DE102013009040A1 - Spindle compressor with high internal compression - Google Patents

Abstract

The invention relates to dry-compressing 2-shaft rotary displacement machines for conveying and compressing gases for applications in vacuum and overpressure. In order to improve efficiency and compression capacity for higher pressure ratios, it is proposed according to the invention that the pair of rotors driven in opposite directions by an external synchronization consisting of 2-tooth (2) and 3-tooth spindle rotors has a non-parallel, that is to say diverging, center distance such that the The inlet-side value a.0 (1.a) is greater than the outlet-side aL (1.b), with an axis crossing angle γΣ (11) and depending on the construction with an axis offset aV (12) with multiple wrapping and small pitch differences as well On the inlet side of the 2-tooth spindle rotor (2) with a flattened portion (21) on the outer tip circle and on the inlet end with a reference surface (22.a, 22b), each spindle rotor having an internal fluid cooling system (8, 9) with cooling fluid Feed (13) at the inlet-side carrier shaft ends, the cooling fluid flow direction (14) being opposite in the rotor cone t to the direction of flow of the pumped medium (15) and on the inlet side via a pitot tube (29) from a collecting channel (28) via cooling fluid outlets (31) to the oil reservoir again.

Description

State of the art:

Drying compressors 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 medium, the known wet-running compressor, such as liquid ring machines, rotary vane pumps and oil or water-injected screw compressors, more and more 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 PCT font WO 00/12899 For a dry-compressing spindle displacement machine, a simple rotor cooling is described by a coolant, preferably oil, is introduced into a conical rotor bore at each rotor to constantly dissipate a portion of the resulting during the compression compression heat. In the protection right PCT / EP2008 / 068364 is in continuation of this approach, the coolant with an internal coolant (oil) pump further still used to cool the pump housing, in a preferably common coolant circuit via a separate heat exchanger, the absorbed amounts of heat from the compression of the fluid and the power losses such to dissipate the clearance play values between the pair of rotors and the surrounding pump housing for all operating conditions. In the patent application DE 10 2012 009 103.6 A description is given of a rotor pairing by means of a 2-toothed and 3-toothed spindle rotor which has been improved with regard to minimizing the internal leakage and design of the working chambers. However, both the compressibility and the power efficiency are still to improve especially for more demanding applications in the pressure with higher pressure ratios, because through internal leakage between the individual working chambers there are still too high losses at the same time often insufficient heat dissipation during compression. This situation needs to be improved.

The object of the present invention is to significantly improve the efficiency and the compressibility for dry compressing 2-shaft rotary displacement machines for conveying and compressing gaseous media for applications in vacuum and in overpressure with pressure ratios between inlet and outlet pressure greater than 3 at the same time simplified spindle rotor production.

