The disclosure relates to a screw pump, a screw rotor, a method of manufac turing a screw rotor and a use of a screw pump or a screw rotor.

Generally, dry running vacuum pumps, such as screw vacuum pumps, have high power consumption when running at low inlet pressure. This can be lowered with the build in volume ratio of the pump. The build in volume ratio is the ratio of the swept volume of the inlet stage to the swept volume of the exhaust stage. The higher the volume ratio, the lower the power consumption.

Generally, screw vacuum pumps comprise a suction chamber in a housing, in which suction chamber two screw rotors are arranged. Each screw rotor com prises at least one displacement element with a helical recess formed by a hel ical protrusion. The helical protrusion preferably forms a plurality of windings. With screw vacuum pumps, particularly with screw fine vacuum pumps, it is often the goal is to achieve an internal volume ratio that is as high as possible. The internal volume ratio is the ratio of the volume at the inlet of the vacuum pump to the volume at the outlet of the vacuum pump. Screw vacuum pumps of the first generation, such as e.g. the pumps LEYBOLD Screwline or BUSCH Cobra, have an internal volume ratio of approx. 3 to 4. With vacuum pumps currently on the market, such as e.g. the screw vacuum pumps LEYBOLD DRYVAC or Edwards GKS, the volume ratio is 5 to 7.

There are basically two known rotor shapes for screw pumps: the parallel, also called cylindrical, rotor and the tapered, also called conical, rotor. For parallel rotors high volume ratio would lead to very small exhaust stages which are difficult to machine. Tapered rotors can be built with high geometrical volume ratio as the exhaust stage can be made small without having to machine small and deep grooves. The drawback is the limited number of the small exhaust stages that can be integrated in tapered rotors. In this case the back leakage through the clearances makes the exhaust stages inefficient in terms of com pression and leaves the compression power at a high value.

To achieve low power consumption on screw pumps, two features must be re alized: high volume ratio, i.e. small exhaust stage, and a sufficient number of exhaust stages to compensate the back leakage. One way of achieving that is the "hybrid" rotor, which is a combination of a tapered rotor section on the inlet side and a parallel rotor section on the exhaust side of the pump. Such a rotor type is described, for example, in DE202017005336U1.

Thermal issues arise in such screw pumps. For example, during operation, e.g. with an inlet pressure between 100-300 mbar, seizures of the rotor with the housing, particularly with the stator, can occur. "Seizures" can also be referred to as "crashes". On the other hand, for example, the pumping efficiency of such screw pumps is relatively low.

It is an object of the disclosure to provide a screw pump preferably a screw vacuum pump, a screw rotor, a method of manufacturing a screw rotor and a use of a screw pump, preferably a screw vacuum pump, with a better seizure safety, particularly additionally with an optimized pumping efficiency.

Objects of the disclosure are achieved with a screw pump as defined by claim 1, a screw rotor as defined by claim 18, a method of manufacturing a screw rotor as defined by claim 19, and a use of a screw pump or screw rotor as defined by claim 21.

