JP7339973B2 - screw vacuum pump - Google Patents

Description

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a screw vacuum pump having a housing and two threaded screw rotors disposed within the housing and engaged with each other. The screw rotor interacts with the housing to repeatedly form a closed process gas delivery volume for carrying the process gas and carries it towards the outlet.

Screw vacuum pumps generally have a complex pumping geometry, in particular a screw profile of a screw rotor and a corresponding housing for the screw rotor. The interaction of these shapes demands high precision in their manufacture. It is therefore extremely cumbersome and expensive to manufacture or control such shapes.

The object of the present invention is to reduce the required manufacturing precision for at least partial regions of the pumping shape.

This problem is solved by a screw pump having the features of claim 1, in particular wherein the screw rotor each has at least two adjacent portions along the screw axis, the screw rotor comprising: Each has a pitch that is at least essentially constant in a first portion located near the inlet, and a smaller pitch in the second portion than the first portion, and where , the first part with respect to the screw axis is longer than the closed conveying volume within the first part.

At the inlet of a screw vacuum pump, each screw profile interacts with the housing and surrounds the conveying volume. If the conveying volume to be closed reaches the second part before it is closed to the inlet, this leads to the process gas being partially discharged back towards the inlet. Hitherto, therefore, high manufacturing precision of the housing in the inlet region was required. This is to accurately define and close the transport volume. Accurate manufacturing of the inlet region is cumbersome, costly, and made difficult by, for example, the wear of the mold for the housing over multiple casting cycles.

Since the first part is longer than the closed transport volume according to the invention, it is ensured that the second part cannot be reached before the transport volume is closed. The manufacturing accuracy of the housing in the region of the inlet is thus reduced, without the risk of backflow of process gas into the inlet.

Due to the reduced manufacturing precision requirements, the manufacture of the vacuum pump is significantly simplified. Therefore, a stable manufacturing process with few erroneous members is possible. This is therefore particularly advantageous as the housing is often difficult to measure in the inlet area and can only be measured with a particularly cumbersome 3D coordinate measuring system. Without measurement, the discrepancies may be discovered only during the final inspection of the screw vacuum pump. This leads to high costs for reassembly in the event of an error. Deviations from the target shape are insignificant or harmless because the invention reduces the impact of manufacturing accuracy.

The invention simplifies not only the manufacture of the housing, but also the manufacture of the screw rotor. This is because larger tools and/or tools with a higher cutting volume can be used for profile production within a relatively long first portion with a relatively large pitch according to the invention. This reduces processing time and associated manufacturing costs. The machining time is furthermore determined, for example, by the respective screw profile of the screw rotor only in the first part, or even in the first part, already during casting, at least with a pre-profile for further machining. It can be reduced by being formed.

It has been shown that an efficient screw profile can be formed under a comparable overall length of the rotor, despite the extended first section. The advantages in terms of manufacturing technology thus far outweigh the limitations of vacuum technology, so that efficient pumps can be manufactured in a particularly simple manner according to the invention.

Furthermore, it has surprisingly been shown that the screw vacuum pump according to the invention has improved suction performance even at relatively high suction pressures. This is due to the fact that the longer first portions also result in longer gaps in the first portions that are formed between the rotors and especially to the housing. Process gas in the region of internal compression occurs in the case of high suction pressures at extra high internal pressures, which means that a longer gap must be crossed for backflow. Backflow is consequently difficult or better sealed. The screw vacuum pump according to the invention is thus not only particularly simple to manufacture given the limitations of possible settings, but also has an even better vacuum-technical performance in a given pressure range.

In general, a closed conveying volume is understood to be an area which is closed except for the inevitable clearances between the rotors and to the housing. The area of this closed conveying volume extends around the rotor along the screw profile between the two threads of each rotor, and outwardly from the housing and along the screw profile of each other. Limited by the rotor. The conveying volume transitions in its profile from inlet to outlet along each screw rotor during the pumping process. This provides a pumping action. The conveying volume can be small in the path between the inlet and outlet. This is called internal compression.

