DE102013009040B4 - Spindle compressor with high internal compression - Google Patents

Abstract

Spindle compressor as a 2-shaft rotary displacement machine operating in the working chamber without operating fluid for conveying and compressing gaseous conveying media for applications in a vacuum and for applications in overpressure for pressure ratios between inlet and outlet pressure greater than 3 with a spindle rotor pair driven in opposite directions by an external synchronization (37) located outside the compressor working chamber in a rotationally accurate manner in a surrounding compressor housing with an inlet (34) and an outlet (35) for the conveying medium, characterized in that the rotor pair consisting of a 2-tooth spindle rotor (2) and a 3-tooth spindle rotor (3) engaging in toothing has a non-parallel, i.e. diverging, axial distance such that on the conveying gas inlet side (34) the axial distance a.0 (1.a) between the two spindle rotors is at least 20%, better 50% and for higher pressure ratios even more than twice as large as the outlet side. Axle distance a.L (1.b), wherein the wrap angle in relation to the 2-tooth spindle rotor is at least 2500 angular degrees, but preferably over 3600 angular degrees, advantageously even more than 4500 angular degrees or even better over 5400 angular degrees and for particularly high pressure differences even over 6300 angular degrees, and in addition, on the spindle rotor pair, the inlet-side slope is less than 20%, for some applications still 60% but at most 120% greater than the outlet-side slope, wherein furthermore each spindle rotor has a rotor internal fluid cooling (8, 9) with the feature that the flow direction (15) of the conveying medium is opposite to the cooling cone flow direction (14) of the cooling fluid, which is fed (13) in the rotor inlet area into the respective central cooling fluid feed hole in each carrier shaft (4, 5) and in each case on its own carrier shaft The non-rotatably mounted spindle rotor (2, 3) is driven by the centrifugal movement of the rotor rotation to the inlet side in the hollow rotor body.

Description

Dry-compressing compressors are becoming increasingly important in industrial compressor technology. Due to increasing obligations in environmental protection regulations and rising operating and disposal costs as well as increased demands on the purity of the pumped medium, the well-known wet-running compressors, such as liquid ring machines, rotary vane pumps and oil or water-injected screw compressors, are increasingly being replaced by dry-compressing 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 requirements in terms of reliability and robustness as well as size and weight while at the same time being low in price and providing satisfactory efficiency.

The well-known dry-compressing spindle compressors are ideal for improving this situation because, as typical 2-shaft displacement machines, they achieve a high compression capacity simply by achieving the necessary multi-stage function as a so-called "conveying thread" by connecting several closed working chambers in series via the number of wraps per displacement rotor, without requiring an operating fluid in the working chamber. In addition, the contactless rolling of the two counter-rotating spindle rotors enables an increased rotor speed, so that the nominal suction capacity and delivery efficiency increase at the same time in relation to the size. Dry-compressing spindle machines can be used for applications in both vacuum and overpressure, although the power requirement is naturally significantly higher in overpressure because in the overpressure range with final pressures well over 2 bar (absolute) up to 15 bar and even higher, significantly larger pressure differences have to be overcome.

In the PCT document WO 00 / 12 899 A1 A simple rotor cooling system is described for a dry-compressing spindle displacement machine, in which a coolant, preferably oil, is introduced into a conical rotor bore in each rotor in order to continuously dissipate part of the compression heat generated during the compression process. In the patent WO 2010 / 006 663 A1 In continuation of this approach, the coolant is also used to cool the pump housing with an internal coolant (oil) pump in order to dissipate the heat absorbed from the compression of the pumped medium and the power losses in a preferably common coolant circuit via a separate heat exchanger in such a way that the clearance values between the rotor pair and the surrounding pump housing are maintained for all operating conditions. In the patent application EN 10 2012 009 103 A1 An improved rotor pairing with a 2-tooth and 3-tooth spindle rotor is described with regard to minimizing internal leakage and design of the working chambers.