The screw pump of the disclosure preferably is a screw vacuum pump. More preferably, the screw pump of the disclosure is a dry running screw vacuum pump. The screw pump comprises a housing defining a chamber. Particularly the inner wall of the chamber corresponds to a stator of the screw pump. The screw pump further comprises two screw rotors. Each screw rotor comprises a rotor shaft, and at least two displacement elements connected with the rotor shaft. Each displacement element has at least one helical protrusion. Preferably, the helical protrusion comprises multiple windings. The helical protrusion par ticularly forms a helical recess therebetween. One of the displacement elements is a suction-side displacement element arranged in a suction-side section of the chamber. Another one of the displacement elements is a pressure-side displace ment element arranged in a pressure-side section of the chamber. The suction- side displacement element is designed tapering in the conveying direction. The clearance between the pressure-side displacement element and the pressure- side section of the chamber decreases in the conveying direction. The clearance is a radial clearance. This pressure-side section of the chamber particularly is a pressure side stator element. The clearance between the pressure-side displace ment element and the pressure-side section of the chamber decreases at least partly, preferably over the entire length between the pressure-side displace ment element and the pressure-side section of the chamber. It is preferred that the clearance between the pressure-side displacement element and the pres sure-side section of the chamber decreases at least in the inlet area of the pres sure-side displacement element. The inlet area of the pressure-side displace ment element is the suction-side area of the pressure-side displacement and the pressure-side section of the chamber. An outlet area is on the opposite site in the conveying direction. It is preferably possible that the clearance between the pressure-side displacement element and the pressure-side section of the chamber is constant, particularly the same, in the outlet area. It was found that the inlet area, particularly the inlet section of the pressure-side displacement element and/or the inlet section of the pressure-side section of the chamber has a high temperature during operation, particularly the highest temperature of the area of the displacement element, or even of the entire screw pump. Par ticularly, for inlet pressures between 100-300 mbar such high temperatures can be observed. Such high temperatures can lead to thermal expansion of the pres sure-side displacement element, particularly of the helical protrusion. It was found that the greatest expansion is in the inlet area of the pressure-sided displacement element. Thus, with the present invention, it is possible to achieve a optimize, preferably a constant, clearance during operation, since the thermal expansion can compensate the different clearances.

It is preferred that the clearance between the pressure-side displacement ele ment and the pressure-side section of the chamber is adjusted such that during operation, particularly during operation in the 100-300 mbar inlet pressure re gion, an essentially uniform gap between the between the pressure-side dis placement element and the pressure-side section of the chamber is formed.

Preferably, the diameter of the pressure-side displacement element increases in the conveying direction. It is preferred that the diameter increases at least partly, preferably over the entire length of the pressure-side displacement ele ment, i. e. from the inlet to the outlet of the pressure-side displacement ele ment. The diameter preferably is a radial diameter of the pressure-side dis placement.

Preferably, the pressure-side displacement element is designed counter-conical to the suction-side displacement element. It is preferred that pressure-side dis placement element is designed at least partly counter-conical, preferably over the entire length of the pressure-side displacement element, i.e. from the inlet to the outlet of the pressure-side displacement element.

Preferably, the clearance between the pressure-side displacement element and the pressure-side section of the chamber at least partly decreases linearly in the conveying direction, particularly over the entire length of the pressure-side displacement element, i.e. from the inlet to the outlet of the pressure-side dis placement element.

Preferably, the radial clearance, particularly at the inlet are, between the pres sure-side displacement element and the pressure-side section of the chamber is 100 pm to 500 pm, preferably 130 pm to 450 pm, more preferably 150 pm to 400 pm. It is preferred that the radial clearance between the pressure-side displacement element and the pressure-side section of the chamber decreases by 20 pm to 150 pm, particularly over the entire length of the pressure-side displacement element, i.e. from the inlet to the outlet of the pressure-side dis placement element.

Preferably, the clearance between the pressure-side displacement element and the pressure-side section of the chamber decreases by 10 % to 50 %, particu larly by 15 % to 30 % in the conveying direction. This decrease particularly defines a decrease over the entire length of the pressure-side displacement el ement, i.e. from the inlet to the outlet of the pressure-side displacement ele ment.

Preferably, the diameter of at least one of the helical protrusions of the pres sure-side displacement element increases in the conveying direction. In other words, it is preferred that the helical protrusion has a conical shape. The diam eter of the helical protrusion preferably is a radial diameter.

Preferably, the diameter of the at least one helical protrusion of the pressure- side displacement element increases by 0.05% to 0.5%, particularly by 0.05% to 0.2%. in the conveying direction. This increase particularly defines an increase over the entire length of the pressure-side displacement element, i.e. from the inlet to the outlet of the pressure-side displacement element.

Preferably, the diameter of the at least one helical protrusion of the suction-side displacement element decreases by 3 % to 40 %, particularly by 5 % to 30 %, more particularly by 15 % to 30 % in the conveying direction. This decrease particularly defines a decrease over the entire length of the suction-side dis placement element, i.e. from the inlet to the outlet of the suction-side displace ment element. Preferably, the incline of the, particularly conical, pressure-side displacement element is smaller than the incline of the, particularly conical, suction-side dis placement element. It is preferred that the incline of the suction-side displace ment element is 2° to 8°. Preferably, the incline of the pressure-side displace ment element is 0.01° to 1°, particularly 0.05° to 1°, more particularly 0.05° to 0.5°..

