U.S. patent application number 16/768017 was filed with the patent office on 2020-11-19 for compressor.
This patent application is currently assigned to Leybold GmbH. The applicant listed for this patent is Leybold GmbH. Invention is credited to Thomas DREIFERT, Bernhard Kliem, Roland MULLER, Kai Nadler.
Application Number | 20200362861 16/768017 |
Document ID | / |
Family ID | 1000005007528 |
Filed Date | 2020-11-19 |
United States Patent
Application |
20200362861 |
Kind Code |
A1 |
DREIFERT; Thomas ; et
al. |
November 19, 2020 |
COMPRESSOR
Abstract
A dry-compressing compressor comprises two screw rotors in a
housing defining a suction chamber. At a compressor inlet of the
compressor preferably atmospheric pressure prevails and at a
compressor outlet of the compressor preferably a pressure of more
than 2 bars (absolute) prevails. For each screw rotor at least one
displacement element including a helical recess defining a
plurality of windings is provided. The at least one displacement
element per screw rotor has a single-pass asymmetrical profile.
Inventors: |
DREIFERT; Thomas; (Kerpen,
DE) ; Nadler; Kai; (Bruhl, DE) ; Kliem;
Bernhard; (Munster, DE) ; MULLER; Roland;
(Koln, DE) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Leybold GmbH |
Koln |
|
DE |
|
|
Assignee: |
Leybold GmbH
Koln
DE
|
Family ID: |
1000005007528 |
Appl. No.: |
16/768017 |
Filed: |
January 4, 2019 |
PCT Filed: |
January 4, 2019 |
PCT NO: |
PCT/EP2019/050145 |
371 Date: |
May 28, 2020 |
Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F04C 2240/30 20130101;
F04C 18/16 20130101; F04C 2240/51 20130101; F04C 2220/12 20130101;
F04C 2240/20 20130101; F04C 25/02 20130101; F04C 29/04
20130101 |
International
Class: |
F04C 18/16 20060101
F04C018/16; F04C 29/04 20060101 F04C029/04 |
Foreign Application Data
Date |
Code |
Application Number |
Jan 12, 2018 |
DE |
20 2018 000178.3 |
Claims
1. A dry-compressing compressor comprising a housing defining a
suction chamber and having a compressor inlet where preferably
atmospheric pressure prevails and a compressor outlet where
preferably a pressure of at least 2 bars (absolute), preferably at
least 5 bars (absolute) prevails, two screw rotors arranged in the
suction chamber and each having at least one displacement element
including a helical recess for defining a plurality of windings,
wherein at least one displacement element per screw rotor has a
single-pass asymmetrical profile, the screw rotors have no internal
cooling of the rotors, and the housing has a mean heat flow density
of less than 80000 W/m.sup.2 in the area of the displacement
elements.
2. The dry-compressing compressor according to claim 1, wherein the
profiles are configured such that not blowhole is formed.
3. The dry-compressing compressor according to claim 1, wherein the
profiles of the at least one displacement element of each screw
rotor are configured a Quimby profile.
4. The dry-compressing compressor according to claim 1, wherein a
displacement element arranged near the outlet of the vacuum pump
has symmetrical profile.
5. The dry-compressing compressor according to claim 1, wherein at
least one displacement element per screw rotor and/or in the case
of a plurality of displacement elements per screw rotor said
displacement elements jointly comprise a number (n) of windings
which is larger than the ratio of outlet pressure (p.sub.out) to
inlet pressure (p.sub.in) such that n > Pout Pin ##EQU00006##
preferably ##EQU00006.2## n > Pout Pin + 4. ##EQU00006.3##
applies . ##EQU00006.4##
6. The dry-compressing compressor according to claim 1, wherein the
installed volume ratio between the delivery volume of the inlet
stage (V.sub.in) and the outlet stage (V.sub.out) is adapted to the
pressure ratio between inlet pressure (p.sub.in) and outlet
pressure (p.sub.out) such that the following applies: V i = Vin
Vout = ( Pout Pin ) 1 / k ##EQU00007## wherein n has a value of
k-0.3 to k+0.3 and k is the isotropic exponent of the gas mixture
to be delivered.
7. The dry-compressing compressor according to claim 1, wherein the
displacement elements include at least one area where the volume of
the inlet stage (V.sub.in) decreases to a precompression volume
(V.sub.VK) in a small rotation angle range, wherein the ratio
between inlet volume (V.sub.in) and the volume of the
precompression (V.sub.VK) is related to the internal volume ratio
(v.sub.i) of the compressor v VK = Vin Vout = ( v i ) 1 / j
##EQU00008## wherein j=2 to 5.
8. The dry-compressing compressor according to claim 7, wherein the
compression from the inlet volume (V.sub.in) to the precompression
volume (V.sub.VK) takes place during one and a half to three rotor
revolutions (windings).
