Note : Les descriptions sont présentées dans la langue officielle dans laquelle elles ont été soumises.
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Dry-compressing vacuum pump
The invention relates to a dry-running vacuum pump, in particular a screw
pump.
Dry-compressing vacuum pumps, such as e.g. screw pumps, comprise two rotor
elements arranged in a suction chamber. With screw pumps, the rotor elements
are formed as helical displacement elements. Each rotor element is supported
by
a rotor shaft. With a screw pump, both rotor elements are arranged in the
suction
chamber formed by the pump housing. Both rotor shafts extend through a housing
wall that defines the suction chamber. Gears are connected with the two rotor
shafts. With screw pumps, the two gears mesh with each other. Thereby, on the
one hand, the two shafts rotating in opposite directions are synchronized and,
on
the other hand, the two shafts are driven. By providing the two meshing gears,
only one of the two shafts has to be driven. For obtaining an efficient
compressing
process and a good volumetric efficiency narrow gaps are necessary between the
rotors which require a very exact synchronizing. Typically, maximum synchroniz-
ing errors or circumferential backlashes between the rotors of 0.25 are
allowable.
With dry-compressing vacuum pumps on the market, this can be achieved by
providing meshing gears on the shaft ends. Due to the precision required and
the
small allowable tolerances, the costs are high.
Further, when providing meshing gears for synchronizing, it is necessary to
provide
for oil lubrication. As a result, a complex and complicated sealing is
required in the
housing wall through which the shaft ends extend.
An electron synchronizing of the two rotor shafts is also know. However, the
same
is also complex and costly. Usually, electric motors are provided as drive
means
for driving the rotor shaft. For increasing the rotary speed of the vacuum
pump,
these may be connected with a frequency converter. This also is a relatively
costly
component. Possibly, intermediate gearings are provided which in turn must
also
be oil-lubricated.
Further, it is known e.g. from DE 38 23 927 to synchronize two rotor shafts of
a
screw pump by means of a toothed belt. However, a corresponding product was
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not technically realized and offered on the market. For constructing screw
pumps
driven by means of a belt drive, with which pumps at least a vacuum of 200
mbar
absolute pressure can be obtained when pumping against atmosphere, corre-
spondingly small synchronizing errors have to be realized between the screw-
shaped rotor elements. With the use of toothed belts, this would only be
possible
if the gears or toothed belt wheels driven by the toothed belts have a very
small
tooth gap clearance. According to ISO 13050 this is less than 0.1 mm or less
than
0.2 mm, depending on the profile. Alternatively, with gears, a larger
effective di-
ameter of typically 0.2 to 0.4 mm would have to be provided. This leads to a
forced
tooth pitch error that compromises the synchronizing. Providing larger
effective
diameters, however, significantly shortens the service life of the toothed
belts
used. This is not accepted in dry-compressing vacuum pumps on the market.
It is an object of the present invention to provide a dry-compressing vacuum
pump
driven by a toothed belt and not having the above disadvantages.
The object is achieved, according to the invention, with the features of claim
1.
The dry-compressing vacuum pump comprises a suction chamber formed by a
pump housing. Two rotor elements are arranged in the suction chamber, the vac-
uum pump particularly being a screw pump. Each rotor element is supported by a
rotor shaft. The two rotor elements extend through a housing wall defining the
suction chamber, so that one shaft end per rotor shaft extends from the
suction
chamber. One toothed belt wheel is arranged on the two shaft ends,
respectively.
Since the two rotor shafts are driven by means of a toothed belt, the two
toothed
belt wheels do not mesh with one another. Further, a drive means such as an
electric motor is provided. Here, a belt pulley is arranged in particular on a
drive
shaft of the electric motor. The toothed belt is connected with the two
toothed belt
wheels and the drive means, in particular the belt pulley of the drive means.
In
order to provide a toothed belt for driving the two rotor shafts, a
circumferential
backlash between the two rotor elements of more than 0.75 , in particular
more
than 1 is provided according to the invention. The use of toothed belts is
possi-
ble only because of the provision of such a large circumferential backlash.
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In order to be able, despite the large circumferential backlash, to achieve a
high
vacuum of in particular less than 200 mbar absolute pressure when compressing
against atmosphere, special designs of the compressing stages, i.e. of the dis-
placer elements arranged on the rotor shaft, are preferred, wherein it is
possible
of course that the displacer elements are formed integrally with the rotor
shaft.
