US20150118093A1 - Spindle compressor - Google Patents

Spindle compressor Download PDF

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US20150118093A1
US20150118093A1 US14/400,134 US201314400134A US2015118093A1 US 20150118093 A1 US20150118093 A1 US 20150118093A1 US 201314400134 A US201314400134 A US 201314400134A US 2015118093 A1 US2015118093 A1 US 2015118093A1
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Prior art keywords
rotor
spindle
toothed
compressor according
spindle rotor
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US14/400,134
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English (en)
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Ralf Steffens
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Individual
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
    • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C18/082—Details specially related to intermeshing engagement type pumps
    • F04C18/084—Toothed wheels
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
    • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C18/12—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type
    • F04C18/14—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
    • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C18/12—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type
    • F04C18/14—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons
    • F04C18/16—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons with helical teeth, e.g. chevron-shaped, screw type
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C29/00—Component parts, details or accessories of pumps or pumping installations, not provided for in groups F04C18/00 - F04C28/00
    • F04C29/04—Heating; Cooling; Heat insulation

Definitions

  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • both spindle rotors have cycloid profile contour flanks which in the two-toothed rotor are designed mainly above its gear-tooth pitch circle.
  • 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.
  • each spindle rotor is 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • each spindle rotor 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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 wrap angle is the determining measure for the number of stages.
  • stage number is the number of closed working chambers in a spindle rotor pair between the rotor inlet side and the rotor outlet side.
  • a stage number consists of a whole number for the rotor length and the selected wrap angle PHI.2.
  • 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.
  • 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.
  • 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.
  • 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 “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.
  • 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.
  • 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 Claim 2 21 and 22 create different bypasses. The characteristics of claims 21 and 22 . This effectively helps to prevent over-compression or under-compression.
  • the diameter of the bypass borehole is not greater than the width of the head, which avoids short circuiting between working chambers.
  • 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.
  • 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 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.
  • 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.
  • the reference numbers for profile contour designs are shown in a transverse-sectional view.
  • the pitch circle radius 6 of the two-toothed spindle rotor 2 is always 40% of axial distance a
  • 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.
  • the cycloid profile contour 38 occurs a total of four times in the two-toothed spindle rotor
  • the profile contour 39 occurs a total of six times in the three-toothed spindle rotor.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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 ⁇ , which initially is only a purely geometrically produced figure.
  • 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.
  • the suction pressure changes due to the evacuation process, which means that a compromise must be found for the internal compression ratio ⁇ . 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.
  • 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.
  • 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 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 is designed such that, as shown in FIG. 7 , the intersection K 3.E of the 3z rotor head line 43 . a with the
  • 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.
  • 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.
  • 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.
  • 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 .
  • 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 .
  • an over-compression flow of conveying gas 55 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
  • FIG. 11 shows an example of an embodiment with which the energy-wasting “over/under compression” can be avoided.
  • 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.
  • 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.
  • 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.
  • 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 ⁇ V.Pi of each input bore hole 60 or 61 is smaller than the width of the head ⁇ m.Ki.
  • the distance as a ⁇ u.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 ⁇ u.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.
  • 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.
  • 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 .
  • the two-toothed spindle rotor 2 these are primarily above its gear-tooth pitch circle 6 and convex, i.e. bulbously raised.
  • 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 .
  • 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 KR ⁇ 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.
  • the suction volume is increased only when the difference in pressure is moderate.
  • this blowhole distance dimension is less than 5% of the axial distance value.
  • the necessary compression capacity is achieved with an accordingly minimized interior leakage.
  • Better than 5% is 3% and even more favourable is 2%.
  • the blowhole distance measure is less than 5% of the axial distance value.
  • 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.
  • 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.
  • 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.
  • all tip circle arcs 36 .K and 37 .K in both spindle rotors 2 and 3 are provided with at least one groove 35 .

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Applications Or Details Of Rotary Compressors (AREA)
  • Rotary Pumps (AREA)
US14/400,134 2012-05-08 2013-05-07 Spindle compressor Abandoned US20150118093A1 (en)

Applications Claiming Priority (3)

Application Number Priority Date Filing Date Title
DE102012009103.6 2012-05-08
DE201210009103 DE102012009103A1 (de) 2012-05-08 2012-05-08 Spindelverdichter
PCT/EP2013/059512 WO2013167605A2 (de) 2012-05-08 2013-05-07 Spindelverdichter

Publications (1)

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US20150118093A1 true US20150118093A1 (en) 2015-04-30

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US14/400,134 Abandoned US20150118093A1 (en) 2012-05-08 2013-05-07 Spindle compressor

Country Status (9)

Country Link
US (1) US20150118093A1 (de)
EP (1) EP2847467B1 (de)
JP (1) JP2015519508A (de)
KR (1) KR20150028961A (de)
CN (1) CN104395609B (de)
DE (2) DE102012009103A1 (de)
ES (1) ES2613246T3 (de)
IN (1) IN2014MN02232A (de)
WO (1) WO2013167605A2 (de)