9. The dry-compressing compressor according to claim 1, wherein at
least one displacement element per screw rotor and/or in the case
of a plurality of displacement elements per screw rotor said
displacement elements jointly have a ratio of length (L) to
diameter (D) for which the following applies L D > Pout 2 Pin -
2 ##EQU00009## and in particular ##EQU00009.2## L D > Pout 2 Pin
- 1 ##EQU00009.3##
10. The dry-compressing compressor according to claim 1, wherein
the pitch of the windings of the displacement elements varies,
preferably changes and particularly preferably decreases from the
compressor inlet to the compressor outlet.
11. The dry-compressing compressor according to claim 1, wherein
the head and the foot diameter of the rotor preferably continuously
changes, wherein the rotor is in particular of a conical
configuration.
12. The dry-compressing compressor according to claim 1, wherein
the pressure ratio Pout Pin ##EQU00010## between outlet and inlet
pressure is at least 5.
13. The dry-compressing compressor according to claim 1, wherein
two screw rotors with parallel axes are provided.
14. The dry-compressing compressor according to claim 1, wherein at
the compressor inlet in particular inside the housing a gas
collection chamber is provided.
15. The dry-compressing compressor according to claim 1, wherein at
the compressor outlet a gas collection chamber is provided in
particular inside the housing.
16. The dry-compressing compressor according to claim 1, wherein in
the housing roller bearings and preferably seals are arranged on
both sides of the two screw rotors.
17. The dry-compressing compressor according to claim 1, wherein
for synchronizing the two screw rotors a synchronization gear is
provided.
18. The dry-compressing compressor according to claim 1, wherein
the speed of the screw rotors is higher than 3 , 000 1 min ,
##EQU00011## 1 min , ##EQU00012## 1 min . ##EQU00013##
19. The dry-compressing compressor according to claim 1, wherein
the one displacement element is configured as a discharge-side
displacement element and for each screw rotor at least one further
displacement element is provided.
20. The dry-compressing compressor according to claim 1, wherein
between an upper surface of the displacement element and an inner
surface of the suction chamber a gap with a height of 0.03 mm to
0.2 mm is formed.
21. The dry-compressing compressor according to claim 1, wherein
the suction-side displacement elements have a constant pitch along
their overall length.
22. The dry-compressing compressor according to claim 1, wherein
each screw rotor comprises a rotor shaft supporting the at least
one displacement element.
23. The dry-compressing compressor according to claim 1, wherein
the displacement elements of a screw rotor are of an integral
configuration.
24. The dry-compressing compressor according to claim 1, wherein
the screw rotors and in particular the at least one displacement
element per screw rotor have a smaller expansion coefficient that
the housing, wherein the expansion coefficient of the housing is in
particular at least larger than that of the screw rotors and/or the
at least one displacement element.
25. The dry-compressing compressor according to claim 1, wherein
the screw rotors do not comprise any ducts through which in
particular a liquid coolant flows.
26. The dry-compressing compressor according to claim 1, wherein
the screw rotors are of a solid configuration.
27. The dry-compressing compressor according to claim 1, wherein a
temperature difference in the area of the discharge-side
displacement elements between the latter and the housing during
normal operation is smaller than 50 K.
28. The dry-compressing compressor according to claim 1, wherein
the distance between the area where 5 .degree. A) to 20% of the
outlet pressure prevails and the last winding of the discharge-side
displacement element is at least 20% to 30% of the rotor
length.
29. The dry-compressing compressor according to claim 1, wherein a
gap between the edges of at least one of the displacement elements
preferably has a gap height of 0.1 to 0.3 mm.
Description
BACKGROUND
1. Field of the Disclosure
[0001] The disclosure relates to a compressor, in particular a
screw compressor.
2. Discussion of the Background Art
[0002] For compressing gases, in particular for providing
compressed air, primarily oil-injected screw compressors are
nowadays used. They can usually perform a compression from 1 bar
(absolute) to 8.5 to 14 bars (absolute) in one compressor stage.
Here, the delivered intake volume flows range from 30 to 5000
m.sup.3/h. Such screw compressors comprise two counter-rotating
screw rotors. The screw rotors each comprise at least one helical
deepened portion such that a displacement element is formed. The
injection of oil into the suction chamber, where the two screw
rotors are arranged, serves for sealing the gaps between the rotors
and the housing and/or the inner wall of the suction chamber. By
providing oil, a sufficient tightness can be attained for realizing
high compression pressures of in particular up to 14 bars in one
compressor stage. In addition, the oil serves for lubricating the
rolling contacts between the two screw rotors. Therefore, a
synchronization gear for the two screw rotors is not required.
Further, the oil serves for discharging compression heat. Only in
this manner, a low temperature can be attained at a high
efficiency. Finally, the oil serves for damping mechanical noise.
