EP2835536A2 - Vacuum pump stage with special surface roughness yielding a lower gas friction - Google Patents

Abstract

The invention relates to a vacuum pumping stage of a threaded or side channel pump having a stator and at least one rotor, in which at least one thread groove in the stator and / or in the rotor or at least one channel is provided in the stator, wherein the rotor a rotor section dips into the channel and a pumping action is achieved by interaction of the rotor section and the channel.

Description

The invention relates to a vacuum pumping stage.

The prior art includes vacuum pumping stages of screw pumps, which essentially consist of two parts, namely a stator and a rotor rotating in the stator. Multi-start threads are mounted on the outer diameter of the rotor and on the inner diameter of the stator.

Side channel pumps, that is to say pumps which have at least one vacuum pump stage in the form of a side channel pumping stage, can be used in a multi-stage construction in the high pressure range up to atmospheric pressure. These can for example be combined well with turbomolecular pumps or other molecular pumps. The rotor parts of both pumps can be accommodated on a shaft, so that both form a structural unit. The side channel pumping stages usually have an impeller, that is to say a rotor, which has peripheral blades in a channel at its edge.

In order to achieve a sufficiently good pumping power in the known from practice pumps, several stages and elaborately designed wheels, for example, the Seitenkanalpumpstufe are usually necessary.

A further embodiment relates to a vacuum pumping stage having an inlet, an outlet and a channel, which has two side walls and a channel bottom, wherein a rotor with a rotor portion dips into the channel and by pumping action of rotor portion and channel is achieved, and with an intermediate Inlet and outlet arranged breaker.

Many industrial processes occur under vacuum conditions in the molecular flow regime. Vacuum pumps or vacuum pump assemblies composed of vacuum pumps are used to generate such vacuum conditions. In the vacuum pumps vacuum pump stages are used according to different principles of action, which are adapted to different pressure ranges to compress gas from the desired final vacuum to the atmosphere.

For example, side channel pumping stages are used to compress the atmosphere. In these blades run around in a channel and promote a vortex-like gas flow between inlet and outlet. The gas stream follows the blades during the circulation and is at a so-called scraper or breaker detached and supplied to the outlet.

In order to achieve a sufficiently good pumping power, several stages and elaborately designed impellers of the side channel pumping stage are usually necessary. The effort to be operated is evident, for example, from the large number of blades which, at least in the case of small quantities, must be laboriously worked out of solid material.

Such side channel pumping stages are for example in the DE 10 2009 021 642 A1 and the DE 10 2010 019 940 A1 disclosed.

These prior art side channel pumping stages can still be improved in terms of pumping power.

The invention further relates to a further vacuum pumping stage in which a breaker is arranged between the inlet and the outlet.

The prior art ( DE 103 34 950 A1 ) includes a side channel compressor having an inlet, an outlet and a rotor and a channel, wherein the rotor is immersed with a rotor portion in the channel and a pumping action is achieved by cooperation of rotor portion and channel. The rotor usually dips into the channel with rotor blades arranged on the rotor. Between the inlet and the outlet, a breaker is arranged. The breaker encloses the rotor on all sides and, as is known in practice, abruptly near the outlet where the side channel ends, as well as near the inlet where the side channel begins.

According to the prior art ( DE 103 34 950 A1 ), the breaker is designed such that the rotor blades uniformly increasingly enclosed, or released evenly decreasing again. The respective rotor blade is thus gradually enclosed by the breaker and constantly, or again continuously released. It does not come to an abrupt, but a continuous and uniform stripping of the compressed gas components of the respective rotor blades. This measure is implemented at the beginning as well as at the end of the breaker, ie at the inlet and at the outlet. As a result, the formation of disturbing sound components in the interrupter area is suppressed and a gas flow at the discharge nozzle is reduced. This leads to an increase in the efficiency. This belonging to the prior art embodiment has the disadvantage that the efficiency is not fully exploited.

The technical problem underlying the invention is to provide an improved vacuum pumping stage for threaded or side channel pumps, which are used in molecular and viscous pressure ranges to achieve an increase in performance of the pump.

This technical problem is solved by a vacuum pumping stage according to claim 1.

The vacuum pumping stage according to the invention of a threaded or side channel pump having a stator and at least one rotor, wherein at least one thread groove in the stator and / or in the rotor or at least one channel is provided in the stator, wherein the rotor with a rotor portion immersed in the channel and by interaction of rotor section and channel pumping action is achieved is characterized in that the at least one channel or the at least one thread groove has at least one surface, are arranged in the grooves and / or surfaces of stators and / or rotors grooves or that the surface of at least one channel or the at least one thread groove and / or the surface of the rotor is coated with a material which has a lower coefficient of friction than a metal surface.

In the sense of a metal surface is to be understood a smooth metal surface, as used in practice in stators and rotors of vacuum pumps.

The three measures mentioned according to the vacuum pumping stage according to the invention reduce the gas friction on the surface of the pump-active surfaces. This increases the speed, circulation flow and intensity of the energy exchange between the rotor and stator. This in turn leads to an increase in compression, reduction of power consumption and increase in pumping speed.

The grooves arranged in the surface act to form less air movement in the grooves so that, although the surface becomes rougher overall and has a larger surface area, there is less friction.

The same effect, namely the reduction of gas friction on the pump-active surfaces is achieved by coating the pump-active surfaces with a material having a lower coefficient of friction than a metal surface. According to the prior art, the pump-active surfaces of the vacuum pumps are made of metal, which have a smooth surface. For example, coating the surface with a coating having a lower coefficient of friction for gases will reduce gas friction. Advantageously, the coating has anti-stick and sliding properties which are better than the non-stick and sliding properties of a smooth metal surface.

According to an advantageous embodiment of the invention, a bottom surface and / or at least one side wall of the thread grooves are formed such that are arranged in these grooves. Advantageously, the grooves are arranged in the bottom surface and the side walls. However, it is also possible to provide only a portion of the surfaces with grooves.

The grooves are arranged either parallel to a flow direction of the gases, transversely to the flow direction of the gases and / or at an angle to the flow direction of the gases. In all embodiments, only a small amount of air movement is formed in the grooves, so that the gas friction on the pump-active surfaces is reduced.

All grooves of a pump-active surface can be aligned parallel or substantially parallel. However, there is also the possibility that on a pump-active surface, the grooves are formed from a combination of the aforementioned orientations. In opposing pump-active surfaces, the grooves may be parallel to each other or at an angle to each other.