Specifically, the closed conveying volume wraps 360 degrees around the screw axis of each rotor.

In one embodiment, the first portion corresponds to at least 1.25 times, especially at least 1.5 times, especially at least 1.75 times, especially at least 2 times the length of the transport volume. The advantages mentioned above are correspondingly enhanced thereby.

In the suction area of the pump, it is possible, for example, to form an end face for a screw rotor. It is shaped to a screw rotor so that each delivery volume can be opened for aspiration and closed for delivery.

The end faces in particular run at least essentially at right angles to the screw axis. The end face, for example, forms the front face, in particular the interior, of the housing. The end face can preferably be designed as a free-form surface and/or at least essentially run parallel to at least one corresponding screw profile. In particular, the end face can be provided with two regions each parallel to the screw profile. These are joined to a raised, especially lower region and/or a lowered, especially upper region, in particular with respect to the screw axis.

The end face can generally be formed in the housing, for example. The housing can be formed, for example, as a cast part. In one embodiment, the end face is formed in the cast part, particularly in the housing, and is unmachined. This completely eliminates the production step of machining the end faces, which further simplifies the production of the housing.

In another embodiment, the screw profile of each screw rotor is formed by a cycloid. Such a screw profile can advantageously be formed as required.

In one development, the screw profile of each screw rotor is double-threaded. A particularly low imbalance of the corresponding screw rotor is thereby achieved. Thus compensating elements (such as compensating masses, which require additional structural space) and/or compensating holes (in which masses can be stored) are omitted. Is possible. On the other hand, higher pressures can be enabled.

In another embodiment, the screw rotors have an at least essentially constant pitch within each second portion. Such screw rotors can be set and manufactured particularly simply. This is because a constant pitch forms a relatively simple shape.

It may be provided that the screw rotors each have at least one third portion and the pitch is smaller in the third portion than in the second portion. Additional internal compression can thereby be achieved.

The pitch in the third portion can in particular be at least essentially constant. In another embodiment, each screw rotor has sections of different pitch over its entire pumping length. The pitch is then constant in all parts, in particular at least regionally or in subparts, in particular respectively. Both of these contribute to easy set-up and manufacturing.

Between two adjacent sections of constant pitch, the pitch of the screw profile varies along the screw axis essentially only in the transition regions between the sections. The transition region is in particular smaller than one portion, in particular adjacent portions. Especially all transition areas are smaller than all parts.

The screw vacuum pump is for example internal compression with a compression ratio of less than 5:1, in particular less than 4:1, in particular less than 3.5:1 and/or greater than 2:1, in particular greater than 3:1. It is possible to be Each screw profile forms or conveys, for example, more than 7, in particular more than 10, in particular more than 12, in particular more than 13, closed conveying volumes. The ratio of the length of the screw profile of each screw rotor to its diameter is in particular at least 2.0, in particular at least 2.5, in particular at least 3.0 and/or at most 5.0, in particular at most 4.0 be.

The invention is explained below, purely by way of example, on the basis of simplified drawings.

Perspective view of screw type vacuum pump A plan view of the screw type vacuum pump in FIG. FIG. 2 is a side view of the screw vacuum pump of FIGS. 1 and 2; FIG. 3 is a cross-sectional view of the screw vacuum pump along line AA shown in FIG. Immersion cooler for the screw vacuum pump of Figures 1 to 4 5 is a cross-sectional perspective view of the suction region of the screw vacuum pump of FIGS. 1-4; FIG. Side view of the screw type vacuum pump of FIG. Perspective view of screw vacuum pump housing Diagram showing two suction performance curves of screw type pumps with machined and unmachined end faces in one chart.

A screw vacuum pump 10 is shown in FIGS. It has a motor 12 , a gearbox 14 , a housing 16 , a bearing shield 18 and a cover 20 . The screw vacuum pump 10 conveys process gas from an inlet 22 to a downward directed outlet 24 visible in FIG.