The publication also reveals EN 10 2012 202 712 A1 that with only one screw spindle compressor machine without operating fluid in the working chamber, higher pressures are achieved at the gas outlet by the fact that the spindle rotor pair, which work in opposite directions and without contact, each has a wrap angle of more than 800 degrees for the outer conveying thread, and that part of the compression heat generated between the gas inlet and the gas outlet chamber is dissipated both at the spindle rotor pair and at the surrounding compressor housing during the operating fluid-free compression, with a cooling fluid flowing through the spindle rotor pair and a cooling fluid flowing through the compressor housing via the cooling fluid guide, dissipating heat, and the size of the heat transfer surfaces touched by the conveying gas is at least 50 cm2 per kW drive power, with the working chamber volumes on the conveying gas inlet side being larger than the working chamber volumes on the conveying gas outlet side, and with the spindle rotor pair, which rotate in opposite directions and without contact to one another, also being operated at an increased speed, so that the circumferential speeds on the spindle rotor outer diameter are at least 30 m/sec. The proportion of the dissipated compression heat that is generated during the compression of the conveying medium between the gas inlet and the gas outlet is determined by the size of the heat transfer surfaces in relation to the compressor output.

Furthermore, the publication discloses EN 10 2011 004 960 A1 a compressed air compressor with the following features: a) a housing having an air inlet communicating with the atmosphere and a compressed air outlet, b) two spindle rotors arranged in the housing so as to be rotatable in opposite directions, each with at least one multi-rotating conveying thread, delimiting a plurality of conveying chambers arranged one after the other in the threads sealed on the housing side, which mesh with one another like a gear, c) a drive motor coupled to the spindle rotors, d) a cooling circuit connected in a heat-conducting manner to the conveying chambers of the dry-compressing spindle rotors, e) the last conveying chamber is open to the compressed air outlet and the operating pressure at the compressed air outlet is more than 3 bar.

In addition, the state of the art includes the publications FR 796 274 A , FR 789 211 A , GB 464 475 A as well as US 3 180 559 A known.

Nevertheless, both the compression capacity and the performance efficiency still need to be improved, particularly for more demanding applications in overpressure with higher pressure ratios, because internal leaks between the individual working chambers still result in excessive losses and, at the same time, often inadequate heat dissipation during compression. This situation needs to be improved.