According to an advantageous embodiment of the invention, the grooves have a width and / or a distance of the grooves from one another to between 1 .mu.m and 1 mm. In these Widths and distances a particularly advantageous effect is achieved.

According to a further advantageous embodiment of the invention, the grooves have a depth of between 1 .mu.m and 100 .mu.m. Here, too, the gas friction is optimally reduced.

The grooves may have a uniform structure according to a first embodiment of the invention. This structure may, for example, have a triangular, rectangular, trapezoidal or other shape in cross section.

According to another advantageous embodiment of the invention, the grooves may be formed as irregular grooves. The grooves may be formed, for example, as with a grindstone introduced into the surface grooves. The grooves are arranged in the surfaces of the pump-active surfaces. The grooves in this case advantageously have a roughness of 0.1 μm to 100 μm.

According to another advantageous embodiment of the invention, the pump-active surfaces of stators and / or rotors are coated with a friction-reducing and / or anti-slip material. Advantageously, the pump-active surfaces of stators and / or rotors are coated with a friction which reduces friction with respect to an uncoated surface and / or a material which increases the sliding properties relative to an uncoated surface. The coating material has a lower coefficient of friction for gases and, moreover, better anti-stick and sliding properties, for example against metal, for example aluminum or stainless steel. Also by this As a result, the gas friction at the pump active surfaces is significantly reduced, thereby increasing the velocity of the circulation flow and the intensity of the energy exchange between the rotor and the stator, which, as already stated, leads to an increase in compression, reduction in power consumption and increase in pumping speed.

The object of the invention is also achieved by a vacuum pumping stage with the features of claim 6.

The vacuum pumping stage according to the invention with an inlet, an outlet, a rotor and a channel, which has two side walls and a channel bottom, wherein the rotor with a rotor portion dips into the channel and by co-operation of the rotor portion and the channel a pumping action is achieved and with an between inlet and outlet arranged breaker, characterized in that at least one side wall of the channel is curved.

This vacuum pumping stage according to the invention has the advantage that the side channel has a significant improvement in the vacuum technical data of side channel pumps compared to a rectangular side channel, as it belongs to the prior art. At the same time, the side channel according to the invention is easy to manufacture.

According to a particularly preferred embodiment of the invention, the curvature of the side walls is convex. This training achieves the best vacuum technical values.

Advantageously, the channel is formed axially symmetrical to a median plane of the rotor. With this design, a good pumping power of the vacuum pumping stage is achieved.

According to a particularly preferred embodiment of the invention, the curvature of the at least one side wall in cross section is in each case semicircular. As a result, the best improvements of the vacuum technical data of the vacuum pumping stage according to the invention are achieved.

According to a further advantageous embodiment of the invention, the rotor blades of the rotors are formed in cross-section V-shaped. This shape of the rotor blades has given the best pumping power with the curved side walls of the channel.

Advantageously, the rotor blades on a blade bottom on a supernatant. The rotor with the rotor blades is designed such that a projection is provided over a blade root of the rotor blades. This means that the material of the rotor blades is not worn down to the bottom of the blade, but that a supernatant is present. This supernatant also has an advantageous effect on the pumping power of the vacuum pumping stage.

According to a preferred embodiment of the invention, the projection above the blade root to the rotor blade center is tapered. The projection above the blade root to the rotor blade center is seen in the axial direction is tapered. This means that at the axial edges of the blades, the blades are worn down to the blade base and that the supernatant over the blade base is maximally formed toward the center.

A further advantageous embodiment provides that the rotor blades are completely in their height in the Side channel are arranged. This also achieves optimized pump performance.

According to a further very advantageous embodiment of the invention, the blade root of the rotor blades and a radially arranged in the direction of the shaft boundary surface of the side channel are arranged in the radial direction at the same height. This means that the rotor blades are completely arranged in the side channel and unfold their full effect there. The boundary surface arranged radially in the direction of the shaft is the surface of the side channel, which is arranged opposite the channel bottom. By this embodiment, the rotor blades run in their full height in the channel.

In other words, a blade root radius and a radius of the boundary surface of the side channel arranged radially in the direction of the shaft have the same size R _{S 1} .

Advantageously, a blade root radius and a radius of the radially arranged in the direction of the shaft boundary surface of the side channel on the same size R _{S 1} . This also significantly increases the pumping effect.

• Δ ≤ 0,3 mm für p ₂ ≤ 10 mbar
• Δ ≤ 0,2 mm für 10 mbar < p ₂ ≤ 100 mbar
• Δ ≤ 0,15 mm für p ₂ > 100 mbar.

According to a further advantageous embodiment, an axial gap (Δ) is provided between the rotor and stator disks, and the axial gap is configured as follows:

• Δ ≤ 0.3 mm for p ₂ ≤ 10 mbar
• Δ ≤ 0.2 mm for 10 mbar < p ₂ ≤ 100 mbar
• Δ ≤ 0.15 mm for p ₂ > 100 mbar.

These values have proven to be particularly advantageous.

A further advantageous embodiment of the vacuum pump stage provides that with increasing speed and increasing peripheral speed of the rotor disks of the side channel radius R _{S 3} and the distance d _{S 1 is} increasingly formed. This also has a positive influence on the pumping power.

Advantageously, a blade height of the rotor blades is 60% to 100% of a rotor disk width. This serves to further improve the pump power.

The optimum blade height is advantageously 60% to 100% of the rotor disk width. In addition, the optimum side channel radius depends on the peripheral speed of the rotor disc. The side channel radius is advantageously formed between 80% to 120% of the rotor disk width.

A width d _{S 1 of} the channel arc is preferably between 20% and 120% of the rotor disc width.

In addition, a blade spacing of the rotor blades is advantageously between 50% and 100% of the rotor disk width.

According to a further advantageous embodiment, the blade clearance is less than or equal to 55% of the rotor disk width with a side channel area which is smaller than 2.5 times the blade area. A blade spacing of 50% of the rotor disk width is particularly advantageous in side channels with a side channel surface which is not is greater than 2.5 times the blade surface. These are small side channels.

Advantageously, the blade spacing is greater than or equal to 85% of the rotor disk width with a side channel area that is greater than 5 times the blade area. These are big side channels.

The optimum number of blades thus becomes smaller with increasing side channels, or the optimum distance between the blades is larger.

The latter measures all serve to improve the pumping capacity of the pump.

Furthermore, it has been found to be advantageous that the minimum web width is formed depending on the manufacturing accuracy and the material strength of the rotor disk. This ensures the stability of the rotor disk.