Due to the large circumferential backlash allowed by the invention, a
comparatively
high return flow occurs in particular at the suction-side compressing
elements, i.e.
the compressing elements downstream of the pump inlet.
With screw-type rotors having a pitch variable in the delivery direction, the
profile
engagement gap in the inlet region is decisive for the maximum allowable syn-
chronizing error, due to the large pitch of the winding of the displacer
elements
existing there. Even comparatively small angular deviations lead to undesired
flank
contacts in the suction-side displacer elements. To avoid this, a large
circumfer-
ential backlash must be selected. For achieving a good volumetric efficiency
of the
pump despite the large gap created thereby, it is preferred to increase the
number
of windings having a large pitch and a large profile gap in the inlet region.
In
particular, two to three windings are preferably provided in this region. In
addition
or as an alternative, the number of windings in the outlet region, i.e. on the
pres-
sure side, may also be increased. This results in a lesser pressure gradient
in the
inlet region and thus also results in a reduce return flow. In the outlet
region, the
windings have a smaller pitch.
Further, in an alternative preferred embodiment, it is possible that the two
screw
rotors have a plurality of rotor or displacer elements or displacer stages.
Prefera-
bly, at least two displacer elements or displacer stages are provided.
Such a vacuum pump screw-type rotor preferably comprises at least two helical
displacer elements arranged on a rotor shaft. The at least two displacer
elements
preferably have different pitches, the pitch being constant for a respective
dis-
placer element. For example, the vacuum pump screw rotor comprises two dis-
placer elements, wherein a first, suction-side displacer element has a larger
con-
stant pitch and a second, pressure-side displacer element has a smaller
constant
pitch. Due to the preferred provision of a plurality of displacer elements
which each
have a constant pitch, manufacture is significantly facilitated.
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Preferably, each displacer element has at least one helical recess having the
same
contour all along its length. Preferably, the contours differ for each
displacer ele-
ment. Each individual displacer element preferably has a constant pitch and an
unvarying contour. This facilitates the manufacture significantly so that the
man-
ufacturing costs can be reduced drastically.
For a further improvement of the suction capacity, the contour of the suction-
side
displacer element, i.e. in particular of the first displacer element seen in
the pump-
ing direction, is asymmetric in shape. Due to the asymmetric design of the
contour
or the profile, the flanks may be designed such that the leakage surfaces, the
so-
called blowholes, are preferably eliminated completely or at least have a
reduced
cross section. A particularly suitable asymmetric profile is the so-called
"Quimby
profile". Such a profile may be relatively difficult to produce, but has the
advantage
that no continuous blowhole exists. A short circuit only exists between two
adja-
cent chambers. Since the profile is an asymmetric profile with different
profile
flanks, at least two work steps are required for the manufacture, since, due
to
their asymmetry, the two flanks have to be made in different work steps.
The pressure-side displacer element, in particular the last displacer element
in the
pumping direction, is preferably provided with a symmetric contour. In
particular,
the symmetric contour has the advantage that the manufacture is simpler.
Specif-
ically, both flanks with a symmetric contour may be produced with a rotating
end
mill or a rotating disk mill in one work step. Such symmetric profiles have
only
small blowholes, which, however, are continuous, i.e. provided not only
between
two adjacent chambers. The size of the blowhole decreases when the pitch is re-
duced. In this respect, such symmetric profiles may be provided in particular
with
the pressure-side displacer element, since the same, in a preferred
embodiment,
has a smaller pitch than the suction-side displacer element and, preferably,
also
than the displacer element arranged between the suction-side and the pressure-
side displacer element. Although such symmetric profiles are slightly less
tight,
these have the advantage that the manufacture is significantly simpler. In
partic-
ular, it is possible to produce the symmetric profile in a single work step
and pref-
erably with a simple end mill or disc mill. This reduces the costs
drastically. A
particularly well suited symmetric profile is the so-called "cycloid profile".