The object of the present invention is to significantly improve the efficiency and compression capacity of dry-compressing 2-shaft rotary displacement machines for conveying and compressing gaseous conveying media for applications in vacuum and overpressure with pressure ratios between inlet and outlet pressure greater than 3, while at the same time simplifying 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 a 2-shaft displacement machine, the rotor pair consisting of a 2-tooth spindle rotor and a 3-tooth spindle rotor engaging in terms of gearing, which are driven in opposite directions by an external synchronization located outside the compressor working chamber, have a non-parallel, i.e. diverging, axis distance such that on the conveying gas inlet side the axis distance between the two spindle rotors is at least 20%, better 50% and for higher pressure ratios even more than twice as large as the outlet-side axis distance, whereby the two rotor axes are designed to intersect with an axis crossing angle γ _Σ that is not equal to zero and lies between 2 and 40 angular degrees, and then has no axis offset a _V if the available space for the pressure-side bearing is sufficient in the structural implementation, whereas to improve the structural Space conditions, the two rotor axes are designed to cross (also referred to as "skew") with an axis crossing angle γ _Σ not equal to zero of 2 to 40 degrees and with an axis offset a _V that is greater than 30 mm and is positioned within ±20% of the total rotor length to the front rotor outlet, but preferably coincides with the front cutting planes of both rotor outlet front sides, and with a wrap angle related to the 2-tooth spindle rotor of at least 2500 degrees, but preferably over 3600 degrees, preferably more than 4500 degrees or even better over 5400 degrees and for particularly high pressure differences even over 6300 degrees, because the higher the compression capacity is to be, the larger the wrap angle must be selected, whereby the inlet-side pitch on the spindle rotor pair is less than 20%, for some applications still 60% but at most 120% greater than the outlet-side pitch, whereby on the inlet side of the 2-tooth spindle rotor, in order to improve the suction capacity, a flattening of the outer tip circle diameter over a length L.2zyl, which corresponds at least to the averaged starting profile pitch length, is carried out in such a way that the inlet-side tip cone inclination angle γ _2K.0 is preferably zero (i.e. cylindrical), but in any case is smaller than the tip cone inclination angle γ _2K.L , which defines the reduction of the rotor outer diameter in the direction of the outlet side for at least 66% of the rotor profile length L _2z , whereby each spindle rotor also 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 is fed into the respective carrier shaft in the rotor inlet area and is rotated by the centrifugal movement of the spindle rotor, which is seated on its own carrier shaft in a rotationally fixed manner. Spindle rotor rotation is driven to the inlet side and is there preferably continuously tapped from a collecting trough via a pitot tube, whereby a particularly high cooling fluid flow rate is set in such a way that the temperature increase in the cooling fluid due to the heat dissipation from each spindle rotor during compression of the conveying medium is less than 6 degrees, better still below 4 degrees and ideally even below 3 degrees, and in addition the cooling cone angle in the smaller diameter area is made at least 10%, better still 25% or ideally more than 50% steeper than in the cooling cone area subsequently flowed through by the cooling fluid, whereby the thermodynamic design is carried out via the spindle rotor pair geometry values in such a way that the specific heat exchanger characteristic value (because ultimately the spindle compressor in question is both a compressor and a heat exchanger at the same time) a heat transfer surface load as compressor performance per working space surface size (i.e. the surface touched by the conveying medium in the working space Compressor working surfaces, namely rotor pair conveying threads and enveloping housing inner surfaces) of less than 50 kW per m ² and for larger machines (i.e. over 55 kW) even less than 30 kW per m ² and with higher efficiency requirements even less than 20 kW per m ² , whereby the design of the conveying thread profile flanks on both spindle rotors is carried out in such a way that in each frontal cutting plane within the rotor profile thread length L.3z on the 3-tooth spindle rotor, a tip circle radius r.3KK with the respective tip 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 met: 0.42· a (z) ≤ r3.KK(z) ≤ 0.72· a (e)
with ä(z) as the respective center distance to the corresponding pairing of the face cutting planes of both rotors, and that in each face cutting plane within the rotor profile thread length L.2z on the 2-tooth spindle rotor a tip circle radius r.2KK with the respective tip 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· a (z) ≤ r.2KK(z) ≤ 0.52· a (z) and at the same time with the additional condition: (r.3KK(z) + r.2KK(z)) > 1.05· a (z) as an intermeshing rotor pair, wherein a bevelled reference stop surface is also formed on the inlet-side face of the 2-tooth spindle rotor, which is inclined to one another by the axis crossing angle for the correct positioning of the toothing planes in the rotor longitudinal axis direction of both spindle rotors, and furthermore the minimization of the torsional flank play in the conical gear pair required for the synchronization of both rotors is achieved by axial displacement of these gears to one another and subsequent fixing, and the axial forces from this synchronous toothing are used specifically to compensate for the gas forces caused by the pressure difference of the conveying medium between the outlet and inlet in conjunction with the known countershaft gear stage to increase the speed, wherein for very high final pressure values (depending on the machine dimensions, for example for more than 12 bar) to reduce the axial bearing load caused by the high pressure differences, a Pressure difference compensation stage is realized at each outlet-side carrier shaft end.

According to the gearing law, the distance between the rotation axes of the two spindle rotors in the z-axis direction is the distance between the rotation axes of the two spindle rotors for each conveyor thread gearing plane, whereby this distance between the axes of rotation as a(z) lies according to the invention between the largest value a.0 at the inlet and the smallest value a.L at the outlet. The resulting gearing with crossing, i.e. "skew" rotation axes is known to no longer belong to the rolling gears but to the helical gears.

The “wrap angle” on the spindle rotor is the sum of all twist angles along the spindle rotor axis between the individual face cut profile contours that result as the z-axis value progresses in the longitudinal direction of the rotor. If the profile face cut at a z-position z _i is compared with the profile face cut at the adjacent position z _i+1 , both face cuts are twisted relative to each other by an angle phii that is known according to the selected z(phi) function for exactly this step from z _i to z _i+1 . The sum of all twist angles for the face cuts along the spindle rotor axis results in the wrap angle, which here refers to the 2-tooth rotor and is referred to as PHI.2 for short. For the 3-tooth rotor, this twist angle must be adjusted by the transmission ratio as a factor in accordance with the gearing law and is therefore inevitably fixed if the spindle rotor length is the same. The wrap angle is the decisive measure for the number of stages.

