Screw pump with a tapered suction-side section and a pressure-side section with a decreasing clearance

Abstract

A screw pump comprising a housing defining a chamber and two screw rotors. Wherein each screw rotor comprises a rotor shaft and at least two displacement elements connected with the rotor shaft. Each displacement element having at least one helical protrusion. 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 displacement 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 at least partly decreases in the conveying direction. Furthermore, a screw rotor, a method of manufacturing a screw rotor and a use of a screw pump or a screw rotor.

Claims

1. A screw pump comprising a housing comprising a wall, the housing defining a chamber and two screw rotors, wherein each screw rotor comprises: a rotor shaft, and at least two displacement elements connected with the rotor shaft, each displacement element having at least one helical protrusion, wherein one of the displacement elements is a suction-side displacement element arranged in a suction-side section of the chamber, and wherein another one of the displacement elements is a pressure-side displacement element arranged in a pressure-side section of the chamber, wherein the suction-side displacement element is designed tapering in a conveying direction, and wherein the clearance between the pressure-side displacement element and the pressure-side section of the chamber at least partly decreases in the conveying direction.

2. The screw pump of claim 1, wherein the clearance between the pressure-side displacement element and the pressure-side section of the chamber is such that during operation in the 100-300 mbar region, a gap between the between the pressure-side displacement element and the pressure-side section of the chamber is formed.

3. The screw pump of claim 1, wherein the diameter of the pressure-side displacement element increases in the conveying direction.

4. The screw pump of claim 1, wherein the pressure-side displacement element is designed counter-conical to the suction-side displacement element.

5. The screw pump of claim 1, wherein 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.

6. The screw pump of claim 1, wherein the clearance between the pressure-side displacement element and the pressure-side section of the chamber decreases by 10% to 50%, in the conveying direction.

7. The screw pump of claim 1, wherein the diameter of the at least one helical protrusion of the pressure-side displacement element increases in the conveying direction.

8. The screw pump of claim 7, wherein the diameter of the at least one helical protrusion of the pressure-side displacement element increases by 0.05% to 0.5% in the conveying direction.

9. The screw pump of claim 1, wherein the inner diameter of the pressure-side section of the chamber decreases in the conveying direction.

10. The screw pump of claim 1, wherein an inner volume ratio of the screw pump is at least 4.

11. The screw pump of claim 1, wherein the suction-side displacement element has a volume ratio of at least 4.

12. The screw pump of claim 1, wherein the pressure-side displacement element has a volume ratio of 1 to 3.

13. The screw pump of claim 1, wherein the diameter of an inner element of the suction-side displacement element increases in the conveying direction.

14. The screw pump of claim 1, wherein the diameter of an inner element of the pressure-side displacement element is essentially constant.

15. The screw pump of claim 1, wherein each displacement element has at least one helical recess.

16. The screw pump of claim 15, wherein 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.

17. The screw pump of claim 1, wherein a further displacement element is provided that is arranged upstream of the suction-side displacement element in the conveying direction, the further displacement element being substantially cylindrical in shape.

18. A screw rotor for a screw pump comprising: a rotor shaft, and at least two displacement elements connected with the rotor shaft, each displacement element having at least one helical protrusion, wherein one of the displacement elements is a suction-side displacement element, and wherein another one of the displacement elements is a pressure-side displacement element, wherein the suction-side displacement element is designed tapering in a conveying direction, and wherein the diameter of the pressure-side displacement element increases in the conveying direction.

19. A method of manufacturing a screw rotor for a screw pump comprising: providing a screw rotor, the 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, wherein one of the displacement elements is a suction-side displacement element, wherein the suction-side displacement element is designed tapering in a conveying direction, wherein another one of the displacement elements is a pressure-side displacement element, wherein the pressure-side displacement element is designed substantially cylindrically; and machining the pressure-side displacement element to have an increasing diameter in the conveying direction.