An essential disadvantage of the use of oil is that the oil enters
the gas to be delivered. The oil must be removed from the
compressed air with the aid of multi-stage separators. As a result,
such compressors are complex and require a large installation
space. The use of oil-injected screw compressors in particular in
areas where a high purity of the compressed air is required, such
as in the field of pharmaceutical or food industry, is not possible
or possible only when using extremely complex multi-stage oil
separators.
[0003] For generating oil-free compressed air, it is known to use
dry-compressing screw compressors. Here, the two screw rotors are
arranged in a contactless manner and synchronized to each other via
an oil-lubricated gear. However, dry-compressing screw compressors
have the drawback that one compressor stage only allows for a
compression to 4 to 5 bars (absolute). The reason for this is in
particular that large leakages occur through the gaps between the
rotors and the housing. For reaching pressures of 9 bars
(absolute), for example, two-stage screw compressors must therefore
be used. Besides the two compressor stages, an intermediate cooling
of the compressed air is necessary, which results in complex
equipment comprising many components and requiring a large
installation space.
[0004] In addition, dry-compressing compressors configured as
so-called rotary tooth compressors are known. These, too, have the
drawback that they must be of a multi-stage configuration for
achieving high pressures of approximately 9 bars (absolute).
[0005] In addition, dry-compressing spindle compressors are known.
These comprise a plurality of closed working chambers arranged one
behind the other along a plurality of windings or loops of a
displacer. Theoretically, high compression pressures are said to be
achieved even with a one-stage design such that the spindle
compressors can substitute multi-stage screw compressors or rotary
tooth compressors. However, spindle compressors are so far not
commercially available such that there is no evidence that high
compression pressures can be reached with a one-stage design.
[0006] Spindle compressors are described in DE 10 2010 064 388, WO
2011/101064, DE 10 2012 202 712 and DE 10 2011 004 960, for
example.
[0007] It is an object of the disclosure to provide a
dry-compressing compressor with the aid of which high pressures of
in particular more than 5 bars (absolute) can be reached even with
a one-stage design.
SUMMARY
[0008] The dry-compressing compressor according to the disclosure
comprises a suction chamber defined by a housing. In the suction
chamber, two screw rotors engaging with each other are arranged.
These are counter-rotated with respect to each other for delivering
the gas. For this purpose, each screw compressor comprises at least
one displacement element having a helical recess for defining the
windings. In particular, for each screw rotor only one displacement
element can be provided which can be integrally formed with a rotor
shaft. Further, the housing comprises a compressor inlet where
preferably atmospheric pressure prevails. At a compressor outlet
preferably a pressure of more than 2 bars (absolute) prevails,
wherein it is particularly preferred that a pressure of more than 5
bars (absolute) prevails at the compressor outlet.
[0009] With the aid of the dry-compressing compressor according to
the disclosure high pressures can be reached with a one-stage
design since, according to the disclosure, the at least one
displacement element per screw rotor is of a single-pass
configuration and has an asymmetrical profile. According to a
particularly preferred embodiment, the asymmetrical profile is
configured such that no or merely a small blowhole occurs. Since no
continuous blowhole exists, in a profile which is preferably
asymmetrical according to the disclosure a short-circuit merely
occurs between two adjacent chambers. According to a particularly
preferred embodiment, the so-called Quimby profile is provided as
the asymmetrical profile. Asymmetrical profiles have two different
profile edges. Although the manufacture is complex due to the
required two separate operating steps, an extremely tight working
chamber can be realized.
[0010] Providing single-pass, possibly even symmetrical rotor
profiles offers the advantage that a larger tightness can be
achieved. In the case of profiles having two more passes of the
respective meshing displacement elements, connections across
several chambers are formed through the gaps such that the leakage
affects the delivered gas flow and the energy conversion
quality.
[0011] According to another preferred embodiment of the
dry-compressing compressor according to the disclosure, the number
of windings of the at least one displacement element or, in the
case of a plurality of displacement elements the sum of the
windings of the displacement elements of a screw rotor is larger
than the ratio of the pressure prevailing at the compressor outlet
to the pressure prevailing at the compressor inlet. The number of
windings thus results from
n > Pout Pin ##EQU00001##
[0012] wherein p.sub.out is the outlet pressure and p.sub.in is the
inlet pressure of the compressor. It is particularly preferred that
the number of windings or loops is calculated as follows
n > Pout Pin + 4. ##EQU00002##
[0013] Due to such a large number of windings or loops per screw
rotor, a continuous but relatively slow compression of the gas is
achieved. Thereby, it is possible to easily discharge heat produced
during the compression.
[0014] In addition, it is preferred that the installed volume ratio
of the dry-compressing screw compressor between the theoretical
delivery volume at the inlet stage (V.sub.in) and the theoretical
delivery volume at the outlet stage (V.sub.out) is adapted to the
pressure ratios at the inlet (p.sub.in) and the outlet (P.sub.out).
