EP2473739A2 - Pompe à vis sèche à compression interne - Google Patents

Pompe à vis sèche à compression interne

Info

Publication number
EP2473739A2
EP2473739A2 EP10743078A EP10743078A EP2473739A2 EP 2473739 A2 EP2473739 A2 EP 2473739A2 EP 10743078 A EP10743078 A EP 10743078A EP 10743078 A EP10743078 A EP 10743078A EP 2473739 A2 EP2473739 A2 EP 2473739A2
Authority
EP
European Patent Office
Prior art keywords
pump according
screw pump
gas
thread
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.)
Granted
Application number
EP10743078A
Other languages
German (de)
English (en)
Other versions
EP2473739B1 (fr
Inventor
Ralf Steffens
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.)
Individual
Original Assignee
Individual
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 Individual filed Critical Individual
Publication of EP2473739A2 publication Critical patent/EP2473739A2/fr
Application granted granted Critical
Publication of EP2473739B1 publication Critical patent/EP2473739B1/fr
Not-in-force legal-status Critical Current
Anticipated expiration legal-status Critical

Links

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
    • F04C18/12—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
    • 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
    • F04C29/00—Component parts, details or accessories of pumps or pumping installations, not provided for in groups F04C18/00 - F04C28/00
    • F04C29/04—Heating; Cooling; Heat insulation
    • 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
    • F04C2220/00—Application
    • F04C2220/10—Vacuum
    • F04C2220/12—Dry running