20. The method of claim 19, wherein the machining is performed by means of turning, and/or milling, and/or grinding.

21. A method of manufacturing a screw pump comprising: providing a screw rotor comprising at least two displacement elements, wherein one of the displacement elements is a suction-side displacement element, wherein the suction-side displacement element is designed tapering in a conveying direction, wherein another one of the displacement elements is a pressure-side displacement element; machining the pressure-side displacement element to have an increasing diameter in the conveying direction; and arranging the screw rotor inside a housing.

Description

BRIEF DESCRIPTION OF DRAWINGS

(1) The disclosure will be explained hereinafter in more detail with reference to preferred examples and to the accompanying drawings.

(2) FIG. 1 a schematic side view of a screw rotor of the state of the art.

(3) FIG. 2 a schematic cutaway side view of a section of a screw vacuum pump of the state of the art.

(4) FIG. 3 a schematic side view of an example of a screw rotor according to the disclosure.

(5) FIG. 4 a schematic cutaway side view of a section of an example of a screw vacuum pump according to the disclosure.

(6) FIG. 5a-5b schematic detail side views of examples based on the detail section V of FIG. 3.

(7) FIG. 6 a schematic cutaway side view of a section of another example of a screw vacuum pump according to the disclosure.

(8) FIG. 7 a graph showing the radial clearance of a screw vacuum pump of the state of the art.

(9) FIG. 8 a graph showing the radial clearance of an example of a screw vacuum pump according to the disclosure.

DETAILED DESCRIPTION

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

(11) The screw rotor 10 illustrated in FIG. 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 manner, i.e. on one side.

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

(13) 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 FIG. 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 FIG. 3). Individual chamber volumes formed by the two meshing screw rotors thus reduce their respective volume in the conveying direction 22.

(14) In the example 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.

(15) In FIG. 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.

(16) Due to the recess 34 eight windings are formed in the pressure-side displacement element 14 in the example illustrated.

(17) FIG. 2 shows a schematic cutaway side view of a section of a screw vacuum pump 100 with the screw rotor 10 of FIG. 1.

(18) 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 chamber 62 within.

(19) 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.

(20) The clearance between the suction-side displacement element 12 and the suction-side section 64 of the chamber 62 corresponds to distance A.sub.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 chamber 62. Preferably, the distance A.sub.R is constant, i.e. the distance A.sub.R hast the same value for the entire suction-side displacement element 12.

(21) 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.sub.P.sup.a between the outer surface 52.sub.a at winding 51.sub.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.sub.P.sup.e between the outer surface 52.sub.e at the winding 51.sub.e of the helical protrusion 50 at the left side of the pressure-side displacement element 14 and the inner surface 70 of the suction-side section 68 of the chamber 62. Preferably, the distances A.sub.R as well as A.sub.P.sup.e and A.sub.P.sup.a are the same.

(22) For the screw pump with a screw rotor 10 as illustrated in FIGS. 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 displacement 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 operation, for example with an inlet pressure between 100-300 mbar, the compression power is high and the compression 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 temperatures.

(23) 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 pressures between 100-300 mbar the compression power is high and the compression 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 temperatures 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.

(24) The impact of the changing radial clearances based on the thermal effect for the vacuum pump 100 of FIGS. 1-2 are shown in FIG. 7.

(25) The graph at the bottom of FIG. 7 shows the radial clearance over the length of the screw rotor 10.

(26) Curve C shows the clearance in the cold condition. In the cold condition, the clearance is constant. Preferably, the clearance corresponds to A.sub.P.sup.e=A.sub.P.sup.a=A.sub.R.

(27) Curve W shows the clearance in the warm condition, preferably at an inlet pressure 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 direction, 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.

(28) It was found that the radial clearance may have relatively higher priority compared to other clearances 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 may be designed for this operating region.

(29) FIG. 3 shows an example of a screw rotor 10 according to the present disclosure.

(30) The example is based on the screw rotor 10 of FIG. 1.

(31) In contrast to the screw rotor 10 of FIG. 1, the screw rotor 10 of FIG. 3 shows a pressure-side displacement element 14 having a helical protrusion 50 with a counter-conical shape compared to the suction-side displacement element 12. In detail, the diameter of the pressure-side displacement element 14 increases in the conveying direction 22. As indicated, the diameter D.sub.P.sup.a at the inlet area of the suction-side displacement element 12 is smaller than the diameter D.sub.P.sup.e at the outlet area of the suction-side displacement element 12. The diameter D.sub.P.sup.a of the helical protrusion 50 is determined at winding 51a, whereas the diameter D.sub.P.sup.e of the helical protrusion 50 is determined at winding 51e.