As an alternative to the increasing diameter of one or more of the helical pro trusions of the pressure-side displacement element, it is possible that one or more helical protrusions of the pressure-side displacement element has a con stant diameter, i.e. a cylindrical shape.

Preferably, the inner diameter of the pressure-side section of the chamber de creases in the conveying direction. This pressure-side section of the chamber corresponds to the stator for the pressure-side displacement element. The de creasing inner diameter of the pressure-side section of the chamber can be im plemented, at least partly, preferably end-to-end, in a linear or curved manner.

Preferably, the inner volume ratio of the screw pump is at least 4, particularly at least 7.

Preferably, the diameter of an inner element of the suction-side displacement element increases, at least partly, particularly end-to-end, in the conveying di rection. It is preferred that the inner element increases in a conical manner. This inner element preferably is part of the rotor shaft.

Preferably, the diameter of an inner element of the pressure-side displacement element is essentially constant. In other words, the inner element of the pres sure-side displacement element has the same diameter, particularly end-to- end. This inner element preferably is part of the rotor shaft. Preferably, each displacement element has at least one helical recess. It is pre ferred that the helical recess is formed by the helical protrusion, particularly between windings of the helical protrusion.

Preferably, the volume of the helical recess of the suction-side displacement element is greater than the volume of the helical recess of the pressure-side displacement element.

Preferably, the displacement elements have substantially the same diameter at the end faces directed towards each other.

Preferably, the diameter, particularly the average or maximum diameter, of the pressure-side displacement element is smaller by 5-35%, particularly by 10- 25%, than the inlet diameter of the suction-side displacement element.

Preferably, the suction-side displacement element has a volume ratio of at least 4, particularly at least 7.

Preferably, the pressure-side displacement element has a volume ratio of 1 to 3, particularly of 1 to 1.5, more particularly of 1.0001 to 1.1.

Preferably, the diameter of the pressure-side displacement element is 70 mm to 200 mm.

Preferably, the diameter of the suction-side displacement element is 80 to 300 mm in the region of the pump inlet.

Preferably, the diameter of the suction-side displacement element is 65 to 180 mm in the transition region to the pressure-side displacement element. The transition region corresponds to the outlet region of the suction-side displace ment element.

Preferably, the diameter of the pressure-side displacement element is 65 to 180 mm in the region of the outlet of the pump and/or outlet of the pressure-side displacement element.

Preferably, the number windings of the helical protrusion of the pressure-side displacement element is at least 3, particularly at least 5, more particularly at least 8. Preferably, the number of gaps between the windings of the pressure- side displacement element is at least 2, particularly at least 4, more particularly at least 7.

Preferably, the number of windings of the helical protrusion of the suction-side displacement element is 3 to 6. Particularly, the number of gaps between the windings of the suction-side displacement element is 2 to 5.

Preferably, a further displacement element is provided that is arranged up stream of the suction-side displacement element in the flow direction, the fur ther displacement element being preferably substantially cylindrical in shape.

The present disclosure further discloses a screw rotor for a screw pump. Prefer ably, the screw rotor of the disclosure is a screw rotor for a screw vacuum pump, more preferably for a dry running screw vacuum pump. The screw rotor com prises a rotor shaft and at least two displacement elements connected with the rotor shaft. Each displacement element having at least one helical protrusion. Preferably, the helical protrusion comprises a plurality of windings. One of the displacement elements is a suction-side displacement element. Another one of the displacement elements is a pressure-side displacement element. The suc tion-side displacement element is designed tapering in the conveying direction. The diameter of the pressure-side displacement element increases in the conveying direction. The suction-side displacement element is preferably adapted to be arranged in a suction-side section of a chamber of the screw pump. The pressure-side displacement element is preferably adapted to be ar ranged in a pressure-side section of the chamber of the screw pump.