According to the invention this object is achieved in that for vacuum and overpressure applications in a dry-compressing spindle compressor as 2-wave displacement machine in opposite directions from an outer, so located outside of the compressor working space, synchronizing rotation angle driven pair of rotors consisting of a 2-toothed spindle rotor and a toothed engaging 3-toothed spindle rotor has a non-parallel, ie diverging center distance such that on the conveying gas inlet side of the axial distance between the two spindle rotors at least 20%, better 50% and for higher pressure ratios even more than twice as large the exhaust-side wheelbase, wherein the two rotor axes are intersecting with an axis crossing angle γ _Σ that is non-zero and between 2 and 40 angular degrees, and then has no axle offset a _V , if available in the design implementation Space for the pressure side bearing sufficient, whereas to improve the constructive space conditions, the two rotor axes crossing (also referred to as "skewed") with an axis crossing angle γ _Σ not equal to zero from 2 to 40 angular degrees and with a Achsversetzung a _{V are} executed, which is greater than 30 mm and positioned within ± 20% of the total rotor length to the front rotor outlet, but preferably coincides with the end cutting planes of both rotor outlet end faces, also at a Wrap angle with respect to the 2-toothed spindle rotor of at least 2500 angular degrees, but preferably over 3600 angular degrees, desirably more than 4500 angular degrees or even better over 5400 angular degrees and for particularly high pressure differences even over 6300 angular degrees, because the higher the compressibility should be, the more greater is to choose the wrap angle, wherein at the spindle rotor pair, the inlet-side slope by less than 20%, for some applications still 60% but at most 120% greater than the outlet-side slope, the inlet side of the 2-toothed spindle rotor to improve the Suction capacity is a flattening at the outer tip diameter over a length L.2zyl, which corresponds at least to the averaged start profile pitch length, such that the inlet-side head cone inclination angle γ _{2K.0 is} preferably zero (ie cylindrical), but in any case smaller than the head cone inclination angle γ _2K.L , which defines for at least 66% of the rotor profile length L _2z, the reduction of the rotor outer diameter toward the outlet side, wherein moreover each spindle rotor has an internal cooling with the feature that the flow direction of the conveying medium is opposite to the cooling cone flow direction of the cooling fluid, which in Rotor inlet area is fed into the respective carrier shaft and at each on its own carrier shaft rotatably seated spindle rotor is driven by the centrifugal movement of the spindle rotor rotation to the inlet side and there is preferably continuously tapped by a pitot tube from a gutter, with a particularly high Cooling fluid flow rate is adjusted so that the temperature increase in the cooling fluid by the heat removal from each spindle rotor during the compression of the pumped medium is less than 6 degrees, better still less than 4 degrees, and ideally even kept below 3 degrees, and also still the K ühlkonus angle in the smaller diameter range by at least 10%, better still 25% or ideally more than 50% is performed steeper than in the below flowed by the cooling fluid cooling cone area, wherein on the spindle rotor geometry geometry values, the thermodynamic interpretation is such that the specific heat exchanger characteristic value (because ultimately the present spindle compressor both compressor and at the same time heat exchanger) a heat transfer surface load as compressor power per workspace surface size (ie the contacted by the pumped medium in the working space compressor working surfaces, namely rotor pair -Fördergewinde and enveloping housing inner surfaces) of less than 50 kW per m ² and for larger machines (ie over 55 kW) still under 30 kW per m ² and higher efficiency requirements even less than 20 kW per m ² , where In addition, the configuration of the conveyor thread profile flanks on both spindle rotors such takes place that in each end section plane within the rotor profile thread length L.3z on 3-toothed spindle rotor a head circle radius r.3KK with the respective head end points E.3a, E.3b, E.3c, E.3d, E.3e and E.3f is chosen such that for at least 80% of the rotor profile total length L.3z the following condition is fulfilled: 0.42 · ā (z) ≤ r.3KK (z) ≤ 0.72 · ā (z)
with ā (z) as the respective center distance to the corresponding pairing of the end cutting planes of both rotors, and that also in each end section plane within the rotor profile thread length L.2z on the 2-toothed spindle rotor a head circle radius r.2KK with the respective head end points E. 2a, E.2b, E.2c and E.2d is selected such that for at least 80% of the rotor profile total length L.2z the following condition is satisfied: 0.25 · ā (z) ≤ r.2KK (z) ≤ 0.52 · ā (z) and at the same time with the additional condition: (r.3KK (z) + r.2KK (z))> 1.05 · ā (z) as an intermeshing rotor pairing, and further at the inlet side End face of the 2-toothed spindle rotor, a tapered reference stop surface is formed, which is inclined for correct positioning of the toothing planes in the rotor longitudinal axial direction of both spindle rotors to each other about the Achskreuzungswinkel, and also minimizing the Verdrehflangenenspiels in the necessary for the synchronization of both rotors Konuszahnradpaa r carried out by axial displacement of these gears to each other with subsequent fixation, and the axial forces from this Synchro.-teeth used in particular by selecting the helix angle and the positioning of the helix angle at crossing axis targeted for the compensation of the gas forces by the pressure difference of the pumped medium between output and inlet are used in conjunction with the known countershaft transmission stage for speed increase, wherein for very high end pressure values (depending on the machine dimensions, for example for more than 12 bar) to reduce the axial bearing load due to the high pressure differences, a pressure differential balancing stage at each outlet side carrier shaft end is realized.