The "number of stages" is the number of closed working chambers on the spindle rotor pair between the rotor inlet side and the rotor outlet side. The aim is to achieve a number of stages that is as whole as possible based on the rotor length and the selected z(phi) function with the total wrap angle PHI.2. The PHI.2 value is preferably rounded up to at least the next 10th place, e.g. from 2411° to 2420°.

A "working chamber" is the tooth gap volume closed for the rotor pair, which is limited by the surrounding compressor housing and the spindle rotor profile gap flanks between the profile contour engagements defined according to the gearing law, whereby these engaging rotor pair profile flanks are considered to be touching, i.e. close with a distance of zero. In practice, however, the engaging rotor pair profile flanks have a certain, albeit minimal, distance, which results in an internal leakage backflow. The "working chamber volume on the inlet side" is the volume of the first closed working chamber on the suction side, and the "working chamber volume on the outlet side" is the volume of the last closed working chamber before the conveying gas outlet. The quotient of these two volumes represents the "internal compression ratio". Values above 3 can be conveniently defined as "higher internal compression ratios". The volume of a working chamber is calculated from the relevant working space cross-sectional area multiplied by the step-by-step working chamber extension in the longitudinal axis direction of the rotor, defined by the spindle pitch.

In particular, a “frontal cut” is defined as any cut perpendicular to the spindle rotor axis of rotation, which is preferably defined as the z-axis, so that the frontal cut lies in the xy-plane of the rectangular Cartesian coordinate system. In the case of crossing axes, the frontal cuts planes of the two spindle rotors the axis crossing angle to each other.

The "pitch" in the spindle rotor conveyor thread is the length progression in the rotor's longitudinal axis direction after exactly one revolution (i.e. 360 degrees) of the wrap angle. Thanks to modern production machines, different pitches in the rotor's longitudinal axis direction can now be realized for each rotor in order to thermodynamically adjust the volume curve as the course of the working chamber volumes.

The "external synchronization" of the two spindle rotors is necessary because the rotor pair in the compressor working chamber works without operating fluid, i.e. is operated in a "dry compression" manner, and because of the high speeds, rotates in opposite directions without contact and with the smallest possible flank distance from each other. In order to ensure that the rotor pair can operate without contact at all times, the two spindle rotors must be driven continuously with a high degree of angle accuracy, in the range of a few minutes of arc, which is known to be achieved via external synchronization. The most common form of external synchronization is by far the use of directly meshing gears. However, there is also the option of electronic rotor pair synchronization, for example, in which each rotor is driven electronically by its own motor with a precise angle of rotation.

The "inlet area" can be described by the wrap angle area with which the first closed working chamber is created on the inlet side by increasing the angle of rotation. This occurs with the 2:3 spindle rotor pair starting from the inlet face cut side after 720 angular degrees plus the tip arc central angle ga.KB2 on the inlet side of the 2-tooth spindle rotor.

In the case of atmospheric intake, “overpressure” is defined as the final pressure during operation as absolute pressure values of at least 2 bar. Usually, 8 bar to 15 bar are used, but with a high number of stages, pressure values of more than 25 bar can also be achieved. In the case of non-atmospheric intake, these values shift accordingly.

Final pressures are considered to be “vacuum” or negative pressure as absolute 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 compared to the outlet pressure, which is in the atmospheric pressure range.

• 1 zeigt beispielhaft für die vorliegende Erfindung eine Schnittdarstellung durch das Spindelrotorpaar mit auseinanderstrebendem Achsabstand, wobei diese Darstellung vereinfachend ist, weil die Drehachsen (6) und (7) nicht in einer Ebene liegen müssen. Zugleich wird in dieser Zeichnung deutlich erkennbar, dass bei sich schneidender Ausrichtung der Drehachsen der Platz bei der konstruktiven Umsetzung für die Auslass-seitige Lagerung (24) knapp wird, weil diese Auslass-Lager von Rotor und Gegenrotor einander zu nahe kommen, wobei für die Trägerwellen eine beidseitige Lagerung bei zugleich möglichst geringem Lagerabstand je Rotor wegen der erforderlich hohen Rotordrehzahlen und der damit verbundenen biegekritischen Drehzahl als unabdingbar anzusehen ist. Die Positionierung der nötigen Synchro.-Räder (37) ist Auslass-seitig günstiger, weil die Baugröße der Zahnräder andernfalls Einlass-seitig zu groß wäre, was nicht nur wegen der enormen Umfangsgeschwindigkeiten aufwändig wäre. Damit ist wegen der Auslass-seitigen konstruktiven Platz-Bedingungen bei dieser Darstellung eine kreuzende Lage der Drehachsen günstiger, wobei deren Achsversetzung in die Nähe der beiden Spindelrotor-Auslass-Stirnseiten (36) sich als vorteilhaft zeigt.