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US20170175744A1 (en) * 2015-12-21 2017-06-22 Ingersoll-Rand Company Compressor with ribbed cooling jacket
ITUA20164368A1 (it) * 2016-06-14 2017-12-14 Settima Meccanica S R L Soc A Socio Unico Pompa a due viti di tipo perfezionato
US11047387B2 (en) * 2017-09-27 2021-06-29 Johnson Controls Technology Company Rotor for a compressor
CN113685348A (zh) * 2020-05-18 2021-11-23 莱斯特里兹泵吸有限责任公司 螺杆泵
US11300123B2 (en) 2016-08-30 2022-04-12 Leybold Gmbh Screw vacuum pump without internal cooling
CN115003914A (zh) * 2020-01-21 2022-09-02 R718主轴公司 R718*压缩机的容积比
GB2607936A (en) * 2021-06-17 2022-12-21 Edwards Ltd Screw-type vacuum pump

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DE102013009040B4 (de) 2013-05-28 2024-04-11 Ralf Steffens Spindelkompressor mit hoher innerer Verdichtung
WO2015109048A1 (en) * 2014-01-15 2015-07-23 Eaton Corporation Method of optimizing supercharger performance
DE102014008288A1 (de) * 2014-06-03 2015-12-03 Ralf Steffens Spindelverdichter für Kompressionskältemaschinen
GB2537635A (en) * 2015-04-21 2016-10-26 Edwards Ltd Pump
CN105971877B (zh) * 2016-07-11 2017-11-14 中国石油大学(华东) 一种锥形螺杆转子及其双螺杆真空泵
US20200386228A1 (en) * 2017-01-17 2020-12-10 Ralf Steffens Steam compressor comprising a dry positive-displacement unit as a spindle compressor
DE202017005336U1 (de) * 2017-10-17 2019-01-21 Leybold Gmbh Schraubenrotor
WO2019083778A1 (en) * 2017-10-24 2019-05-02 Carrier Corporation LUBRICANT SUPPLY PASSING FOR COMPRESSOR BOTTOM
CN108955795B (zh) * 2018-08-10 2024-01-26 金迈思液压设备（天津）有限公司 一种双螺杆转子型线和基于该型线的双螺杆式流量计
ES2883556T3 (es) * 2018-09-11 2021-12-09 Common Spolka Akcyjna Medidor de flujo rotativo para medir el flujo de gas
DE102019205258A1 (de) * 2019-04-11 2020-10-15 Gardner Denver Nash Llc Schraubenverdichter
CN110879159B (zh) * 2019-12-27 2022-11-15 长安大学 一种高温高湿度气溶胶采样装置及采样方法
CN114483585A (zh) * 2022-03-01 2022-05-13 德斯兰压缩机(上海)有限公司 一种螺杆转子及使用该螺杆转子的空压机
JP2023140880A (ja) * 2022-03-23 2023-10-05 株式会社豊田自動織機 ルーツポンプ

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Publication number Priority date Publication date Assignee Title
US20170175744A1 (en) * 2015-12-21 2017-06-22 Ingersoll-Rand Company Compressor with ribbed cooling jacket
US10718334B2 (en) * 2015-12-21 2020-07-21 Ingersoll-Rand Industrial U.S., Inc. Compressor with ribbed cooling jacket
ITUA20164368A1 (it) * 2016-06-14 2017-12-14 Settima Meccanica S R L Soc A Socio Unico Pompa a due viti di tipo perfezionato
US11300123B2 (en) 2016-08-30 2022-04-12 Leybold Gmbh Screw vacuum pump without internal cooling
US11047387B2 (en) * 2017-09-27 2021-06-29 Johnson Controls Technology Company Rotor for a compressor
CN115003914A (zh) * 2020-01-21 2022-09-02 R718主轴公司 R718*压缩机的容积比
US12012961B2 (en) 2020-01-21 2024-06-18 R-718 Spindel Gbr Volume ratio for a R718* compressor
CN113685348A (zh) * 2020-05-18 2021-11-23 莱斯特里兹泵吸有限责任公司 螺杆泵
GB2607936A (en) * 2021-06-17 2022-12-21 Edwards Ltd Screw-type vacuum pump

Also Published As

Publication number Publication date
DE102013210817A1 (de) 2014-11-13
CN104395609B (zh) 2017-04-12
DE102013210817B4 (de) 2024-04-25
EP2847467B1 (de) 2016-11-30
CN104395609A (zh) 2015-03-04
JP2015519508A (ja) 2015-07-09
DE102012009103A1 (de) 2013-11-14
ES2613246T3 (es) 2017-05-23
WO2013167605A3 (de) 2014-01-03
IN2014MN02232A (de) 2015-07-24
KR20150028961A (ko) 2015-03-17
EP2847467A2 (de) 2015-03-18
WO2013167605A2 (de) 2013-11-14

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