SPINDLE COMPRESSOR

Abstract

A spindle rotor pair of a spindle compressor has a two-toothed spindle rotor and a three-toothed spindle rotor meshing the other without contact. The wrap angle related to the two-toothed spindle rotor is at least 800 angular degrees. A range of at least 30 m/sec is achieved as the mean circumferential speed of the rotor head. In the transverse section, both spindle rotors have arc sectors and cycloid profile contour flanks. In the case of the two-toothed spindle rotor, they are primarily above its gear-tooth pitch circle and of convex design, however, in the case of the three-toothed spindle rotor, they are below its ear-tooth pitch circle and of concave, i.e., hollow, design. Preferably, the transverse sections of each spindle rotor are symmetrical in a way that in each transverse section, the centre of gravity of the profile surfaces comes to lie on the respective rotor pivot point.

Description

BACKGROUND OF THE INVENTION

Dry-running compressors are becoming increasingly important in industrial compressor technology. Thanks to increasing commitments under environmental protection regulations, and thanks to the rising cost of operation and disposal as well as higher demands for the purity of the conveying medium, the known wet-running compressors such as liquid ring compressors, rotary vane pumps and oil or water injected screw compressors are increasingly replaced by dry-running machines. These machines include dry-running screw compressors, claw pumps, diaphragm pumps, piston pumps, scroll machines and vacuum roots pumps. However, these machines have in common that they still do not meet today's expectations in terms of reliability and robustness as well as size and weight at a low price level and satisfactory compressor efficiency.

To improve this situation, the known dry-running spindle compressors are an alternative because as typical two-shaft displacement machines they can provide a high compression capacity simply by achieving the required multiple stages in an extremely unelaborate manner as so-called “pumping screws”, with several series-connected closed working chambers over the number of wraps per displacement rotor, but without requiring an operating fluid medium in the working chamber. Furthermore, the non-contact rolling-off of the two counter-rotating spindle rotors allows for a higher rotor speed such that—related to size —there is an increase in nominal suction capacity and delivery rate. Dry-running spindle compressors can be used for vacuum as well as positive pressure applications; their power consumption with positive pressure applications is naturally significantly higher because in the positive pressure range, with final pressures clearly above 2 bar (absolute), up to 15 bar and even higher, much greater pressure differences have to be overcome.

In the PCT patent document WO 00/12899 there is described a simple rotor cooling system for a dry-running spindle displacement machine where a conical rotor bore hole is provided in each rotor into which a coolant, preferably oil, is introduced to continuously remove some of the compression heat generated in the compression process. In the patent document PCT/EP2008/068364, in continuation of this approach, the coolant is conveyed with an internal coolant (oil) pump to cool the pump housing, creating a preferably common coolant cycle via a separate heat exchanger to remove the absorbed amount of heat from the compression process of the conveying medium and to remove the dissipation loss such that the clearance values between the rotor pair and the surrounding pump housing is maintained for all operating conditions. These patent documents advantageously effect heat dissipation via the heat balance of the relevant working chamber/core components during compression, thus considerably improving effectiveness and reliability. Nevertheless, compression performance as well as capacity can still be improved—and not only for the more sophisticated applications in dry-running displacement machines—because the losses caused by internal leakages among individual series-connected working chambers between the inlets and discharge outlets of the conveying gas are presently still too high. This situation has to be improved.

The object of the present invention is to significantly improve the effectiveness and compression efficiency of dry-running two-shaft rotary displacement machines for transporting and compressing gaseous conveying media for vacuum pressure and positive pressure applications.

SUMMARY OF THE INVENTION

According to the embodiments of the present invention, this object is achieved in that in a dry-running spindle compressor as a two-shaft displacement machine for vacuum pressure and positive pressure application, the rotor pair, driven true to the rotational angle in counter-rotating directions by a synchronization arrangement situated outside the compressor working chamber consists of a two-toothed spindle rotor and a meshing three-toothed spindle rotor with a wrap angle of at least 800 degrees, but preferably more than 1160 degrees, most advantageously more than 2600 degrees and for particularly high pressure differences even above 3500 degrees. This is mainly due to the fact that the greater the compression capacity the higher should be the wrap angle, whereby the high-speed spindle rotors are operated such that as a mean rotor head circumferential speed, a range of at least 30 m/sec, better 45 m/sec, but most advantageously above 60 m/sec or even better more than 80 m/sec is achieved.

Because the greater the circumferential speed, the greater is the degree of effectiveness of the spindle compressor, whereby both spindle rotors have cycloid profile contour flanks which in the two-toothed rotor are designed mainly above its gear-tooth pitch circle. In addition, they are of convex shape, i.e. raised in bulbous fashion, and in the three-toothed rotor they are designed mostly below its gear-tooth pitch circle, and they are of concave shape, i.e. hollow, whereby the transverse sections of each spindle rotor are preferably symmetrical such that in each transverse section the centre of gravity lies on the rotor's pivot point, whereby the working chamber volume, as the so-called inner compression ratio, is larger on the inlet side than on the outlet side. This is achieved when on the spindle rotor pair either the inlet-side transverse section has a larger working chamber cross section than that on the outlet side′ or the spindle pitch at the rotor pair decreases so much that the increase at the inlet is greater than at the discharge outlet, whereby for higher inner compression conditions, i.e. more than about 3 times, the reduction of the transverse section areas is combined with the pitch reduction.