Active fluid cooling is provided for the motor 12 . It emerges from the motor 12 housing. Active fluid cooling is likewise provided for the screw rotors 28 and 30 which are provided inside the housing 16 and which are visible in FIG. It has two cooling lines. These are not represented in FIG. However, its extension is represented by corresponding grooves 32 in housing 16 . Cooling pipes are fitted in these grooves. Further active fluid cooling is provided in the gearbox 14 and in the cover 20 and here each formed as an immersion cooler 34 . These are explained in detail below on the basis of FIG.

As can be seen in FIGS. 1-4, the housing 16 of the screw vacuum pump 10 has a waist 36 . A constriction 36 is provided in the area of the outlet 24 .

FIG. 4 shows the screw vacuum pump 10 in cross section. The section corresponds to line AA in FIG. Two screw rotors 28 and 30 can be seen. These each have a double threaded, inwardly and outwardly meshing screw-like profile 38 and 40, respectively. They are built with a cycloidal profile and have a cylindrical basic shape with a screw-like base and a cylindrical sleeve profile. The screw profiles 38 and 40 interact with the housing 16 and form the pumping areas of the screw vacuum pump 10 . This then repeatedly conveys the closed conveying volume of process gas from inlet 22 to outlet 24, ie from left to right in FIG.

The pump performance of the screw vacuum pump 10 depends on the size and aspect of various gaps in the pumping area. This is unavoidable due to the relative motion of rotors 28, 30 and housing 16, but should be kept small and as constant as possible for good pump performance. Temperature changes in the members involved lead to their deformation. The measures described here for avoiding, dissipating and generally suppressing heat within the pump 10 thus achieve the least possible deformations and thus the most controllable clearances possible. The clearance can thus be precisely designed, which improves the pump performance or its efficiency.

The screw rotor 28 is driven by the motor 12 directly, i.e. without an intermediate connection. On the other hand, the screw rotor 30 is driven at a predetermined angular ratio with respect to the screw rotor 28 by the gear 43 via the synchronous gear 42 .

Motor 12 has a housing 44 . The housing is made, for example, of aluminum and has formed therein cooling lines 26 for the active fluid cooling. Motor 12 also includes a wound stator 46 . This stator together with the magnet carrier 48 provided at the shaft end of the screw rotor 28 forms the electric motor and direct drive for the screw rotor 28 . A screw rotor 28 forms the rotor of the motor 12 . Magnet carrier 48 has a plurality of permanent magnets. The motor 12 thus forms a permanent magnet synchronous machine with integrated multiple magnets, also referred to as IPMSM.

The stator 46 is arranged in a casting 50 which insulates the non-detailed electrical conductors in the stator 46 and guides them to the substrate 52 in an insulating manner. The casting 50 forms a vacuum-tight connection of the motor 12 together with the substrate 52 here. This connection is to the control electronics provided in the area of atmospheric pressure. For example, an external frequency converter for motor 12 could be provided. Alternatively or additionally, at least part of the control electronics for motor 12 may be provided on substrate 52 .

A synchronous gear 42 is provided in the gearbox 14 . Oil is also arranged in the gearbox 14 as a lubricant. It is distributed over the synchronous gear 42 by the splash disc 54 and over the adjacent bearings 56 .

Constriction 36 forms a shield or thermal barrier. This is particularly due to the heat generated during pump operation, especially in the area of the screw rotors 28,30. Heat from the screw motor (heat that would otherwise spread within the housing 16) is prevented from reaching other areas by leaving less material cross-section and by altering the heat path due to deformation. . Thus, in particular the oil in the gearbox 14 and the bearings 56 are protected from too high temperatures. An immersion cooler 14 located within the gearbox 14 likewise contributes to the temperature reduction. It is located in an oil bath (not shown) in the gearbox 14 and thus directly cools the oil.

