EP3088743B1 - Side-channel vacuum pump stage with a stripper that is slanted on the suction side - Google Patents

Description

The invention relates to a vacuum pump stage.

The prior art includes vacuum pump stages of threaded pumps, which essentially consist of two parts, namely a stator and a rotor rotating in the stator. Multi-start threads are attached to the outside diameter of the rotor and to the inside 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 pump stage, can be used in a multi-stage construction in the high pressure range up to atmospheric pressure. These can be easily combined with turbomolecular pumps or other molecular pumps, for example. The rotor parts of both pumps can be accommodated on a shaft, so that both form a structural unit. The side channel pump stages usually have an impeller, that is to say a rotor, which has blades running around its edge in a channel.

In order to achieve a sufficiently good pump performance in the pumps known from practice, several stages and elaborately designed impellers, for example the side channel pump stage, are generally necessary.

A further embodiment relates to a vacuum pump stage with an inlet, an outlet and a channel, which has two side walls and a channel bottom, wherein a rotor with a rotor section is immersed in the channel and a pumping action is achieved by the interaction of the rotor section and channel, and with a between Inlet and outlet arranged breaker.

Many industrial processes run under vacuum conditions in the molecular flow range. Vacuum pumps or vacuum pumping stations composed of vacuum pumps are used to generate such vacuum conditions. Vacuum pump stages are used in the vacuum pumps according to different operating principles, which are adapted to different pressure ranges in order to compress gas from the desired final vacuum to the atmosphere.

Side channel pumping stages are used to compress the atmosphere. In these, blades circulate in a channel and promote a vortex-like gas flow between the inlet and outlet. The gas flow follows the blades as it circulates and is used on a wiper or breaker detached and fed to the outlet.

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

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

These side channel pump stages belonging to the prior art can still be improved with regard to their pumping performance.

State of the art ( DE 33 17 868 A1 ) belongs to a friction pump, in which at least some of the pump-active surfaces have surface areas with different roughness, such that the roughness of the surface areas facing away from the conveying direction is greater than the roughness of the surface areas facing towards the conveying direction.

This friction pump, which belongs to the prior art, can be further improved with regard to the pumping effect.

Furthermore, the state of the art ( JP H 01 267390 A ) a side channel pump, in which several side channels are arranged, which are designed to interact with pump-active surfaces of the rotor. This too state-of-the-art vacuum pump can be further improved in terms of pump performance.

In addition, the state of the art ( DE 39 32 288 A1 ) a turbo vacuum pump with a side channel pump stage. This side channel pump stage has an inlet oriented in the radial direction. A bevel of the interrupter provided between the inlet and the outlet is arranged on an inner radius of the side channel of the inlet. This vacuum pump, which belongs to the state of the art, can be further improved with regard to avoiding turbulence of the inflowing gas.

Furthermore, the state of the art ( US 2005/0118013 A1 ) a side channel pump, which also has an inlet arranged in the radial direction and an outlet arranged in the radial direction. This pump can also be further improved with regard to the pumping effect.

The invention relates to a vacuum pump stage in which an interrupter is arranged between the inlet and the outlet.

State of the art ( DE 103 34 950 A1 ) belongs to a side channel compressor, which has an inlet, an outlet and a rotor, and a channel, the rotor being immersed in the channel with a rotor section and a pumping effect being achieved by the interaction of the rotor section and channel. The rotor usually dips into the channel with rotor blades arranged on the rotor. An interrupter is arranged between the inlet and the outlet. The interrupter encloses the rotor on all sides and, as is known from practice, abruptly in the vicinity of the Outlet where the side channel ends, as well as near the inlet where the side channel begins.