The former is achieved in at least one, but preferably both spindle rotors in the rotor's longitudinal direction through a predetermined shortening of the root circle radii with a resulting increase of the engaging root circle radii. In the latter case, the changes in cross section in longitudinal rotor direction are preferably made such that the outer rotor diameters take on a conical shape with at least one constant right-angle bevel value per spindle rotor, whereby in the inlet region preferably a cylindrical region with a constant diameter value must be provided in each spindle rotor. In the inlet region, the profile contour flanks are preferably designed such that the profile contour flanks on the three-toothed spindle rotor are extended in length, preferably cycloid, also above its gear-tooth pitch circle, by which means—under the gear tooth system—the profile contour flanks on the two-toothed rotor must also be extended in length below its gear-tooth pitch circle. Also, preferably the spindle rotors are designed with an internal rotor fluid cooling arrangement for heat dissipation, and the compressor housing is also provided with fluid cooling for heat dissipation whereby the coolant for the rotor pair as well as for the compressor housing is used preferably in a common cooling circuit.

The spindle rotor design parameters such as the angular pitch of the head profile and the tip radius of each rotor are designed such that the mean rotor temperature of the two-toothed spindle rotor deviates by less than 25%, better yet by less than 10% from the mean rotor temperature of the three-toothed spindle rotor. This is achieved with the rotor parameter design when thermodynamically for each rotor the heat balance is established via the heat-absorbing surfaces on the gas side. The heat transfer in the material and the heat-dissipating coolant-contacting internal rotor cooling cone surfaces causes a mean rotor temperature in each rotor to deviate by less than 25% from the temperature of the surrounding compressor housing, and better yet by less than 10% from the highest mean temperature of the spindle rotor. Thereby, this mean housing temperature depends on the size of the coolant-contacting surfaces of the compressor housing and on the coolant flow parameters, especially with regard to the coolant mass flow and the coolant temperature level, and to achieve the desired level and better minimization of the temperature differences through adaptation to the mean spindle rotor temperatures.

Aside from the path to each cooling cone diameter and the regulation of mass flow regulation, there is an additional possibility to specifically influence heat conduction at each spindle rotor by optionally providing thread-like recesses profile-symmetrically in each boring hole of the internal rotor cooling cone. In this way, the recesses are below the respective spindle rotor teeth, which can be reliably produced in manufacturing by means of drilling. According to the embodiments of the present invention, it is also recommended that when the tip radii are selected, via the angular pitch of the head profile, the rotor's angular pitch elbow angle on the two-toothed spindle rotor is preferably designed such that this angular pitch elbow angle is greater than the aperture angle of each rotor's two-sided compressor housing.

Also according to the embodiments of the present invention, each spindle rotor is rigidly mounted on its own carrier shaft, whereby the functions of each carrier shaft include the supply of coolant, the external synchronization and the mounting. If synchronization takes place via the spur gears, the invention also recommends to design the outer diameter of the gear-side rotor mounting on the two-toothed spindle rotor is greater than the outside diameter of the synchronization gear of the two-toothed spindle rotor, such that the two-toothed spindle rotor as a rotational unit can be completely mounted and finally balanced. Manufacturing of the profile contour flanks, which differ in particular in the rotor's longitudinal direction, is done successively by turning individual point-sequence helix lines in the rotor's longitudinal direction on a lathe, which in combination finally result in the profile flanks. Based on experience, to reduce the weight and for better heat dissipation during compression, it is recommended that the spindle rotor pair is made from a material with high heat conduction, preferably an aluminum alloy, on a steel carrier shaft, whereby the compressor housing is also preferably an aluminum alloy.

BRIEF DESCRIPTION OF THE DRAWINGS

The present disclosure is described in conjunction with the appended figures:

FIG. 1 illustrates an exemplary sectional view of an embodiment of a spindle rotor pair.

FIG. 2 illustrates an enlarged individual transverse-sectional view of an embodiment of a spindle with a compressor housing.

FIG. 3 illustrates profile contour designs of the enlarged individual transverse-sectional view shown in FIG. 2.

FIG. 4 illustrates an exemplary sectional view of an entire spindle compressor having two unequal taper angles.

FIG. 5 illustrates an exemplary transverse section of a spindle rotor pair.

FIG. 6 illustrates the spindle rotor, shown in FIG. 4, in greater detail.

FIG. 7 illustrates an example of provisional head/root line configuration.

FIG. 8 illustrates an exemplary provisional head line in a two-toothed spindle rotor.

FIG. 9 illustrates an exemplary provisional head line in a three-toothed spindle rotor.

FIG. 10 illustrates an actual configuration of the exemplary provisional headlines in both two-toothed and three-toothed spindle rotors.

FIG. 11 illustrates an exemplary embodiment of a spindle rotor to avoid energy wasting.

FIG. 12 illustrates an exemplary embodiment of a chamber bore holes foro a spindle rotor.

In the appended figures, similar components and/or features may have the same reference label. Further, various components of the same type may be distinguished by following the reference label by a dash and a second label that distinguishes among the similar components. If only the first reference label is used in the specification, the description is applicable to any one the similar components having the same first reference label irrespective of the second reference label.