Some terms are briefly explained:

According to the gearing law, the distance between the axes of rotation of the two spindle rotors per conveyor screw gearing plane in the z-axis direction is considered to be the axial distance between the largest value a.0 at the inlet and the smallest value a.L at the outlet. In this case, the resulting toothing in crossing, that is "skewed" axes of rotation is no longer part of the Wälz- but to the screw drives.

The "wrap angle" on the spindle rotor is the sum of all angles of rotation along the spindle rotor axis between the individual cross-sectional profile contours that occur as the z axis axis progresses. Overall result in rotor longitudinal axis direction. Thus, if the profile endcut at a z-position z _{i is} compared to the profile-endcut at the adjacent position z _{i + 1} , then both endcuts are one to another according to the selected z (phi) function for that particular step of z _i twisted to z _{i + 1} known angle phi _i . The sum of all angles of rotation for the end cuts along the spindle rotor axis gives the wrap angle, which is here referred to the 2-toothed rotor, and is referred to briefly as PHI.2. For the 3-toothed rotor, this angle of rotation is to be adjusted by the gear ratio as a factor according to the gearing law and is therefore necessarily determined at the same spindle rotor length. The wrap angle is the decisive measure for the number of stages.

The number of completed working chambers at the spindle rotor pair between the rotor inlet side and the rotor outlet side is considered as "number of stages". The rotor length and selected z (phi) function with overall wrap angle PHI.2 should be aimed at as even an integer number of stages as possible. In this case, preferably the PHI.2 value is rounded up to at least the next 10th digit, ie z. From 2411 ° to 2420 °.

A "working chamber" is closed for the rotor pair tooth space volume, which is bounded by the surrounding compressor housing and the spindle rotor profile back flanks between the defined according to gearing law profile contour interventions, said engaging rotor pair profile edges are considered as touching, so close to zero distance , Practically, however, the engaging rotor pair profile flanks have a certain, albeit minimal, distance, resulting in an internal leakage return flow. The "working chamber volume on the inlet side" is the volume of the first working chamber closed on the suction side, and the "working chamber volume on the outlet side" is corresponding to the volume of the last closed working chamber in front of the conveying gas outlet. The quotient of these two volumes represents the "internal compression ratio". Values over 3 can be suitably defined as "higher internal compression ratios". The volume of a working chamber is calculated from the respective working area cross-sectional area multiplied by the stepwise working chamber extent defined by the spindle pitch in the rotor longitudinal axis direction.

In particular, per spindle rotor, each cut is perpendicular to the spindle rotor rotation axis, which is preferably defined as a z-axis, so that the end cut lies in the x-y plane of the right-angled Cartesian coordinate system. When the axis position is crossed, the end section planes of the two spindle rotors have the axis crossing angle to each other.

As "pitch" in the spindle rotor conveying thread is the length progress in Rotorlängsachsrichtung after exactly one revolution (ie 360 degrees angle) of the wrap. Thanks to state-of-the-art production machines, different gradients in the rotor's longitudinal axis can be realized today with each rotor in order to set the volumetric curve in a thermodynamically targeted manner as a progression of the working chamber volumes.

The "external synchronization" of the two spindle rotors is required because the rotor pair operates in the compressor working space without operating fluid, ie "dry compressing" is operated, and therefore because of the high speeds rotates without contact with the least possible edge distance to each other in opposite directions. In order for this non-contact operation of the rotor pair can be constantly ensured, the two spindle rotors are constantly driven with high, accurate in the range less minute minutes rotation angle accuracy, which is known to be carried out via external synchronization. By far the most common version for external synchronization is via directly meshing gears. But there is also, for example, the possibility of electronic rotor pair synchronization by each rotor is driven by his own engine electronically true to the angle of rotation.