According to the state of the art ( DE 103 34 950 A1 ) the interrupter is designed in such a way that the rotor blades are increasingly enclosed in a uniform manner or are released again in a uniformly decreasing manner. The respective rotor blade is thus gradually and continuously enclosed by the interrupter, or continuously released again. This does not result in an abrupt, but a continuous and even stripping of the compressed gas components from the respective rotor blades. This measure is implemented at the beginning as well as at the end of the breaker, that is, at the inlet and at the outlet. As a result, the formation of disturbing sound components in the interrupter area is suppressed and gas congestion at the pressure port is reduced. This leads to an increase in efficiency. This embodiment, which belongs to the prior art, has the disadvantage that the efficiency has not yet been fully exhausted.

State of the art ( EP 1 541 871 A1 ) includes a side channel pump stage with rotor elements and stator elements attached to a shaft, which are connected to the fixed parts of the pump. Concentric channels are incorporated in the stator element. The blades of the rotor disc engage in these channels and, during rotation, promote the gas. There are barriers in the channels, which have the effect that the gas is not continuously conveyed in a circle, but rather leaves the channels and is conveyed into the channels adjacent in the axial direction. This side channel pump stage, which belongs to the prior art, can be improved in terms of its performance.

Furthermore, the state of the art ( DE 691 04 749 T2 ) a turbomolecular pump with a side channel pump stage. In this side channel pump stage, too, a baffle is provided, which closes the channel between an intake opening and an outflow opening. This pump can also be improved in terms of its performance.

In addition, the state of the art ( EP 0 767 308 A1 ) a side channel compressor with at least one side channel formed in the compressor housing and an interrupter separating the suction and pressure side, the boundary sides of which extend in the radial direction with respect to the side channel have a course deviating from the axial direction of the compressor, in which the boundary sides of the interrupter over their full radial height are provided with an at least approximately heart-shaped incision, and in which the incisions provided on the suction and pressure sides point towards one another with their tips. This side channel compressor, which belongs to the prior art, can also be further improved in terms of performance.

State of the art ( DE 24 09 183 A1 ) also includes a ring compressor, in which an interrupter is provided between an inlet and an outlet opening of the side channel. A pocket-shaped nozzle is provided on the interrupter. The gas in the blade cells, which is transported via the interrupter, enters this and relaxes in the nozzle. This ring compressor, which belongs to the prior art, can also be further improved in terms of performance.

The state of the art also includes ( DE 103 34 950 A1 ) a side channel blower and a Method of operating a side channel blower. This side channel compressor, which belongs to the prior art, has an impeller with blades arranged on the circumference of the impeller. This side channel compressor, which belongs to the prior art, can be further improved with regard to the pump performance.

The technical problem on which the invention is based is to provide an improved vacuum pump stage for side channel pumps which are used in molecular and viscous pressure ranges in order to achieve an increase in performance of the pump.

The technical problem is solved by a vacuum pump stage with the features according to claim 1.

The vacuum pump stage according to the invention is defined by claim 1.

The rotor has a base body and rotor blades arranged on the base body. The rotor is arranged in the side channel pumping stage in such a way that two axial sealing surfaces of the side channel pumping stage are assigned to the base body of the rotor. These axial sealing surfaces expand into the so-called side channel in which the rotor blades rotate. The interrupter interrupts the side channel between the inlet and outlet. The bevel of the interrupter 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 pump stage, seen in the axial direction of the pump stage. This means that the bevel is designed to run into the side channel.

It is advantageous if the bevel in the area of the interrupter is only provided at the inlet. This increases the efficiency of the pump stage compared to the state of the Technology in which the bevel is additionally arranged in the area of the outlet.

According to a particularly advantageous embodiment, which is not part of the invention, it is provided that the bevel has an opening angle β of the bevel that is formed between the at least one bevel and a central axis, and that one with the opening angle β corresponding supplementary angle α corresponds to a blade angle α of the rotor blades.

This represents a further significant innovation to the prior art. This measure ensures that the vortex flow is formed immediately after the inlet, that is to say immediately after the outlet from the interrupter, and not only later in the side channel.