(32) Preferably, the diameter of the helical protrusion 50 of the pressure-side displacement 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.

(33) As for FIG. 1, the inner diameter D.sub.RP 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.sub.RP has the same value over the entire length of the pressure-side displacement element 14 in the conveying direction 22.

(34) Also, the inner and outer diameter of the suction-side displacement element 12 in FIG. 3 are the same as for FIG. 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 DR at a first area on the right side of the helical protrusion 36 is indicated, which is greater than the diameter DR 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 D.sub.RS.sup.1 at a first area on the right side of the helical recess 26 is indicated, which is smaller than the diameter D.sub.RS.sup.2 at a second area on the left side of the helical recess 26.

(35) FIG. 3 shows an example of a screw vacuum pump 100 according to the present disclosure. The screw rotor 10 of the example corresponds to the screw rotor 10 of FIG. 3.

(36) The right side of FIG. 3, i.e. the section of the suction-side displacement element 12, corresponds to the example of FIG. 2. In some examples, as illustrated in the middle of FIG. 3, a further displacement element 13 is provided that is arranged upstream of the suction-side displacement element 12 in the conveying direction 22. The further displacement element 13 may be substantially cylindrical in shape.

(37) 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 displacement element 14 and the pressure-side section of the chamber 68 decreases in the conveying direction 22. This is indicated by the distance A.sub.P.sup.a 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.sub.P.sup.e 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.

(38) 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 displacement element 14, i.e. from the inlet to the outlet of the pressure-side displacement element 14.

(39) FIG. 5a shows a detail side view of section V of FIG. 3.

(40) 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.

(41) The diameter D.sub.P.sup.d2 on the left side of the left winding 51d is greater than the diameter D.sub.P.sup.d1 on the right side of the left winding 51d. The minimum diameter D.sub.P.sup.d1 of the left winding 51d is greater than the maximum diameter D.sub.P.sup.c2 of the middle winding 51c.

(42) 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 FIG. 5a.

(43) FIG. 5b shows an alternative example of the pressure-side displacement element 14 based on the FIG. 5a.

(44) 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 FIG. 5a. Thus, particularly, the surfaces 52b, 52c of the windings 51b, 51c have a conical form.

(45) The left winding 51d of the helical protrusion 50 however has a cylindrical form, thus a constant diameter. In this example, only a section of the helical protrusion 50 has an increasing diameter. In other words, the diameter of the helical protrusion 50 increases partly over the length of the pressure-side displacement element 14.

(46) FIG. 6 shows another example of a screw vacuum pump 100 according to the present disclosure. The screw rotor 10 of the example corresponds to the screw rotor 10 of FIG. 1.

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

(48) 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 A.sub.P.sup.a between the helical protrusion 50 and the pressure-side section 68 on the right side is greater than the distance A.sub.P.sup.e between the helical protrusion 50 and the pressure-side section 68 on the left side.

(49) In the shown examples 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.

(50) The examples 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.

(51) The screw rotor 10 of the examples 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 examples 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.

(52) FIG. 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 disclosure.

(53) The graph at the bottom of FIG. 8 shows the radial clearance over the length of the screw rotor 10 of FIG. 3.

(54) Curve C shows the clearance in the cold condition.

(55) Curve W shows the clearance in the warm condition, preferably at an inlet pressure 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 direction.

(56) In contrast to the pronounced clearance drop (area Q) in the vacuum pumps of the state of the art (see FIG. 7), the screw rotor 10 and/or screw vacuum pump 100 according to the present disclosure shows an essentially constant change in the clearance, particularly in area Q.

(57) Thus, the seizure risk can be reduced. On the other hand, with the present disclosure it is possible to optimize, particularly minimize, the radial clearance to achieve an optimal pumping efficiency.