 4 in a known manner in such a way that the fixed bearing 60 is on the pressure side 10 and the floating bearing 61 is on the suction side 5 of the screw compressor system 1.
  • the system 1 is driven via a coupling (not shown in more detail) on the suction-side shaft end 62 of the main rotor 2.
  • the secondary rotor 3 is supported in a bearing shell 51 inserted in the housing bore 50 by cladding, this bearing shell. 51 is mounted both on the pressure-side housing end face 63 and in the suction-side housing cover 64 up to the cutting edge 33 of the housing bore.
  • FIG. 8 shows a schematic longitudinal section through a further screw compressor system 1 according to the invention, the main rotor 2 being rotatably mounted in the housing 4 in a pressure-side fixed bearing 60 and a suction-side floating bearing 61.
  • the secondary rotor 3 is also rotatably mounted in a pressure-side fixed bearing 65 and a suction-side floating bearing 66.
  • the system 1 is driven via a coupling, not shown, which is flanged on the suction side shaft end 67 of the secondary rotor 3.
  • the drive torque is transmitted to the main rotor 2 via the secondary rotor teeth 13 and the gas sucked in is compressed in the tooth gaps and pushed out on the pressure side 10 into a pressure container (not shown).

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Applications Or Details Of Rotary Compressors (AREA)
  • Compressors, Vaccum Pumps And Other Relevant Systems (AREA)
  • Compressor (AREA)
  • Toys (AREA)
EP84890238A 1983-12-14 1984-12-06 Compresseur à piston rotatif Withdrawn EP0149446A3 (fr)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
AT434883 1983-12-14
AT4348/83 1983-12-14

Publications (2)

Publication Number Publication Date
EP0149446A2 true EP0149446A2 (fr) 1985-07-24
EP0149446A3 EP0149446A3 (fr) 1985-08-07

Family

ID=3562842

Family Applications (1)

Application Number Title Priority Date Filing Date
EP84890238A Withdrawn EP0149446A3 (fr) 1983-12-14 1984-12-06 Compresseur à piston rotatif

Country Status (4)

Country Link
US (1) US4614484A (fr)
EP (1) EP0149446A3 (fr)
JP (1) JPS60147590A (fr)
DE (1) DE8434596U1 (fr)

Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0334040A1 (fr) * 1988-03-23 1989-09-27 Robert Bosch Gmbh Machine à engrenages (pompe ou moteur)
DE19502323A1 (de) * 1995-01-26 1996-08-01 Guenter Kirsten Verfahren zur Herstellung von Rotoren von Schraubenverdichtern

Families Citing this family (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US9267504B2 (en) 2010-08-30 2016-02-23 Hicor Technologies, Inc. Compressor with liquid injection cooling
US8794941B2 (en) 2010-08-30 2014-08-05 Oscomp Systems Inc. Compressor with liquid injection cooling
US9057373B2 (en) 2011-11-22 2015-06-16 Vilter Manufacturing Llc Single screw compressor with high output
DE102014000911B4 (de) * 2014-01-28 2016-01-28 Klaus Union Gmbh & Co. Kg Förderschraube als Läufer oder Gegenläufer eines Förderschraubenpaares einer Schraubenpumpe

Citations (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE652216C (de) * 1935-02-15 1937-10-27 Gustaf Zakarias Goeransson Umlaufverdichter oder Umlaufantriebsmaschine mit Schraubenraedern
US2873909A (en) * 1954-10-26 1959-02-17 Svenska Rotor Maskiner Ab Rotary devices and casing structures therefor
FR1330352A (fr) * 1962-05-18 1963-06-21 Svenska Rotor Maskiner Ab Machine rotative à rotors hélicoïdaux engrenants
US3307777A (en) * 1963-12-23 1967-03-07 Svenska Rotor Maskiner Ab Screw rotor machine with an elastic working fluid
FR1544017A (fr) * 1966-07-29 1968-10-31 Svenska Rotor Maskiner Ab Paire de rotors coopérants pour machine rotative
DE2429856A1 (de) * 1973-06-27 1975-01-16 Joy Mfg Co Rotorpaar fuer eine schraubenrotormaschine und verfahren zu dessen herstellung
EP0077031A1 (fr) * 1981-10-09 1983-04-20 Technika Beteiligungsgesellschaft mbH Compresseur à pistons rotatifs