When the rotor blades leave a standard breaker that has an abrupt start of the side channel, the compressed gas flows out of the blades into the side channel in all directions. As a result, the formation of the vortex flow is not favored, but rather adversely affected. In an interrupter designed according to the invention, the gas only flows in the direction in which the bevel is arranged, since the channel surrounds the rotor blades on the other sides. In the side channel, the gas flows into this area in the area of the bevel and is sucked in by the rotor blades on the sides and expelled again on the side opposite the bevel. As a result, a vortex flow is formed immediately after the rotor blades emerge from the interrupter, which has a very advantageous effect on the compression and the pumping speed. The effect of the vortex flow is optimized by the formation of the complementary angle β and thus the opening angle α of the bevel, analogous to the angle of attack of the rotor blades.

The gas can even reach a supersonic speed at the outlet of the interrupter. The interrupter shape according to the invention acts like a Laval nozzle. The gas in the blades expands towards the bevel and becomes cooled. The gas pressure at the inlet and in particular the gas pressure in the rotor blades decrease. This favors gas intake on the sides of the rotor blades. Due to the ordered gas flow above the rotor blades due to the shape of the interrupter according to the invention, a rapid formation of an ordered vortex flow is achieved immediately after the interrupter. 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, which, however, is not part of the invention, the interrupter has the bevel only at the inlet. This means that the interrupter at the outlet area has the abrupt end known from the prior art. To increase the pumping speed and the compression of the pump, it is sufficient or even beneficial to only bevel the channel at the inlet area.

According to a further advantageous embodiment, which, however, is not part of the invention, the interrupter has a length d ₁ which corresponds to at least one blade length. By interrupter is meant the area that has a reduced cross section compared to the channel.

It has turned out 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 ensures optimal pumping speed and optimal compression.

The invention optionally provides that the at least one bevel is linear is trained. This optimally favors the formation of the vortex flow.

Another advantageous embodiment, which is not part of the invention, provides that the channel has a round or angular cross section or an angular cross section with rounded corners. With both cross-sectional shapes, the chamfer towards the inlet is advantageous in order to optimize the pump performance.

According to a further advantageous embodiment, which is not part of the invention, the channel has at least one side wall and the at least one side wall of the channel is curved. This design of the side channel is particularly advantageous.

According to a further advantageous embodiment, which, however, is not part of the invention, the bevel is formed in the radial direction of the vacuum pump stage during the transition into the side channel over the entire width of the channel. This also has a positive influence on the formation of the vortex flow.

Another advantageous embodiment, which is however not part of the invention, provides that the bevel in the radial direction of the vacuum pump stage has only a part of the entire width of the channel during the transition into the side channel.

According to a further advantageous embodiment, which is not part of the invention, the bevel has a maximum depth that corresponds to the axial depth of the channel. This embodiment allows a good formation of the vortex flow.

According to another advantageous embodiment, which, however, is not part of the invention, the bevel has a maximum depth up to an axial sealing surface of the rotor disk in the area without blades. This embodiment also allows a sufficient formation of the vortex 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 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 Unterbrecher in Seitenansicht und in Draufsicht (schematisch) ;

Fig. 26

eine Rotorschaufel in Seitenansicht zur Darstellung des Anstellwinkels a;

Fig. 27

ein geändertes Ausführungsbeispiel;

Fig. 28

eine Darstellung einer Kompression einer Seitenkanalstufe mit Standardunterbrecher und mit einem hier gezeigten Unterbrecher;

Fig. 29

eine Darstellung des Saugvermögens einer Seitenkanalstufe mit Standardunterbrecher und mit einem hier gezeigten Unterbrecher;

Fig. 30

eine Statorscheibe mit einem erfindungsgemäßen Unterbrecher in axialer Draufsicht.

