U.S. patent application number 14/400134 was filed with the patent office on 2015-04-30 for spindle compressor.
The applicant listed for this patent is Ralf Steffens. Invention is credited to Ralf Steffens.
Application Number | 20150118093 14/400134 |
Document ID | / |
Family ID | 48446290 |
Filed Date | 2015-04-30 |
United States Patent
Application |
20150118093 |
Kind Code |
A1 |
Steffens; Ralf |
April 30, 2015 |
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.
Inventors: |
Steffens; Ralf; (Esslingen,
DE) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Steffens; Ralf |
Esslingen |
|
DE |
|
|
Family ID: |
48446290 |
Appl. No.: |
14/400134 |
Filed: |
May 7, 2013 |
PCT Filed: |
May 7, 2013 |
PCT NO: |
PCT/EP2013/059512 |
371 Date: |
November 10, 2014 |
Current U.S.
Class: |
418/91 ;
418/194 |
Current CPC
Class: |
F04C 18/14 20130101;
F04C 18/084 20130101; F04C 18/16 20130101; F04C 29/04 20130101 |
Class at
Publication: |
418/91 ;
418/194 |
International
Class: |
F04C 18/08 20060101
F04C018/08; F04C 29/04 20060101 F04C029/04; F04C 18/14 20060101
F04C018/14 |
Foreign Application Data
Date |
Code |
Application Number |
May 8, 2012 |
DE |
10 2012 009 103.6 |
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 OV.Pi of the working chamber bore holes and are smaller
than the width of the spindle rotor head .DELTA.m.Ki in the
respective transverse section.
Description
[0001] This application is a national phase application under 35
U.S.C. .sctn.371 of International Application Serial No.
PCT/EP2013/059512, filed on May 7, 2013, and claims the priority
under 35 U.S.C. .sctn.119 to German Patent Application No. 10 2012
009 103.6, filed on May 8, 2012, which are hereby expressly
incorporated by reference in their entirety for all purposes.
BACKGROUND OF THE INVENTION
[0002] 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.
[0003] 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.
[0004] 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.
[0005] 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
[0006] 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.
[0007] 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.
[0008] 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.
[0009] 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.
[0010] 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.
[0011] 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
[0012] The present disclosure is described in conjunction with the
appended figures:
[0013] FIG. 1 illustrates an exemplary sectional view of an
embodiment of a spindle rotor pair.
[0014] FIG. 2 illustrates an enlarged individual
transverse-sectional view of an embodiment of a spindle with a
compressor housing.
[0015] FIG. 3 illustrates profile contour designs of the enlarged
individual transverse-sectional view shown in FIG. 2.
[0016] FIG. 4 illustrates an exemplary sectional view of an entire
spindle compressor having two unequal taper angles.
[0017] FIG. 5 illustrates an exemplary transverse section of a
spindle rotor pair.
[0018] FIG. 6 illustrates the spindle rotor, shown in FIG. 4, in
greater detail.
[0019] FIG. 7 illustrates an example of provisional head/root line
configuration.
[0020] FIG. 8 illustrates an exemplary provisional head line in a
two-toothed spindle rotor.
[0021] FIG. 9 illustrates an exemplary provisional head line in a
three-toothed spindle rotor.
[0022] FIG. 10 illustrates an actual configuration of the exemplary
provisional headlines in both two-toothed and three-toothed spindle
rotors.
[0023] FIG. 11 illustrates an exemplary embodiment of a spindle
rotor to avoid energy wasting.
[0024] FIG. 12 illustrates an exemplary embodiment of a chamber
bore holes foro a spindle rotor.
[0025] 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
[0026] 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.
[0027] 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.sub.i is compared
with the profile's transverse section in the neighbouring position
z.sub.i+1, both transverse sections are twisted in relation to each
other by an angle phi.sub.i known for exactly this step from
z.sub.i to z.sub.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.
[0028] 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.degree. to 2420.degree..
[0029] 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.
[0030] 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.
[0031] "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.
[0032] 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.
[0033] 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.
[0034] 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.
[0035] 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.
[0036] 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 Claim2 21 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.
[0037] 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.
[0038] 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.
[0039] 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.
[0040] 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.
[0041] 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.
[0042] 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.
[0043] 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.
[0044] 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.
[0045] 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 .PI., 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.
[0046] 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 .PI..
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.
[0047] 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.
[0048] 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.sub.3.E of the 3z rotor head line 43.a with the 3z
pitch circle 7 defines a length of L.sub.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=0.5D.3K.inand after a monotonically falling
course at the outlet, the value of
R.3K(z=L.ges)=R.3K.out=0.5D.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.
[0049] 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.
[0050] 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.
[0051] 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.sub.2.c and
K.sub.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.
[0052] 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.sub.3.C and K.sub.3.F and K.sub.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.sub.3.B and K.sub.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.sub.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.
[0053] 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.
[0054] As is well known in practical compressor operations, a
difference must be made between the geometrical internal
compression ratio .PI..sub.Geo and the actual internal compression
ratio .PI..sub.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 .PI..sub.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 .PI..sub.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 .PI..sub.IST can be made
adaptable.
[0055] To ensure that the actual internal compression ratio
.PI..sub.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.
[0056] 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.
[0057] 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.
[0058] 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.
[0059] 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.
[0060] 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
OV.Pi of each input bore hole 60 or 61 is smaller than the width of
the head .DELTA.m.Ki. The distance as a .DELTA.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 .DELTA.u.3i is between the KB.i(z) head arc value and the
FB.i(z) gap arc value.
[0061] 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.
[0062] 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.sub.KRO values are permissible, since higher h.sub.KRO values and
thus a high suction capacity are advantageous for large cross
sections.
[0063] 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.sub.KRO values
(ideally zero) to minimize the internal leakage backflow.
[0064] 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%.
[0065] 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.
[0066] 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.
[0067] 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.
[0068] 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.
* * * * *