The technical problem is also solved by a vacuum pumping stage having the features according to claim 11.

The vacuum pumping stage according to the invention with an inlet, an outlet, a rotor and a channel, wherein the rotor with a rotor portion dips into the channel and by co-operation of the rotor portion and the channel pumping action is achieved in which rotor blades are arranged on the rotor, and the one Rotor disposed between inlet and outlet, characterized in that the breaker has a bevel on a side facing the inlet.

The rotor has a main body and arranged on the main body rotor blades. The rotor is arranged in the side channel pumping stage such that the main body of the rotor is associated with two axial sealing surfaces of the side channel pumping stage. These axial sealing surfaces expand to the so-called side channel in which the blades of the rotor rotate. The breaker interrupts the side channel between inlet and outlet. The chamfering according to the invention on the side facing the inlet advantageously has a depth which is greater than the depth of the axial sealing surface of the pumping stage in the axial direction of the pumping stage. This means that the chamfer is designed to run into the side channel.

The vacuum pumping stage according to the invention has the advantage that the bevel in the area of the breaker is provided only at the inlet. This increases the efficiency of the pumping stage over the prior art, in which the bevel is additionally arranged in the region of the outlet.

According to a particularly advantageous embodiment of the invention, it is provided that the chamfer has an opening angle β of the chamfer formed between the at least one bevel and a central axis, and that an additional angle α corresponding to the opening angle β corresponds to a blade angle α of the rotor blades.

This represents a further significant innovation in the prior art. By this measure, it is achieved that the turbulence flows just after the inlet, the means immediately after the exit from the breaker, forms, and not later in the side channel.

When the rotor blades leave a standard breaker having an abrupt beginning of the side channel, then the compressed gas from the blades flows into the side channel in all directions. As a result, the formation of the vortex flow is not favored, but adversely affected. In an inventively designed breaker, the gas flows only in the direction in which the chamfer is arranged, since on the other sides of the channel surrounds the rotor blades. In the side channel, the gas flows in the region of the chamfer in this area and is sucked by the rotor blades on the sides and ejected again on the side opposite the chamfer. As a result, a vortex flow is formed immediately after the exit of the rotor blades from the breaker, which has a very beneficial effect on the compression and the pumping speed. By the formation of the complementary angle β and thus the opening angle α of the chamfer analogous to the angle of attack of the rotor blades, the effect of the turbulence is optimized.

At the outlet of the breaker, the gas can even reach a supersonic speed. The breaker mold according to the invention acts like a Laval nozzle. The gas in the blades expands in the direction of the chamfer and is cooled. The gas pressure at the inlet and in particular the gas pressure in the rotor blades decreases. This favors the gas suction on the sides of the rotor blades. Due to the ordered gas flow above the rotor blades because of the breaker shape according to the invention, a rapid formation of an ordered vortex flow immediately after reached the breaker. This in turn leads to an increase in the pumping speed and an increase in the compression of the pump.

According to a particularly preferred embodiment of the invention, the breaker has the bevel only at the inlet. This means that the breaker at the outlet region has the abrupt end known from the prior art. For increasing the pumping speed and the compression of the pump, it is sufficient or even conducive to bevel the duct only at the inlet area.

According to a further advantageous embodiment of the invention, the breaker has a length d ₁ , which corresponds to at least one blade length. Breaker means the area which has a reduced cross-section with respect to the channel.

It has been found to be particularly advantageous that the at least one bevel has a length d ₂ which corresponds to at least one or two blade lengths. This also achieves optimum pumping speed and optimum compression.

A further advantageous embodiment of the invention provides that the at least one bevel is linear. As a result, the formation of the vortex flow is optimally favored.

A further advantageous embodiment of the invention provides that the channel has a round or polygonal cross-section or a polygonal cross-section with rounded corners. In both cross-sectional shapes is the chamfer towards the inlet advantageous to optimize the pump performance.

According to a further advantageous embodiment of the invention, the channel has at least one side wall and the at least one side wall of the channel is formed curved. This design of the side channel is particularly advantageous.

According to a further advantageous embodiment of the invention, the bevel is formed in the radial direction of the vacuum pumping stage in the transition into the side channel over the entire width of the channel. This also positively influences the formation of the turbulent flow.

Another advantageous embodiment of the invention provides that the chamfer in the radial direction of the vacuum pump stage at the transition into the side channel only has a part of the entire width of the channel.

According to a further advantageous embodiment of the invention, the chamfer has a maximum depth corresponding to the axial depth of the channel. This embodiment allows a good formation of the vortex flow.

According to another advantageous embodiment of the invention, the bevel has a maximum depth up to an axial sealing surface of the rotor disk in the blade-less region. This embodiment also already allows a sufficient formation of the turbulent flow.

Fig. 1

einen Längsschnitt durch eine Vakuumpumpe mit Seitenkanalpumpstufen;

Fig. 2

eine schematische Darstellung einer bogenförmigen Rillenstruktur im Querschnitt oder Längsschnitt;

Fig. 3

eine schematische Darstellung einer trapezförmigen Rillenstruktur im Querschnitt oder Längsschnitt;

Fig. 4

eine schematische Darstellung einer dreieckförmigen Rillenstruktur im Querschnitt oder Längsschnitt;

Fig. 5

eine schematische Darstellung einer rechteckförmigen Rillenstruktur im Querschnitt oder Längsschnitt;

Fig. 6

eine schematische Darstellung einer dreieckförmigen Rillenstruktur im Querschnitt oder Längsschnitt;

Fig. 7

einen Querschnitt oder Längsschnitt durch eine unregelmäßige Rillenstruktur;

Fig. 8

eine beschichtete Gewindenut im Querschnitt;

Fig. 9

einen Längsschnitt durch eine Vakuumpumpe mit einer Seitenkanalpumpstufe;

Fig. 10

einen Schnitt quer zur Wellenachse durch die Seitenkanalpumpstufe gemäß Fig. 9 entlang der Linie I-I;

Fig. 11

einen Teilquerschnitt durch einen erfindungsgemäßen Seitenkanal;

Fig. 12

eine Darstellung eines Vergleiches der Kompressionen von rechteckigen und kreisförmigen Seitenkanälen mit V-förmigen Rotorschaufeln bei 800 Hz und 1000 Hz Drehfrequenz;

Fig. 13

eine Darstellung der Abhängigkeit des Kompressionsfaktors vom Axialspalt zwischen Rotor und Statorscheiben bei 217 m/s Rotorumfangsgeschwindigkeit;