Further features and advantages emerge from the associated drawing, in which several exemplary embodiments of a vacuum pump stage are only shown by way of example, and the invention in Fig. 30 is illustrated. The drawing shows:

Fig. 1

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

Fig. 10

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

Fig. 11

a partial cross section through a side channel;

Fig. 12

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

Fig. 13

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

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;

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;

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;

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 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

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

Fig. 22

a modified embodiment;

Fig. 23

a modified embodiment;

Fig. 24

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

Fig. 25

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

Fig. 26

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

Fig. 27

a modified embodiment;

Fig. 28

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

Fig. 29

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

Fig. 30

a stator with an interrupter according to the invention in axial plan view.

Fig. 1 shows a vacuum pump with a housing 1 and three pump 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 stationary gas-producing components. The rotating components are mounted one behind the other on a shaft 6 in the axial direction. The operation of the shaft 6 includes a drive system 8 and bearing elements 10 and 12. The fixed components are firmly connected to the housing 1.

A pump unit 14 facing the gas inlet opening is designed as a turbomolecular pump. The pump unit 16 following in the direction of the gas flow consists of several sub-units 16a, 16b, 16c. These each have one or more molecular pump stages of the Gaede type, hereinafter referred to as Gaede stages. The Gaede stages are connected in parallel within the subunits. The subunits themselves are connected in series. This means that connecting elements 34a for subunit 16a, or 34b for subunit 16b, connect the input sides and on the other side the output sides of the Gaede stages in such a way that a parallel gas flow in the individual subunits is made possible. The subunits are connected by connecting elements 36a, 36b and 36c such that the output side of one subunit is connected to the input side of the following subunit. The pump unit 18 facing the gas outlet opening is designed as a multi-stage side channel pump. In the Fig. 1 The pump shown is only shown as an example.

The invention relates to all vacuum pumps in which side channel pump stages are provided.

It is provided, for example, 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.

The 2 to 6 show possible structures that are evenly mounted in a surface 41, for example a thread groove of a side channel or on a rotor.

Fig. 2 shows a structure with grooves 40 which have a rounded bottom. The grooves 40 are arcuate. Fig. 3 shows a trapezoidal structure with a tapered cross section while Fig. 4 shows a triangular structure with a tapered cross-section. In Fig. 5 a rectangular structure is shown. Fig. 6 again shows a triangular structure, which has an asymmetrical configuration.

The depth of the grooves 40 can 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 can be machined into the surface 41 along the direction of flow, transverse to the direction of flow and at an angle to the direction of flow of the gas.

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

According to Fig. 8 a thread groove 50 of a thread pump is shown. The thread groove 50, which is arranged, for example, in a stator 51, as well as the adjacent 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. This measure also reduces the gas friction on the channel surface, which gives rise to the advantages mentioned above.

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 constructed from four housing parts 120, 121, 122, 123, which accommodate the components of the vacuum pump 100.

Gas entering the vacuum pump 100 through the gas inlet 102 first reaches a molecular stage 105. This has an inner stator 505, which is provided with an inner thread groove 507, and an outer stator 506, which is provided with an outer thread groove 508. A cylinder 502 with a smooth surface is provided between the inner stator and the outer stator and is connected to the rotor 500. The molecular stage 105 is thus designed as a Holweck stage. In the Fig. 9 Holweck stage shown is constructed symmetrically with a second cylinder 502 'surrounded by stator components and therefore works in two stages.

The rotor is connected to a shaft 108 which is rotatably mounted in roller bearings 110 and 111. Instead of the roller bearings 110, 111, passive and active magnetic bearings can also be used. At least one permanent magnet 113 is arranged on the shaft 108, which interacts with a stationary coil 112 and forms a drive 107 together with the latter. The roller 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 contains a side channel pump stage 104. The side channel pump stage 104 is formed by a side channel 401 and an impeller 400, at least one blade 402 being arranged on the impeller 400, 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 a further transfer channel 125.

