EP2835536B1 - Vacuum pump stage with special surface roughness yielding a lower gas friction - Google Patents

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.

Another example of the prior art relates to a vacuum pumping stage having an inlet, an outlet, and a channel having two sidewalls and a channel bottom, wherein a rotor having a rotor portion is inserted into the channel and a pumping action is achieved by cooperation of the rotor portion and the channel , and with a breaker arranged between inlet and outlet.

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 prior art ( DE 33 17 868 A1 ) includes a friction pump in which surface areas with different roughness are present at least in a part of the pump-active surfaces, such that the roughness of the area facing away from the conveying direction surface areas is greater than the roughness of the surface areas facing the conveying direction.

This prior art friction pump can be further improved in terms of pumping action.

Furthermore belongs to the state of the art ( JP H 01 267390 A ) a side channel pump, in which a plurality of side channels are arranged, which are designed to interact with pump-active surfaces of the rotor. These too The prior art vacuum pump can be further improved in terms of pumping power.

In addition, the state of the art ( DE 39 32 288 A1 ) a turbo vacuum pump with a side channel pumping stage. This side channel pumping stage has a radially directed inlet. At an inner radius of the side channel of the inlet, a chamfer of the interrupter provided between inlet and outlet is arranged. This prior art vacuum pump can be further improved in terms of avoiding turbulence of the incoming gas.

Furthermore belongs to the state of the art ( US 2005/0118013 A1 ) a side channel pump, which also has a radially disposed inlet and a radially arranged outlet. Also, this pump can be further improved in terms of the pumping action.

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 in the vicinity of the rotor 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 wherein the at least one channel or the at least one thread groove has at least one surface in which grooves are arranged and / or wherein pumping surfaces of stators and / or rotors have grooves, wherein the grooves have a depth between 1 .mu.m and 100 .mu.m , wherein the grooves have a width and / or a distance between the grooves of between 1 .mu.m and 1 mm, is characterized in that the grooves of two opposite pump-active surfaces are arranged at an angle to each other.

The 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.

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 are formed according to the invention at an angle to each other.

According to the invention, the grooves have a width and / or a distance between the grooves of between 1 μm and 1 mm. In these Widths and distances a particularly advantageous effect is achieved.

According to the invention, the grooves have a depth of between 1 μm and 100 μm. 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.

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 Beispiel eines Querschnittes eines Seitenkanales;

Fig. 19

ein Beispiel eines Querschnittes eines Seitenkanales;

Fig. 20

ein Beispiel 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 Beispiel;

Fig. 23

ein Beispiel;

Fig. 24

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

Fig. 25

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

Fig. 26

eine Rotorschaufel in Seitenansicht zur Darstellung des Anstellwinkels a;

Fig. 27

ein Beispiel;

Fig. 28

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

Fig. 29

eine Darstellung des Saugvermögens einer Seitenkanalstufe mit Standardunterbrecher und mit geändertem 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 Examples of a vacuum pumping stage are shown. 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;

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

an example of a cross section of a side channel;

Fig. 19

an example of a cross section of a side channel;

Fig. 20

an example 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

an example;

Fig. 23

an example;

Fig. 24

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

Fig. 25

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

Fig. 26

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

Fig. 27

an example;

Fig. 28

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

Fig. 29

a representation of the pumping speed of a side channel stage with standard breaker and with changed 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 pump-active 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, 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 according to the invention must vary from 1 μm to 100 μm. The groove width, or the distance between the individual grooves 40, according to the invention, must 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 side channel pumping stage, in which case the geometry of the arranged on the impeller 600 and rotating in the side channel 601 blades 602 deviates from 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.

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

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

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. The measure R _{S 2} can be calculated with the following three equations: $${\mathrm{<figure-callout\; id="2"\; label="R\; S"\; filenames="imgf0001.png,imgf0016.png"\; state="\{\{state\}\}">R\; </figure-callout>}}_S={\mathrm{<figure-callout\; id="1"\; label="R\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">R\; </figure-callout>}}_S+\sqrt{R_{S3}^2-\frac{1}{4}{\left(d_{S2}-{\mathrm{<figure-callout\; id="1"\; label="d\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">d\; </figure-callout>}}_S\right)}^2}$$

$${\mathrm{<figure-callout\; id="2"\; label="d\; S"\; filenames="imgf0001.png,imgf0016.png"\; state="\{\{state\}\}">d\; </figure-callout>}}_S=d_{R1}+2\cdot \mathrm{<figure-callout\; id="1"\; label="\Delta \; R\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">\Delta \; </figure-callout>}$$

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

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

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

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

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 design of the side channel compared to 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.

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 up to 1000 Hz increases, the optimal distance is depending on the pressure range 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. $$A_{\mathit{SK}}=R_{S3}^2\cdot \left(\pi -\alpha \right)+{\mathrm{<figure-callout\; id="1"\; label="d\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">d\; </figure-callout>}}_S\cdot \left(R_{S3}+C\right)+C\cdot \frac{1}{2}\cdot \left(d_{S2}-{\mathrm{<figure-callout\; id="1"\; label="d\; S"\; filenames="imgf0009.png,imgf0010.png"\; state="\{\{state\}\}">d\; </figure-callout>}}_S\right)$$

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

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

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

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 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 concave side walls 421, 422. The channel bottom 420 is flat.

In the side channel configurations of the 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 another example 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 example, but consists of two circular sections with a radius R _{S. 3}

Fig. 23 shows another example of an embodiment of the side channel 401. The side channel 401 has curved side surfaces 421, 422 and a non-planar formed 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 configuration.

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 illustrates another example. The breaker 404 formed in the stator 700 has the taper 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 (3)

1. A vacuum pump stage of a screw thread or side channel pump having a stator (51) and at least one rotor, in which a thread channel (50) is provided in the stator (51) and/or in the rotor or at least one duct is provided in the stator (51), wherein the rotor extends into the duct by means of a rotor portion, and a pumping action is achieved by the interaction of the rotor portion and the duct,

   - wherein the least one duct or the at least one thread channel (50) has at least one surface (41) in which there are arranged grooves (40),

   - and/or wherein pump-active surfaces (41) of stators and/or rotors have grooves (40),

   - wherein the grooves (40) have a depth of between 1µm and 100µm.

   - wherein the grooves (40) have a width and/or a spacing of the grooves (40) from each other of between 1µm and 1mm,

   characterised in that the grooves (40) of two opposing, pump - active surfaces are arranged at an angle to each other.
2. A vacuum pump stage according to claim 1, characterised in that a bottom surface and/or at least one side wall of the thread channel has grooves (40).
3. A vacuum pump stage according to claim 1 or 2, characterised in that the grooves (40) in a pump-active surface are arranged along a flow direction of the gases, transverse to the flow direction of the gases and/or at an angle to the flow direction of the gases.

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