For each screw rotor 28 and 30, a lubricant entrainment device formed as a deflector 58 is provided adjacent to the bearing 56 (which forms the stationary bearing here). Each deflector 58 forms a barrier for oil within the gearbox. Thus, it does not reach the area where the oil has a pumping effect or the vacuum area, especially the outlet area. Deflector 58 has a centrifugal edge for the oil that cannot be seen in detail. A flow-off groove is formed in the housing 16 against the centrifugal edge. This flow-off groove receives the oil to be centrifuged and directs it into the gearbox 14 or to an oil sump therein. The oil conveyed or distributed by the splash disc 54 to the gear 42 and to the bearing 56 is thus discharged again from the rotor 28 or 30 by the deflector 58 .

Piston rings are mounted on piston ring carriers 60 as dynamic fluid seals. This forms a contactless seal and thus prevents frictional heating. The deflector 58 returns as much oil as possible to the gearbox 14 so that as little oil as possible is already in the piston rings. Overall acceptable sealing properties are thus achieved with particularly low heat generation.

Screw rotors 28 and 30 have three sections of different pitch in their respective screw profiles 38 or 40 . The first portion 62 in the pump direction, on the left in FIG. 4, forms the suction area and has a constant and largest pitch of the three portions. The first portion 62 is longer than the closed conveying volume within the first portion with respect to the screw axis 63 (which extends along each rotor 28 or 30). The second portion 64 has multiple sub-portions. These are not referenced in detail. The lower part has a different but constant screw profile 38 or 40 pitch respectively than the first part. The pitch is then smaller than in the first portion. The second portion 64 now forms the longest portion. A third portion 66 having a smaller pitch forms the discharge portion. The third part is again constant pitch. Internal compression is provided by a pitch that decreases along the pump direction. This already compresses the pump gas before it is discharged.

Rotors 28, 30 or screw profiles 38, 40 can be designed and manufactured particularly simply due to the presence of certain parts. As can be seen in FIG. 4, the extended first portion 62 leads into the gap between the correspondingly enlarged screw profiles 28, 30 and the housing 16, so that the portions 62 and 64 from the inner seals are removed. The distance or gap from the internal compression at the transition to the suction space or suction area 67 is longer. Correspondingly, the sealability of the gap is also enhanced. This provides an improved seal against the suction area 67 of the inner compression section, especially at high pressure differentials.

The screw vacuum pump 10 has internal seals. The screw rotors 28, 30 of the pump 10 again surround the conveying volume that interacts with the housing 16 and is repeatedly closed. Its size is larger at the inlet-side end or in portion 62 than at the outlet-side end or in portion 66 . The size of the conveying volume is determined by the cross section of the screw profiles 38, 40 and by their pitch.

The size of the delivery volume on the inlet side, or portion 62, determines the theoretical suction performance of the screw pump 10. The pitch of the screw profiles 38, 40 is constant over the portion 62 on the inlet side, so that the conveying volume is compressed downstream of the closed portion by the internal compression section. If the rotors 28, 30 encircle their respective conveying volumes too early or too late, or if the internal compression starts too early, the theoretical suction performance of the pump will be degraded.

The size of each delivery volume on the outlet side, or portion 66, determines the power consumption of the pump during operation at the achievable final pressure. The ratio of the sizes of the delivery volumes on the inlet side and on the outlet side or in portions 62 and 66 corresponds to the ratio of compression inside the pump.

In portion 66 the pitch is constant over multiple revolutions of the screw profile 38,40. The pitch then corresponds to approximately the minimum pitch achievable with a given machining tool and is thus attributed to the manufacturing technology, especially considering costs. Due to the fact that multiple rotations, i.e. multiple closed conveying volumes, are provided in the part 66, the backflow as a result of the pressure difference between the gaps is compensated. Overall, in particular the course of the overall pitch of the rotors 28, 30 and the size of the gap formed between the rotors 28, 30 and the rotors 28, 30 and the housing 16 are the vacuum-technical performance data of the pump. , which in particular determines the suction performance and the achievable final pressure.

The screw profiles 38, 40 have a particularly low imbalance due to their two-start configuration. Thus, for example, no compensating elements (such as compensating masses, which require additional construction space) and/or compensating holes (in which masses can be stored) are not required. By means of the double cycloidal screw profile 38, 40 the pump can be operated in a further speed range, in particular by means of speed control and/or in standby operating mode.