Definitions

  • Dry-compressing pumps are gaining in importance in industrial compressor technology, because of increasing obligations in environmental regulations and rising operating and disposal costs and increased demands on the purity of the pumped medium, the known wet-running pump systems, such as liquid ring machines, rotary vane pumps and oil or water-injected screw compressors, increasingly replaced by dry-compressing pumps. These machines include dry screw compressors, claw pumps, diaphragm pumps, piston pumps, scroll machines and Roots pumps. However, these machines have in common that they still do not meet today's demands in terms of reliability and robustness and size and weight while maintaining low price level and satisfactory compressor efficiency.
  • the object of the present invention is to improve for dry-2-W ⁇ fill-Rotat ⁇ ons-V ⁇ rdrängermaschinen to the demand and compression of gaseous fluids the efficiency and to reduce the expenses for higher discharge pressures and the load of the gas forces in axial and radial direction significantly to diminish.
  • this object is achieved in that in a dry-compressing screw pump with a Rotorinnenk ⁇ hlung for the spindle rotor pair and a cooling for the spindle rotors surrounding pump housing, the multi-stage screw spindle rotor pair has an interlocking main F ⁇ rdergewinde and additionally an interlocking secondary conveyor thread, wherein at each screw spindle rotor In the case of the main conveyor thread, the outside tip circle diameter of the conveyor thread decreases in the gas outlet direction, whereas the root diameter increases correspondingly more suitably, and the additional secondary conveyor thread on each screw spindle rotor is provided with a threading thread whose nominal thickness is opposite to the nominal conveying direction at the main conveying thread of the same spindle rotor, so that the secondary conveying thread has exactly the reverse conveying direction than the main conveying towers.
  • the main conveying thread is initially made cylindrical for the outer diameter for about the same time until the working chambers which open on the inlet side to scoop the conveying medium and enclose the conveying medium are approximately closed again.
  • "swallowing" Schopf s further increases the input side slope of the main conveyor thread to close this ansaugenden working chambers to a maximum value to be then reduced in the direction of the gas outlet plenum to implement the apphkationsspezifisch optimal internal compression ratio then again.
  • the inner so-called "built-in" compression ratio is carried out as desired in the parameter design of the individual working chamber volumes by changing the pitch and diameter values.
  • the housing-fixed gas access for the delivery medium is designed such that this gas inlet to the suctioning working chambers takes place both in the radial and in the axial direction.
  • this machine has only one permissible direction of rotation, this is in addition to the largest possible cross-sectional design in Gasemlass- room with frontally comprehensive rotor access via additional housing-fixed inlet control edges, which extend specifically into an initial part of the cylindrical receiving bore for the spindle rotors.
  • the gas outlet collecting space for the conveying medium to be compressed is located between the main conveying thread and the secondary conveying thread, wherein the main Fordergewmde in the axial direction has a greater length than the secondary conveying thread.
  • the secondary conveyor thread is designed for a significantly lower flow volume flow, preferably at least factor 5 to 10 times smaller than the nominal volume flow of the main conveyor belt.
  • the secondary conveying thread is positioned between the gas outlet collecting space of the main conveyor thread and the working space shaft passages to the gear compartment, where the oil-lubricated bearing and the oil-lubricated drive gears and the entry of the coolant (oil) into each of the two Screw spindle rotors are located
  • the Forder for the secondary F ⁇ rdergewinde according to the invention is now carried out such that the actually occurring in operation gas flow over this secondary conveyor thread adjusts in the opposite direction to the nominal conveying direction of this secondary conveyor thread, so that the resulting, in operation actually resulting Gasstr ⁇ mung is directed away from the gas outlet plenum via this secondary conveying thread and points in the direction of the working space Wellen pressf ⁇ hrung to ol-lubricated gear chamber.
  • this Gasstr ⁇ mung thus represents a leakage or loss Gasstr ⁇ mung, but in this way reliably ensures the desired freedom from oil for the delivery medium in the pump working space.
  • the compressive capacity of the sub-conveyor thread is made slightly weaker than the compression capacity of the main conveyor thread by selecting the design parameters for this sub-conveyor thread such that the actual gas flow is directed away from the gas outlet plenum of the main conveyor thread and In the direction of the working-space shaft feed-through, for example, the respective working chamber sizes are reduced in relation to the play values and the number of stages is changed such that the actually resulting compression behavior of the secondary conveying thread is slightly weaker than the compression behavior of the main conveying thread.
  • the amount of this leakage or loss gas flow through the secondary conveyor thread is determined application specific, which in applications with low demands on the freedom from oil in the working space this leakage gas flow, which flows away via the secondary conveyor thread from the gas outlet plenum, is made smaller as in applications with high demands on the freedom from oil in the
  • the pressure in the gear chamber thus corresponds approximately to the ambient pressure and the two shaft passages to the pump working space can be designed as a non-contact labyrinth seal with series connection of several separation chambers with the largest possible buffer volumes.
  • the secondary threading receives a Gas shutt ⁇ tts opening between the ant ⁇ ebs- side edge of its own conveyor thread and the working space shaft passages to the gear compartment.