Here, p.sub.in and p.sub.out are defined as absolute pressures.
Preferred is a volume ratio V.sub.i of
V i = Vin Vout = ( Pout Pin ) 1 / k ##EQU00003##
[0015] wherein n has a value of k-0.3 to k+0.3 and preferably a
value between k-0.1 and k+0.1. Here, k is the isotropic exponent of
the gas mixture to be delivered.
[0016] According to another preferred embodiment, the displacement
elements comprise at least one area or portion where the chamber
volume Vin de- creases to an intermediate volume V.sub.VK.
[0017] According to another preferred or alternative embodiment,
the decrease of the delivery volume of the stages (working
chambers) from the large inlet volume (V.sub.in) to the smaller
outlet volume (V.sub.out) is divided into two areas. Here, it is
particularly preferred that in the first area the working chamber
closed towards the suction side is reduced to a specific volume
(volume of the precompression V.sub.VK) within a small rotation
angle range. Here it is preferred that
V.sub.VK=x V.sub.in
[0018] wherein x=0.1 to 0.5, in particular x=0.2 to 0.4, and
particularly preferred x=0.3. Due to the compression operation, the
precompression raises the temperature of the gas to a moderate
value of 150.degree. C.-200.degree. C. In the second area of the
compression, depending on the rotation angle, the working chamber
volume decreases to a considerably smaller extent than in the first
area. The rotation angle and thus the number of stages in the
second area is considerably larger than in the first area. Due to
the moderate temperature rise in the first area, the large housing
surface in the second area and the relatively long dwell time of
the gas in the second area due to the larger rotation angle, in the
second area another temperature rise of the gas due compression can
be avoided to a large extent by heat transport into the
housing.
[0019] The compression of the gas is selected such that the
produced compression heat can be easily discharged via the side
walls of the housing such that the temperature of the gas does not
rise or rises only to a small extent. Here, the maximum temperature
change is preferably less than 50.degree. C., and particularly
preferably less than 30.degree. C.
[0020] A particular advantage of the selected division of the
volume decrease is that a largely homogeneous temperature
distribution in the component is achieved. Thereby, thermal peak
loads and the associated large component expansions can be
avoided.
[0021] The ratio between the inlet volume (V.sub.in) and the volume
of the precompression (transition from the first to the second area
V.sub.VK) can be related to the internal volume ratio v, of the
compressor
v VK = Vin Vout = ( v i ) 1 / j ##EQU00004##
[0022] wherein j=2 to 5, in particular j=2.5 to 3.5, and
particularly preferred j=3.
[0023] According to a particularly preferred embodiment, the
precompression is performed in the described first area at 1.5 to 3
rotor revolutions (windings).
[0024] According to a preferred embodiment, the inventive large
number of windings in the second area can be realized by a single
displacement element for each rotor. However, it is also possible
to provide a corresponding number of windings in this
discharge-side area by two displacement elements, for example. By
providing an inventive large number of windings in this area,
where, according to the disclosure, preferably the medium to be
delivered is only compressed to a small extent per winding, it is
possible to do without internal cooling of the rotors. The reason
for this is in particular that due to the relatively small extent
of compression in this area the temperature increase of the
displacement element caused by compression is small. In addition,
in this area, due to the high density of the delivered medium, a
good heat dissipation from the displacement element into the
compressor housing via the medium is realized.
[0025] Preferably, the screw rotors and the at least one provided
displacement element are configured such that between an area where
5%-20% of the outlet pressure prevails, and the discharge-side
rotor end at least 6, in particular at least 8, and particularly
preferably at least 10 windings are provided. Here, the
discharge-side rotor end is the area of the compressor outlet.
Here, according to a preferred embodiment, the inventive large
number of windings in this area can be provided at a single
discharge-side displacement element provided per rotor. However, it
is also possible to provide a corresponding number of windings in
this discharge-side area at two displacement elements, for example.
By providing an inventive large number of windings in an area
where, according to the disclosure, the medium to be delivered is
only compressed to a relatively small extent, it is possible to do
without internal cooling of the rotors. The reason for this is in
particular that due to the relatively small extent of compression
in this area the temperature increase of the displacement element
caused by compression is smaller. In addition, in this area, due to
the high density of the delivered medium, good heat dissipation
from the displacement element into the compressor housing via the
medium is realized.
[0026] Moreover, due to the preferably large number of winding, a
large surface area for heat exchange to the housing is
available.
[0027] It is particularly preferred that the preferably at least 6,
in particular at least 8, and particularly preferably at least 10
windings are provided in a discharge-side displacement element.
[0028] In addition, for configuring screw rotors without internal
cooling according to the disclosure, it is preferred that the
discharge-side displacement element has a mean working pressure of
more than 2 bars (absolute) at at least 6, in particular at least
8, and particularly preferably at least 10 windings. In particular,
it is intended to realize a flat pressure gradient inside the
compressor. Therefore, the pressure should slowly rise across many
windings, in particular 6 to 10 windings.