Fig. 14a

eine Darstellung des Kompressionsfaktors in Abhängigkeit von Auslassdruck p ₂, Drehfrequenz f und Seitenkanaldurchmesser R _S3 bei 1000 Hz;

Fig. 14b

eine Darstellung des Kompressionsfaktors in Abhängigkeit von Auslassdruck p ₂, Drehfrequenz f und Seitenkanaldurchmesser R _S3 bei 800 Hz;

Fig. 15a

eine Darstellung des Kompressionsfaktors in Abhängigkeit von Auslassdruck p ₂, Drehfrequenz f und Abstand d _s1 bei 1000 Hz;

Fig. 15b

eine Darstellung des Kompressionsfaktors in Abhängigkeit von Auslassdruck p ₂, Drehfrequenz f und Abstand d _s1 bei 800 Hz;

Fig. 16

eine Draufsicht auf eine Rotorscheibe mit V-förmigen Schaufeln;

Fig. 17

eine Seitenansicht der Rotorscheibe gemäß Fig. 16;

Fig. 18

ein geändertes Ausführungsbeispiel eines Querschnittes eines Seitenkanales;

Fig. 19

ein geändertes Ausführungsbeispiel eines Querschnittes eines Seitenkanales;

Fig. 20

ein geändertes Ausführungsbeispiel eines Querschnittes eines Seitenkanales;

Fig. 21a

eine Darstellung der Verringerung der Seitenkanalfläche von oben;

Fig. 21b

eine Darstellung der Verringerung der Seitenkanalfläche von unten;

Fig. 21c

eine Darstellung der Verringerung der Seitenkanalfläche von oben und von unten;

Fig. 22

ein geändertes Ausführungsbeispiel;

Fig. 23

ein geändertes Ausführungsbeispiel;

Fig. 24

einen zum Stand der Technik gehörenden Unterbrecher in Seitenansicht und in Draufsicht (schematisch);

Fig. 25

einen erfindungsgemäßen Unterbrecher in Seitenansicht und in Draufsicht (schematisch);

Fig. 26

eine Rotorschaufel in Seitenansicht zur Darstellung des Anstellwinkels α;

Fig. 27

ein geändertes Ausführungsbeispiel;

Fig. 28

eine Darstellung einer Kompression einer Seitenkanalstufe mit Standardunterbrecher und mit erfindungsgemäßem Unterbrecher;

Fig. 29

eine Darstellung des Saugvermögens einer Seitenkanalstufe mit Standardunterbrecher und mit erfindungsgemäßem Unterbrecher;

Fig. 30

eine Statorscheibe mit Unterbrecher in axialer Draufsicht.

Further features and advantages of the invention will become apparent from the accompanying drawings, in which several embodiments a vacuum pumping stage according to the invention are shown only by way of example. In the drawing show:

Fig. 1

a longitudinal section through a vacuum pump with side channel pumping stages;

Fig. 2

a schematic representation of an arcuate groove structure in cross section or longitudinal section;

Fig. 3

a schematic representation of a trapezoidal groove structure in cross-section or longitudinal section;

Fig. 4

a schematic representation of a triangular groove structure in cross section or longitudinal section;

Fig. 5

a schematic representation of a rectangular groove structure in cross-section or longitudinal section;

Fig. 6

a schematic representation of a triangular groove structure in cross section or longitudinal section;

Fig. 7

a cross section or longitudinal section through an irregular groove structure;

Fig. 8

a coated thread groove in cross section;

Fig. 9

a longitudinal section through a vacuum pump with a side channel pumping stage;

Fig. 10

a section transverse to the shaft axis through the side channel pumping stage according to Fig. 9 along the line II;

Fig. 11

a partial cross section through a side channel according to the invention;

Fig. 12

a plot of a comparison of the compressions of rectangular and circular side channels with V-shaped rotor blades at 800 Hz and 1000 Hz rotation frequency;

Fig. 13

a representation of the dependence of the compression factor of the axial gap between the rotor and stator discs at 217 m / s rotor peripheral speed;

Fig. 14a

a representation of the compression factor as a function of outlet pressure p ₂ , rotational frequency f and side channel diameter R _{S 3} at 1000 Hz;

Fig. 14b

a representation of the compression factor as a function of outlet pressure p ₂ , rotational frequency f and side channel diameter R _{S 3} at 800 Hz;

Fig. 15a

a representation of the compression factor as a function of outlet pressure p ₂ , rotational frequency f and distance d _{s 1} at 1000 Hz;

Fig. 15b

a representation of the compression factor as a function of outlet pressure p ₂ , rotational frequency f and distance d _{s 1} at 800 Hz;

Fig. 16

a plan view of a rotor disk with V-shaped blades;

Fig. 17

a side view of the rotor disk according to Fig. 16 ;

Fig. 18

a modified embodiment of a cross section of a side channel;

Fig. 19

a modified embodiment of a cross section of a side channel;

Fig. 20

a modified embodiment of a cross section of a side channel;

Fig. 21a

a representation of the reduction of the side channel area from above;

Fig. 21b

a representation of the reduction of the side channel area from below;

Fig. 21c

an illustration of the reduction of the side channel area from above and below;

Fig. 22

a modified embodiment;

Fig. 23

a modified embodiment;

Fig. 24

a prior art breaker in side view and plan view (schematic);

Fig. 25

a breaker according to the invention in side view and in plan view (schematically);

Fig. 26

a rotor blade in side view to illustrate the angle of attack α;

Fig. 27

a modified embodiment;

Fig. 28

a representation of a compression of a side channel stage with standard breaker and with inventive breaker;

Fig. 29

a representation of the pumping speed of a side channel stage with standard breaker and with inventive breaker;

Fig. 30

a stator with breaker in axial plan view.

Fig. 1 shows a vacuum pump with a housing 1 and three pumping units 14, 16, 18. The housing 1 is provided with a gas inlet opening 2 and a gas outlet opening 4. The pump units consist of rotating and fixed gas-conveying components. The rotating components are mounted on a shaft 6 in the axial direction one behind the other. To operate the shaft 6 includes a drive system 8 and bearing elements 10 and 12. The fixed components are firmly connected to the housing 1.

