DE102017106781A1 - Rotor edge pairings - Google Patents

Abstract

The invention relates to the rotor flanks in dry-compressing 2-shaft rotary displacement machines for the promotion and compression of gases for applications in vacuum and in overpressure including refrigerant compressor. In order to improve the efficiency in spindle compressors, the invention proposes that in the respective end cutting preferably identical spindle rotor pair (5) with a Rotorzähnezahlverhältnis of 1: 1 or 2: 2 or 3: 3 or 4: 4 or even higher at a wrap angle of at least 1000 ° each spindle rotor (5) has a sickle edge (1) and the other profile edges as balance edge (2) and minor edges (3) are designed such that by "edge profile adjustment" and / or each positioned unbalance compensation groove (7) in the carrier wave ( 6) and / or multiple rotor head bores (8) and / or also a partial residual bending moment support on the support shaft (6) the centroid FS in each endcut according to the desired balancing quality comes close enough to lie at the rotor pivot point OR, where at the same time Rib efficiency per profile tooth with regard to greater heat dissipation and the design on the spindle rotor opf (11) with respect to less gap leakage can be performed.

Description

Drying compressors are gaining in importance in industrial compressor technology, because increasing obligations in environmental regulations and rising operating and disposal costs and increased demands on the purity of the medium, the known wet-running compressors, such as liquid ring machines, rotary vane pumps and oil or water-injected screw compressors, increasingly replaced by dry compacting machines. 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 with low price level and satisfactory efficiency.

To improve this situation, offer the known dry-compressing spindle compressors because they realize a typical high performance as a typical 2-wave displacement machines simply by the fact that they have the necessary multistage as so-called. "Conveyor thread" by connecting several closed working chambers on the number of wraps per displacement rotor reach extremely uncomplicated, but without requiring a working fluid in the workspace. In addition, an increased rotor speed is made possible by the non-contact rolling of the two oppositely rotating spindle rotors, so that based on the size of the same nominal suction and delivery rate increase. Here, dry-compacting spindle machines can be used both for applications in vacuum and for overpressure, the power requirement in the overpressure is naturally significantly higher, because in the overpressure range with final pressures well above 2 bar (absolute) up to 15 bar and even higher significantly greater pressure differences to be overcome.

In the protection right PCT / EP2014 / 060851 For a dry compressing spindle compressor, a high internal compression is realized while cooling the working space components, in order to constantly dissipate part of the heat of compression generated during the compression process. These work space components are the pair of spindle rotors with the surrounding compressor housing, wherein the distance between the spindle rotors on the conveying gas inlet side is greater than on the conveying gas outlet side. However, the desired improvement in compressor efficiency is, in addition to the most efficient heat dissipation during compression, also decisively influenced by the internal leakage currents. Because a 2-shaft positive displacement machine as a dry-running machine has in principle the difficulty that the inner gap leakage between the working space components, ie rotor pair with surrounding compressor housing, a detrimental effect on the efficiency, because a certain amount of gas constantly between the working chambers with different Pressure per working chamber flows back and thereby arise according to the second law of thermodynamics entropy now losses. Therefore, these inner gap leakage should be kept as low as possible. At the same time, it is desirable to achieve good compressor efficiency so that the compression heat generated during compression is best dissipated during the working process. For thermodynamic reasons logically sufficiently large heat transfer surfaces of the working space components are required. Application-specific there are the state of the art as well as the property right PCT / EP2014 / 060851 Goal conflicts with regard to minimizing the internal gap leakage between the work space components and simultaneously desired maximization of the heat transfer surfaces in these work space components. For this application-specifically a fundamentally better solution is desirable.