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Providing at least two such displacer elements results in the corresponding
screw
vacuum pump to be able to generate low inlet pressures at low power consump-
tion. Further, the thermal load is also low. Arranging at least two such
displacer
elements, designed in a preferred manner, with a constant pitch and an
unvarying
contour in a vacuum pump leads to essentially the same results as obtained
with
a vacuum pump having a varying pitch. In case of high built-in volume ratios,
three or four displacer elements may be provided per rotor.
In a particularly preferred embodiment, a pressure-side, i.e. in particular a
last
displacer element in the pumping direction, has a large number of windings for
reducing the obtainable inlet pressure and/or for reducing the power
consumption
and/or the thermal load. By a large number of windings, a larger gap between
the
screw-type rotor and the housing may be accepted, while the performance
remains
the same. Here, the gap may have a cold gap width of 0.05 - 0.3 mm. A large
number of outlet windings or a large number of windings in the pressure-side
dis-
placer element may be produced in an economic manner, since this displacer ele-
ment has a constant pitch and in particular also has a symmetric contour. This
allows for a simple and economic manufacture, so that providing a larger
number
of windings is acceptable. Preferably, this pressure-side or last displacer
element
has more than 6, in particular more than 8 and particularly preferred more
than
windings. In a particularly preferred embodiment, the use of symmetric
profiles
has the advantage that both flanks of the profile can be cut simultaneously
using
a mill. Here, the mill is also supported by the respective opposite flank, so
that a
deforming or bending of the mill during the milling operation and inaccuracies
caused thereby are avoided.
For a further reduction of the manufacturing costs, it is particularly
preferred to
form the displacer elements and the rotor shaft as one piece.
In another preferred embodiment, the change in the pitch between adjacent dis-
placer elements is discontinuous or erratic. Possibly, the two displacer
elements
are arranged at a distance from each other in the longitudinal direction, so
that a
circumferential, cylinder ring-shaped chamber is formed between two displace
el-
ements, which serves as a tool run-out. This is advantageous in particular
with
integrally formed rotors, since the tool producing the helical line can be
guided out
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in a simple manner in this region. If the displacer elements are manufactured
in-
dependently and are thereafter mounted on a shaft, providing a tool run-out,
in
particular such a circular cylindrical region, is not required.
In a preferred development of the invention, no tool run-out is provided at
the
change in pitch between two adjacent displacer elements. In the region of the
change in pitch, preferably both flanks have a discontinuity or recess for
guiding
out the tool. Such a discontinuity has no decisive influence on the
compression
performance of the pump, since it is a discontinuity or recess that is very
limited
locally.
The vacuum pump screw rotor preferably comprises a plurality of displacer ele-
ments. These may each have the same or different diameters. It is preferred in
this respect that the pressure-side displacer element has a smaller diameter
than
the suction-side displacer element.
With displacer elements manufactured independently of the rotor shaft, these
are
mounted on the shaft e.g. by press fitting. In this respect, it is preferred
to provide
elements such as dowel pins to fix the angular position of the displacer
elements
relative to each other.
In particular in case of the integral design of the screw-type rotor, but also
in case
of a multi-part design, it is preferred to manufacture the same from aluminum
or
from an aluminum alloy. It is particularly preferred to manufacture the rotor
from
aluminum or an aluminum alloy, in particular AlSi9Mg or AlSi17Cu4Mg. The alloy
preferably has a high proportion of silicon of preferably more than 9%, in
particular
more than 15%, in order to reduce the expansion coefficient.
In a further preferred development of the invention, the aluminum used for the
rotors has a low expansion coefficient. It is preferred for the material to
have an
expansion coefficient of less than 22 = 10-6/K, in particular less than 20 =
10-6/K.
In another preferred embodiment, the surface of the displacer elements is
coated,
wherein in particular a coating against wear and/or corrosion is provided. In
this
respect it is preferred to provide an anodic coating or another suitable
coating,
depending on the application.
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The vacuum pump has at least two compression stages.
Further, it is preferred for the dry-compressing vacuum pump of the present in-
vention that the vacuum pump has a maximum volumetric efficiency of at least
75%, in particular at least 85%. The volumetric efficiency is the quotient of
the
real maximal obtained volume flow and the theoretically possible volume flow
in a
loss-free pump relative to the suction chamber geometry and the operating
speed.
The maximum volumetric efficiency is usually reached in a range between 1 and
mbar.