The present invention is further explained in the following illustrations:

• 1 shows, as an example of the present invention, a sectional view through the spindle rotor pair with diverging center distances, whereby this view is simplified because the axes of rotation (6) and (7) do not have to lie in the same plane. At the same time, it is clearly visible in this drawing that when the axes of rotation intersect, space in the structural implementation for the exhaust-side bearing (24) becomes tight because these exhaust bearings of the rotor and counter rotor come too close to one another, whereby for the carrier shafts, bearings on both sides with the smallest possible bearing distance per rotor is essential due to the necessary high rotor speeds and the associated bending-critical speed. The positioning of the necessary synchronizer wheels (37) is more favorable on the exhaust side because the size of the gear wheels on the inlet side would otherwise be too large, which would be complex not only due to the enormous peripheral speeds. Therefore, due to the structural space conditions on the outlet side, a crossing position of the rotation axes is more favorable in this representation, whereby their axis offset in the vicinity of the two spindle rotor outlet front sides (36) proves to be advantageous.

Furthermore, this illustration clearly shows that in the conveyor thread area (16 and 17) the rotor outer diameter and thus also the conveyor thread profile height from h.0 to h.L follow the variable center distance a(z), which can be directly understood from the conditions mentioned above when selecting the head radius values. In this way, the desired high internal compression ratio is achieved in a particularly advantageous manner compared to the state of the art, with a higher number of stages and associated larger working chamber surfaces for heat dissipation during compression in order to improve the efficiency of compression via a lower polytropic exponent in accordance with the laws of thermodynamics. The conditions mentioned when determining the conveyor thread profile flanks also reduce internal leakage through blowhole-free compression.

In addition, the flattening (21) according to the invention on the outer diameter of the 2-tooth spindle rotor improves the suction capacity of the spindle compressor and no longer has to be achieved by changing the conveyor thread pitch as was previously the case. This increases the efficiency improvement.

The cooling fluid supply (31) takes place at each inlet-side end of the support shafts (4 and 5) in their central supply bore, so that a Operation-dependent pressure difference load in the oil reservoir and thus also for the drive shaft feedthrough is advantageously avoided.

2 shows, as an example, the two rotor axes of rotation (6, 7) in a crossing (also called "skew") design with the axis offset aV (12) as the shortest distance between the two axes of rotation (and thus perpendicular to both axes) at the axis crossing angle γ _Σ (11). Advantageously (as shown here), the outlet-side axis distance aL (1.b) and the axis offset aV (12) are the same and at the same time the starting point on the respective axis of rotation for each running parameter z.2 and z.3 to describe the gearing conditions over the respective frontal cutting planes. The conveyor thread for the spindle rotor pair is ultimately just a gearing according to the gearing law, namely that at every flank profile contact point the speed components perpendicular to the flank surface of each rotor must be the same. Otherwise, the two flank surfaces would lift off or penetrate each other at the contact point. With crossing axes of rotation as shown in the present illustration, it can be seen that the previous rolling gear for the spindle rotor pair conveyor thread now becomes a helical gear with the general center distance ä(z) and the instantaneous axis (10) [cf. Technical book by Karlheinz Roth: “Gear technology, special involute gears”, ISBN 3-540-64236-6, Springer-Verlag 1998 and dissertation by Tsai, S.-J. “Unified system of involute gears”, TU-Braunschweig 1997 ]