• a) die Verluste durch die inneren Spalt-Leckagen zwischen den Arbeitsraum-Bauteilen minimiert werden, wobei insbesondere die Anzahl der über die Leckage-Spalte miteinander kommunizierenden Arbeitskammern zu minimieren ist,
• b) die Wärmetransfer-Flächen bei den Arbeitsraum-Bauteilen maximiert werden, um zur Verbesserung des Verdichter-Wirkungsgrades möglichst viel Wärme während der Verdichtung abführen zu können,
• c) sichere Drehzahlfestigkeit, also Minimierung der Unwuchten, wobei die Wuchtgüten der Rotore wie bei heutigen Rotoren zu realisieren sind (beispielsweise entsprechend der Norm DIN-ISO 1940 )
• d) die Lastaufteilung zwischen den beiden Spindelrotoren möglichst ähnlich ist, um beispielsweise Temperatur-Unterschiede zu minimieren
• e) die Unterschiede bei den Leistungswerten sowie sonstigen Merkmalen (wie Geschwindigkeitswerte, Geometrie und Ausführung) möglichst gering sind,
• f) bei gleichzeitig vereinfachter Fertigung, insbesondere für das Spindelrotorpaar
• g) die Wärmeabführung über die Arbeitsraum-Bauteile verbessert wird
• h) und die Stufenzahl (also die Anzahl abgeschlossener Arbeitskammern zwischen Ein- und Auslass) abhängig vom gewünschten Betriebsdruckbereich bestgeeignet ausgeführt wird: So ist bei Spindelverdichtern für höhere Betriebsdrücke (z.B. über 8 bar abs.) prinzipiell eine höhere Stufenzahl (z.B. über 6 Stufen) erforderlich als bei geringeren Betriebsdrücken (z.B. kleiner 6 bar) mit Stufenzahlen vorzugsweise unter 5 Stufen

The object of the present invention is to improve the efficiency of the compressor for dry-compressing 2-shaft rotary displacement machines according to the screw compressor principle with a number of steps greater than 2 by the inner gap leakage between the working space components (ie rotor pair and surrounding Compressor housing) is limited and at the same time the heat transfer surfaces of the work space components are maximized. In this case, application-specific, different number of stages are as simple as possible to implement in order to meet the respective requirements as best as possible. In principle, this is initially determined application-specifically via the specific execution of the respectively most suitable screw compressor profile pairing and is to be carried out in such a way that

• a) the losses due to the internal gap leakage between the working space components are minimized, wherein in particular the number of working chambers communicating with one another via the leakage gap is to be minimized,
• b) the heat transfer surfaces are maximized in the work space components in order to dissipate as much heat during compression to improve the compressor efficiency,
• c) safe speed stability, so minimizing imbalances, the balancing qualities of the rotors as in today's rotors are to be realized (for example, according to the Standard DIN-ISO 1940 )
• d) the load distribution between the two spindle rotors is as similar as possible, for example, to minimize temperature differences
• e) the differences in performance and other characteristics (such as speed, geometry and design) are as small as possible,
• f) at the same time simplified production, especially for the spindle rotor pair
• g) the heat dissipation over the working space components is improved
• h) and the number of stages (ie the number of completed working chambers between inlet and outlet) depends on the desired operating pressure range is carried out most suitable: For spindle compressors for higher operating pressures (eg over 8 bar abs.) In principle, a higher number of stages (eg over 6 stages) required as at lower operating pressures (eg less than 6 bar) with levels preferably less than 5 levels

In order to be able to optimally fulfill these different requirements in each case in an application-specific manner, a novel design is required for the spindle compressor profile pairings. Until now, the number of stages and the size of the heat transfer surfaces were linked. However, since the heat transfer surfaces should be as large as possible and the number of stages varies according to entropy balances and application-specific requirements, a decoupling of the number of stages and heat transfer surfaces is highly desirable.