Preferably, the screw rotor for a screw pump comprises one or more features as defined for the above screw pump of the present disclosure.

Preferably, the screw rotor for a screw pump is a screw rotor for the above defined screw pump.

The present disclosure further discloses a method of manufacturing a screw rotor fora screw pump. Preferably, the method is a method of manufacturing a screw rotor for a screw vacuum pump, more preferably for a dry running screw vac uum pump. The method comprises the step of providing a screw rotor. The provided screw rotor comprising a rotor shaft and at least two displacement elements connected with the rotor shaft. Each displacement element having at least one helical recess. Preferably, the helical protrusion comprises multiple windings. One of the displacement elements is a suction-side displacement. An other one of the displacement elements is a pressure-side displacement ele ment. The pressure-side displacement element is designed substantially cylin- drically. Another step of the method comprises machining the pressure-side dis placement element to achieve an increasing diameter of the pressure-side dis placement element in the conveying direction.

Preferably, the machining is performed in such a way to achieve a pressure-side displacement element with one or more features as defined for the above pres sure-side displacement element of the screw pump of the present disclosure.

Preferably, the machining is performed by means of turning, and/or milling, and/or grinding. The present disclosure further discloses a method of manufacturing a screw pump. Preferably, the method is a method of manufacturing a screw vacuum pump, more preferably a dry running screw vacuum pump. The method com prises the steps of the above defined method of manufacturing a screw rotor for a screw pump. Another preferred step of the method is to arrange at least one, preferably two screw rotors manufactured with the steps of the of the above defined method of manufacturing a screw rotor for a screw pump inside a housing of a pump.

The present disclosure further discloses a use of a screw pump as defined above to generate a vacuum, or the use of a above defined screw rotor, particularly of two of the above defined screw rotors, in a screw pump, preferably in a screw vacuum pump (100), more preferably in a dry running screw vacuum pump (100), to generate a vacuum.

The disclosure will be explained hereinafter in more detail with reference to pre ferred embodiments and to the accompanying drawings.

The drawings show:

Fig. 1 a schematic side view of a screw rotor of the state of the art,

Fig. 2 a schematic cutaway side view of a section of a screw vac uum pump of the state of the art,

Fig. 3 a schematic side view of an embodiment of a screw rotor according to the disclosure,

Fig. 4 a schematic cutaway side view of a section of an embodi ment of a screw vacuum pump according to the disclosure, Fig. 5a-5b schematic detail side views of embodiments based on the detail section V of Fig. 3,

Fig. 6 a schematic cutaway side view of a section of another em bodiment of a screw vacuum pump according to the disclo sure,

Fig. 7 a graph showing the radial clearance of a screw vacuum pump of the state of the art, and

Fig. 8 a graph showing the radial clearance of an embodiment of a screw vacuum pump according to the disclosure.

Similar or identical components or elements are identified with the same refer ence signs or variations thereof (e.g. 51 and 51a-51e) in the figures.

The screw rotor 10 illustrated in Figure 1 preferably corresponds to the screw rotor of DE202017005336U1. The screw rotor 10 comprises a rotor shaft 11 supporting two displacement elements 12, 14. The two cylindrical ends 16, 18 of the rotor shaft serve to receive bearings for supporting the screw rotor in a pump housing. It is also possible to support the rotor shaft in an overhung man ner, i.e. on one side.

The displacement element 12 on the right in Figure 1 is conical and tapers in the conveying direction 22 from a pump inlet 20 which is arranged on the right in Figure 1 but not illustrated therein, towards a pump outlet 24 which is ar ranged on the left in Figure 1 but not illustrated therein.

The helical recess 26 of the conical suction-side displacement element 12 is designed such that the volume decreases. This is achieved on the one hand due to the conical outer shape of the displacement element 12. The conical outer shape of the displacement element 12 is achieved by a decreasing diameter of the helical protrusion 36 of the displacement element (see also Figure 3). On the other hand, the decreasing volume is achieved due to the inner portion 28 of the displacement element 12 widening in the conveying direction (see also Figure 3). Individual chamber volumes formed by the two meshing screw rotors thus reduce their respective volume in the conveying direction 22.