Compression of the process gas generally generates heat. Heat is cooled in the screw pump 10, particularly by fluid cooling. A groove 32 provided for this purpose can be seen in FIG. The cooling lines for the fluid cooling section extend here and preferably longitudinally over another region of the screw profile, in particular over half the length of the screw profile. In particular, the fluid cooling is arranged at or near the area of inner compression.

A bearing shield 18 is fixed to the inlet-side end of the housing 16 . It carries in particular another bearing part with a bearing 68 . This forms a free-side bearing. With respect to the bearing shield 70 (formed integrally with the housing 16, but could be formed separately) at the opposite outlet end of the housing, the bearing shield 68: It is constructed as a separate part, but it can also be constructed in one piece.

On the inlet side there is likewise provided a splash disk 54, a deflector 58 and a piston ring carrier 60 with a plurality of piston rings. These work correspondingly with the outlet side devices. On the inlet side, another separately formed oil reservoir is provided in the cover 20 . An immersion cooler 34 is also provided for this oil bath. Alternatively or additionally, it is also possible, for example, for cooling lines to be provided, in particular cast, in the walls of the bearing shield 18 and/or the cover 20 .

At the beginning of the evacuation pumping process by pump 10, inlet 22 is normally at essentially the same pressure as the outlet. In contrast, during the exhaust pumping process, the pressure at inlet 22 drops to the final pressure. The final pressure is essentially zero with respect to the force produced. The pressure at outlet 24 thus exerts a force on rotor 28 . This is different than at the beginning of the exhaust pumping process. To compensate for this force, it is possible, for example, to provide a preloading device, in particular a spring. This is provided in particular at the free side bearing of the rotor and/or on the inlet side. The preloading device absorbs the forces exerted on the rotor by, for example, the gears that mesh obliquely and/or, in general, preloads the bearings under consideration regardless of operating conditions at varying pressures or pressure ratios. guaranteed.

FIG. 5 shows how the immersion cooler 34 is arranged within the gearbox 14 or within the cover 20 of the screw vacuum pump 10 . In this embodiment, the immersion coolers 34 are identically formed, which leads to less material duplication and lower manufacturing costs.

The immersion cooler 34 has cooling lines 72 . It extends through the cooling body 74 . The cooling body has structuring to increase the surface area of the cooling body and to optimize heat transfer. The immersion cooler 34 also has a flange 76 . The immersion cooler 34 is thereby secured.

In FIG. 6, the suction area 67 of the screw vacuum pump 10 is represented in a perspective sectional view. In the suction region 67 the suction-side ends of the screw profiles 38 and 40 respectively are arranged. As a result of their rotation and interaction with the end face 78, they repeatedly close the delivery volume and deliver along the axis of rotation to the outlet 24. FIG.

The end face 78 is essentially formed as a free-form surface, ie has a relatively complex shape as opposed to a flat surface. For example, the end face 78 has a roof-like lower region 80 and a valley-like or groove-like upper region 82 . In other words, machining the end face 78 is extremely cumbersome and expensive, especially when high precision is required. On the other hand, the end face 78 is formed without machining. It is formed by a corresponding mold during casting of the housing 16 . This produces a relatively imprecise end face 78, which is essentially acceptable due to the relatively long first portion 62 according to the invention. Machining can thus advantageously be omitted.

The cross-sectional view of FIG. 6 is shown in side view in FIG. This is for further clarifying the extension of the end surface 78 . Only the screw rotor 30 is then visible. 2, since the screw rotor 28 is covered in this view.

As can be seen in FIG. 7, the end face 78 extends essentially parallel to or corresponding to the screw profile 40 in the area surrounding the screw rotor 30 . When the rotor 30 is rotated counterclockwise with respect to FIG. 6, the spacing between the screw profile 40 and the end face 78 is reduced until the spacing is zero and the conveying volume is closed within the corresponding profile intermediate chamber 84. Become.