The "inlet area" can be described by the wrap angle area with which the first closed working chamber on the inlet side is formed by progressive twist angle. This happens in the case of the 2: 3 spindle rotor pair starting from the inlet end cutting side after 720 degrees of angle plus the head arc angle ga.KB2 on the inlet side of the 2-toothed spindle rotor.

In the case of atmospheric suction, the term "overpressure" is defined as absolute pressures of at least 2 bar, usually 8 bar to 15 bar, but with a high number of stages, pressure values of more than 25 bar can be achieved. With non-atmospheric aspiration, these values shift accordingly.

The term "vacuum" or negative pressure is defined as the ultimate pressure values of less than 50 mbar, better still less than 1 mbar and with a corresponding number of stages even less than 0.01 mbar absolute versus outlet pressure, which is in the atmospheric pressure range.

The present invention will be further explained by means of the following illustrations:

1 shows an example of the present invention is a sectional view through the spindle rotor pair with auseinanderstrebendem axial distance, this representation is simplistic, because the axes of rotation ( 6 ) and ( 7 ) do not have to lie in one plane. At the same time, it can be clearly seen in this drawing that, when the axes of rotation intersect, the space in the structural conversion for the outlet-side bearing (FIG. 24 ) becomes scarce, because these outlet bearings of the rotor and counter rotor come too close to each other, for the carrier waves a two-sided storage at the same time the lowest possible bearing distance per rotor because of the required high rotor speeds and the associated critical speed is considered essential. The positioning of the necessary synchro-wheels ( 37 ) On the outlet side is cheaper, because the size of the gears would otherwise be too large inlet side, which would be expensive not only because of the enormous peripheral speeds. This is due to the outlet-side constructive space conditions in this representation, a crossing position of the axes of rotation cheaper, the axial offset in the vicinity of the two spindle rotor outlet end faces ( 36 ) shows itself to be advantageous.

Furthermore, it is clear in this illustration that in the conveying thread area ( 16 and 17 ) follow the rotor outer diameter and thus also the conveying thread profile height of h.0 to hL the variable center distance a (z), which is directly traceable on the conditions mentioned above according to the invention in the choice of head-radius values. In this way, compared with the prior art, the desired high internal compression ratio is achieved in a particularly high number of stages and associated larger working chamber surfaces for heat dissipation during compression to improve the efficiency of compression over a lower Polytropenexponenten according to the laws of thermodynamics , Due to the conditions mentioned in determining the conveying thread profile flanks also the inner leakage is reduced by blow hole-free compression.

In addition, the flattening invention ( 21 ) on the outer diameter of the 2-toothed spindle rotor, the suction capacity of the spindle compressor and must not be achieved as previously by changing the Fördergewinde pitch. This enhances the efficiency improvement.

The cooling fluid supply ( 31 ) takes place at each inlet end of the carrier waves ( 4 and 5 ) in the central feed bore, so that an operation-dependent pressure difference load in the oil reservoir and thus also for the drive shaft implementation is advantageously avoided.

2 shows by way of example the two rotor axes of rotation ( 6 . 7 ) in crossing (also called "skewed") version with the axle offset aV ( 12 ) as the shortest distance between the two axes of rotation (and thus perpendicular to both axes) below the axis crossing angle γΣ ( 11 ). Advantageously, (as also shown here) the outlet-side center distance aL ( 1.b ) and the axle offset aV ( 12 ) equal and at the same time starting point on the respective axis of rotation for each run parameters z.2 and z.3 for the description of the toothing conditions on the respective end cutting planes. The delivery thread for the spindle rotor pair is ultimately only a toothing according to the gearing law, namely that in each edge profile contact point the speed components perpendicular to the flank surface of each rotor must be equal. Otherwise, it would come at the touch point for lifting or penetration of both flank surfaces. When crossing axes of rotation according to the present representation will be seen that from the previous Wälzgetriebe for the spindle rotor pair conveyor thread now a helical gear is with the general center distance ā (z) and the instantaneous axis ( 10 ) Reference book by Karlheinz Roth: "Gear Technology, involute special gearings", ISBN 3-540-64236-6, Springer-Verlag 1998 and dissertation by Tsai, S.-J. "Unified System of Evolventic Gears", TU-Braunschweig 1997 ]