  • a Verölung the working space is reliably avoided, because the suction side bearing of the screw spindle rotors is executed with life fat lubricated bearings, running for special applications even as a hybrid bearings, so with ceramic balls on steel bearing rings.
  • the pitch of the thread is carried out according to the invention both on the main thread as well as on the secondary thread so that the highest possible number of steps is achieved as a number of wraps of the conveyor thread on the screw spindle as so-called "multi-stage".
  • This Umschli ⁇ gu ⁇ gsa ⁇ kohl or number of steps to describe this multi-stage should be at least 800 ° as the degree of angle, but preferably values above 1100 ° are more advantageous as wrap angle.
  • flank pitch for the conveyor thread are designed significantly lower than in today's screw compressor Verdrangerrotoren.
  • the flank pitch for the conveying thread on the pitch circle should be well below 20 ° in the region of the diameter values changing according to the invention.
  • the quotient of average tooth gap depth to middle tooth space width lies in a range between 1, 5 and up to 4 as dimensionless ratio. The entire length of the spindle rotors is carried out as long as possible, in that practically only the critical bending speed or the bending load is taken as the design limit.
  • the internal compression is then achieved via the change in diameter according to the invention on the main Forder- thread.
  • This diameter change can be linear or via any other function.
  • the heat exchange surfaces are increased, thereby improving the compressor efficiency, as well as increasing the compressibility of the machine and reducing the gas forces in the radial direction, as well as reducing the requirements for the Verfwinkel synchronization accuracy in the driving gearing.
  • the screw spindle rotor pair is preferably mirror-inverted identical and 2-gang ⁇ g performed, so has two teeth in the frontal section, so that in each end section plane are always working chambers equal pressure in equilibrium opposite each other
  • the outer diameter for the secondary conveying thread is chosen such that the gas forces in the axial direction, which arise due to the pressure difference between the suction pressure at the gas inlet and the final compression pressure at the gas outlet collecting space, are effectively compensated or advantageously absorbed by the bearing ,
  • the internal rotor cooling preferably realized in accordance with PCT publication WO 00/12899 as a conical rotor bore into which the coolant (usually oil) is permanently introduced, extends in particular in the region of the increasing root diameter of the main conveyor thread.
  • the angles of the cone cooling bore for the rotor internal cooling and the angle of the changing root diameter on the main delivery thread have the same orientation, so that increases with increasing bore diameter in the conical rotor cooling bore and the root diameter at the main delivery thread.
  • To improve the heat dissipation of the cooling cone angle can be increased thanks to the conical design on the main conveyor thread with running in the same direction diameter changes.
  • This internal rotor cooling is also sufficient for sufficient heat dissipation at the secondary conveyor thread, the cone cooling angle in this area does not need to be so steep because the cooling oil arrives at high speed for good heat transfer rates at the same time large cooling hole diameter values and thus a safe heat dissipation is guaranteed.
  • the secondary conveying thread is preferably made cylindrical, that is, with constant values for the outer end circle diameter as well as root diameter, and can also be made with a constant pitch for the conveying thread for ease of manufacture.
  • a delivery medium partial flow is branched off from a transport region at a higher pressure and cooled back to one or more regions of the main delivery thread at a lower pressure via a heat exchanger. This process should be referred to as "return gas cooling”.
  • the gas cooler for heat dissipation from this branched conveying medium partial flow can of course also be designed as a separate heat exchanger, which also with the
  • Coolant is operated, which is already used for internal rotor cooling and heat dissipation for the pump housing.
  • the inner compression ratio of the compressor according to the invention by relative displacement in the axial direction between the screw spindle rotor pair and the surrounding pump housing at the main conveying thread such that this inner compression ratio of the compressor application specific different operating pressures optimally adapted by this axial displacement between screw shaft pair and pump housing different high compression capacity due to the variable gas backflow only in the conical region of the main conveyor thread through which vary due to this axial displacement gap heights between the rotor heads and the surrounding pump housing in the conical region of the main conveyor thread, so that the internal compression ratio of the Compressor by this shift the prevailing conditions best possible follows.
  • the relative displacement is controlled by the axial gas pressure differential forces and / or the power consumption. At the same time a reduction of the nominal pumping speed is avoided by the gas inlet according to the invention cylindrical design of the main conveyor thread in the described axial displacement, because this axial displacement causes virtually no changes in the gas transport volume in the cylindrical region.
  • this axial displacement takes place, for example, as well as the design for suction-side storage on the fixed bearing, but to maintain the axial position of the two spindle rotors to each other, the fixed bearings are to be moved together as a unit.
  • this axial displacement can also be realized in the pump housing, which is facilitated by the cooling oil design, for example via an axially displaceable housing insert.
  • FIG. 1 shows an exemplary embodiment of the present invention with a section through the entire screw pump