[0029] According to the disclosure, it is thus preferably possible
to provide a cold gap having a height of 0.03 mm-0.2 mm, and in
particular 0.05 mm-0.1 mm between the surface of the at least one
displacement element and the inside of the section chamber, in
particular in the discharge-side area, even in the case of rotors
without internal cooling of the rotors or a housing of aluminum or
an aluminum alloy. Such a relatively large gap height can be
provided due to the inventive configuration of the particularly 6,
preferably 8, and particularly preferably 10 last windings, as
described above.
[0030] According to another preferred embodiment of the disclosure,
a relative long screw rotor relative to the diameter is selected.
In particular, the at least one displacement element per screw
rotor or, in the case of a plurality of displacement elements per
screw rotor, said plurality of displacement elements jointly have a
ratio of length L to diameter D where the following applies:
L D > Pout 2 Pin - 2 ##EQU00005## and in particular
##EQU00005.2## L D > Pout 2 Pin - 1 ##EQU00005.3##
[0031] By providing a long rotor having in particular many
chambers, the area usable for heat dissipation is increased. Due to
the resulting good heat exchange, the gas temperatures of the
compressed gas are relatively low. Providing many chambers further
offers the advantage that the pressure differences between adjacent
chambers are small and thus a large tightness can be achieved. Due
to such a reduction of the delivery volume per stage from the inlet
to the outlet side, the compression process becomes particularly
effective in terms of thermodynamics and the gas temperatures
remain relatively low. Here, it is particularly preferred that the
internal volume ratio is adapted to the ratio of outlet to inlet
pressure such that neither overcompression or compression by
re-aeration occurs.
[0032] The internal volume ratio can be attained by varying the
pitch of the windings. Preferably, the pitch of the windings is in
particular changed such that it is decreased and/or becomes steeper
from the compressor inlet to the compressor outlet. The pitch can
be changed continuously and/or stepwise.
[0033] In addition to or instead of the variation of the pitch, the
head or foot diameter of the profile can be changed continuously or
stepwise. Again, a continuous change of the head or foot diameter
is particularly preferred such that the rotor has a conical
configuration, in particular in combination with a continuous
change of the pitch.
[0034] According to a particularly preferred embodiment, the
pressure ratio between the outlet pressure and the inlet pressure
is at least 5. According to a particularly preferred embodiment,
the outlet pressure is at least 2 bars (absolute), in particular at
least 5 bars.
[0035] According to another particularly preferred embodiment, the
dry-compressing compressor comprises at the compressor inlet and/or
at the compressor outlet a respective gas collection chamber
preferably inside the compressor housing.
[0036] Moreover, it is preferred that the dry-compressing
compressor is a compressor having two shafts. The latter are
preferably supported on both sides such that narrow gaps can be
realized both between the displacement elements and between the
displacement elements and the inner wall of the suction chamber.
Preferably, the two rotor shafts are synchronized by a
synchronization gear preferably arranged outside the suction
chamber. The bearings can be lubricated by grease and/or oil.
Likewise, the gear can be lubricated by grease and/or oil. This is
possible since both the bearings and the synchronization gear are
preferably arranged outside the suction chamber and it is thus
avoided that the gas to be delivered is contaminated by oil.
[0037] Preferably, the housing is made from aluminum or an aluminum
alloy. Here, an aluminum alloy AlSi7Mg or AlMg07,5Si for the
housing is particularly preferred. In particular, the heat
expansion coefficient (expansion coefficient) of the material of
the screw rotors is smaller than the expansion coefficient of the
material of the housing. It is particularly preferred that the
expansion coefficient of the screw rotors is smaller than
12*10.sup.-61/K. This can be achieved with rotors made from iron or
steel materials.
[0038] The two screw rotors arranged in the suction chamber
comprise at least one displacement element having a helical recess.
The helical recesses define several windings. According to the
disclosure, the at least one displacement element is made from a
steel or iron alloy. It is thus particularly preferred that the
screw rotors including the displacement elements are made from a
steel or iron alloy. The housing is also made from a steel or iron
alloy or from aluminum or an aluminum alloy.
[0039] Preferably, each displacement element comprises at least one
helical recess having the same contour along its overall length.
Preferably, the contours are different for each displacement
element. Thus the individual displacement element preferably has a
constant pitch and an unvarying contour. Thereby, the manufacture
is considerably simplified such that the manufacturing costs can be
largely reduced.