3 shows an example of the present invention, a representation of the profile design, this representation is simplified by the two actually inclined to each other end face planes are shown per rotor in a common plane. Nevertheless, the essential features for generating the profile flanks ( 38 and 39 ). Recognize the head profile end points E.2a, E.2b, E.2c and E.2d defined per z position over the selected angle be.KB2 (z) and over the likewise selected tip circle radius r.2KK (z) in its rotational movement with ω.2 about its axis of rotation ( 6 ) for the 2-toothed spindle rotor point-wise the counter-profile flank ( 39 ) of the 3-toothed spindle rotor which rotates about its axis of rotation ( 7 ) rotates with ω.3, both axes of rotation ( 6 . 7 ) according to 2 by axis crossing angle γ _Σ ( 11 ) and (if so selected) axis offset aV (12) are defined to each other in space. The rotational speeds ω.2 and (0.3 are of course defined according to the gear ratio by the number of rotor teeth.) Similarly, with the head circle radius r.3KK (z) selected, the head profile end points E.3a, E.3b, E.3c, E.3d, E .3e and E.3f point-wise the counter-profile flank ( 38 ) determined at the 2-toothed spindle rotor, wherein the blow hole freedom is achieved in particular in the range of large working chamber volume changes, characterized in that the course of the central angle in polar coordinate representation to the determined in this way Conveyor thread flank ( 38 ) undergoes no sign inversion. If this condition of the monotonous central angle curve is violated, the selected value for the tip circle radius r.3KK (z) must be reduced at this z position until the central angle profile is monotone.

In the mathematical determination of the respective counter-profile flank is cleverly proceeded according to the "principle of kinematic inversion" by each of the spindle rotor with the generating head profile end points its fixed counter-rotor according to Achskreuzungswinkel and (if selected) axle offset from each other in space, so that for this Consideration of the actually housing-fixed center distance rotates. The stationary observer then sees the respective counter-profile arising for "his" standing rotor.

In 4 is detailed too 1 the inlet side cooling fluid supply ( 13 ) via the housing-fixed feed tube ( 40 ), from which the cooling fluid (usually oil) passes into the central bore of each carrier shaft, where it is pressed against the bore wall due to the centrifugal forces during rotation and due to the overflow bushing ( 41 ) via the transfer holes ( 42 ) according to 1 gets to the inner cone of the respective spindle rotor. There it is centrifugally conditioned according to the flow direction ( 14 ) to the inlet-side rotor end and there in a rotor-fixed and sealed gutter ( 28 ) as a rotating fluid ring ( 27 ) in order to gain access to 29 ) to be tapped. About the rack-solid collecting space ( 30 ) the cooling fluid passes through discharge ( 31 ) in the oil reservoir, which is not shown here.

5 shows in a section perpendicular to the axis of rotation in a plan view of the operation of the cut-open pitot tube ( 29 ), whose openings in the rotating cooling fluid collecting ring ( 27 protrude). The kinetic energy of the rotating cooling fluid ring ( 27 ) the cooling fluid through the pitot tubes ( 29 ) in the fixed collecting space ( 30 ). Since only one direction of rotation is permissible and clearly defined for the spindle rotor pair, the pitot tubes must ( 29 ) even in one direction only. In this case, the size of the pitot tube opening cross-sections should be designed so that always a safe discharge of the amount of cooling fluid is guaranteed, so oversize as a precaution. In addition to the pitot tube opening cross-sections, therefore, the number of Pitot tubes to be increased, which for the purpose of safe skimming are advantageously also arranged offset from each other axially. For simpler applications, this Pitot tube solution can also be replaced by simple cooling fluid collecting cushions as simple (steel) wool packages (accumulations).