  • the mirror image identical rotor pair (1) rotates in a pump housing (3) with a gas inlet (8) and a gas outlet plenum (9)
  • Rotor pair is driven by a crown or bevel gear (12) with drive shaft such that the Verdrangerrotore rotate in opposite directions in the pump working space.
  • the Ant ⁇ ebsmotor is not shown separately
  • Each screw spindle rotor is on the gas inlet side in life fat lubricated roller bearings (14) as a floating bearing and Antnebs side held in olschmm investigating roller bearings (13) as a fixed bearing for securing the axial rotor position between the pump working space and the oil-lubricated space of the gear box (11) are the working space -Wellenabdichteptept (20) for both Spmdelrotor shaft ends
  • the inner cooling (2) for each screw rotor is shown in dashed lines and is shown in Fig.
  • the secondary Fordergewinde (7) the nominal Forder ⁇ chtung (23) from the gas passage (9) to the gas outlet plenum (9), whereas the actual gas flow (24) on the secondary threading thread (7) goes in the other direction as the so-called “leakage or loss gas flow”.
  • the gas outlet collecting space (9) is under the compressor pressure p u
  • the gas inlet (8) is under the pressure p em and both the gas passage (10) for the Mau-Fordergewinde (7) and the space in the gearbox (1 1) under ambient pressure, simply registered as "atm" stand
  • the main threading thread (5) has clearly visible a cylindrical portion (6) and a conical part with change of tip diameter (28) and root diameter (27)
  • the coolant (26) at the same time also lubricant in the gear compartment, from the Forderpumpe (19) to the spindle rotor mecanickuhlung (2) and on the Warmeabbigungsflachen (4) of the Pumpengehauses (2) required.
  • the embarrassedsge enthusiasm high number of stages on the spindle rotor provides on the one hand for the desired high compression capacity of p em on p u and sufficient heat exchange surfaces together with the Pumpengehausekuhlung to dissipate during the gas transport appreciably compression heat, which is known to increase the compressor efficiency as desired
  • Fig. 2 shows for the present invention by way of example an embodiment for the targeted reduction of the gas temperatures during compression via the "return gas Kuhlung" With branch (29) a Fordermedium subset and feedback (31) of this subset in the pump working space is in the upper area this return gas cooling represented by a separate heat exchanger (30), while at the bottom of the recirculated into the working space delivery medium subset (31) is diverted directly from the conveyor medium (32), which must be cooled anyway in the nachtern ⁇ cooler (33) in general , In this case, this cooler (33) is usually operated with an external cooling fluid (34), which is usually air or (more rarely) water.
  • the return (31) of the cooled delivery flow subset into the pump working space takes place symmetrically for both screw spindle rotors (1) at the same longitudinal axial position.
  • Fig. 3 shows for the present invention by way of example a more detailed sectional drawing as a practically complete construction design.
  • the representation of the rotor internal cooling (2) is shown in more detail, as well as the design of the rotor fixed gear of the crown gear drive (12) in terms of improved Biegesteif candy on anthebs districten shaft end of the screw spindle rotors (1).
  • a wave-shaped thread for increasing the cooling surface is shown by way of example only on one side of the conical cooling bore for illustrative purposes.
  • Fig. 4 shows in more detail the design for suction-side mounting (14) with improvement of the desired floating bearing function on the axially displaceable support bushing (15) with oil-filled recess space (16) and the seals (17).
  • the bearing outer ring of the suction-side mounting (14) sits firmly in the carrier bushing (15).
  • a potential co-rotation of the carrier bushing (15) can be prevented, for example, via end-face pins, which are not shown here in detail.
  • the friction of the two sealing rings (17) should reliably prevent unwanted co-rotation of the carrier bushing (15).
  • paired roller bearings for the suction-side mounting (14) it is possible to choose between O and X arrangements, with both embodiments being shown in this figure for the sake of simplicity.
  • a control gas delivery stage (39) additionally provided, which is attached to the two inlet-side spindle rotor shaft ends.
  • the control gas delivery stage (39) has a nominal delivery direction (41) opposite to the nominal delivery direction (21) of the main delivery stage (5).
  • the control gas delivery stage (39) is located in a control gas housing (38) and is operated on the one side in the collecting space (37) with the Druckp aj and on the other side with the pressure p ⁇ , wherein the inlet Pressure /? ", -" via a bypass opening (42) is ensured.
  • the invention also proposes that the resulting axial forces are reduced by a non-linear course of the diameter change in the conical region of the main conveyor stage (5) by corresponding to the pressure increase along the conical main conveyor stage, the diameter change is such that the axial working chamber gas forces are minimized:
  • the parameters for the additional stages are preferably designed "motor".
  • the pressure difference along the F ⁇ rdergewinde this additional stages is used to drive the rotors of these additional stages.
  • These additional stages therefore do not act as a pump that require a Fordermedium of a lower to a higher pressure, but these additional stages act as a motor that relaxes the available gas with higher pressure to a lower pressure level while driving the Verdrängerrotore the additional stages.
  • the execution is carried out such that the parameters for these additional stages are carried out so that the idling speed, ie the speed of these additional stage Verdrängerrotore without mechanical losses, above the later actual operating speed of the main Forderhave.
  • the gas quantity for each additional stage is minimized, for example, by increasing the number of stages and / or reducing the play values
  • the additional stages are preferably monodentate, also called "catchy".