In FIG. 8, the housing 16 of the screw vacuum pump 10 is shown in perspective view without the screw rotors 28,30. It is further revealed here that the complex end face 78 can only be given the desired shape precisely at high cost, in particular by cutting methods, for example by milling. Due to its complex shape, the end face 78 cannot be measured or can only be measured with great effort (cost) and is therefore an additional risk for mass production. This is because defects without costly measurements may only be discovered within the framework of the final inspection of the pump. This entails high assembly costs or manufacturing costs and expensive rejects. By means of the present invention, the influence of the accuracy of the end face 78 on vacuum-technical performance data is generally reduced.

FIG. 9 shows on the horizontal axis the pressure p of the process gas in hectopascals at the inlet of the screw vacuum pump tested and on the vertical axis the suction performance S of the pump in cubic meters per hour. Two curves 86 and 88 of suction performance as a function of inlet pressure are shown.

The dashed line transition 86 is based on an exemplary screw vacuum pump having an inlet region or end face machined to the target contour by milling. The length of the first portion of constant pitch then basically corresponds to the length of the closed transport volume in the first portion. The exemplary screw vacuum pump of transition 86 is consequently not formed in accordance with the present invention.

A transition 88 represented by a solid line was recorded for another exemplary screw vacuum pump. In this vacuum pump, allowances provided for machining of the end faces are left unmachined. Otherwise, the screw vacuum pump of transition 88 is constructed in the same way as the screw vacuum pump of transition 86 . Both pumps, or end faces, differ only by certain tolerances. A tolerance is provided in the direction of the axis of rotation. Thus, in a stationary rotor, each conveying volume is closed early so that the first portion is longer than the conveying volume. An exemplary screw vacuum pump of transition 88 is consequently formed in accordance with the present invention.

Both curves 86 and 88 have suction performance curves typical of screw vacuum pumps with internal compression. The suction performance is then greatest in the central pressure region. Conversely, suction performance is low at high pressures, or early in the evacuation pumping process (right side of FIG. 9). The suction performance drops to zero upon reaching the final pressure (left side of FIG. 9).

In the central region transition 86 exhibits higher suction performance. It can be seen here that the conveying volume is precisely closed due to the end face being manufactured exactly to the target contour. Arrow 90 indicates this effect. The suction performance according to curve 86 here basically corresponds to the theoretical suction performance. Backflow as a result of compression, in particular backflow into the second part, has no dominant effect in this central pressure region.

Arrow 92 indicates that the achievable final pressure is lower or better for a screw vacuum pump with tolerance, ie transition 88 . This is attributed to the overall pumping length of the screw rotor being longer by tolerance. Correspondingly, the gap between the rotor and the housing is also long (large) so that overall better sealing is achieved against backflow from the outlet to the inlet.

Surprisingly, experiments on which FIG. 9 is based have shown that screw vacuum pumps, or transitions 88, with tolerances also have improved suction performance at high pressures. This is indicated by arrow 94 . This means that there is a gap between the compression and suction areas inside the pump, extended (extended) by a tolerance, which again leads to a correspondingly improved sealing against backflow. attributed to This is particularly visible in areas of high pressure. This is because internal compression leads to a maximum internal pressure, or pressure differential acting across the gap.

It has been shown that the screw vacuum pump according to the invention, in particular its housing, can be manufactured in a particularly simple manner. This is because the effect of the corresponding surface accuracy has been reduced. In particular, this makes it possible to dispense with costly machining on the end face or on the housing in the inlet region. The negative impact on vacuum performance is then less, and in many pressure ranges the invention even leads to improved performance data.