DETAILED DESCRIPTION OF THE INVENTION

The ensuing description provides preferred exemplary embodiment(s) only, and is not intended to limit the scope, applicability or configuration of the disclosure. Rather, the ensuing description of the preferred exemplary embodiment(s) will provide those skilled in the art with an enabling description for implementing a preferred exemplary embodiment(s) of the disclosure. It should be understood that various changes may be made in the function and arrangement of elements without departing from the spirit and scope of the invention as set forth in the appended claims.

In the following, some of technical terms used in the disclosure is first described. The “wrap angle” on the spindle rotor is defined as the sum of all torsional angles along the spindle rotor axis between the individual transverse-section profile contours which result altogether when the z-axis value in the rotor's longitudinal direction increases. Thus, when the profile's transverse section in a z position z_i is compared with the profile's transverse section in the neighbouring position z_i+1, both transverse sections are twisted in relation to each other by an angle phi_i known for exactly this step from z_i to z_i+1 according to the selected function of z(phi). The sum of all torsional angles for the transverse sections along the spindle rotor axis equals the wrap angle, which is here related to the to-toothed rotor and is abbreviated as PHI.2. For the three-toothed rotor, this torsional angle must be adapted to the transmission ratio—as a requirement under the gear tooth system—and it is thus a given factor for spindle rotors of equal length. The wrap angle is the determining measure for the number of stages.

The “stage number” is the number of closed working chambers in a spindle rotor pair between the rotor inlet side and the rotor outlet side. Preferably, a stage number consists of a whole number for the rotor length and the selected wrap angle PHI.2. Preferably, the PHI.2 value is rounded up at least to the next ten, i.e. for example from 2411° to 2420°.

A “working chamber” is the volume of the closed space between teeth of a rotor pair that is limited by the surrounding compressor housing and the spindle rotor profile gap flanks between the profile contour engagements defined in the law of gearing, whereby these engaging rotor pair profile flanks are regarded as contacting, i.e. close to zero clearance. However, in practice, the engaging rotor pair profile flanks do have a certain clearance, albeit as minimal as possible, which results in an interior leakage backflow. The “working chamber volume on the inlet side” is the volume of the first closed working chamber on the pumping side, and accordingly the “working chamber volume on the outlet side” is the volume of the last working chamber before the outlet for the conveying gas. The quotient of these two volumes is the “internal compression ratio”. For practical purposes, values above 3 can be determined as “higher interior compression ratios”. The volume of a working chamber is calculated from the respective working chamber cross-sectional area multiplied by the step-by-step extent of the working chamber in the longitudinal direction of the rotor axis defined by the spindle pitch.

Particularly for the spindle rotor pair, the “transverse section” is defined as each section through the spindle rotor pair vertical to the spindle rotor axis, which is preferably determined as z axis, such that the transverse section lies in the x-y plane of the rectangular Cartesian coordinate system. The spindle rotor pair axes are always parallel with a constant distance, which—as the so-called “axial distance”—represents an important parameter of the spindle compressor.

“External synchronization” of the two spindle rotors is required because the rotor pair works in the compressor's working chamber without operating fluid medium, i.e. it is a “dry-running”, and due to its high speed it runs without contact, with the rotors counter-rotating in relation to each other with the smallest possible flank clearance. To constantly ensure this non-contact operation of the rotor pair, the two spindle rotors must always be driven at a high rotational-angle accuracy within the range of a few angular minutes, which is known to work through external synchronization. The by far most common way to achieve external synchronization is via directly engaging spur gears whose pitch circles are just as large as the gear-tooth pitch circles of the respective spindle rotor pumping screws. However, there is also the possibility, for example, of electronic rotor pair synchronization, where each rotor is driven electronically by its own motor, true to the rotational angle.

The “inlet region” can be described by means of the wrap angle region, with which on the inlet side, the first closed working chamber is created by continuous torsional angles. In the spindle rotor pair according to the invention, this begins at the inlet transverse section side after 720 degrees plus the tip circle arc central angle ga.KB2 on the inlet side of the two-toothed spindle rotor.

In atmospheric suction, “positive pressure” means final pressures in operation as absolute pressure values of more than 25 bar; mostly 8 bar to 15 bar are common, but at a high number of stages, pressures of more than 25 bar can be reached. In non-atmospheric suction, these values shift accordingly. Final pressures as absolute pressures of under 50 mbar, better yet under 1 mbar, are regarded as vacuum or negative pressures, and with the respective number of stages even below 0.01 mbar absolute against outlet pressure in the atmospheric pressure range.

Die above named “desirable minimization of temperature differences” is based on the circumstance that the core components active in the compressor's working chamber, i.e. the rotor pair in the surrounding compressor housing, should work with as little clearance as possible in relation to each other to keep the internal backflow reasonably low. While the dry-running displacement machine is going through different operating processes, for example from a thermally usually cold state at start-up to a hotter state at a certain point of operation, the differences in thermal expansion should be kept as low as possible for the said core components to keep the backflow through the gap under control. However, since with the present geometry, thermal expansion is determined substantially by the component temperatures, the temperature differences between core components must be kept as low as possible.