The toothed belt used preferably not only serves for driving, but also for
synchro-
nizing the rotor shafts. With screw pumps, the rotor shafts rotate in opposite
di-
rections. Thus, in a preferred embodiment, the toothed belt is designed as a
dou-
ble-sided toothed belt. In top plan view, the toothed belt thus preferably
extends
between the two toothed belt wheels connected with the shaft end.
In a preferred embodiment comprising the above described rotor, tooth gap
clear-
ances of the two toothed belt wheels of more than 0.10 mm can be accepted.
Here, the tooth gap clearance is defined by the combination of the tooth shape
of
the toothed belt wheels used and the tooth shape and size of the teeth of the
toothed belt. Due to the relatively large tooth gap clearance, the service
life of the
toothed belts is significantly extended.
For achieving a further extension of the service life of the toothed belts, it
is further
preferred that the effective diameter is not enlarged and that thus no forced
tooth
pitch error occurs.
Providing a toothed belt for driving and synchronizing the two rotor shafts in
par-
ticular has the advantage that no oil lubrication has to be provided. This has
the
particular advantage that the sealing of the shaft ends with respect to the
suction
chamber may be designed in a significantly more economic manner. Moreover, it
is possible to use grease-lubricated roller bearings. In particular, the two
shafts
are supported in the housing wall through which the shaft ends are passed,
wherein these bearings may be grease-lubricated bearings. The opposite shaft
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ends supported in the region of the inlet side are preferably supported in
grease-
lubricated bearings, but oil-lubricated bearings may also be used.
Further, a belt tensioning means may be provided to constantly keep the belt
taut.
Preferably, this is an automatic tensioning means in which the tension is
generated
e.g. by a spring or the like or a fixed bias is applied during assembly.
Likewise, it
is possible to tension the belt by configuring the drive motor such that it is
dis-
placeable.
Another advantage of the toothed-belt drive according to the invention is that
a
variation of the vacuum pump speed is possible in a simple manner. For that
pur-
pose, it is merely necessary to exchange the toothed belt pulley connected
with
the drive means. When the toothed belt pulley is exchanged, the toothed belt
must
be exchanged as well, as needed.
The invention will be explained hereinafter in detail with reference to a
preferred
embodiment and the accompanying drawings.
In the Figures:
Fig. 1 is a schematic longitudinal section through a screw-type vacuum
pump,
Fig. 2 is a schematic illustration of the drive of the vacuum pump,
Fig. 3 is a schematic illustration of a combination of a toothed belt and a
toothed belt disc with a tooth gap,
Fig. 4 is a schematic illustration of a combination of a toothed belt and a
toothed belt disc without a tooth gap,
Fig. 5 is a schematic top plan view of a first preferred embodiment of a
vac-
uum pump screw-type rotor,
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Fig. 6 is a schematic top plan view of a second preferred embodiment of a
vacuum pump screw-type rotor,
Fig. 7 is a schematic sectional view of displacer elements with an asymmet-
ric profile, and
Fig. 8 is a schematic sectional view of displacer elements with an asymmet-
ric profile.
Fig. 1 is a greatly simplified schematic illustration of a pump housing 10. A
suction
chamber is formed inside the pump housing 10, in which chamber two rotor ele-
ments 14 are arranged. In the embodiment illustrated the rotor elements 14 are
screw-type rotors. The screw-type rotors 14 have helical compression elements
that mesh with each other. The two screw-type rotors 14 are driven in opposite
directions. In the embodiment illustrated the two screw-type rotors 14 have
two
pump stages 16, 18.
The two rotor elements are respectively arranged on a rotor shaft 22. On the
suc-
tion side, the two rotor shafts 22 are supported in a housing cover 24 via
bearing
elements 26. On the opposite side, shaft ends 28 extend through a housing wall
30. The two rotor shafts 22 are supported in the housing wall 30 by grease-
shaped
bearings 32.
The dry-compressing vacuum pump convey a medium through an inlet 34 to an
outlet 36.
For driving the two rotor elements 14, the two shaft ends 28 are each
connected
with a respective toothed belt wheel 38, wherein the two toothed belt wheels
38
do not mesh with each other. Synchronizing is effected via a toothed belt 40
(Fig.