LIST OF REFERENCE NUMBERS

1

Center distance a (z) as a non-parallel, ie diverging distance per tooth plane in the z-axis direction between the axes of rotation of the two spindle rotors within the limits: 1.a. Intake-side center distance a.0 as the largest conveying thread-toothing distance 1.b Outlet-side center distance aL as the smallest conveyor thread-toothing distance

2

2-toothed spindle rotor, referred to for short as "Rotor-2", with the overall thread length L.2z

3

3-tooth spindle rotor, referred to for short as "Rotor-3", with the overall length of the conveyor thread L.3z

4

Support shaft for the rotor-2, which is rotatably connected to the rotor-2 (preferably pressed) with central cooling fluid supply hole

5

Support shaft for the rotor-3, which is non-rotatably connected to the rotor-3 (preferably pressed) with central cooling fluid supply hole

6

Rotation axis for the rotor-2 with the angular velocity ω.2

7

Rotation axis for the rotor-3 with the angular velocity ω.3 opposite to ω.2 while maintaining the gear ratio between the 2-toothed and the 3-toothed rotor

8th

Rotor internal fluid cooling for the rotor-2 according to PCT script WO 00/12899

9

Rotor internal fluid cooling for the rotor-3 according to PCT script WO 00/12899

10

Current axis of the spindle rotor toothing with crossing axes of rotation

11

Axis crossing angle γ _Σ

12

Axle displacement a _V

13

Cooling fluid supply

14

Coolant flow direction in the cooling cone of each rotor

15

Flow direction of the pumped medium

16

Rotor 2 Feed Screws

17

Rotor 3 Feed Screws

18

Conveying thread engagement area for the spindle rotor pair

19

_{Head taper} inclination angle γ _3K.L at rotor-3

20

Head cone inclination angle γ _2L.L at the rotor-2

21

Inlet-side flattening at the outer tip circle diameter at the inclination angle γ _2k.0 for the length L.2zyl at the rotor-2

22

Reference position area optionally executed as: 22.a Reference position area on the inlet end of the rotor-2 or 22.b Reference position surface at the inlet end of the rotor-3

23

Inlet side rotor bearing

24

Outlet-side rotor bearing

25

Shaft seal for the compressor working space

26

neutral space between oil-free compressor working space and oil-lubricated storage space

27

rotating cooling fluid collecting ring, held by the centrifugal forces in the gutter (28)

28

Pitot tube, rotatably rotating with each spindle rotor body

29

Pitot tube, fixed and housing-fixed with exhaust cavities

30

Cooling fluid collecting chamber, fixed and housing-fixed

31

Cooling fluid discharge

32

Counter profile-generating head end points on rotor-2, namely points E.2a, E.2b, E.2c and E.3d together on circle with radius r.2KK (z)

33

Counter profile-generating head end points on rotor-3, namely the points E.3a, E.3b, E.3c, E.3d, E.3e and E.3f together on circle with radius r.3KK (z)

34

Inlet space for the pumped medium

35

Discharge space for the pumped medium

36

Spindle rotor outlet face

37

Conical gear pair (also called "beveloid") for synchronization for the spindle rotor pair

38

Conveyor thread flanks on the rotor-2, generated by the head end points of the rotor-3

39

Conveyor thread flanks on the rotor-3, generated by the head end points of the rotor-2

40

Housing / stationary cooling fluid supply tube

41

Carrier shaft fixed overflow socket

42

Cooling fluid transfer bores

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

• WO 00/12899 [0003, 0025, 0025]
• EP 2008/068364 [0003]
• DE 102012009103 [0003]

Cited non-patent literature

• Reference book by Karlheinz Roth: "Gear Technology, involute special gearings", ISBN 3-540-64236-6, Springer-Verlag 1998 and dissertation by Tsai, S.-J. "Unified System of Evolventic Gears", TU-Braunschweig 1997 [0021]

Claims (12)