3 shows an example of the present invention for the profile design, whereby this representation is simplified in that the two frontal cutting planes, which are actually inclined to one another, are shown in a common plane for each rotor. Nevertheless, the essential features for generating the profile flanks (38 and 39) are recognizable: The head profile end points E.2a, E.2b, E.2c and E.2d, defined for each z-position via the selected angle be.KB2(z) and via the also selected tip circle radius r.2KK(z), generate the counter profile flank (39) of the 3-tooth spindle rotor point by point during their rotational movement with ω.2 around their axis of rotation (6) for the 2-tooth spindle rotor, which rotates around its axis of rotation (7) with ω.3, whereby both axes of rotation (6, 7) according to 2 by the axis crossing angle γ _Σ (11) and (if so selected) axis offset aV (12) relative to each other in space. The rotational speeds ω.2 and ω.3 are of course defined according to the transmission ratio by the number of rotor teeth. Likewise, with the selected tip circle radius r.3KK(z), the counter profile flank (38) for the 2-tooth spindle rotor is determined point by point via the tip profile end points E.3a, E.3b, E.3c, E.3d, E.3e and E.3f, whereby the blow hole freedom is achieved, particularly in the area of large working chamber volume changes, by the fact that the course of the central angle in polar coordinate representation does not undergo a sign reversal for the conveyor thread profile flank (38) determined in this way. If this condition of the monotonic central angle progression is violated, the selected value for the tip circle radius r.3KK(z) must be reduced at this z-position until the central angle progression is monotonic.

When mathematically determining the respective counter profile flank, the "principle of kinematic reversal" is cleverly used, whereby the spindle rotor with the generating head profile end points rotates around its stationary counter rotor in space according to the axis crossing angle and (if selected) axis offset from each other, so that for this observation the actually fixed housing axis distance rotates. The stationary observer then sees the respective counter profile being created for "his" stationary rotor.

In 4 is detailed to 1 the inlet-side cooling fluid supply (13) can be seen via the housing-fixed supply pipe (40), from which the cooling fluid (usually oil) enters 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), is forced to flow via the transfer bores (42) according to 1 to the inner cone of the respective spindle rotor. There, due to centrifugal forces, it will reach the inlet-side rotor end in accordance with the flow direction (14) and collect there in a rotor-fixed and sealed collecting channel (28) as a rotating fluid ring (27) in order to be tapped off via frame-fixed pitot tubes (29). The cooling fluid reaches the oil storage chamber (not shown here) via the frame-fixed collecting chamber (30) via a discharge (31).

5 shows in a section perpendicular to the axis of rotation in a top view the operation of the cut-open pitot tube (29), the openings of which protrude into the rotating cooling fluid collecting ring (27). The kinetic energy of the rotating cooling fluid ring (27) drives the cooling fluid via the pitot tubes (29) into the fixed collecting chamber (30). As only one direction of rotation is permitted and clearly defined for the spindle rotor pair, the pitot tubes (29) must also only point in one direction. The size of the pitot tube opening cross-sections must be designed in such a way that a safe discharge of the cooling fluid quantity is always guaranteed, i.e. over-dimensioned as a precaution. In addition to the pitot tube opening cross-sections, the number of pitot tubes must also be increased, which for the purpose of For safe skimming, they can also be arranged axially offset from one another.

For simpler applications, this pitot tube solution can also be replaced by simple cooling fluid collection pads as simple (steel) wool packages (accumulations).

List of reference symbols:

1

Axial distance a(z) as non-parallel, i.e. diverging distance per gear plane in the z-axis direction between the rotation axes of the two spindle rotors within the limit values:

1.a

Inlet-side center distance a.0 as the largest conveyor thread toothing distance

1.b

Outlet-side center distance a.L as smallest conveyor thread toothing distance

2

2-tooth spindle rotor, referred to as “Rotor-2” for short, with the total conveyor thread length L.2z

3

3-tooth spindle rotor, referred to as “Rotor-3” for short, with the total conveyor thread length L.3z

4

Support shaft for the rotor-2, which is connected to the rotor-2 in a rotationally fixed manner (preferably pressed on) with a central cooling fluid supply hole

5

Support shaft for the rotor-3, which is connected to the rotor-3 in a rotationally fixed manner (preferably pressed on) with a central cooling fluid supply hole

6

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

7

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

8th

Rotor interior fluid cooling for rotor-2 according to PCT document WO 00 / 12899 A1

9

Rotor interior fluid cooling for rotor-3 according to PCT document WO 00 / 12899 A1