The characteristic of claim 5 has the advantage that the blow hole quickly becomes smaller when compression begins. This results in a high suction volume. The characteristic of claim 11 leads to better heat dissipation. This is an advantage if the rotors are manufactured and machined by turning on a lathe. The characteristic of claim 12 leads to an improvement with regard to internal leakage; tightness is improved. The characteristic of claim 13 leads to an improvement in mounting the finished rotor unit. This is particularly important for the faster of the two rotors.

The characteristic of claim 14 provides a suitable manufacturing process for the rotors. It has been found that it is not feasible to produce the rotors with a form cutter. The characteristic of claim 16 leads to good heat dissipation. The characteristic of claim 17 leads to a resistance for leakages by disturbing the course of leakage flows. The characteristic of claim 18 leads to improved heat dissipation. The characteristic of claim 19 leads to a kind of elbow, which makes better access below the pitch circle line. Reference is made to FIGS. 7 and 9, where this is explained. The characteristics of claim 20 makes manufacturing easier. The characteristics of Claim221 and 22 create different bypasses. The characteristics of claims 21 and 22. This effectively helps to prevent over-compression or under-compression. According to a characteristic of claim 23, the diameter of the bypass borehole is not greater than the width of the head, which avoids short circuiting between working chambers.

In what follows, embodiments of the present invention are further described with reference to the appended figures. FIG. 1 shows an exemplary sectional view of the spindle rotor pair according to the present invention with a total of 4 transverse section views at different z positions in the direction of the rotor's longitudinal axis. In this embodiment, the reduction of the working chambers' cross-sectional areas 40 between inlet 18 and outlet 19 becomes just as clear as the declining spindle pitch m(z) in the direction of the rotor's longitudinal axis, whereby both measures are designed to achieve a higher internal compression ratio, in this case more than threefold. The term SE.z=0 marks the respective transverse section plane at the longitudinal-axis position z=0. The outer diameters of the spindle rotors change after the cylindrical inlet region 41 such that in this example a constant taper angle ga.2Ke or ga.3Ke is formed in each spindle rotor.

Also shown is the uncooled cylindrical inlet region 41 with profile extensions beyond the respective pitch circles, as well as the rigid connection 17a between the spindle rotor and the respective carrier shaft, whereby the second rigid connection 17b between spindle rotor and carrier shaft at the outlet-side transverse section at SE.z=L.ges can be seen together with the cooling fluid passages. The other transverse section drawings show the cooling arrangement for the interior rotor 8 and 9 and the cooling arrangement for the housing 12. Here, external synchronization is provided by the spur gears 14 and 15, whereby on the two-toothed rotor the outer diameter of the gear-side mounting 13 is greater than the outer diameter of the synchronization gear 14 to allow for the complete installation of this rotational unit of the two-toothed spindle rotor 2, and to allow balancing and only then the subsequent installation in the spindle compressor.

FIG. 2 shows an example of an enlarged individual transverse-sectional view of the present invention with the compressor housing 1, the rotor pair consisting of the two-toothed spindle rotor 2 and the three-toothed spindle rotor 3 with complete fluid cooling for the rotor pair and for the compressor housing 1 and also the working chamber cross-sectional areas 40 in this transverse section whose change in size leads to the next transverse section showing the internal compression by reducing the content of the working chamber volume.

In FIG. 3 the reference numbers for profile contour designs are shown in a transverse-sectional view. Thus, the pitch circle radius 6 of the two-toothed spindle rotor 2 is always 40% of axial distance a, and the pitch circle radius 7 of the three-toothed spindle rotor 3 is accordingly constant for all transverse sections at 60% of the a value. With the preferably symmetrical profile contour design (for better balancing quality) the cycloid profile contour 38 occurs a total of four times in the two-toothed spindle rotor, while the profile contour 39 occurs a total of six times in the three-toothed spindle rotor. By changing the tip radii R.2(z) and R.3(z) and the angular pitch of the head profile ga.K2(z), these profile contours change. Formation of the working chamber is controlled by running the four angular-pitch end points E.2a, E.2.b, E.2.c and E.2.d of the two-toothed spindle rotor 2 two-toothed spindle rotors 2 through the M.2-M.3 centre-to-centre connection line.

FIG. 4 shows an example of a sectional view of the present invention through the entire spindle compressor with two unequal taper angles ga.G2.ke1 and ga.G.2.ke2 in the two-toothed rotor 2 with rotor length sections L.zyl via L.2.ke1 and L.2.ke2 for a total length of L.ges between inlet 18 and outlet 19. The Die rotor pair synchronization via the spur gear pair 14 and 15 is shown as well as the internal rotor cooling arrangement 8 and 9 including the cooling fluid supply 22 and the fluid cooling arrangement 12 for the housing.

FIG. 5 shows an example of a transverse section of the present invention with the spindle rotor pair to explain the thermal balance to be established, for in the direction of the rotor's longitudinal axis the design parameters such as the angular pitch of the rotor head profile 34 and the tip radii 30 and 31 per rotor 2 and 3 must be implemented such that the mean rotor temperature of the two-toothed rotor 2 deviates by less than 25%, better yet less than 10% from the mean rotor temperature of the 3-toothed rotor. For that purpose, the temperature of each component is determined and compared and the working chamber regions AK.ij, AK.ji, AK.ii and AK.jj for each component are determined and compared with each other according to the indicated thermal-flow arrows via heat absorption on the conveying-gas side (24, 25 and 28), heat conduction in the material, and heat dissipation (26, 27 and 29) through the cooling fluid in a thermodynamic thermal-balance calculation. With iterative parameter adaptation, especially also with regard to the cooling fluid parameters such as coolant mass flow and coolant temperature level, the component temperature differences of the core components, i.e. for rotor 2 and rotor 3 and for the housing, can be minimized, such that the reliability of the spindle compressor is improved, because with minimal temperature differences, the danger of a thermal reduction of clearance is avoided.