From the side channel pump stage 104, the gas passes through the transfer channel 125 into a forevacuum stage 106 likewise designed as a side channel pump stage, the geometry of the blades 602 arranged on the impeller 600 and rotating in the side channel 601 deviating from the geometry of the blades 402 here. From this pump stage 106, the gas from the vacuum pump 100 is expelled through the gas outlet 103.

There are narrow gaps between the impellers 400 and 600 and the housing parts 121, 122 and 123. These allow the impeller in question to rotate freely, but are so narrow that no disruptive gas flows occur.

Fig. 10 shows a section through the housing part 122 along the line II of Fig. 9 , The impeller 400 is seated on the shaft 108. This has an edge 403, on which blades 402 are evenly distributed along the circumference. The side channel 401 surrounds the impeller, the side channel essentially radially surrounding the blade area of the impeller in the radial direction. Only a part of the circumference is close to the impeller. This section forms an interrupter 404, which separates the suction and discharge sides from one another, and at which the gas stream 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 illustrated, the side channel 401 has a channel base 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 protrude completely into the side channel 401. A radius R _{S 1 of} a blade base 423 is the same size as the radius R _{S 1} radially in the direction of the Shaft arranged boundary surface 424 of the side channel 401.

This means that the blades 402 are completely immersed in the side channel 401.

The pumping performance of the side channel pump stage is significantly improved by the curved side surfaces 421, 422. The web between the blades is advantageously made as small as possible (not shown). The bucket volume filled with gas should be as large as possible.

These measures significantly improve the vacuum properties of the pump.

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

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

Improvements in the vacuum data are also achieved by an optimized setting of 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 optimal radius R _{S 3} and distance d _{S 1} depend on the circumferential 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. The dimension R _{S 2} can be calculated using the following three equations: $${\mathrm{<figure-callout\; id="2"\; label="R\; S"\; filenames="imgf0001.png"\; state="\{\{state\}\}">R\; </figure-callout>}}_S={\mathrm{<figure-callout\; id="1"\; label="R\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">R\; </figure-callout>}}_S+\sqrt{R_{S3}^2-\frac{1}{4}{\left(d_{S2}-{\mathrm{<figure-callout\; id="1"\; label="d\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">d\; </figure-callout>}}_S\right)}^2}$$

$${\mathrm{<figure-callout\; id="2"\; label="d\; S"\; filenames="imgf0001.png"\; state="\{\{state\}\}">d\; </figure-callout>}}_S=d_{R1}+2\cdot \mathrm{<figure-callout\; id="1"\; label="\Delta \; R\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">\Delta \; </figure-callout>}$$

$$R_S$$$$1_{}=R_{R1}-H$$

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

$$\mathrm{\Delta }\le \mathrm{0,2}\mathrm{mm}\mathrm{für}10\mathrm{mbar}<{\mathit{p}}_2\le 100\mathrm{mbar}$$

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

Δ denotes the axial gap between the rotor and the stator disk. The axial gap Δ can preferably be from 0.01 mm to 0.5 mm. Small axial gaps on the exhaust side and large axial gaps on the intake side make sense. If a labyrinth seal is used on the axial surface between the rotor and stator disks, the axial gap can be more than 0.5 mm. The guide values for the axial gaps can be selected as follows: $$\mathrm{\Delta }\le 0.3\mathrm{mm}\mathrm{For}{\mathit{<figure-callout\; id="2"\; label="p"\; filenames="imgf0001.png"\; state="\{\{state\}\}">p</figure-callout>}}_2\le 10\mathrm{mbar}$$

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

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

In Fig. 12 is a comparison of side channels rectangular in cross section and side channels with two side walls semi-circular in cross section with V-shaped rotor blades at 800 Hz and 1000 Hz rotational frequency in comparison. 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 rotational frequency of 800 Hz. A side channel with semicircular side walls has a higher compression (curve 718) on as a prior art cross-sectionally rectangular channel (curve 719). The two upper curves 716, 717 relate to a rotational frequency of 1000 Hz. The upper curve 716 represents the compression as a function of the pressure for a side channel with side walls which are semicircular in cross section. Here too, the compression due to the design of the side channel shown here is significantly increased compared to a side channel with a rectangular cross section (curve 717). It can be spotted, that the side channels with two side walls that are semicircular in cross section have a significantly 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 recorded. The compression factor k ₀ is greater, the smaller the axial gap.