According to the invention this object is achieved in that for vacuum and overpressure applications incl. Refrigerant compressor for dry (ie working in the working space between inlet and outlet without operating fluid) compression of gaseous media in a spindle compressor as a 2-wave displacement machine that of an outer (ie outside from the working space) mechanical or electronic synchronization with the transmission ratio of 1: 1 in opposite directions rotation angle-driven spindle rotor pair ( 5 ) is performed with a (preferably) identical Rotorstirnschnittprofilpaarung with Rotorzähnezahlverhältnissen of 1: 1 or 2: 2 or 3: 3 or 4: 4 or even higher, the most suitable version depending on the application-specific requirement profile (eg pressure level, working range, size etc.) and the respective leakage gap losses (keyword entropy and respective manufacturing accuracies) is determined, wherein the distance between the rotor axes of rotation at the gas inlet side (preferably) at least 15% greater than on the gas outlet side and the rotor axes of rotation lie (preferably) in a common plane and each spindle rotor has a wrap angle (ie the sum of the face-cut twist angle) of at least 1000 ° (in words "one thousand degrees of angle"): By turning 360 ° one turn, that is more than two Turns around which the respective face cuts between inlet and outlet side zue are twisted in each other), and that each rotor has a flank as a sickle flank ( 1 ) is executed, which is characterized in that the engagement line (ie, the fixed frame location of all flank contact points with the counter rotor) to the intersection edge G of the compressor housing ( 4 ) theoretically (practically because of the non-contact rotor pair Flankenabwälzung known course with clearance and various tolerances) is sufficient, so that according to the invention a "blowhole-free rotor profile pair" arises, and that in addition the other profile flanks ( 2 and 3 ) are carried out in such a way that at the desired high rotor speeds (ie rotational speed values such that the average circumferential speed values at the rotor heads reach at least 40 m / sec) the respectively required balancing quality (for example according to DIN-ISO 1940 ), wherein for Rotorzähnezahlverhältnisse of 2: 2 and higher the balance edge ( 2 ) and the minor flanks ( 3 ) According to the known gearing law rolling off each other according to the invention are carried out such that the centroid FS in each endcut according to the desired balancing quality (for example, according to DIN-ISO 1940 ) comes to lie so close enough to the rotor pivot point OR, so that the rotors can be safely operated in the desired speed range, the specific calculation for the respective edge profile profile for the balance edge ( 2 ) and for the minor flanks ( 3 ) as so-called. "Flank profile adjustment", for example, can be carried out by "profile slope function" ( Source: Steffens, Ralf: "The profile slope function, a new way of analytical determination and optimization of general profile flank pairings", Dissertation TU-Baunschweig, 1993, Report No. 40 ), where mathematically only the crescent side ( 1 ) is present clearly in each end cut and over the respective profile flank pairings to the balance flank ( 2 ) and to the minor flanks ( 3 ) the position of the centroid FS is influenced in such a way that this centroid FS comes to lie sufficiently close enough to achieve the desired balancing quality at the rotor pivot point OR, so that in each end section the centroid FS by the respective execution of the edge profiles to the balance edge ( 2 ) and to the minor flanks ( 3 ) is moved close to the rotor pivot point OR. By eliminating minor side edges (1: 1 rotor teeth ratio) 3 ) are present, there is to achieve the desired balancing quality at least for these monorail rotors according to the invention in addition to this flank profile adaptation (for example by means of "profile slope function") for the balance edge ( 2 ) as further measures in the spindle rotor design according to DE 10 2015 012867.1 [alternatively abbreviated as "(aluminum) rotor against rotation on (steel) carrier shaft with final gas conveyor thread production"] a respective positioned unbalance compensation groove ( 7 ) in the carrier wave ( 6 ) and / or multiple rotor-head bores ( 8th ) as well as a residual bending moment in sections of the carrier shaft ( 6 ), wherein the sum of the residual bending moments for the final finished (as ready for final assembly) is balanced overall rotor, which takes place over the total rotor length by means of wrap, where true, any combination of these measures to the desired balancing quality, for example DIN ISO1940 , ie: • "Flank profile adjustments" (sufficient for rotor tooth number ratios greater than 1: 1) and / or • "Imbalance compensation groove" ( 7 ) (especially if the number of rotor teeth ratio is 1: 1 required) and / or • "Rotorkopfbohrungen-Serie" ( 8th ) (especially if the number of rotor teeth is required to be 1: 1) and / or • "residual bending moments" (especially necessary for rotor tooth ratio of 1: 1) is possible. In this case, according to the invention, the flank profile adjustments are made such that the heat conduction is performed by each rotor tooth according to the known fin efficiency, which is well understood in the rotor axis section: Specifically, this means, for example, that relevant to the heat conduction rotor tooth cross section in the axial section representation to smaller Rotor diameter values must not be reduced, whereby not only the heat conduction through the rotor tooth material (with good thermal conductivity, for example as aluminum alloy), but also the heat absorption at the rotor surface has to be taken into account, which today thanks to modern PC computer performance, for example by means of FEM technology. Simulation calculations is well optimized. Here is the spindle rotor head, on one side of the sickle flank ( 1 ) ends, so as to minimize the internal leakage currents: for example, over the gap height (to minimize) and the length (to maximize) of this gap to the surrounding housing and over sharp edges against the leakage flow, because only this rotor head sees the cause Pressure difference between the directly adjacent working chambers, in turn, at the same time the known fin efficiency for each rotor tooth is observed. In addition, the rotor cooling is preferably according to DE 102015012 867.1 combined with cooling for the compressor housing ( 4 ), wherein the respective cooling fluid flows are conducted via a control unit to virtually every operating / operating point of the compressor "intelligent", where "intelligent" means that each individual cooling fluid flow from the control unit based on the respective compressor state so it is regulated that the thermal expansions due to the working operation of the screw compressor do not differ more than a factor of 2 from the gap values set during assembly (as so-called "cold gap dimensions").