One of the gas inlet opening facing pump unit 14 is formed as a turbomolecular pump. The following in the direction of gas flow pump unit 16 consists of several subunits 16a, 16b, 16c. These each have one or more molecular pumping stages according to the type of Gaede, hereinafter called Gaede stages. Within the subunits, the Gaede stages are connected in parallel. The subunits themselves are connected in series. This means that connecting elements 34a for the subunit 16a and 34b for the subunit 16b, the input sides and on the other side, the output sides of Gaede stages together so that a parallel gas flow in the individual subunits is made possible. The subunits are interconnected by connecting members 36a, 36b and 36c so that the output side of one subunit is connected to the input side of the following subunit, respectively. The gas outlet opening facing pump unit 18 is formed as a multi-stage side channel pump. In the Fig. 1 shown pump is shown only by way of example.

The invention relates to all vacuum pumps in which side channel pumping stages and / or screw pumps are provided.

According to the invention it is provided that grooves are arranged in the surface of thread grooves and / or that grooves are arranged in the surfaces of stators and / or rotors.

These grooves can have a structure as in Fig. 2 shown, have.

The Fig. 2 to 6 show possible structures that are uniformly mounted in a surface 41, such as a thread groove of a side channel or on a rotor.

Fig. 2 shows a structure with grooves 40 having a rounded bottom. The grooves 40 are arcuate. Fig. 3 shows a trapezoidal structure with a conically tapering cross section, while Fig. 4 shows a triangular structure with a conically tapering cross section. In Fig. 5 is shown a rectangular structure. Fig. 6 again shows a triangular structure having an asymmetric configuration.

The depth of the grooves 40 may vary from 1 μm to 100 μm. The groove width, or the distance between the individual grooves 40, can vary from 1 μm to 1 mm. The grooves 40 may be machined into the surface 41 along the flow direction, transverse to the flow direction, and at an angle to the flow direction of the gas.

As in Fig. 7 shown, the grooves 40 can also be generated with a grindstone in a surface 41. The grooves 40 have an irregular structure in this case. The rough surface should have a roughness of 0.1 .mu.m to 100 .mu.m, preferably from 2 .mu.m to 100 .mu.m. In all profiles in the Fig. 2 to 7 are shown formed in the grooves 40 standing air, so that the gas friction on the surface 41 is reduced. By this effect, the sliding of gas layers affected. By influencing these so-called boundary layer forces, a sliding of the gases on the surface of the pump-active surfaces is favored. As a result, the speed of the circulation flow and the intensity of the energy exchange between the pump-active surfaces of the rotor and stator is increased. This leads to an increase in compression, a reduction in power consumption and an increase in pumping speed.

According to Fig. 8 is a thread groove 50 of a threaded pump shown. The thread groove 50, which is arranged, for example, in a stator 51, as well as the adjoining surfaces of the thread groove 50 are coated with a coating 52 which reduces the friction and improves the sliding properties of the surface compared to an uncoated surface, for example a metal surface, for example aluminum or stainless steel. Also by this measure, the gas friction is reduced at the channel surface, whereby the above advantages occur.

Fig. 9 shows a vacuum pump 100 with a gas inlet 102 and a gas outlet 103 and a housing 101. The housing 101 is composed of four housing parts 120, 121, 122, 123, which receive the components of the vacuum pump 100.

Gas entering the vacuum pump 100 through the gas inlet 102 first enters a molecular stage 105. It has an inner stator 505 provided with an internal thread groove 507 and an outer stator 506 provided with an outer thread groove 508. Between inner stator and outer stator, a smooth surface cylinder 502 is provided, which is connected to the rotor 500. The molecular step 105 is thus designed as a Holweck stage. In the Fig. 9 illustrated Holweck level is symmetrical with a surrounding of stator components second cylinder 502 'and therefore operates in two stages.

The rotor is connected to a shaft 108, which is rotatably mounted in rolling bearings 110 and 111. Instead of rolling bearings 110, 111, passive and active magnetic bearings can also be used. On the shaft 108 at least one permanent magnet 113 is arranged, which cooperates with a stationary coil 112 and forms a drive 107 together with this. The rolling bearing 110, the drive 107 and the molecular stage 105 are arranged in the housing parts 120, 121.

The shaft 108 passes through the housing part 122, which includes a side channel pumping stage 104. The side channel pumping stage 104 is formed by a side channel 401 and an impeller 400, wherein on the impeller 400 at least one blade 402 is arranged, which rotates in the side channel by the rotation of the shaft 108 and thus generates the pumping action. Gas passes through a transfer channel 124 from the molecular stage 105 into the side channel stage 104 and is expelled through another transfer channel 125.

From the side channel pumping stage 104, the gas passes through the transfer channel 125 in a Vorvakuumstufe 106. This is also designed as a Seitenkanalpumpstufe, in which case the geometry of the arranged on the impeller 600 and rotating in the side channel 601 blades 602 of the geometry of the blades 402. From this pumping stage 106, the gas from the vacuum pump 100 is discharged through the gas outlet 103.

Between the wheels 400 and 600 and the housing parts 121, 122 and 123 are narrow gaps. These allow a free rotation of the impeller concerned, but are designed so tight that no disturbing gas flows occur.

Fig. 10 shows a section through the housing part 122 along the line II of Fig. 9 , On the shaft 108, the impeller 400 sits. This has an edge 403, on which along the circumference evenly distributed blades 402 are arranged. The side channel 401 surrounds the impeller, wherein the side channel in the radial direction surrounds the blade region of the impeller in a substantially annular manner. Only over part of the circumference of the housing adjacent to the impeller. This section forms a breaker 404 which separates the suction and discharge sides and at which the gas flow, which forms in the side channel and follows the rotation of the impeller, is detached therefrom and transferred to the transfer channel 125.

As in Fig. 11 1, the side channel 401 has a channel bottom 420 and two side walls 421, 422. The side walls 421, 422 are curved. That is, they have a concave shape. The blades 402 of the impeller or rotor 400 project completely into the side channel 401. A radius R _{S 1 of} a blade root 423 is equal to the radius R _{S 1 of} a radially arranged in the direction of the shaft boundary surface 424 of the side channel 401st

That is, the vanes 402 are completely immersed in the side channel 401.

Due to the curved side surfaces 421, 422, the pumping power of the side channel pumping stage is significantly improved. Advantageous is the bridge between the blades as low as possible (not shown). The volume of gas filled with gas should be as large as possible.

These measures considerably improve the vacuum technical properties of the pump.