[0040] For further improving the suction capacity, the contour of
the suction-side displacement element, that is in particular the
first displacement element as seen in the pumping direction, is
preferably of an asymmetrical configuration. Due to the
asymmetrical configuration of the contour and/or the profile, the
edges can be configured such that the leakage areas, the so-called
blowholes, can in particular completely disappear or have at least
a smaller cross-section. A particularly suitable asymmetrical
profile is the so-called "Quimby" profile. Although such a profile
is relatively difficult to produce, it offers the advantage that no
continuous blowhole exists. A short-circuit occurs only between two
adjacent chambers. Since this is an asymmetrical profile having
different profile edges, at least to working steps are required for
the production since the two edges have to be produced in two
different working steps due to their asymmetry.
[0041] The discharge-side displacement element, in particular the
last displacement element as seen in the pumping direction,
preferably has a symmetrical contour. The symmetrical contour in
particular offers the advantage that it is easier to produce. In
particular, the two edges having a symmetrical contour can be
produced in one working step using a rotating end milling cutter or
a rotating side milling cutter. Although such symmetrical profiles
have blowholes, these are continuous, i.e. do not only exist
between two adjacent chambers. The size of the blowhole decreases
with decreasing pitch. Thus, such symmetrical profiles can in
particular be provided for the discharge-side displacement element
since, according to a preferred embodiment, it has a smaller pitch
than the suction-side displacement element and preferably also than
the displacement element arranged between the suction-side and the
discharge-side displacement element. Even though the tightness of
such symmetrical profiles is somewhat smaller, they offer the
advantage that they are considerably easier to produce. In
particular, it is possible to produce the symmetrical profile in a
single working step and preferably using a simple end milling
cutter or side milling cutter. Thereby, the costs are considerably
reduced. A particularly suitable symmetrical profile is the
so-called "cycloidal profile".
[0042] Providing at least two such displacement elements results in
the corresponding screw compressor being capable of generating high
outlet pressures at a low power consumption. Further, the thermal
load is small. Arranging at least two displacement elements having
the configuration according to the disclosure with a constant pitch
and an unvarying contour in a compressor leads to essentially the
same results as with a compressor having a displacement element
with a varying pitch. At high installed volume ratios three or four
displacement elements per rotor can be provided.
[0043] According to a particularly preferred embodiment, for
increasing the attainable outlet pressure and/or for reducing the
power consumption and/or the thermal load, a discharge-side
displacement element, that is in particular the last displacement
element as seen in the pumping direction, comprises a large number
of windings. A large number of windings allows for accepting a
larger gap between the screw rotor and the housing at constant
performance. Here, the gap can have a cold-gap width of 0.05-0.3
mm. A large number of outlet windings or of windings of the
discharge-side displacement element is inexpensive to produce
since, according to the disclosure, this displacement element can
have a constant pitch and preferably also a symmetrical contour. On
the outlet side an asymmetrical profile can be used. This allows
for an easy and inexpensive production such that it is acceptable
to provide a larger number or windings. Preferably, this
discharge-side or last displacement element has more than 6, in
particular more than 8, and particularly preferably more than 10
windings. According to a particularly preferred embodiment, the use
of symmetrical profiles offers the advantage that both edges of the
profile can be simultaneously cut with a milling cutter. Here, the
milling cutter is supported by the respective opposite edge such
that deformation or distortion of the milling cutter during the
milling operation and resultant inaccuracies are avoided.
[0044] For further reducing the manufacturing costs it is
particularly preferred to integrally form the displacement elements
and the rotor shaft.
[0045] According to another preferred embodiment, the change of
pitch between adjacent displacement elements is inconsistent or
erratic. Possibly, the two displacement elements are arranged at a
distance to each other in the longitudinal direction such that
between two displacement elements a circular cylindrical chamber is
defined which serves as a tool outlet. This is in particular
advantageous for manufacturing integrally formed rotors since the
tool producing the helical line can be easily removed in this area.
If the displacement elements are manufactured separately from each
other and are then mounted to a shaft, a tool outlet, in particular
such an annularly cylindrical area need not be provided.
[0046] According to a preferred aspect of the disclosure, no tool
outlet is provided between two adjacent displacement elements at
the location where the pitch changes. In the area of the change of
pitch both edges preferably have a discontinuity or recess for
removing the tool. Such a discontinuity has no notable influence on
the compression capacity of the compressor since it is a localized
discontinuity or recess.
[0047] The compressor screw rotor according to the disclosure in
particular comprises a plurality of displacement elements. These
may have the same or a different diameter. Here, it is preferred
that the discharge-side displacement element has a smaller diameter
than the suction-side displacement element.
[0048] In the case of displacement elements manufactured separately
from the rotor shaft, the former are mounted to the shaft by
press-fitting. Here, it is preferred to provide elements, such as
locating pins, for defining the angular position of the
displacement elements relative to each other.
[0049] It is particularly preferred that the screw rotor is formed
integrally in particular from a steel or an iron alloy. The screw
rotor can also comprise a rotor shaft which supports the at least
one displacement element. In particular when providing a plurality
of displacement elements, this offers the advantage that these can
be manufactured separately from each other and then be connected to
the rotor shaft in particular by press-fitting or shrink-fitting.