The terms and terminology used here have been explained several times in my previous patent rights, so that a repetition is omitted here, but these explanations can of course be submitted later on request.

General explanation for understanding:

The desired improvement in compressor efficiency mainly depends on dissipating as much heat as possible during compression as much as possible. The design and size of the so-called "heat transfer surfaces" are crucial, especially in the spindle rotors. At the same time, the number of stages (ie the number of completed working chambers between inlet and outlet) is important for the reliable fulfillment of the application-specific requirements in the parameter design. Previously, the number of stages and heat transfer surfaces were directly coupled with the same machine geometry, i. at a high number of stages, the heat transfer surfaces were automatically larger than at a lower number of stages. However, this disadvantageously reduces the heat transfer surfaces with a small number of stages, as a result of which the removal of heat during compaction is inevitably reduced and consequently also the compressor efficiency with respect to polytropic exponential change deteriorates. The relevant heat transfer surfaces are generated by the spindle rotors, the rotor production on the so-called "comb border" (which profile gap depth can still be achieved at a desired profile gap width of the machining tool for the correct generation of the profile edges) the feasible slope and thus the size of the Heat transfer surfaces determined. So far, however, the size of the heat transfer surfaces was determined by the number of stages and thus the wrap angle, so that often difficult compromises had to be found. With the present invention, this disadvantage is finally avoided, because regardless of the application-specific desired number of stages, the heat transfer surfaces can be maximized and are practically only dependent on the manufacturing "comb border". In the frontal section, the rotor end profile pairings are identical, with respect to the pitch curve, because of the opposing slope orientation, a mirror-image identical spindle rotor is created. The term "rotor tooth" is to be understood as meaning the rotor material between the rotor profile flanks in the rotor longitudinal axis direction.

Explanations to the exemplary illustrations in 1 to 3 :

1 shows an example of the present invention, a frontal section simplified in terms of the change in the axis distance as a planar representation of an inventive rotor flank pair of execution with the Number of teeth ratio of 2: 2, wherein in this rotor position just the sickle edges ( 1 ) of each rotor are engaged with each other. It is easy to see how the end point E0 of a sickle flank ( 1 ) the course of the other crescent flank ( 1 ) on the engaging counter spindle rotor ( 5 ) generated as a counter-flank. The internal rotor cooling preferably in accordance with industrial property rights DE 10 2015 012867.1 is as a dashed line between rotor ( 5 ) and carrier wave ( 6 ). The relevant profile flank contours are shown as thicker lines. The centroid FS of the hatched rotor front-face interface is due to the edge adjustment to achieve the desired balancing such close enough to the rotor fulcrum OR, that the difference in such a representation is practically unrepresentable, because these two points in a total scale practically to each other come.

2 shows an example of the present invention, an end section simplified in terms of the change in the axial distance as a planar representation of a rotor flank pair of embodiments according to the invention with the teeth ratio of 1: 1, wherein the rotor head bores ( 8th ) were projected for convenient representation in this one end section plane, because their actual bore line is dependent on the pitch curve of the spindle rotor pretty much in the middle on the respective Spindelrotorkopflinie ( 11 ) lie in each endcut, and would have, for example, a constant pitch course a virtually equal distance from the spiral line, the sum of all end points E0 of the crescent edge ( 1 ) is produced. The unbalance compensation groove ( 7 ) in the carrier wave ( 6 ) may look in the present cross-section as shown, in axial section this unbalance compensation groove ( 7 ) in the carrier wave ( 6 ) then corresponds to the typical groove shape by the inclination gradient of the spindle rotor is always made exactly positioned in the region of the rotor tooth profile, or to the protection right DE 102015012867.1 rotatably on the carrier shaft ( 6 ) applied spindle rotor ( 5 ) fits. The internal rotor cooling preferably in accordance with industrial property rights DE 10 2015 012867.1 is as a dashed line between rotor ( 5 ) and carrier wave ( 6 ). The relevant profile flank contours are shown as thicker lines.