Spindle compressor as a 2-shaft rotary displacement machine operating in working space without operating fluid for conveying and compressing gaseous media for applications in vacuum and for applications in overpressure for pressure ratios between inlet and outlet pressure greater than 3 with an opposite from an outer, ie outside from the compressor workspace located synchronization ( 37 ) rotatably driven spindle rotor pair in a surrounding compressor housing with an inlet ( 34 ) and an outlet ( 35 ) characterized in that the rotor pair consisting of a 2-toothed spindle rotor ( 2 ) and a toothed engaging 3-toothed spindle rotor ( 3 ) has a non-parallel, ie diverging center distance such that on the conveying gas inlet side ( 34 ) the axial distance a.0 (1.a) between both spindle rotors is at least 20%, better 50% and for higher pressure ratios even more than twice as large as the outlet-side axial distance aL (1.b), wherein also the wrap angle related on the 2-toothed spindle rotor at least 2500 angular degrees, but preferably over 3600 angular degrees, desirably even more than 4500 degrees, or even better over 5400 degrees and even for particularly high pressure differences even over 6300 degrees, because the higher the compressibility is, the greater In addition, the wrap angle on the spindle rotor pair is less than 20% larger than the outlet side slope for some applications, but still 60% but not more than 120% for some applications, and each spindle rotor further includes rotor internal fluid cooling (FIG. 8th . 9 ) has the feature that the flow direction ( 15 ) of the pumped medium takes place opposite to the cooling cone flow direction ( 14 ) of the cooling fluid, which in the rotor inlet region into the respective central cooling fluid supply hole in each carrier shaft ( 4 . 5 ) is supplied ( 13 ) and at each on its own carrier shaft rotatably seated spindle rotor ( 2 . 3 ) is driven by the centrifugal movement of the rotor rotation to the inlet side in the hollow rotor body. Spindle compressor according to claim 1, characterized in that the diverging rotor axes of rotation ( 6 . 7 ) from the 2-toothed ( 2 ) and the 3-toothed ( 3 ) Spindle rotor are designed to be cutting with an axis crossing angle γ _Σ ( 11 ), which lies between 2 and 40 angular degrees, and without axial offset a _V for those constructions in which the outlet side of the respective compressor machine still sufficient space, in particular for the rotor bearing is present. Spindle compressor according to claim 1, characterized in that the diverging rotor axes of rotation ( 6 . 7 ) from the 2-toothed ( 2 ) and the 3-toothed ( 3 ) Spindle rotor are designed to be crossing with an axis crossing angle γ _Σ ( 11 ) between 2 and 40 angular degrees and with an axial offset a _V ( 12 ), which is greater than 30 mm and within ± 20% of the total length of the conveying thread L.2z or L.3z to the front-side rotor outlet ( 36 ) is positioned, but preferably with the end section planes of both spindle rotor outlet end faces ( 36 ) coincides. Spindle compressor according to one of the preceding claims, characterized in that the inlet side on the 2-toothed spindle rotor ( 2 ) a flattening ( 21 ) at the outer tip diameter over a length L.2zyl that corresponds at least to the mean start pitch length, such that the inlet-side _{tip taper} angle γ _{2K.0 is} preferably zero and thus cylindrical, but in any case is less than the head _taper inclination angle γ _2K.L ( 20 ) defining for at least 66% of the rotor profile length L _2z the reduction of the rotor outer diameter toward the outlet side. Spindle compressor according to one of the preceding claims, characterized in that either on the inlet end face of the 2-toothed spindle rotor ( 2 ) a tapered reference position surface ( 22.a ) or at the inlet end of the 3-toothed spindle rotor ( 3 ) a tapered reference position surface ( 22.b ), which are each about the axis crossing angle γ _Σ ( 11 ) is inclined for the correct positioning of the toothing planes in the rotor longitudinal axis of both spindle rotors to each other. Spindle compressor according to one of the preceding claims, characterized in that the cooling fluid supply ( 13 ) is carried out at each inlet-side carrier shaft end in the central feed bore and thereby a particularly high cooling fluid flow rate is set such that the temperature increase in the cooling fluid by the heat removal from each spindle rotor during the compression of the pumped liquid is less than 6 degrees, better still less than 4 degrees and ideally even less than 3 degrees. Spindle compressor according to claim 6, characterized in that in the rotor internal fluid cooling ( 8th . 