FIG. 6 shows the representation of FIG. 4 in detail, namely the specific design of the spindle rotor's tip circle arcs as grooves 35 which are preferably turned on a lathe when the rotors are manufactured, as a helically circulating groove in the tip circle to increase the flow resistance of the housing-to-rotor head leakage flow, thus reducing internal leakage.

In FIG. 3 and FIG. 5, in the transverse section, the profile lines are shown which form the working chambers to transport the conveying medium, i.e. 36.F and 38 as well as 37.F and 39, for the spindle rotor pair in relation to the coolant-contacted heat-dissipating lines 26 and 27 as the straight length. In each spindle rotor, this relationship changes in the direction of the rotor's longitudinal axis such that when compression begins, the lines on the working chamber side are longer than those on the coolant side, and the closer each working chamber comes to the outlet, the larger do the lines on the coolant side become, while the lines on the working chamber side become shorter. According to the embodiment of the presen invention, the spindle rotors, at least positive-pressure applications, must be designed such that on the outlet side and thus on the compression end, the coolant-side lines are longer than the lines on the working chamber side.

According to the embodiment of the present invention, the working chamber volumes formed by the spindle rotor pair decrease between the inlet and the outlet. The quotient from the largest to the smallest working chamber volume is called the internal compression ratio Π, which initially is only a purely geometrically produced figure. As is well-known, any compressor performs at its ideal operating point when the “last” working chamber directly before it opens toward the outlet has through internal compression reached exactly the pressure that exists at the outlet.

However, in most vacuum pressure applications, the suction pressure changes due to the evacuation process, which means that a compromise must be found for the internal compression ratio Π. Since this value is relatively low for the majority of vacuum pressure applications (the value is often below 3), it is enough for most vacuum spindle compressors if according to the invention, the internal compression ratio is implemented only by increasing the pitch with constant radius values, such that for many vacuum pressure applications at least one spindle rotor is designed with a simply cylindrical diameter.

However, in most positive pressure applications, higher values must be aimed at for the internal compression ratio, which according to the invention is done by changing the pitch as well as geometrically by reducing the cross-sectional areas in the direction of the rotor's longitudinal axis. At the same time, when the working chambers are transported from the inlet to the outlet in the direction of the rotor's longitudinal axis, the internal backflow, the so-called “internal leakage” between the individual working chambers, must be minimized, while the aim must be for the working chambers on the inlet side to have the largest possible suction volume. For large suction volumes, the spindle rotor's outer diameter must be enlarged, such that the tip radius of the three-toothed spindle rotor becomes larger than the pitch circle of the three-toothed spindle rotor and is preferably designed cylindrically constant in the inlet region.

According to the embodiments of the present invention, the outer diameter of the three-toothed spindle rotor, as the course for the R.3K(z) value 31, in the direction of the rotor's longitudinal axis is designed such that, as shown in FIG. 7, the intersection K_3.E of the 3z rotor head line 43.a with the 3z pitch circle 7 defines a length of L_dicht.Knick 50.a, which is larger than half the overall length of the rotor profile 66. The 3z rotor head line 43.a at the inlet has the in some sections preferably cylindrically constant value of R.3K(z=0)=R.3K.in=0.5·D.3K.inand after a monotonically falling course at the outlet, the value of R.3K(z=L.ges)=R.3K.out=0.5·D.3K.out with R as the radius and D as the diameter. The two head lines 42 and 43 must be designed as continuously monotonically falling, whereby for practical purposes, the angles of incline for the respective head lines are selected.

FIG. 7 shows only the “provisional” head/root line configuration at the beginning of the design, for in terms of manufacturing technology, special adaptations are provided for optimal tool movement, to receive in the end the “actual” head/root line configuration according to FIG. 10 for the spindle rotor pair.

It is well known that in spindle rotor pairing with a constant axial distance, the 2z head line 42, through mirroring at the rotational axes, directly and unequivocally leads to the complete 3z root line 45, just as the 2x root line 44 unequivocally results from the 3z head line. As shown in FIGS. 8 and 9, it is enough to look only at the head configuration for each of the two spindle rotors to completely and unequivocally describe all rotor radius lines.

FIG. 8 shows in a two-toothed spindle rotor the provisional 2z head line 42.a of FIG. 7, simplified with the cylindrical inlet part of length L.2K.zyl and between the points K_2.c and K_2.E with the monotonically continuous configuration to the discharge outlet. According to the invention, for the actual 2z head line 42.b there is a curvature-constant transition whose length L.2b defines the tool movement in spindle rotor manufacturing according to the permissible load limits. With this actual 2z head line 42.b, the actual 3x root line 45.b is also completely and unequivocally defined.