Different rotor disks of a multi-stage side channel pump with the same blade size have the same speed, but can have different peripheral speeds depending on the rotor disk diameter R _{R 1} . 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 the optimal side channel radius R _{S 3} and the distance d _{S 1} increase with increasing speed and consequently increasing peripheral speed of rotor disks. The side channel size with the best compression factor is said to be optimal. The pumping speed and power consumption increase proportionally to the side channel area.

In the 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 one in the 14a and 14b The compression factor shown is given for the following example:
For a rotor disc 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, 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 optimal side channel radius R _{S 3 is optimal} = 5.3 mm. With increasing speed f and peripheral speed V, the optimal side channel radius will continue to increase, or decrease with decreasing rotational frequency and peripheral speed.

In the 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 optimal distance d _{S 1} at a speed of f = 800 Hz, depending on the pressure range, is either d _{S 1} = 1.2 mm or d _{S 1} = 3.6 mm. If the speed f increases up to 1000 Hz, the optimal distance, depending on the pressure range, will be either d _{S 1} = 3.6 mm or d _{S 1} = 4.8 mm. A tendency to increase the optimal distance d _{S 1} with increasing speed f can be seen.

The above dependencies only apply to rotor disks with V-shaped blades, as described in Fig. 16 are shown. Fig. 16 shows the impeller 400 with the blades 402. The blades 402 are V-shaped. The blade base has a projection in the area of a central plane 425 of the impeller 400, which protrudes from the edges 426, 427 of the blade base and tapers to the central plane 425. The impeller 400 rotates in the direction of arrow A.

Fig. 17 shows the impeller 400 according to FIG 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. A projection 428 protrudes above the blade base 423.

In general, the following design guidelines should be observed when designing side channel pumps. An optimal blade height is 60% to 100% of the rotor disc width. An optimal side channel radius depends on the circumferential 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 can 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 optimal 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 50% to 100% of the rotor disk width, the optimal distance between the blades is less than or equal to 55% for small side channels (side channel area is not greater than 2.5 times the blade area) and is greater than or equal to 85% for large side channels (side channel area not less than 5 times the blade area). The optimal number of blades becomes smaller as the side channels increase, or the optimal distance between blades increases. The side channel area A _SK and the blade area A _Sch can be calculated using equations 4 to 7. $$A_{\mathit{SK}}=R_{S3}^2\cdot \left(\pi -\alpha \right)+{\mathrm{<figure-callout\; id="1"\; label="d\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">d\; </figure-callout>}}_S\cdot \left(R_{S3}+C\right)+C\cdot \frac{1}{2}\cdot \left(d_{S2}-{\mathrm{<figure-callout\; id="1"\; label="d\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">d\; </figure-callout>}}_S\right)$$

$$\alpha =\arcsin \left(\frac{d_{S2}-{\mathrm{<figure-callout\; id="1"\; label="d\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">d\; </figure-callout>}}_S}{2\cdot R_{S3}}\right)$$

$$C=\sqrt{R_{S3}^2-{\left(\frac{d_{S2}-{\mathrm{<figure-callout\; id="1"\; label="d\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">d\; </figure-callout>}}_S}{2}\right)}^2}$$

$$A_{\mathit{Sch}}=d_{R1}\cdot \left(R_{R1}-{\mathrm{<figure-callout\; id="1"\; label="R\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">R\; </figure-callout>}}_S\right)$$

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 disc.