Here, it is possible to provide fitting keys or the like for
defining the angular position of the individual displacement
elements.
[0050] If a plurality of displacement elements per screw rotor are
provided, it is possible to integrally form the displacement
elements.
[0051] According to the disclosure, it is preferred that the screw
rotors have no internal cooling. Hence it is particularly preferred
that the screw rotors do not have any ducts through which in
particular liquid coolant flows. However, the screw rotors can
comprise boreholes or ducts for the purpose of weight reduction,
for balancing or the like, for example. It is particularly
preferred that the screw rotors are of a solid configuration.
[0052] In addition, it is preferred that the housing has a mean
heat flow density in the area of the displacement elements of less
than 80,000 W/m.sup.2, preferably less than 60,000 W/m.sup.2, and
in particular less than 40,000 W/m.sup.2.
[0053] The mean heat flow density is the ratio of the compression
capacity to the wall surface of the compression area.
[0054] In the dry-compressing screw compressor according to the
disclosure a gas aftercooler and/or a condensate separator for
separating the condensate produced by compression and/or a silencer
may additionally be provided at the compressor outlet. Further, it
is possible to provide an inlet air filter or an inlet silencer at
the compressor inlet.
[0055] Particularly preferably, with the aid of the compressor
according to the disclosure a volumetric efficiency of at least 70
percent, preferably at least 85 percent for at least one operating
point of the compressor can be achieved. A decisive factor is the
ratio of theoretically possible and practically achieved volume
flow. The high volumetric efficiency adapted to be achieved by the
compressor according to the disclosure is an indication of the good
tightness of the compressor.
[0056] Further, the compressor according to the disclosure
preferably has a high isothermal efficiency factor of at least 45
percent, preferably at least 60 percent. The isothermal efficiency
factor is the ratio of ideal isothermal compression capacity and
real compression capacity. The isothermal efficiency factor is also
an indication of good tightness and good cooling of the
compressor.
[0057] In addition, it is preferred that the dry-compressing
compressor is operated by a motor at a mean speed. In particular,
the speed is higher than 3,000 1/min, and particularly preferred
more than 4,000 1/min. On the other hand, the speed is preferably
lower than 10,000 1/min.
[0058] At relatively low speeds in the range of 3,000 1/min of
conventional asynchronous motors, for example, large rotor
diameters must be used. This results in unfavorable ratios of
delivered gas volume and leakage areas. This is approximately
proportional to the rotor diameter. On the other hand, very high
speeds of more than 10,0000 1/min entail very high demands on the
balancing of the rotors or the displacement elements. This is
difficult to achieve in the case of single-pass screw threads. In
addition, with increasing power density due to high speeds, it
becomes more and more difficult to cool the compressor. Another
drawback of very high speeds with very small tooth gaps is the high
gas friction in the gas paths. Thereby, the energy efficiency
decreases. At mean speeds according to the disclosure a good
compromise between tightness, balancing, gas friction and heat
transfer or temperature level can be achieved.
[0059] Preferably, the housing is intensively cooled for keeping
the gas and the components cool. In the embodiment of the
compressor according to the disclosure, this can possibly also be
achieved without internal cooling of the rotors. Low gas
temperatures cause a reduction of the compression operation and
thus have a positive effect on the power consumption of the
compressor.
[0060] According to a preferred aspect of the disclosure, the
rotors and/or the displacement elements can be coated with layers
on the basis of PTFE or molybdenum sulfide, for example, in order
to decrease the gap heights without affecting the operational
safety.
BRIEF DESCRIPTION OF THE DRAWINGS
[0061] Hereunder the disclosure will be explained in detail on the
basis of a preferred embodiment with reference to the accompanying
drawings in which:
[0062] FIG. 1 shows a schematic top view of a preferred embodiment
of a screw rotor of the screw compressor according to the
disclosure,
[0063] FIG. 2 shows a schematic sectional view of displacement
elements having an asymmetrical profile,
[0064] FIG. 3 shows a schematic sectional view of displacement
elements having a symmetrical profile, and
[0065] FIG. 4 shows a schematic sectional view of a screw
compressor.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT
[0066] The screw rotors illustrated in FIGS. 1 to 3 can be used in
a screw compressor according to the disclosure as shown in FIG.
4.
[0067] According to a preferred embodiment of the screw compressor,
the rotor has a pitch changing and/or variable in the direction of
compression, i.e. from left to right in FIG. 1. In a first
suction-side area 10 defining a first displacement element a large
pitch of approximately 50-150 mm/revolution is provided. Here, the
pitch changes in the area 10, i.e. in the precompression area, to
55-65% of the inlet pitch, i.e. approximately 30-100 mm/revolution.