FIG. 9 shows in the three-toothed spindle rotor the provisional 3z head line 43.a of FIG. 7, simplified with the cylindrical inlet part of length L.3K.zyl and between the points K_3.C and K_3.F and K_3.H with the monotonically continuous 3z head line configuration to the discharge outlet whereby the 3z pitch circle line 7 is cut such that the sealing surface L.dicht.Knick 50.a is at least half as long as the overall rotor profile length as L.ges 66. Experience has shown that for the actual 3z head line 43.b there is a curvature-constant transition between points K_3.B and K_3.G, preferably with a turning point whose length L.3b defines the tool movement in spindle rotor manufacturing according to the permissible load limits of the processing machine. Via intersection K_3.D with the 3z pitch circle line 7 the actual sealing surface L.dicht.IST 50.b results unequivocally, which is at least half as long as the overall rotor profile length as L.ges value 66. With this actual 3z head line 43.b, the actual 2z root line 44.b is also completely and unequivocally defined.

FIG. 10 finally shows the actual configurations of the 2z head line 42.b and the 3z head line 43.b which—via the overall length L.ges 66 unequivocally define the actual configurations of the engaging 2z root line 44.b and the 3z root line 45.b per axial distance, whereby the pumping screw 46 of the two-toothed spindle rotor is shown as a cross-hatched section, and the pumping screw of the three-toothed spindle rotor 47 is shown as an area with triangular hatching as well as the meshing pumping screw 48. Furthermore, the inner rotor cooling 8 and 9 for each spindle rotor is shown, as well as the respective pitch circle lines 6 and 7.

As is well known in practical compressor operations, a difference must be made between the geometrical internal compression ratio Π_Geo and the actual internal compression ratio Π_IST, for only with isothermal compression (i.e. without temperature change during compression) are both values identical. However, while in the spindle compressor the temperature of the conveying medium increases during compression, the actual internal compression ratio Π_Geo depends on temperature change, which, as is well known, must be calculated with the polytropic exponent. As mentioned above, it should be endeavoured for the ideal operation of a compressor that every “last” working chamber of a compressor directly prior to its opening toward the discharge outlet has reached exactly that pressure through internal compression which exists at the outlet, such that any energy-wasting “over” or “under” compression is avoided. However, while in the finished machine, the geometrical internal compression ratio Π_Geo is already determined by the factual design of the parts, and the polytropic exponent is subject to fluctuations due to the different heat dissipation depending on the application (for example already in a hot/cold environment), and since the operating final pressure will be variable, it would be advantageous if the actual internal compression ratio Π_IST can be made adaptable.

To ensure that the actual internal compression ratio Π_IST can be ideally adapted to the specific application conditions, it is also recommended according to the invention that in case of “over compression” (when pressure in the spindle rotor's working chamber already exceeds the operating pressure ahead of the discharge outlet), an over-compression flow of conveying gas 55, controlled by a control means 56 via additional input bore holes 60, is provided as a partial conveying-gas flow besides the main conveying-gas flow 52, and that in case of “under compression” (when pressure in the spindle rotor's working chamber ahead of the discharge outlet does not reach operating pressure) an under-compression conveying gas flow 57, controlled by a regulating means 58 is provided as a partial conveying-gas flow besides the main conveying-gas flow 62 after leaving the conveying-gas aftercooler is provided, such that in case of “under compression”, cooled conveying gas under operating pressure flows into the working chambers with insufficient pressure, whereby the pressure in the outlet chamber 19 is approximately the same as the operating pressure.

To explain further, it should be mentioned that—as is well known—“under compression” leads to isochoric surplus compression when the last working chamber volume must be pushed out against the higher outlet pressure without a change in volume, which of course is a disadvantage in terms of energy consumption.

While every working chamber extends over the two-toothed as well as the three-toothed spindle rotor, the input as well as the output of the conveying-gas equalization partial flow during over- or under-compression depends only on the z.Pi position as the transverse section plane, as FIG. 12 shows this in addition.

FIG. 11 shows an example of an embodiment with which the energy-wasting “over/under compression” can be avoided. During compression, due to the rotation of the spindle rotors, the working chambers come close to the outlet chamber 19, and due to a reduction of the working chamber volumes, pressure rises in the working chamber. While every working chamber passes the bore holes 60 and 61, it is found directly by how much the working chamber pressure deviates from the operating pressure, such that either the over-compression conveying gas flow 55 is triggered by the regulating means 56 or the under-compression conveying gas flow 57 is triggered by the regulating means 58, whereby the bore holes (54, 55 and 60, 61) can naturally be advantageously distributed on the circumference.

Furthermore, the drill holes 54 and 59 as well as 60 and 61 can of course be used in both flow directions, such that the two regulating means 56 and 58 can be combined in one regulating means which, depending on the pressure in the working chamber, conducts the conveying-gas partial flow either as an over-pressure conveying gas flow 55 to the conveying-gas aftercooler 53 or as an under-pressure conveying gas flow 57 to the conveying-gas aftercooler 53 into the working chamber. In an alternative embodiment, the regulating means 56 and 58 can also be designed as simple no-return valves.