Improvements in vacuum technology data are also achieved by optimizing the side channel radius R _{S 3} (80% to 120% of the rotor width) and the distance between two centers of the side channel semicircles d _{S 1} (20% to 120% of the rotor width). The optimum radius R _{S 3} and distance d _{S 1} depend on the peripheral speed of the rotor disk and on the blade size. The dimensions R _{R 1} , R _{R 3} , d _{R 1} , blade height h and blade angle α are predetermined.

$$d_{S2}=d_{R1}+2\cdot \mathrm{\Delta }$$

$$R_{S1}=R_{R1}-h$$

The measure R _{S 2} can be calculated with the following three equations: $$R_{S2}=R_{S1}+\sqrt{R_{S3}^2-\frac{1}{4}{\left(d_{S2}-d_{S1}\right)}^2}$$

$$d_{S2}=d_{R1}+2\cdot \mathrm{\Delta }$$

$$R_{S1}=R_{R1}-H$$

The dimension R _{S 1} is predetermined by the lower blade edge of the rotor disk.

$$\mathrm{\Delta }\le 0,2\mathrm{mm\; für}10\mathrm{mbar}<p_2\le 100\mathrm{mbar}$$

$$\mathrm{\Delta }\le 0,15\mathrm{mm\; für}{p}_2>100\mathrm{mbar}$$

Δ denotes the axial gap between the rotor and the stator disc. The axial gap Δ may preferably be from 0.01 mm to 0.5 mm. Small axial gaps are useful on the discharge side and large axial gaps on the suction side. If a labyrinth seal is used on the axial surface between rotor and stator discs, the Axial gap more than 0.5 mm. The guideline values for the axial gaps can be selected as follows: $$\mathrm{\Delta }\le 0.3\mathrm{mm\; for}{\mathrm{<figure-callout\; id="2"\; label="p"\; filenames="imgf0001.png,imgf0016.png"\; state="\{\{state\}\}">p</figure-callout>}}_2\le 10\mathrm{mbar}$$

$$\mathrm{\Delta }\le 0.2\mathrm{mm\; for}10\mathrm{mbar}<{\mathrm{<figure-callout\; id="2"\; label="p"\; filenames="imgf0001.png,imgf0016.png"\; state="\{\{state\}\}">p</figure-callout>}}_2\le 100\mathrm{mbar}$$

$$\mathrm{\Delta }\le 0.15\mathrm{mm\; for}{\mathrm{<figure-callout\; id="2"\; label="p"\; filenames="imgf0001.png,imgf0016.png"\; state="\{\{state\}\}">p</figure-callout>}}_2>100\mathrm{mbar}$$

In Fig. 12 a comparison of rectangular in cross-sectional side channels and side channels with two semi-circular in cross-section side walls with V-shaped rotor blades at 800 Hz and 1000 Hz rotational frequency in comparison is shown. The curves 716, 717, 718, 719 represent the course of the compression as a function of the pressure. The lower two curves 718, 719 refer to a rotation frequency of 800 Hz. A side channel with semicircular side walls has a higher compression (curve 718). on as a prior art belonging in the cross section rectangular channel (curve 719). The two upper curves 716, 717 refer to a rotational frequency of 1000 Hz. The upper curve 716 represents the compression as a function of pressure for a side channel with semi-circular semicircular side walls. Again, the compression is significantly increased by the inventive design of the side channel against a side channel with a rectangular cross-section (curve 717). It can be seen that the side channels with two semi-circular in cross-section side walls have a much better compression.

In Fig. 13 the dependence of the compression factor on the axial gap is shown. Like the legend in Fig. 13 As can be seen above, axial gaps between 0.15 mm and 0.4 mm have been detected. The compression factor k ₀ is greater, the smaller the axial gap is.

Different rotor discs of a multi-stage side channel pump with the same blade size have the same speed, but depending on the rotor disc diameter R _{R 1 have} different peripheral speeds. For this reason, rotor disks with different diameters R _{R 1} and the same blade size should have side channels with different radii R _{S 3} and distances d _{S 1} .

Measurements have shown that with increasing speed and consequently increasing peripheral speed of rotor disks, the optimum side channel radius R _{S 3} and the distance d _{S 1} increase. The optimum is the side channel size with the best compression factor. The pumping speed and power consumption increase in proportion to the side channel area.

• Für eine Rotorscheibe mit einem Radius R _R1 = 69 mm, Breite d _R1 = 5 mm und Schaufelhöhe R _R1 - R _S1 = 4 mm beträgt der optimale Seitenkanalradius bei einer Drehzahl f = 800 Hz und einer Umfangsgeschwindigkeit V = 173 m/sec gleich R _{S3 optimal} = 5 mm. Für eine Drehzahl f = 1000 Hz und eine Umfangsgeschwindigkeit V = 217 m/sec beträgt der optimale Seitenkanalradius R _{S3 optimal} = 5,3 mm. Mit steigender Drehzahl f und Umfangsgeschwindigkeit V wird der optimale Seitenkanalradius weiter zunehmen, beziehungsweise mit fallender Drehfrequenz und Umfangsgeschwindigkeit abnehmen.

In the Fig. 14a and 14b the compression factor is given as a function of the outlet pressure p ₂ , rotational frequency f and side channel diameter R _{S 3} . The in the Fig. 14a and 14b The compression factor shown is given for the following example:

• For a rotor disk with a radius R _{R 1} = 69 mm, width d _{R 1} = 5 mm and blade height R _{R 1} - R _{S 1} = 4 mm is the optimum side channel radius at a speed f = 800 Hz and a peripheral speed V = 173 m / sec equal to R _{S 3 optimal} = 5 mm. For a speed f = 1000 Hz and a peripheral speed V = 217 m / sec, the optimum side channel radius R _{S 3 is optimal} = 5.3 mm. With increasing speed f and peripheral speed V, the optimum side channel radius will continue to increase, or decrease with decreasing rotational frequency and peripheral speed.

In the Fig. 15a and 15b the compression factor is shown as a function of the outlet pressure p ₂ , rotational frequency f, distance d _{S 1} .

The optimum distance d _{S 1} is at a speed of f = 800 Hz depending on the pressure range either d _{S 1} = 1.2 mm or d _{S 1} = 3.6 mm. When the rotational speed f increases up to 1000 Hz, the optimum distance, depending on the pressure range, becomes either d _{S 1} = 3.6 mm or d _{S 1} = 4.8 mm. There is a tendency to increase the optimum distance d _{S 1} with increasing speed f to recognize.

The above dependencies apply only to rotor disks with V-shaped blades, as they are in Fig. 16 are shown. Fig. 16 shows the impeller 400 with the blades 402. The blades 402 are V-shaped. The blade ground has in the region of a median plane 425 of the impeller 400 a projection that rises from edges 426, 427 of the blade root to the median plane 425. The impeller 400 rotates in the direction of arrow A.