In a second discharge-side area 12 corresponding to a second
displacement element 12 the pitch is considerably smaller. In this
area the pitch is in the range of 10-30 mm/revolution. In the
illustrated embodiment, the at least one displacement element per
screw rotor is thus defined by a screw rotor having a variable,
preferably continuously changing pitch. This corresponds to a
plurality of displacement elements arranged one behind the other as
seen in the direction of delivery.
[0068] In the illustrated preferred embodiment, both in the inlet
area and the outlet area a gas collection chamber 14 each is
provided.
[0069] Further, the integral screw rotor comprises two bearing
seats 16 and a shaft end 18. The shaft end 18 has connected thereto
a gearwheel for driving purposes, for example.
[0070] Likewise, it is possible that the individual displacement
elements 10, 12 are manufactured separately from each other and are
separately affixed to the rotor shaft by pressing, for example.
Here, the bearing seats 16 and the shaft ends 18 can be integral
components of the shaft 20. Here, the continuous shaft 20 can be
made from a material differing from that of the displacement
elements 10, 12.
[0071] In addition, conical rotors can be provided. According to
the disclosure, they comprise a plurality of displacement elements.
Here, too, it is particularly preferred that the plurality of
displacement elements are realized by a variable pitch. Conical
rotors, too, are of a single-pass configuration.
[0072] FIG. 2 shows a schematic sectional view of an asymmetrical
profile (e.g. a Qumiby profile). The illustrated asymmetrical
profile is a so-called Quimby profile. The sectional view shows two
screw rotors which mesh with each other and whose longitudinal
direction is perpendicular to the drawing plane. The
counter-rotation of the rotors is indicated by two arrows 15.
Relating to a plane 17 extending perpendicularly to the
longitudinal axis of the displacement elements, the profiles of the
edges 19 and 21 are of different configuration for each rotor. The
opposing edges 19, 21 must thus be manufactured separately from
each other. However, this somewhat more complex and difficult
manufacture offers the advantage that no continuous blowhole exists
but a short-circuit occurs merely between two adjacent
chambers.
[0073] Preferably, such an asymmetrical profile is provided for the
suction-side displacement element 10.
[0074] The schematic sectional view in FIG. 3 shows a cross-section
of two displacement elements and/or two screw rotors which are
again counter-rotating (arrows 15). Relating to the symmetry axis
17, the edges 23 of each displacement element are of a symmetrical
configuration. The preferred exemplary embodiment of a symmetrical
contour illustrated in FIG. 4 is a cycloid profile.
[0075] A symmetrical profile, as illustrated in FIG. 3, is
preferably provided for the discharge-side displacement elements
12.
[0076] Further, it is possible that more than two displacement
elements are provided. They can possibly have different head
diameters and corresponding foot diameters. Here, it is preferred
that a displacement element having a larger head diameter is
arranged at the inlet, i.e. on the suction side, for realizing a
larger suction capacity in this area and/or increasing the
installed volume ratio. Further, combinations of the embodiments
described above are possible. For example, one or a plurality of
displacement elements can be integrally formed with the shaft, or
an additional displacement element can be separately manufactured
and then mounted to the shaft.
[0077] In the schematic view of a preferred embodiment of a screw
compressor according to the disclosure illustrated in FIG. 4, two
screw rotors, as illustrated in FIG. 1, are arranged in a housing
26. The compressor housing 26 comprises an inlet 28 through which
gas is taken in in the direction indicated by an arrow 30. Further,
the compressor housing 26 comprises a discharge-side outlet 32
through which the gas is discharged in the direction indicated by
an arrow 38. Preferably, the screw compressor according to the
disclosure compresses air in a compressed air chamber.
[0078] Between upper surfaces 42 of the two displacement elements
12 and an inner surface 44 of a suction chamber 46 defined by the
compressor housing 26, a gap is formed whose height preferably lies
in the range of 0.03 mm-0.2 mm and in particular in the range from
0.05 mm-0.1 mm.
[0079] The gap between the edges of the displacement elements
preferably has a gap height of 0.1-0.3 mm.
[0080] In the illustrated exemplary embodiment, the compressor
housing 26 is closed by two housing covers 47. The left housing
cover 47 in FIG. 4 comprises two bearing supports where a ball
bearing 48 each for supporting the two rotor shafts is arranged. On
the right side in FIG. 4, journals 50 of the two screw rotor shafts
protrude through the covers 47. On the outside a respective
gearwheel 52 is arranged on the two shaft journals 50. In the
illustrated exemplary embodiment, the two gearwheels 52 mesh with
each other for synchronizing the two screw rotors with each other.
Further, in the right cover 47 in FIG. 4, two bearings 48 for
supporting the screw rotors are arranged. In the housing walls 47 a
seal not illustrated is provided in addition to the bearings
48.
[0081] The lower shaft in FIG. 4 is a drive shaft connected to a
drive motor not illustrated.
* * * * *