3 shows an example of the present invention, an endcut simplified with respect to the Achsabstandsänderung a flat representation of a Rotorflankenpaar execution according to the invention with the teeth ratio of 3: 3, for a preferably identical Rotorstirnschnittprofilpaarung the pitch angle between the Flankenwälzpunkten at this teeth ratio is logically 60 °.

The flanks shown in the following Fig.-drawings were calculated according to profile slope function and meet in the endcut the desired condition that the centroid FS to achieve the desired balance good enough close enough to be at the rotor pivot point OR, the deviation are well below a hundredth of a millimeter ,

LIST OF REFERENCE NUMBERS

• 1. sickle flank as an extended epitrochoid whose line of action extends to the intersection edge G of the two housing bores and mathematically generated by the fact that the end point E0 of a rotor in the rotor pair rotation leaves this lane as a locus on the respective other rotor and simplified as a planar frontal view is preferably described in parameter form via the following equations in the rotor-fixed xy coordinate system:
  
  (Index SF stands for Sickle Flank.)
  
  with: a = center distance of both rotors (simplified here for the plane case) R = head radius of the spindle rotor and running variable τ _SF according to:
  
• 2. Balance edge, the sickle flank ( 1 ) and like the sickle flank ( 1 ) occurs for all rotor tooth number ratios exactly once on each spindle rotor
• 3. Secondary flanks So all profile flanks between the crescent flank ( 1 ) and the balance edge ( 2 ), wherein for a selected rotor tooth number ratio z, the number of secondary edges is given by the following equation:
  
  
  and thus then applies: if z = 2, there are therefore 2q minor edges at z = 3, there are therefore 4 minor edges at z = 4, there are therefore 6 minor edges and so on. whereas at z = 1 there is no side-edge
• 4. compressor housing with outer cooling fins with a guided by CU volume flow of a cooling fluid, which in the circuit (preferably) flows through each spindle rotor.
• 5. Spindelrotorpaar, as an intermeshing rotor pair once right and left once again, preferably mirror-inverted identical and with internal rotor cooling ( 9 ) preferably according to property rights DE 10 2015 012867.1 executed.
• 6. carrier shaft of a material with high bending stiffness and higher density than the rotor body (ie, for example, carrier shaft of a steel material and each rotor body made of aluminum material)
• 7. Imbalance compensation groove in the (steel) carrier shaft ( 6 ) positioned in each endcut as unbalance compensation to the rotor profile, in particular at the rotor teeth ratio of 1: 1 in accordance with DE102015012 867.1 rotatably mounted (aluminum) rotor body with subsequent processing of the outer gas supply thread, so that the unbalance compensation is exactly positioned to the rotor body with his percutaneous specific Gasfördergewindeprofilverlauf
• 8. Rotor head bores serially several times in succession
• 9. internal rotor cooling, preferably according to DE 102015012867.1
• 10. Central cooling fluid supply bore in each carrier shaft ( 6 )
• 11. Spindelrotorkopf, on one side of the sickle flank ( 1 ) ends

List of terms:

• a

  Axial distance (simplified in the plane representation to the rotor pair endcut)

  r

  Pitch circle radius, which corresponds exactly to half the center distance because of the 1: 1 gear ratio

  R

  Tip circle radius

  z

  Rotor tooth number ratio generally as z: z with z as natural number, ie all positive integers, ie: z = 1, 2, 3, 4, 5 ...

  E ₀

  End point of the sickle flank ( 1 ), which in rotor pair-rolling respectively the sickle flank ( 1 ) generated on the counter rotor.

  C

  Rolling point of the toothing with the balance edge and with the minor sidewalls, the pitch point always lies on the pitch circle of the toothing and represents the transition between head flank and foot flank.

  G

  Point of the intersection edge of the two holes in the compressor housing ( 4 ) for receiving the spindle rotors ( 5 )

  O _R

  Spindle rotor pivot point

  F _S

  Center of gravity in each endcut, which is sufficiently dense (ie, according to the desired balancing quality, whereby the distance between O _R and F _{S is} generally significantly smaller than 1 mm) by appropriate execution of the flank profile profiles for balance edge and minor flanks to the spindle rotor pivot point O to bring is. In the drawings, simplified on the hatching of the face cutting surface registered.