Fig. 17 shows the impeller 400 according to Fig. 16 in side view in the direction of arrow B. The impeller 400 carries the V-shaped blades 402. The blades have a blade base 423 on. Above the blade bottom 423 is a projection 428 over.

In general, the following design guidelines should be followed when designing side channel pumps. An optimal blade height is 60% to 100% of the rotor disk width. An optimum side channel radius depends on the peripheral speed of the rotor disk 400 and can be from 80% to 120% of the rotor disk width. The distance d _{S 1 also} depends on the peripheral speed of the rotor disk and may be from 20% to 120% of the rotor disk width.

$$\alpha =\arcsin \left(\frac{d_{S2}-d_{S1}}{2\cdot R_{S3}}\right)$$

$$C=\sqrt{R_{S3}^2-{\left(\frac{d_{S2}-d_{S1}}{2}\right)}^2}$$

$$A_{\mathit{Sch}}=d_{R1}\cdot \left(R_{R1}-R_{S1}\right)$$

The optimum number of blades or the optimal distance between the blades does not depend on the speed. The optimal distance between the blades is proportional to the blade size and also depends on the side channel size. It is from 5o% to 100% of the rotor disk width, the optimum spacing between the blades is less than or equal to 55% for small side channels (side channel area no greater than 2.5 times the blade area) and greater than or equal to 85% for large side channels (side channel area not less than 5 times the blade area). The optimum number of blades thus becomes smaller with increasing side channels, or the optimum distance between blades is larger. The side channel area A _SK and the blade area A _Sch can be calculated using equations 4 to 7. $$A_{\mathit{SK}}=R_{S3}^2\cdot \left(\pi -\alpha \right)+d_{S1}\cdot \left(R_{S3}+C\right)+C\cdot \frac{1}{2}\cdot \left(d_{S2}-d_{S1}\right)$$

$$\alpha =\arcsin \left(\frac{d_{S2}-d_{S1}}{2\cdot R_{S3}}\right)$$

$$C=\sqrt{R_{S3}^2-{\left(\frac{d_{S2}-d_{S1}}{2}\right)}^2}$$

$$A_{\mathit{Sch}}=d_{R1}\cdot \left(R_{R1}-R_{S1}\right)$$

The web width of the blades should be as small as possible. The minimum web width is limited by the manufacturing accuracy and the material strength of the rotor disk.

The Fig. 18 to 20 show further design possibilities of a side channel. In Fig. 18 the side channel 401 is circular in shape. The side channel 401 has no plan side channel bottom, but overall a circular cross-section.

According to Fig. 19 the side channel 401 is also circular. However, the radius of the side channel 401 is smaller than in FIG Fig. 18 shown.

According to Fig. 20 The side channel 401 has convex side walls 421, 422. The channel bottom 420 is flat.

In the embodiments of the side channels of Fig. 18 to 20 the side channel cross-sectional diameter is advantageously formed constant over the entire circumference of the side channel. There is also the possibility that the side channel cross-sectional diameter decreases from an inlet 124 to an outlet 125. According to Fig. 9 For example, the inlet 124 and the outlet 125 are diametrically opposed. However, it is also an arrangement in a side channel pumping stage possible, as in Fig. 10 has been drawn by dashed lines. Here an inlet 124 'is drawn. In this configuration, it is possible for the side channel cross-sectional diameter to decrease from the inlet 124 'to the outlet 125. This reduction can be linear with the circumferential angle. It can also be another function of the circumferential angle.

In the Fig. 21a to 21c a side channel surface is shown with a centerline 126 of the side channel as a function of radius and angle φ .

The reduction of the side channel area may, as in Fig. 21a shown, done from above. It can also be done from below, as in the illustration Fig. 21b shown. However, it can also be done from above and from below, as in the illustration Fig. 21c shown. The side channel diameter may also be reduced from one side or both sides along the side channel from the inlet 124 'to the outlet 125. The inlet 124 'is in Fig. 10 shown.

Fig. 22 shows a further embodiment of a side channel 401. The side channel 401 has side walls 421, 422 which are formed in a circular section. The channel bottom 420 is also not shown plan in this embodiment, but consists of two circular sections with a radius R _{S. 3}

Fig. 23 shows a further embodiment of an embodiment of the side channel 401. The side channel 401 has curved side surfaces 421, 422 and a not plan trained channel bottom 420. The curved side surfaces 421, 422 in this case do not correspond to circular sections.

A breaker 404 is in Fig. 10 shown. The breaker is in the side channel pumping stage 104 of Fig. 9 arranged. The description of the figure Fig. 9 and 10 are fully transferable to the present invention.

Fig. 24 FIG. 12 shows a prior art breaker 404 having an inlet 701 and an outlet 702. The breaker 404 as well as the inlet 701 and the outlet 702 are part of a stator 700. The top view in FIG Fig. 24 shows a side view of the breaker 404. The bottom view shows a plan view of the breaker 404. A rotor 703 is shown in dashed lines in the upper illustration. The rotor 703 rotates at a rotational speed v. As in Fig. 24 ₁ , the prior art breaker 404 has a region d ₁ in which the breaker 404 completely encloses the rotor 703. In the region of the inlet 701 and in the region of the outlet 702, a side channel 704 ends abruptly. It comes here to disturbing sound components as well as to a gas flow at the discharge nozzle 702.

Fig. 25 shows the breaker 404 disposed in the stator 700. In the stator 700, an inlet 701 and an outlet 702 are arranged for the side channel 704. In the stator, a rotor 703 rotates at a speed v.

Again Fig. 25 can be seen in the upper part, the breaker 404 over a length d _{1 on} a region in which the rotor 703 is completely enclosed by the breaker 404. In a region over a length d ₂ , the breaker has a bevel 705. In the region of this bevel 705, the side channel 701 widens continuously to its total width outside the range d ₂ . Rotor blades 706 are arranged on the rotor 703, only shown schematically. The length d _{1 of} the breaker is greater than a blade length. Also the length d _{2 of} the bevel 705 is longer than a blade length.

The channel 701 may have a shape as shown in FIG Fig. 11 for the channel 401 is shown. The rotor 400 is bounded by a sealing surface 707 of the stator. This sealing surface 707 is arranged in the blade-less region of the rotor 400.

In the lower part of the Fig. 25 The chamfer 705 tapers in the direction of the area d _{2 of} the interrupter 404, in which the interrupter 404 completely surrounds the rotor 703. An angle β indicates the opening angle of the bevel 705. An angle α is a complementary angle to the angle β , that is, the sum of the angles α and β together make 180 °. The angle α corresponds to a blade angle of the rotor blades 706 of the rotor 703, as in FIG Fig. 26 shown.