10

Instantaneous axis of the spindle rotor pair gearing with crossing rotation axes

11

Axis crossing angle γ _Σ

12

Axis offset a _V

13

Cooling fluid supply

14

Cooling fluid flow direction in the cooling cone of each rotor

15

Flow direction of the pumped medium

16

Conveying thread of rotor-2

17

Conveying thread of rotor-3

18

Conveyor thread engagement area for the spindle rotor pair

19

Head cone inclination angle γ _3K.L on rotor-3

20

Head cone inclination angle γ _2K.L on rotor-2

21

Inlet-side flattening on the outer tip circle diameter at the inclination angle γ _2K.0 for the length L.2zyl on rotor-2

22

Reference position surface optionally designed as:

22.a

Reference position surface on the inlet face of Rotor-2 or

22.b

Reference position surface on the inlet face of Rotor-3

23

Inlet-side rotor bearing

24

Exhaust-side rotor bearing

25

Shaft seal for the compressor working chamber

26

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

27

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

28

Pitot tube collecting trough, non-rotatably rotating with each spindle rotor body

29

Pitot tube, fixed and housing-fixed with discharge cavities

30

Cooling fluid collection chamber, fixed and housing-fixed

31

Cooling fluid discharge

32

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

34

Inlet space for the conveying medium

35

Outlet space for the pumped medium

36

Spindle rotor outlet face

37

Cone gear pair (also called “beveloid”) for synchronization for the spindle rotor pair

38

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

39

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

40

Housing/fixed cooling fluid supply pipe

41

Carrier shaft fixed overflow bushing

42

Cooling fluid transfer holes

Claims (12)