In the embodiment illustrated in which only two displacement elements 12, 14 are provided, an end face 30 of the displacement element 12, which is directed towards the pump outlet 24 or towards the pressure side of the pump, abuts on an end face 32 of the pressure-side displacement element 14. The end face 32 is directed towards the pump inlet or in the direction of the suction side of the vacuum pump. The diameters of the two displacement elements 12, 14 are preferably substantially the same in the region of the end faces 30, 32.

In Figure 1, the pressure-side displacement element 14 has a cylindrical shape. The pressure-side displacement element 14 also has a helical protrusion 36 forming a helical recess 34.

Due to the recess 34 eight windings are formed in the pressure-side displace ment element 14 in the embodiment illustrated.

Figure 2 shows a schematic cutaway side view of a section of a screw vacuum pump 100 with the screw rotor 10 of Figure 1.

Only one screw rotor 10 of the screw vacuum pump 100 and a section of a wall 60 of a housing of the vacuum pump 100 is shown. The housing forms a cham ber 62 within. The right section of the shown wall 60 corresponds to the suction-side section 64 of the chamber 62, which forms the stator for the suction-side displacement element 12. On the other hand, the left section of the shown wall 60 corresponds to the pressure-side section 68 of the chamber 62, which forms to the stator for the pressure-side displacement element 14.

The clearance between the suction-side displacement element 12 and the suc tion-side section 64 of the chamber 62 corresponds to distance A _R between the outer surface 38 of the helical protrusion 36 of the suction-side displacement element 12 and the inner surface 66 of the suction-side section 64 of the cham ber 62. Preferably, the distance A _R is constant, i.e. the distance A _R hast the same value for the entire suction-side displacement element 12.

The clearance between the pressure-side displacement element 14 and the pressure-side section 68 of the chamber 62 is constant. Thus, the distance A between the outer surface 52 _a at winding 51 _a of the helical protrusion 50 at the right side of the pressure-side displacement element 14 and the inner surface 70 of the suction-side section 68 of the chamber 62 is the same as the distance A ^e _P between the outer surface 52 _e at the winding 51 _e of the helical protrusion 50 at the left side of the pressure-side displacement element 14 and the inner sur face 70 of the suction-side section 68 of the chamber 62. Preferably, the dis tances A _R as well as A$ and Ap are the same.

For the screw pump with a screw rotor 10 as illustrated in Figures 1-2, when running at low inlet pressure the compression is mainly done in the pressure- side displacement element 14 which gives the low power consumption for the screw pump. But when operating, for example in the 100-300 mbar range, the compression is done in the tapered suction-side displacement element 12 and the gas reaches atmospheric pressure before entering the pressure-side dis placement element 14. The pressure-side displacement element 14 does not contribute to the compression and only convey the gas while acting as rotating cooling fins, i.e. they cool both gas and rotor. When running at low inlet pressure the power consumption is low and rotor temperature is also low. During opera tion, for example with an inlet pressure between 100-300 mbar, the compres sion power is high and the compression is done in the tapered suction-side dis placement element 12. This causes a high-power density in this region and con sequently leads to high rotor temperatures.

Challenging in terms of the clearances is the temperature profile of the pump 100. When running at low inlet pressure the power consumption is low due to the high volume ratio and rotor temperature is also low. This causes a relatively small thermal reduction of radial clearances. But, for example, for inlet pres sures between 100-300 mbar the compression power is high and the compres sion is done in the tapered suction-side displacement element 12. This causes a high power density in this region and consequently leads to high rotor tem peratures and a high thermal reduction of the radial clearances, particularly in the transition area between the displacement element 12 and the pressure-side displacement element 14.

The impact of the changing radial clearances based on the thermal effect for the vacuum pump 100 of Figures 1-2 are shown in Figure 7.

The graph at the bottom of Figure 7 shows the radial clearance over the length of the screw rotor 10.