The 18 to 20 show further design options of a side channel. In Fig. 18 the side channel 401 is generally circular. The side channel 401 does not have a flat side channel bottom, but rather an overall 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 concave side walls 421, 422. The channel floor 420 is flat.

In the embodiments of the side channels of the 18 to 20 the side channel cross-sectional diameter is advantageously constant over the entire circumference of the side channel. There is also a possibility that the side channel cross-sectional diameter may decrease from an inlet 124 to an outlet 125. According to Fig. 9 are the entrance 124 and the outlet 125 arranged diametrically opposite. However, an arrangement in a side channel pump stage, as shown in FIG Fig. 10 has been drawn in dashed lines. An inlet 124 'is drawn here. With 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 take place linearly with the circumferential angle. It can also represent another function of the circumferential angle.

In the 21a to 21c a side channel surface having a center line 126 of the side passage as a function of radius and φ from the angle.

The reduction in the side channel area can, as in Fig. 21a shown, done from above. It can also be done from below, as shown Fig. 21b shown. However, it can also be done from above and from below, as shown Fig. 21c shown. The side channel diameter can also be reduced from one or both sides along the side channel from inlet 124 'to 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 designed in the form of a circular section. The channel floor 420 is also not shown flat in this exemplary 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 one not flat channel floor 420 on. The curved side surfaces 421, 422 do not correspond to circular sections in this case.

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

Fig. 24 shows a prior art breaker 404 having an inlet 701 and an outlet 702. The interrupter 404 as well as the inlet 701 and the outlet 702 are part of a stator 700. The upper illustration in FIG Fig. 24 shows a side view of the interrupter 404. The lower illustration shows a top view of the interrupter 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 can be seen, the interrupter 404 belonging to the prior art has a region d ₁ in which the interrupter 404 completely surrounds the rotor 703. In the area of the inlet 701 and in the area of the outlet 702, a side channel 704 ends abruptly. There are disturbing sound components and a gas jam at the pressure port 702.

Fig. 25 10 shows the breaker 404 located in the stator 700. An inlet 701 and an outlet 702 are arranged in the stator 700 for the side channel 704. A rotor 703 rotates in the stator at a speed v.

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

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

In the lower part of the Fig. 25 the interrupter 404 is shown with the bevel 705. The bevel 705 tapers in the direction of the region 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 results in 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 A rotor blade 706 is shown in section and the angle of attack α. D is the blade height.

Fig. 27 represents a further embodiment. The interrupter 404, which is formed in the stator 700, has the bevel 705. An additional bevel 706 is provided in the direction of the side channel 704. This additional bevel, which has a length d ₃ , results in an even higher compression and a higher pumping speed.

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

According to Fig. 29 the pumping speed of a side channel pump stage is shown. One can clearly see that the according Fig. 25 used breaker shape leads to a higher pumping speed than a breaker shape belonging to the prior art.

Fig. 30 corresponds to the invention and shows the stator disk 700 with a side channel 704 and an outlet 702. The interrupter 404 borders with a surface 708 while maintaining a narrow gap (not shown) on blades of the rotor, which is also not shown here. The interrupter 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, which results in the edge or surface 708. The bevel 705 firstly represents a radial opening of the interrupter 404 and also an axial depression 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 disc

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 (2)

1. A vacuum pump stage with an inlet (701), an outlet (702), a rotor (703) and a channel (704), wherein a portion of the rotor (703) extends into the channel (704) and a pumping action is achieved by the interaction of the rotor portion and the channel (704), in which rotor blades are arranged on the rotor (703), and the stage comprising a stripper (404) arranged between the inlet (701) and the outlet (702), wherein the stripper (404) has a bevel (705) which is on a side facing the inlet (701) and which runs into the side channel (704), wherein the bevel (705) is arranged on an outer radius of the side channel (704),
   characterised in that
   the bevel (705) extends up to a side of the stripper facing the outlet (702).
2. A vacuum pump stage according to claim 1, characterised in that the bevel is configured to be linear.

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