Spindle compressor as a 2-shaft rotary displacement machine operating in the working chamber without operating fluid for conveying and compressing gaseous conveying media for applications in a vacuum and for applications in overpressure for pressure ratios between inlet and outlet pressure greater than 3 with a spindle rotor pair driven in opposite directions by an external synchronization (37) located outside the compressor working chamber in a rotationally accurate manner in a surrounding compressor housing with an inlet (34) and an outlet (35) for the conveying medium, characterized in that the rotor pair consisting of a 2-tooth spindle rotor (2) and a 3-tooth spindle rotor (3) engaging in toothing has a non-parallel, i.e. diverging, axial distance such that on the conveying gas inlet side (34) the axial distance a.0 (1.a) between the two spindle rotors is at least 20%, better 50% and for higher pressure ratios even more than twice as large as the outlet side. Axle distance aL (1.b), wherein the wrap angle in relation to the 2-tooth spindle rotor is at least 2500 angular degrees, but preferably over 3600 angular degrees, advantageously even more than 4500 angular degrees or even better over 5400 angular degrees and for particularly high pressure differences even over 6300 angular degrees, and in addition, on the spindle rotor pair, the inlet-side slope is less than 20%, for some applications still 60% but at most 120% greater than the outlet-side slope, wherein furthermore each spindle rotor has a rotor internal fluid cooling (8, 9) with the feature that the flow direction (15) of the conveying medium is opposite to the cooling cone flow direction (14) of the cooling fluid, which is fed (13) in the rotor inlet area into the respective central cooling fluid feed hole in each carrier shaft (4, 5) and in each case on its own carrier shaft The non-rotatably mounted 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 rotation axes (6, 7) of the 2-tooth (2) and the 3-tooth (3) spindle rotor are designed as intersecting with an axis crossing angle γ _Σ (11) which lies between 2 and 40 angular degrees, and without axis offset a _V for those designs in which there is still sufficient space on the outlet side of the respective compressor machine, in particular for the rotor bearing. Spindle compressor according to Claim 1 , characterized in that the diverging rotor rotation axes (6, 7) of the 2-tooth (2) and the 3-tooth (3) spindle rotor are designed as crossing with an axis crossing angle γ _Σ (11) between 2 and 40 angular degrees and with an axis offset a _V (12) which is greater than 30 mm and is positioned within ±20% of the total conveyor thread length L.2z or L.3z to the front rotor outlet (36), but preferably coincides with the front cutting planes of both spindle rotor outlet front sides (36). Spindle compressor according to one of the preceding claims, characterized in that on the inlet side of the 2-tooth spindle rotor (2) a flattening (21) on the outer tip circle diameter over a length L.2zyl, which corresponds at least to the average starting profile pitch length, is carried out in such a way that the inlet-side head cone inclination angle γ _2K.0 is preferably zero and thus cylindrical, but in any case is smaller than the head cone inclination angle γ _2K.L (20), which defines the reduction of the rotor outer diameter in the direction of the outlet side for at least 66% of the rotor profile length L _2z . Spindle compressor according to one of the preceding claims, characterized in that either on the inlet face of the 2-tooth spindle rotor (2) there is a beveled reference position surface (22.a) or on the inlet face of the 3-tooth spindle rotor (3) there is a beveled reference position surface (22.b) made of which is inclined by the axis crossing angle γ _Σ (11) for the correct positioning of the tooth planes in the rotor longitudinal axis direction of both spindle rotors to each other. Spindle compressor according to one of the preceding claims, characterized in that the cooling fluid supply (13) takes place at each inlet-side carrier shaft end in its central supply bore and a particularly high cooling fluid flow rate is set such that the temperature increase in the cooling fluid due to the heat dissipation from each spindle rotor during compression of the conveying medium is less than 6 degrees, better still below 4 degrees and ideally even below 3 degrees. Spindle compressor according to Claim 6 , characterized in that in the case of the rotor internal fluid cooling (8, 9) the cooling cone angle in the smaller diameter region is made at least 10%, better still 25% or ideally more than 50% steeper than in the cooling cone region subsequently flowed through by the cooling fluid. Spindle compressor according to one of the preceding claims, characterized in that the cooling fluid arriving on the inlet side during operation via rotor internal fluid cooling (8, 9) is collected at each spindle rotor (2, 3) in a spindle rotor-fixed pitot tube collecting channel (28) and is continuously tapped from there via housing-fixed pitot tubes (29). Spindle compressor according to one of the preceding claims, characterized in that the thermodynamic design is carried out via the spindle rotor pair geometry values in such a way that the specific heat exchanger characteristic value (because ultimately the present spindle compressor is both a compressor and a heat exchanger at the same time) has a heat transfer surface load of less than 50 kW per m ² and for larger machines (i.e. over 55 kW) even less than 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 flank play in the conical gear pair (37) necessary for the synchronization of both spindle rotors (2, 3) takes place by axial displacement of these conical gears relative to one another with subsequent fixing. Spindle compressor according to one of the preceding claims, characterized in that the design of the conveyor thread profile flanks (38 and 39) on both spindle rotors is such that in each frontal section plane within the rotor profile thread length L.3z on the 3-tooth spindle rotor (3) a tip circle radius r.3KK(z) with the respective tip 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 met: 0.42-a(z) ≤ r.3KK(z) ≤ 0.72-ä(z) for: 0 ≤ z ≤ L.3z where in the range z < (L.2z - L.2zyl) the upper limit preferably applies: r.3KK(z) ≤ 0.6- a (z) with ä(z) as the respective center distance to the corresponding pairing of the toothing planes of both rotors, and that in each face cutting plane within the rotor profile thread length L.2z on the 2-tooth spindle rotor (2) a tip circle radius r.2KK(z) with the respective tip 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- a (z) ≤ r.2KK(z) ≤ 0.52· a (z) for: 0 ≤ z ≤ L.2z with the additional condition: (r.3KK(z) + r.2KK(z)) > 1 ,05- a (z) as an intermeshing rotor pair. Spindle compressor according to Claim 4 and Claim 11 , characterized in that for at least 66% and, with higher demands on the compression capacity, even better more than 80% of the spindle rotor length (L.2z - L.2zyl) when represented by polar coordinates for the conveyor thread profile flank line (38) on the 2-tooth spindle rotor, the central angle at adjacent successive flank profile points runs monotonically without a change of sign, in that the head profile end points E.3a, E.3b, E.3c, E.3d, E.3e and E.3f of the 3-tooth spindle rotor selected by the tip circle radius r.3KK(z) generate the counter profile flank (38) on the 2-tooth spindle rotor point by point, which is preferably determined by kinematic reversal, whereby by reducing the value for the tip circle radius r.3KK(z) per z-gearing plane, this change of sign for the central angle is avoided, and in addition the value also selected for each gearing plane for the tip circle radius r.2KK(z) deviates only by ± 20% from the value for the tip circle radius r.3KK(z).

Images