In Fig. 26 are a rotor blade 706 in section and the angle of attack α shown. D denotes the blade height.

Fig. 27 FIG. 12 illustrates another embodiment of the invention. The breaker 404 formed in the stator 700 has the bevel 705. In the direction of the side channel 704, an additional bevel 706 is provided. This additional bevel, which has a length d ₃ , an even higher compression and a higher pumping speed are achieved.

In Fig. 28 is the compression of a side channel pumping stage shown. The curves show, on the one hand, the values for a standard breaker and, on the other hand, for a breaker form according to FIG Fig. 25 , It can be seen that the compression in accordance with the breaker shape Fig. 25 is increased.

According to Fig. 29 the suction capacity of a side channel pumping stage is shown. It can be seen clearly that the according to Fig. 25 used breaker form leads to a higher suction capacity than a prior art breaker form.

Fig. 30 shows the stator 700 with a side channel 704 and an outlet 702. The breaker 404 is adjacent to a surface 708 while maintaining a narrow gap (not shown) on blades of the rotor, which is also not shown here. The breaker has the bevel 705, which widens in the direction of the channel 704. A sealing surface 707 has a lower level than a surface 709 of the stator 700, resulting in the edge or surface 708. The bevel 705 represents on the one hand a radial opening of the interrupter 404 and also an axial recess of the sealing surface 707.

The stator 700 has a bore 710 for the passage of a shaft of the rotor (not shown).

reference numerals

1

casing

2

Gas inlet opening

4

gas outlet

6

wave

8th

drive system

10

bearing element

12

bearing element

14

pump unit

16

pump unit

16a

Pump subunit

16b

Pump subunit

16c

Pump subunit

18

pump unit

32

connecting channels

34a

fasteners

34b

fasteners

36a

fasteners

36b

fasteners

36c

fasteners

38

connecting channels

40

groove

41

surface

42

connecting line

50

thread groove

51

stator

52

coating

100

vacuum pump

101

casing

102

gas inlet

103

gas outlet

104

Side channel pump stage

105

molecular level

106

fore-vacuum

107

drive

108

wave

110

roller bearing

111

roller bearing

112

Kitchen sink

113

permanent magnet

120

housing parts

121

housing parts

122

housing parts

123

housing parts

124

Intake / transfer channel

125

Outlet / transfer channel

126

center line

400

Impeller / rotor

401

side channel

402

shovel

403

edge

404

breaker

420

channel bottom

421

Side wall of the side channel

422

Side wall of the side channel

423

blade base

424

Boundary surface of the side channel

425

midplane

426

Edge of the impeller / rotor

427

Edge of the impeller / rotor

428

Got over

500

rotor

502

cylinder

505

internal stator

506

outer stator

507

thread groove

508

thread groove

600

Wheel

601

side channel

602

shovel

700

stator

701

inlet

702

outlet

703

rotor

704

side channel

705

bevel

706

bevel

707

axial sealing surface of the rotor disk

708

area

709

area

710

drilling

711

Curve

712

Curve

713

Curve

714

Curve

715

Area

716

Curve

717

Curve

718

Curve

719

Curve

d ₁

length

d ₂

length

v

speed

α

Angle of attack rotor blade and additional angle

β

Opening angle bevel 705

A

arrows

B

arrows

R

angle

φ

radius

Claims (14)

Vacuum pumping stage of a threaded or side channel pump comprising a stator and at least one rotor, wherein in the stator and / or in the rotor at least one thread groove or at least one channel in the stator is provided, wherein the rotor with a rotor portion in the Dips in the channel and a pumping action is achieved by interaction of rotor section and channel,
characterized, in that the at least one channel or the at least one thread groove has at least one surface (41) in which grooves (40) are arranged, and / or that surfaces (41) of stators and / or rotors have grooves (40), - Or that the surface of the at least one channel or the at least one thread groove (50) and / or the surface of the rotor is coated with a material (52) having a lower coefficient of friction than a metal surface. Vacuum pumping stage according to claim 1, characterized in that a bottom surface and / or at least one side wall of the thread groove grooves (40). Vacuum pumping stage according to claim 1 or 2, characterized in that the grooves (40) are arranged in a pump-active surface along a flow direction of the gases, transversely to the flow direction of the gases and / or at an angle to the flow direction of the gases. Vacuum pumping stage according to one of the preceding claims, characterized in that the grooves (40) of two opposite pump-active surfaces are arranged parallel or at an angle to each other. Vacuum pumping stage according to one of the preceding claims, characterized in that the grooves (40) have a depth of between 1 μm and 100 μm. Vacuum pumping stage with an inlet, an outlet, a rotor and a channel having two side walls and a channel bottom, wherein the rotor with a rotor portion is immersed in the channel and a pumping action is achieved by interaction of the rotor portion and the channel, and with a between inlet and Outlet arranged breakers,
characterized in that at least one side wall (421, 422) of the channel (401) is curved. Vacuum pumping stage according to claim 6, characterized in that the curvature of the at least one side wall (421, 422) is convex. Vacuum pumping stage according to claim 6 or 7, characterized in that the channel (401) is formed axially symmetrical to a center plane (425) of the rotor (400). Vacuum pumping stage according to one of claims 6 to 8, characterized in that the curvature of the at least one side wall (421, 422) is formed in each case in cross-section semicircular. Vacuum pumping stage according to one of claims 6 to 9, characterized in that rotor blades (402) of the rotors (400) are formed in cross-section V-shaped. Vacuum pumping stage with an inlet, an outlet, a rotor and a channel, wherein the rotor with a rotor portion immersed in the channel and by cooperation of the rotor portion and the channel a pumping action is achieved, are arranged on the rotor rotor blades, and one between inlet and outlet having arranged breakers,
characterized in that the breaker (404) has a bevel (705) on a side facing the inlet (701). Vacuum pumping stage according to claim 11, characterized in that the chamfer (705) has an opening angle β of the chamfer (705) formed between the at least one chamfer (705) and a central axis, and that an additional angle α corresponding to the opening angle β Blade angle α of the rotor blades (706) corresponds. Vacuum pumping stage according to claim 11, characterized in that the breaker (404) has the bevel (705) only at the inlet (701). Vacuum pumping stage according to claim 1, 12 or 13, characterized in that the breaker (404) has a length d ₁ , which corresponds to at least one blade length.

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