FIG. 12 shows the working chamber bore holes 60 or 61 for a spindle rotor 60 or 61. While the spindle rotor heads 63 closely pass the working chamber bore holes 60 or 61 during rotation of the spindles and thus effect their permanent opening and closing, advantageously at least two input bore holes 60 or 61 should be provided per equalization conveying-gas partial flow 55 or 57 to avoid unpleasant gas pulsations by the equalization conveying-gas partial flows 55 or 57. In this transverse section, the diameter ØV.Pi of each input bore hole 60 or 61 is smaller than the width of the head Δm.Ki. The distance as a Δu.2i value for 2 input bore holes 60 or 61 must be smaller than the head arc length KB.i(z) and should preferably be about half as long as the known KB.i(z) value. In case of three input bore holes, the distance value Δu.3i is between the KB.i(z) head arc value and the FB.i(z) gap arc value.

The wrap angle related to the two-toothed spindle rotor is preferably more than 1160 degrees, favourably more than 1700 degrees and even more favourably more than 2600 degrees, and for especially high-compression requirements even more than 3500 degrees. Preferably, the mean rotor head's circumferential speed is in the range of at least 45 m/sec, favourably however above 60 m/sec and for an even greater effect more than 80 m/sec. In transverse section, both spindle rotors have circle arc sectors (36.K and 36.F, as well as 37.K and 37.F) and cycloid profile contour flanks 38 and 39. In the two-toothed spindle rotor 2 these are primarily above its gear-tooth pitch circle 6 and convex, i.e. bulbously raised. In the three-toothed spindle rotor 3 they are primarily below its gear-tooth pitch circle 7 and concave, i.e. hollow. In both cases, primarily means at least 80% of the profile depth, whereby the profile depth is the distance between the tip circle and the root circle of the two-toothed spindle rotor 2 or the three-toothed spindle rotor 3.

In the inlet region there are only minor conveying-gas pressure differences, and the greatest possible volume is to be pumped per rotation. This means that in the inlet region, higher h_KRÖ values are permissible, since higher h_KRÖ values and thus a high suction capacity are advantageous for large cross sections.

In the outlet region, the working chamber volumes are smaller by the so-called “internal compression ratio”, and there are great differences in pressure, such that the rotor pairing should be as tight as possible, i.e. with minimal h_KRÖ values (ideally zero) to minimize the internal leakage backflow.

A blowhole distance dimension is introduced between the housing intersection edge and the rotor pair engagement line. The value for this blowhole distance dimension is preferably at about 5 to 10% of the axial distance value, whereby the situation is as follows in longitudinal axial direction: in the inlet region, the blowhole distance dimension is preferably more than 5% of the axial distance value. Thus, the suction volume is increased only when the difference in pressure is moderate. In the outlet region, preferably this blowhole distance dimension is less than 5% of the axial distance value. Thus, the necessary compression capacity is achieved with an accordingly minimized interior leakage. Better than 5% is 3% and even more favourable is 2%.

Advantageously, on at least 50% of the compression length (seen in conveying direction toward the outlet) the blowhole distance measure is less than 5% of the axial distance value. Advantageously, the profile contour flanks of the two-toothed spindle rotor are completely above its pitch circle, and the profile contour flanks of the three-toothed rotor dare completely below its pitch circle.

The compression length is defined as the length in direction of the rotor's longitudinal axis (commonly Cartesian as z axis), where the size of the working chamber volumes decreases, which means that the so-called “interior compression” occurs as well as dissipation of compression heat via the rotor cone interior cooling. The compression length equals the major portion of the overall rotor length: only on the suction side is there the input length where the working chambers are formed and the suction volumes are generated. The engagement line is the fixed place of all engagement points of the two spindle rotors. The housing intersection edge is the line of all intersections of the two rotor tip circles in the compressor housing. There are always two housing intersection edges opposite each other.

Preferably each spindle rotor 2 and 3 is rigidly mounted via connection contacts 17, preferably as 17.a and 17.b on its own carrier shaft 4 and 5, preferably pressed on, and that the manufacturing or machining of the spindle rotor profile contours 36, 37, 38 and 39 is only done subsequently. Preferably the spindle rotor pair 2 and 3 consists of a material with high thermal conductivity, preferably an aluminum alloy, and that the compressor housing 1 is also made of an aluminum alloy. Preferably all tip circle arcs 36.K and 37.K in both spindle rotors 2 and 3 are provided with at least one groove 35.

While the principles of the disclosure have been described above in connection with specific apparatuses, it is to be clearly understood that this description is made only by way of example and not as limitation on the scope of the invention.

Claims

1-23. (canceled)

24. A spindle compressor operating in a working chamber without operating fluid medium as a two-shaft rotary displacement machine to convey and compress gaseous media for vacuum pressure and positive pressure applications, the spindle compressor comprises a spindle rotor pair driven true to the rotational angle and counter-rotating with an external synchronization arrangement outside the compressor's working chamber in a surrounding compressor housing with an inlet and a discharge outlet for the conveying medium, wherein the two spindle rotors are designed with a different number of teeth: one two-toothed spindle rotor and one three-toothed spindle rotor meshing the other without contact, with a wrap angle related to the two-toothed spindle rotor of at least 800 angular degrees, whereby the spindle rotors rotate at high speed such that a range of at least 30 msec is reached as the rotor head's mean circumferential speed, that both spindle rotors in the transverse section are provided with arc sectors and with cycloid profile contour flanks which in the two-toothed spindle rotor are primarily above gear-tooth pitch circle with a convex design, and that in the three-toothed spindle rotor they are primarily below its gear-tooth pitch circle with a concave, i.e. hollow design, and that the transverse sections of each spindle rotor are preferably of symmetrical design, such that in every transverse section the profile area's center of gravity comes to lie at the respective rotor's pivot point.