Curve C shows the clearance in the cold condition. In the cold condition, the clearance is constant. Preferably, the clearance corresponds to A ^e _P = = A _R.

Curve W shows the clearance in the warm condition, preferably at an inlet pres sure of 100-300 mbar, particularly at 200 mbar. In the warm condition, the clearance decreases of the length of the screw rotor 10 in the conveying direc tion, whereby a minimum clearance is reached in area Q. A bulge in the clearance can be found in this area Q. This reduced clearance poses a seizure risk.

It was found that the radial clearance is by far the most important clearance for the performance of the pump. It should be as small as possible but allow safe operation of the pump under all conditions including the 100-300 mbar region where the rotor reaches the highest temperature. Thus, the radial clearance must be design for this operating region.

Figure 3 shows an embodiment of a screw rotor 10 according to the present disclosure.

The embodiment is based on the screw rotor 10 of Figure 1.

In contrast to the screw rotor 10 of Figure 1, the screw rotor 10 of Figure 3 shows a pressure-side displacement element 14 having a helical protrusion 50 with a counter-conical shape compared to the suction-side displacement ele ment 12. In detail, the diameter of the pressure-side displacement element 14 increases in the conveying direction 22. As indicated, the diameter Df at the inlet area of the suction-side displacement element 12 is smaller than the di ameter Df at the outlet area of the suction-side displacement element 12. The diameter Dp of the helical protrusion 50 is determined at winding 51a, whereas the diameter Df, of the helical protrusion 50 is determined at winding 51e.

Preferably, the diameter of the helical protrusion 50 of the pressure-side dis placement element 14 increases by 0.05 % to 0.5 %, particularly by 0.05 % to 0.2 % in the conveying direction. This increase particularly defines an increase over the entire length of the pressure-side displacement element 14, i.e. from the inlet to the outlet of the pressure-side displacement element 14.

As for Figure 1, the inner diameter DRP of the pressure-side displacement element 14, which corresponds to the outer diameter of the inner element 54 of the pressure-side displacement element 14 or the diameter of the helical recess 34 is preferably constant for the pressure-side displacement element 14. Thus, the diameter D _RP has the same value over the entire length of the pres sure-side displacement element 14 in the conveying direction 22.

Also, the inner and outer diameter of the suction-side displacement element 12 in Figure 3 are the same as for Figure 1. The diameter at the inlet area of the suction-side displacement element 12 is smaller than the diameter at the outlet area of the suction-side displacement element 12. Exemplary, the diameter Dp at a first area on the right side of the helical protrusion 36 is indicated, which is greater than the diameter D _p at a second area on the left side of the helical protrusion 36. The inner diameter of the suction-side displacement element 12, which corresponds to the outer diameter of the inner element 42 of the suction- side displacement element 12 or the diameter of the helical recess 26 preferably increases in the conveying direction 22. Exemplary, the diameter at a first area on the right side of the helical recess 26 is indicated, which is smaller than the diameter at a second area on the left side of the helical recess 26.

Figure 3 shows an embodiment of a screw vacuum pump 100 according to the present disclosure. The screw rotor 10 of the embodiment corresponds to the screw rotor 10 of Figure 3.

The right side of Figure 3, i. e. the section of the suction-side displacement element 12, corresponds to the embodiment of Figure 2.

On the left side, i. e. the section of the pressure-side displacement element 14, due to the increasing diameter of the helical protrusion 50 of the pressure-side displacement element 14, the clearance between the pressure-side displace ment element 14 and the pressure-side section of the chamber 68 decreases in the conveying direction 22. This is indicated by the distance Ap between the helical protrusion 50 and the pressure-side section 68 of the chamber 62 at an area on the right side of the pressure-side displacement element 14, which is greater than the distance A between the helical protrusion 50 and the pressure- side section 68 of the chamber 62 at an area on the left side of the pressure- side displacement element 14.