9 ) of the cooling cone angle in the smaller diameter range by at least 10%, better still 25% or ideally more than 50% is performed steeper than in the subsequently flowed through by the cooling fluid cooling cone area. Spindle compressor according to one of the preceding claims, characterized in that the inlet side by Rotorinnen-fluid cooling ( 8th . 9 ) In operation, incoming cooling fluid at each spindle rotor ( 2 . 3 ) in a spindle rotor-solid pitot tube ( 28 ) and from there via housing-fixed Pitot tubes ( 29 ) is constantly tapped. Spindle compressor according to one of the preceding claims, characterized in that on the spindle rotor geometry values, the thermodynamic design is such that the specific heat exchanger characteristic (because ultimately the present spindle compressor both compressor and at the same time heat exchanger) a heat transfer surface load of less than 50 kW per m ² and for larger machines (ie more than 55 kW) still under 30 kW per m ² and with higher efficiency requirements even less than 20 kW per m ² . Spindle compressor according to one of the preceding claims, characterized in that the minimization of the torsional backlash in the for synchronization of both spindle rotors ( 2 . 3 ) necessary Konuszahnradpaar ( 37 ) takes place by axial displacement of these conical gears to each other with subsequent fixation. Spindle compressor according to one of the preceding claims, characterized in that the configuration of the conveyor thread profile flanks ( 38 and 39 ) takes place on both spindle rotors in such a way that in each end section plane within the rotor profile thread length L.3z on the 3-toothed spindle rotor ( 3 ) a head circle radius r.3KK (z) with the respective head end points E.3a, E.3b, E.3c, E.3d, E.3e and E.3f is selected such that for at least 80% of the rotor profile Total length L.3z the following condition is satisfied: 0.42 · ā (z) ≤ r.3KK (z) ≤ 0.72 · ā (z) for: 0 ≤ z ≤ L.3z where in the range z <(L. 2z-L.2zyl) as the upper limit preferably: r.3KK (z) ≤ 0.6 · ā (z) with ā (z) as the respective center distance to the corresponding pairing of the tooth planes of both rotors, and that also within each end section plane within the rotor profile thread length L.2z on the 2-toothed spindle rotor ( 2 ) a head circle radius r.2KK (z) with the respective head end points E.2a, E.2b, E.2c and E.2d is selected such that for at least 80% of the rotor profile total length L.2z the following condition is met : 0.25 · ā (z) ≤ r.2KK (z) ≤ 0.52 · ā (z) for: 0 ≤ z ≤ L.2z with the additional condition: (r.3KK (z) + r.2KK (z))> 1.05 · ā (z) as an intermeshing rotor pairing. Spindle compressor according to claim 4 and claim 11, characterized in that for at least 66% and at higher demands on the compressibility even better than 80% of the spindle rotor length (L.2z-L.2zyl) when represented by polar coordinates for the conveyor thread profile edge line ( 38 ) on the 2-toothed spindle rotor, the center angle at adjacent successive flank profile points monotonically without sign change runs by the per head radius r.3KK (z) selected head profile end points E.3a, E.3b, E.3c, E.3d, E.3e and E.3f of the 3-tooth spindle rotor point-wise the counter-profile flank ( 38 ) in the 2-toothed spindle rotor, which is preferably determined by kinematic inversion, which is avoided by reducing the value to the tip circle radius r.3KK (z) per z-tooth plane this sign change in Zentriwinkel, and also the value also selected for each toothing plane for the Tip circle radius r.2KK (z) deviates by only ± 20% from the value for tip circle radius r.3KK (z).

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