2) not illustrated in Fig. 1. The toothed belt is designed as a double-sided
toothed
belt and, for synchronizing the two toothed belt wheels 38 or the two shaft
ends
28 connected with the toothed belt wheels, is passed between these. Further, a
drive means 42 is provided whose drive shaft 44 is connected with a toothed
belt
disc 46.
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Fig. 3 schematically illustrates teeth of a toothed belt disc 38 or 46 in
connection
with a toothed belt 40. A tooth 48 of the toothed belt 40 is designed such
that gap,
illustrated in hatched lines, is formed opposite a tooth interstice 50 of two
adjacent
teeth 52 of the toothed belt wheel 38. Thereby, a certain play exists between
the
toothed belt 40 and the toothed belt wheel 38. The synchronizing of the two
rotor
shafts 22 may be somewhat compromised thereby, but the service life of the
toothed belt 48 is extended.
As an alternative, a toothed belt may be provided, as schematically
illustrated in
Figure 4. The same shows no distances between the tooth interstice 50 and the
tooth 48 of the belt 40, which is referred to as a zero gap.
In the first preferred embodiment (Fig. 5) of the vacuum pump screw-type
rotor,
the rotor has two displacer elements 110, 112 forming the two pump stages 16,
18. A first, suction-side displacer element 110 has a large pitch of about 50 -
150
mm/rotation. The pitch is constant throughout the displacer element 110. The
contour of the helical recess is constant as well. The second pressure-side
displacer
element 112 also has a constant pitch and a constant contour of the recess
over
its length. The pitch of the pressure-side displacer element 112 is preferably
in the
range of 10 - 30 mm/rotation. An annular cylindrical recess 114 is provided be-
tween the two displacer elements. The same serves to realize a tool run-out,
due
to the integral design of the screw-type rotor illustrated in Fig. 5.
Further, the integrally formed screw-type rotor has two bearing seats 116 and
a
shaft end 118. For example, a gear is connected with the shaft end 118 for
driving.
In the second preferred embodiment illustrated in Fig. 6, the two displacer
ele-
ments 110, 112 are manufactured separately and are then fixed on a rotor shaft
120, e.g. by pressing. This way of manufacturing may be somewhat more complex,
but the cylindrical distance 114 between two adjacent displacer elements 110,
112
is not required as a tool run-out. The bearing seats 116 and the shaft ends
118
may be an integral part of the shaft 120. A continuous shaft 120 may also be
made
of another material different from that of the displacer elements 110, 112.
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Fig. 7 illustrates a schematic sectional view of an asymmetric profile (e.g. a
Quimby profile). The asymmetric profile illustrated is a so-called "Quimby
profile".
The sectional view shows two screw-type rotors meshing with each other, their
longitudinal direction being perpendicular to the drawing plane. The
oppositely di-
rected rotation of the rotors is indicated by the two arrows 115. With
reference to
a plane 117 perpendicular to the longitudinal axis of the displacer elements,
the
profiles of the flanks 119 and 121 are designed differently per rotor. The
opposing
flanks 119, 121 thus have to be manufactured independently. The manufacture
which is therefore somewhat more complex and complicated, has the advantage,
however, that no continuous blowhole exists and a short circuit exists merely
be-
tween two adjacent chambers.
Such an asymmetric profile is preferably provided in the suction-side
displacer
element 110.
The schematic sectional view in Fig. 8 again shows a cross section through two
displacer elements or two screw-type rotors which again rotate in opposite
direc-
tions (arrows 115). With reference to the axis of symmetry 117, the flanks 123
of
each displacer element are symmetrically designed. The preferred embodiment of
a symmetrically designed contour illustrated in Fig. 8 is a cycloid profile.
A symmetric profile, as illustrated in Fig. 8, is preferably provided in the
pressure-
side displacer elements 112.
Further, it is possible that more than two displacer elements are provided.
These
may possibly also have different head diameters and corresponding base diame-
ters. In this respect it is preferred that a displacer element with a larger
head
diameter is arranged at the inlet, i.e. at the suction side, so as to realize
a higher
suction capacity in this region and/or to increase the built-in volume ratio.
Further,
combinations of the above described embodiments are possible. For example, one
or a plurality of displacer elements may be manufactured integrally with the
shaft
or an additional displacer element may be manufactured independently of the
shaft
and may then be mounted on the shaft.