Preferably, the clearance between the pressure-side displacement element 14 and the pressure-side section 68 of the chamber 62 decreases by 10 % to 50 %, particularly by 15 % to 30 % in the conveying direction 22. This decrease particularly defines a decrease over the entire length of the pressure-side dis placement element 14, i.e. from the inlet to the outlet of the pressure-side dis placement element 14.

Figure 5a shows a detail side view of section V of Figure 3.

As indicated by the inclined reference line 58, the helical protrusion 50 of the pressure-side displacement element 14 increases in diameter in the conveying direction. Here, as illustrated, a linear increase is implemented.

The diameter Dp ² on the left side of the left winding 51d is greater than the diameter Dp ¹ on the right side of the left winding 51d. The minimum diameter Dp ¹ of the left winding 51d is greater than the maximum diameter Dp ² of the middle winding 51c.

The diameters are measured between the outer surfaces 51d, 51c, 51b of the helical protrusion 50, whereby the outer surfaces on the other side of the helical protrusion 50 are not shown in Figure 5a.

Figure 5b shows an alternative embodiment of the pressure-side displacement element 14 based on the Figure 5a. The right winding 51b and middle winding 51c of the helical protrusion 50 have an increasing diameter in the conveying direction, as also implemented in Figure 5a. Thus, particularly, the surfaces 52b, 52c of the windings 51b, 51c have a conical form.

The left winding 51d of the helical protrusion 50 however has a cylindrical form, thus a constant diameter. In this embodiment, only a section of the helical pro trusion 50 has an increasing diameter. In other words, the diameter of the hel ical protrusion 50 increases partly over the length of the pressure-side displace ment element 14.

Figure 6 shows another embodiment of a screw vacuum pump 100 according to the present disclosure. The screw rotor 10 of the embodiment corresponds to the screw rotor 10 of Figure 1.

The wall 60 of the housing of the screw vacuum pump 100 is based on the embodiment of Figure 2. However, in contrast to Figure 2, the pressure-side section 68 of the chamber 62 has an inner diameter which decreases in the conveying direction 22.

The clearance between the pressure-side displacement element 14 and the pressure-side section 68 of the chamber 62 again decreases in the conveying direction 22. The distance Ap between the helical protrusion 50 and the pres sure-side section 68 on the right side is greater than the distance A _P ^e between the helical protrusion 50 and the pressure-side section 68 on the left side.

In the shown embodiments a constant pitch for the changing diameter of the helical protrusion 36 and/or the helical protrusion 50 is implemented. However, it is possible to have a changing pitch, for example an increasing or decreasing pitch in the conveying direction 22. The embodiments only show one screw rotor 10. It is preferred that the screw pump of the present disclosure has a second screw rotor, preferably identical in terms of the clearance to the screw rotor 10 as defined here.

The screw rotor 10 of the embodiments is a screw rotor 10 for a screw vacuum pump, preferably for a dry running screw vacuum pump. However, it is also possible that the screw rotor 10 of the present disclosure, particularly as shown in the figures is a screw rotor 10 for a general screw pump. The screw pump 100 of the embodiments is a screw vacuum pump 100, preferably a dry running screw vacuum pump 100. However, it is also possible that the screw pump 100 of the present disclosure, particularly as shown in the figures is a general screw pump.

Figure 8 shows the effect of the decreasing clearance between the pressure-side displacement element 14 and the pressure-side section 68 of the chamber 62, according to the invention.

The graph at the bottom of Figure 8 shows the radial clearance over the length of the screw rotor 10 of Figure 3.

Curve C' shows the clearance in the cold condition.

Curve W' shows the clearance in the warm condition, preferably at an inlet pres sure of 100-300 mbar, particularly at 200 mbar. In the warm condition, the clearance decreases of the length of the screw rotor 10 in the conveying direc tion.

In contrast to the pronounced clearance drop (area Q) in the vacuum pumps of the state of the art (see Figure 7), the screw rotor 10 and/or screw vacuum pump 100 according to the present invention shows an essentially constant change in the clearance, particularly in area Q'. Thus, the seizure risk can be reduced. On the other hand, with the present invention it is possible to optimize, particularly minimize, the radial clearance to achieve an optimal pumping efficiency.