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Applications Or Details Of Rotary Compressors (AREA)

Abstract

L'invention concerne des machines volumétriques rotatives à deux arbres à compression sèche en tant que pompes à vis pour le transport et la compression de gaz, les rotors hélicoïdaux (1) disposant d'un refroidissement interne (2) déjà connu. L'invention vise à améliorer le rendement du compresseur et à permettre des pressions de service plus élevées avec une machine seulement. A cet effet, la paire de rotors à vis hélicoïdale (1) est réalisée à plusieurs étages et comprend un filet transporteur principal à engrènement (5) et un filet transporteur auxiliaire à engrènement (7), ainsi qu'un espace collecteur de sortie de gaz (9) entre ces deux filets transporteurs, le diamètre de sommet (28) du filet transporteur principal (5) diminuant en direction de la sortie de gaz alors que son diamètre de fond (27) augmente de manière correspondante pour un rapport de compression "intégré" interne convenant à l'application, et le sens de transport nominal (23) du filet transporteur auxiliaire (7) est opposé au sens de transport nominal (21) du filet transporteur principal (5). Le filet transporteur auxiliaire (7) est réalisé de sorte que le flux de gaz (24) effectivement obtenu sur le filet transporteur auxiliaire (7) est orienté en sens opposé à son sens de transport nominal (23) et s'éloigne ainsi toujours de l'espace collecteur de sortie de gaz (9), avec un passage de gaz (10) sous pression ambiante tout comme la chambre dans la boîte d'engrenages (11). Pour des pressions de service très élevées, la pompe selon l'invention comprend en outre un étage de transport de gaz de commande (39) dont la pression de commande p ax dans l'espace collecteur (37) est réglée par le biais d'un dispositif de régulation (36) à partir du fluide refoulé refroidi, l'étage de transport du gaz de commande (39) et/ou le filet transporteur auxiliaire (7) fonctionnant au moyen d'un moteur.
EP10743078.7A 2009-08-31 2010-08-04 Pompe à vis sèche à compression interne Not-in-force EP2473739B1 (fr)

Applications Claiming Priority (3)

Application Number Priority Date Filing Date Title
DE102009029047 2009-08-31
DE102009051096 2009-10-28
PCT/EP2010/061363 WO2011023513A2 (fr) 2009-08-31 2010-08-04 Pompe volumétrique à compression interne

Publications (2)

Publication Number Publication Date
EP2473739A2 true EP2473739A2 (fr) 2012-07-11
EP2473739B1 EP2473739B1 (fr) 2014-03-26

Family

ID=43628467

Family Applications (1)

Application Number Title Priority Date Filing Date
EP10743078.7A Not-in-force EP2473739B1 (fr) 2009-08-31 2010-08-04 Pompe à vis sèche à compression interne

Country Status (4)

Country Link
US (1) US8876506B2 (fr)
EP (1) EP2473739B1 (fr)
DE (1) DE112010003504A5 (fr)
WO (1) WO2011023513A2 (fr)

Families Citing this family (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102012009103A1 (de) * 2012-05-08 2013-11-14 Ralf Steffens Spindelverdichter
DE102014008288B4 (de) * 2014-06-03 2025-11-27 Steffen Klein Spindelverdichter für Kompressionskältemaschinen
CN110591740B (zh) * 2019-09-30 2024-09-10 招远市汇潮新能源科技有限公司 排氧送料装置及裂解设备
DE102020109189A1 (de) * 2020-04-02 2021-10-07 Andreas Hettich Gmbh & Co. Kg Einsatz für einen Rotor einer Zentrifuge

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GB1220054A (en) * 1967-02-06 1971-01-20 Svenska Rotor Maskiner Ab Two-stage compressor of the meshing screw rotor type
CH635403A5 (de) * 1978-09-20 1983-03-31 Edouard Klaey Schraubenspindelmaschine.
DE19839501A1 (de) 1998-08-29 2000-03-02 Leybold Vakuum Gmbh Trockenverdichtende Schraubenspindelpumpe
US6129534A (en) * 1999-06-16 2000-10-10 The Boc Group Plc Vacuum pumps
CZ2000581A3 (cs) * 2000-02-18 2001-04-11 Perna Vratislav Zařízení se šroubovými zuby ve vzájemné interakci
AU2002213243A1 (en) * 2000-10-18 2002-04-29 Leybold Vakuum Gmbh Multi-stage helical screw rotor
JP2005002872A (ja) * 2003-06-11 2005-01-06 Taiko Kikai Industries Co Ltd スクリュウ型流体機械およびそのスクリュウロータの形成方法
JPWO2007000815A1 (ja) * 2005-06-29 2009-01-22 株式会社前川製作所 二段スクリュー圧縮機の給油方法、装置及び冷凍装置の運転方法
GB0525378D0 (en) * 2005-12-13 2006-01-18 Boc Group Plc Screw Pump
JP2008196390A (ja) * 2007-02-13 2008-08-28 Toyota Industries Corp 容積変動型流体機械

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Title
See references of WO2011023513A2 *

Also Published As

Publication number Publication date
WO2011023513A2 (fr) 2011-03-03
US8876506B2 (en) 2014-11-04
WO2011023513A3 (fr) 2011-09-29
EP2473739B1 (fr) 2014-03-26
US20120171068A1 (en) 2012-07-05
DE112010003504A5 (de) 2012-11-22

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