10 Screw vacuum pump 12 Motor 14 Gearbox 16 Housing 18 Bearing shield 20 Cover 22 Inlet 24 Outlet 26 Cooling pipe 28 Screw rotor 30 Screw rotor 32 Groove 34 Immersion cooler 36 Constriction 38 Screw profile 40 Screw profile 42 Synchronous gear 43 gear 44 housing 46 stator 48 magnet carrier 50 casting 52 substrate 54 splash disk 56 bearing 58 deflector 60 piston ring carrier 62 first portion 63 screw shaft 64 second portion 66 third portion 6 7 suction area 68 bearing 70 Bearing shield 72 Cooling pipe 74 Cooling body 76 Flange 78 End surface 80 Lower region 82 Upper region 84 Profile intermediate space 86 Suction performance transition 88 Suction performance transition 90 Arrow 92 Arrow 94 Arrow p Inlet pressure S Suction performance

Claims (12)

A screw vacuum pump (10) having a housing (16), two screw rotors (28, 30) disposed within the housing (16) and in engagement with each other, the screw rotors being adapted to process For gas transport, it interacts with the housing (16) to repeatedly form a closed transport volume of the process gas and transport it towards the outlet (24), with screw rotors (28, 30) each have at least two portions (62, 64) adjacent along the screw axis (63), wherein the screw rotors (28, 30) each extend into the inlet (22). has an at least essentially constant pitch in the first portion (62) placed in the vicinity of the With a pitch, in the suction area (67) of the screw vacuum pump (10), the end face (78) for the screw rotor (28, 30) is the end face inside the housing and the screw rotor (28 , 30), which end face is screw -shaped in the direction of the screw axis so that each conveying volume can be opened for suction and closed for conveying. formed as two inclined planes and adapted to the screw profile of each screw rotor (28, 30), and the closed conveying volume is then the closed area, which closed The region extends around the rotor along the screw profile (38, 40) between the two threads of each rotor (28, 30) and outwardly by the housing and by the screw profile. bounded by each other rotor along (38, 40),
the conveying volume then transitions in its profile along each screw rotor (28, 30) from the inlet (22) to the outlet (24) during the pumping process, thereby providing a pumping action;
And then, in the direction of the screw axis (63), the length of the first part (62) is closed by the screw rotors (28, 30) and the housing (16) in the first part (62). the length of the volume, i.e. longer than the length of one pitch of the rotor in the first portion (62);
A screw vacuum pump (10) characterized in that the end face (78) is formed in the cast member (16) and is unmachined. Screw vacuum pump (10) according to claim 1, characterized in that the closed conveying volume covers an angle of 360 degrees around the screw axis (63) of each rotor (28, 30). . Screw vacuum pump (10) according to claim 1 or 2, characterized in that the first portion (62) corresponds to at least 1.25 times the length of the conveying volume. Screw vacuum pump (10) according to any one of the preceding claims, characterized in that the first portion (62) corresponds to at least 1.5 times the length of the conveying volume. Screw vacuum pump (10) according to any one of the preceding claims, characterized in that the first portion (62) corresponds to at least 1.75 times the length of the conveying volume. Screw vacuum pump (10) according to any one of the preceding claims, characterized in that the first portion (62) corresponds to at least twice the length of the conveying volume. Screw vacuum pump (10) according to any one of the preceding claims, characterized in that the screw profile (38, 40) of each screw rotor (28, 30) is formed by a cycloid. . 8. Screw vacuum pump (10) according to any one of the preceding claims, characterized in that the screw profile (38, 40) of each screw rotor (28, 30) is double-threaded. . 9. Claims 1 to 8, characterized in that the screw rotors (28, 30) each have an at least essentially constant and at least partially constant pitch in the second portion (64). A screw vacuum pump (10) according to any one of the preceding claims. the screw rotor (28, 30) having at least one third portion (66) and the pitch being smaller in the third portion (66) than in the second portion (64); A screw vacuum pump (10) according to any one of claims 1 to 9. 11. Screw vacuum pump (10) according to claim 10, characterized in that the pitch in the third part (66) is at least essentially constant. A screw rotor (28, 30) each has a plurality of portions (62, 64, 66) of different pitch over its entire pumping length, wherein the pitch is 12. Screw vacuum pump (10) according to any one of claims 1 to 11, constant at (62, 64, 66).

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