Family Cites Families (14)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB254986A (en) * 1925-10-06 1926-07-15 Alexander Johan Mollinger Improvements in or relating to screw pumps
AT238362B (de) * 1961-05-20 1965-02-10 Svenska Rotor Maskiner Ab Schraubenradmaschine
US3666384A (en) * 1970-10-20 1972-05-30 Pavel Evgenievich Amosov Screw-rotor machine for compressible fluids
FR2163935A5 (fr) * 1971-12-07 1973-07-27 Rhone Poulenc Sa
US4028026A (en) * 1972-07-14 1977-06-07 Linde Aktiengesellschaft Screw compressor with involute profiled teeth
DE2234777C3 (de) * 1972-07-14 1980-10-30 Linde Ag, 6200 Wiesbaden Verdichter
DE2302902A1 (de) * 1973-01-22 1974-07-25 H & H Licensing Corp Schraubenkompressor mit rotorabschnitten
US3865523A (en) * 1973-03-16 1975-02-11 Samuel J Baehr Continuous flow rotary pump
US4140445A (en) * 1974-03-06 1979-02-20 Svenka Rotor Haskiner Aktiebolag Screw-rotor machine with straight flank sections
DD122841A1 (fr) * 1975-11-07 1976-11-05
DD128035B1 (de) * 1976-09-27 1979-12-27 Dieter Mosemann Schraubenrotoren
US4210410A (en) * 1977-11-17 1980-07-01 Tokico Ltd. Volumetric type flowmeter having circular and involute tooth shape rotors
DE3140107A1 (de) * 1981-10-09 1983-04-28 Technika Beteiligungsgesellschaft mbH, 4800 Bielefeld Drehkolbenverdichter
US4445831A (en) * 1982-12-15 1984-05-01 Joy Manufacturing Company Screw rotor machine rotors and method of making

Patent Citations (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE652216C (de) * 1935-02-15 1937-10-27 Gustaf Zakarias Goeransson Umlaufverdichter oder Umlaufantriebsmaschine mit Schraubenraedern
US2873909A (en) * 1954-10-26 1959-02-17 Svenska Rotor Maskiner Ab Rotary devices and casing structures therefor
FR1330352A (fr) * 1962-05-18 1963-06-21 Svenska Rotor Maskiner Ab Machine rotative à rotors hélicoïdaux engrenants
US3307777A (en) * 1963-12-23 1967-03-07 Svenska Rotor Maskiner Ab Screw rotor machine with an elastic working fluid
FR1544017A (fr) * 1966-07-29 1968-10-31 Svenska Rotor Maskiner Ab Paire de rotors coopérants pour machine rotative
DE2429856A1 (de) * 1973-06-27 1975-01-16 Joy Mfg Co Rotorpaar fuer eine schraubenrotormaschine und verfahren zu dessen herstellung
EP0077031A1 (fr) * 1981-10-09 1983-04-20 Technika Beteiligungsgesellschaft mbH Compresseur à pistons rotatifs

Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0334040A1 (fr) * 1988-03-23 1989-09-27 Robert Bosch Gmbh Machine à engrenages (pompe ou moteur)
DE19502323A1 (de) * 1995-01-26 1996-08-01 Guenter Kirsten Verfahren zur Herstellung von Rotoren von Schraubenverdichtern

Also Published As

Publication number Publication date
US4614484A (en) 1986-09-30
JPS60147590A (ja) 1985-08-03
EP0149446A3 (fr) 1985-08-07
DE8434596U1 (de) 1985-02-21

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Inventor name: RIEGLER, GEROLD, DIPL.-ING. DR.