QUOTES INCLUDE IN THE DESCRIPTION

This list of the documents listed by the applicant has been generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Cited patent literature

• EP 2014/060851 [0003, 0003]
• DE 102015012867 [0006, 0006, 0010, 0011, 0011, 0013, 0013, 0013]

Cited non-patent literature

• Standard DIN-ISO 1940 [0004]
• DIN-ISO 1940 [0006]
• DIN-ISO 1940 [0006]
• Source: Steffens, Ralf: "The profile slope function, a new way to analytically determine and optimize general profile edge pairing", Dissertation TU-Baunschweig, 1993, Report No. 40 [0006]
• DIN-ISO1940 [0006]

Claims (5)

Rotor flank pairing for a spindle compressor as a 2-shaft rotary displacement machine operating in the working area without operating fluid for conveying and compressing gaseous media for applications in vacuum and for applications in overpressure including refrigerant compressor with a pair of spindle rotors ( 5 ), which in opposite directions with a ratio of 1: 1 in a surrounding compressor housing ( 4 ), characterized in that the rotor edge pairing is performed as a (preferably) mirror-image identical profile pairing with a rotor teeth ratio of 1: 1 or 2: 2 or 3: 3 or 4: 4 or even higher, the most suitable version depending on the respective application-specific Requirement profile (for example, pressure level, working range, size, etc.) and the respective leakage gap losses (keyword entropy and respective manufacturing accuracies) is determined, each spindle rotor has a wrap angle of at least 1000 °, and that in each rotor a flank as sickle flank ( 1 ), characterized in that the engagement line is theoretically up to the intersection edge G of the compressor housing ( 4 ), so that a "blowhole-free rotor profile pair" is formed, and in addition that the other profile edges ( 2 and 3 ) be carried out such that at the desired high rotor speeds (ie speed values such that the average peripheral speed values at the rotor heads reach at least 40 m / sec) the respective required balancing quality (for example, according to DIN-ISO 1940) is achieved by for Rotorzähnezahlverhältnisse of 2: 2 and higher the balance edge ( 2 ) and the minor flanks ( 3 ) according to the known Verzahnungsgesetz abrolling each other via a so-called. "Flank profile adjustment" be carried out such that the centroid FS in each endcut corresponding to the desired balancing quality (eg according to DIN-ISO 1940) is so close enough to the rotor fulcrum O _R , so that the rotors can be operated safely in the desired speed range. Spindle compressor according to claim 1, characterized in that at a rotor tooth ratio of 1: 1 in addition to this "edge profile adjustment" for the balance edge ( 2 ) as further measures a respective positioned unbalance compensation groove ( 7 ) in the carrier wave ( 6 ) and / or multiple rotor-head bores ( 8th ) and / or also a partial residual bending moment support on the carrier shaft ( 6 ), wherein the sum of the residual bending moments for the final finished (as ready for final assembly) is balanced overall rotor, which takes place over the total rotor length by means of wrap, where truly any combination of these measures to the desired balancing quality, for example, according to DIN-ISO1940, So: • "Flank profile adjustments" (is sufficient for rotor tooth number ratios greater than 1: 1) and / or • "unbalance compensation groove" ( 7 ) (especially necessary for rotor number ratio of 1: 1) and / or • "Rotorkopfbohrungen-Serie" ( 8th ) (especially necessary for the rotor teeth ratio of 1: 1) and / or • "residual bending moments" (especially necessary for the rotor teeth ratio of 1: 1) is possible. Spindle compressor according to claim 1 and 2, characterized in that the flank profile adjustments are made such that the heat conduction is performed optimally by each rotor tooth according to the known fin efficiency (= description in the text and in the generally available literature) in terms of heat absorption and heat conduction according to the known rules, for example, in the rotor axis section, the tooth profile cross-section does not decrease in the direction of the rotor axis of rotation. Spindle compressor according to one of the preceding claims, characterized in that the spindle rotor head ( 11 ), on one side of the sickle flank ( 1 ) is performed so that the internal leakage flows are minimized by at this gap to the surrounding compressor housing ( 4 ), for example, the gap height minimized and the length can be maximized and the leading edges against the leakage flow are made sharp-edged. Spindle compressor according to one of the preceding claims, characterized in that the rotor cooling preferably according to DE 10 2015 012867.1 combined with cooling for the compressor housing ( 4 ), wherein the respective cooling fluid flows are conducted via a control unit to virtually every operating / operating point of the compressor "intelligent", where "intelligent" means that each individual cooling fluid flow from the control unit based on the respective compressor state is regulated so that the thermal expansions by the working operation of the screw compressor no more than a factor of 2 from the set during installation gap values (as so-called. "Cold gap") differ.

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