25. The spindle compressor according to claim 24, wherein the working chamber volume on the inlet side is larger than the working chamber volume on the discharge outlet side.

26. The spindle compressor according to claim 25, wherein the transverse section on the inlet side has a larger cross-sectional area than the transverse section of the outlet side, which is achieved in at least one, but preferably in both spindle rotors in the direction of the rotor's longitudinal axis by means of a specific preferably continuously monotonic shortening of the tip radii by more than 3% and at most 20% with the corresponding increase in the respectively engaging root circle radii.

27. The spindle compressor according to claim 25, wherein the spindle pitch m(z) of the rotor pair decreases in the direction of the rotor's longitudinal axis such that the spindle pitch at the inlet is at least 1.5 times and at most 4 times greater than the spindle pitch at the outlet.

28. The spindle compressor according to claim 24, wherein with the change of the outer rotor diameters, a conical outer shape results for each spindle rotor with at least one right-angle bevel value per spindle rotor, and that preferably in the inlet region a cylindrical region with a constant outer diameter of the rotor head is provided for each spindle rotor.

29. The spindle compressor according to claim 24, wherein in the Inlet region, the profile flanks are designed such that in the three-toothed spindle rotor, the profile contour flanks are also extended in length above its pitch circles, preferably cycloid, which means that according to the gear tooth system, the profile flanks in the two-toothed spindle rotor must also be extended in length below its pitch circles.

30. The spindle compressor according to claim 24, wherein the spindle rotors are both designed and operated with a conical interior rotor fluid cooling system via a coolant fluid.

31. The spindle compressor according to claim 24, the compressor housing is also provided with a fluid cooling system for heat dissipation, which is operated with an interior rotor fluid cooling system for the spindle rotors, preferably jointly in a cycle via a coolant fluid.

32. The spindle compressor according to claim 24, wherein in the direction of the rotor's longitudinal axis, the rotor design parameters such as the angular pitch of the rotor head profile and the tip radii per spindle rotor are designed such that the mean rotor temperature of the two-toothed spindle rotor deviates by less than 25%, better yet less than 10% from the mean rotor temperature of the three-toothed spindle rotor.

33. The spindle compressor according to claim 24, wherein the mean temperature of the surrounding compressor housing over the size of the coolant-contacted surfaces of the compressor housing and over the coolant flow parameters, especially pertaining to the coolant mass flow and the coolant temperature level deviates by less than 25%, better yet by less than 10% from the highest mean spindle rotor temperature.

34. The spindle compressor according to claim 24, wherein Thread-like recesses are provided profile-symmetrically in the respective internal rotor cooling cone bore holes such that these recesses are below the respective spindle rotor teeth.

35. The spindle compressor according to claim 24, wherein the rotor's tip circle central angle in the two-toothed spindle rotor in preferably every transverse section is greater than the respective compressor housing opening angle in Rotor.

36. The spindle compressor according to claim 24, wherein the outer diameter of the rotor mounting on the gear side in the two-toothed spindle rotor is configured larger than the outer diameter of the synchronization gear of the two-toothed spindle rotor.

37. The spindle compressor according to claim 24, wherein manufacturing of the different profile contours, in particular in the direction of the rotor's longitudinal axis, is done successively by turning the individual point-sequence helix lines on a lathe in the direction of the rotor's longitudinal axis, which in combination are then resulting in the outer profile contour flanks.

38. The spindle compressor according to claim 24, wherein preferably for positive-pressure applications the coolant-touching lines in the outlet-side transverse section for the spindle rotor pair are at least 5% and no more than 100% greater than the working-chamber lines on the conveying-medium side.

39. The spindle compressor according to claim 24, wherein for positive-pressure applications in the three-toothed spindle rotor, there is an intermediate range with greater decrease in tip radius values which, with values greater than the pitch circle radius of the three-toothed spindle rotor with a preferably cylindrical beginning at the inlet in the direction of the outlet chamber declines continuously monotonically within the first half of the total length of the spindle rotor's pumping screw.

40. The spindle compressor according to claim 24, wherein the actual rotor head lines have a plane and curvature-constant configuration.

41. The spindle compressor according to claim 24, wherein a regulating means R and additional bore holes and are provided, and in case of “over-compression”, i.e. when the pressure in the working chambers prior to opening at the outlet is greater than the pressure in the outlet chamber, an over-pressure conveying gas flow is conducted to the conveying-gas aftercooler.

42. The spindle compressor according to claim 24, wherein a regulating means and additional bore holes are provided, and in case of “under-compression”, i.e. when the pressure in the working chambers prior to opening at the outlet is smaller than the pressure in the outlet chamber, an under-pressure conveying gas flow, which preferably has already been cooled by the conveying gas aftercooler, is conducted via the regulating means and the at least one additional bore hole.

43. The spindle compressor according to claim 24, wherein the diameters ØV.Pi of the working chamber bore holes and are smaller than the width of the spindle rotor head Δm.Ki in the respective transverse section.