DRY-COMPRESSING VACUUM PUMP

Abstract

A dry-compressing vacuum pump, in particular a screw pump, has two rotor elements (14) which are arranged in a pump chamber (12) and which are each carried by a rotor shaft (22). Two shaft ends of the rotor shafts (22) project through a side wall (28) of the pump housing (10). On the two shaft ends (28) there is arranged in each case one toothed belt pulley (38). Furthermore, a drive device and an electric motor for driving the rotor shaft (22) are provided. According to the invention, the rotor shafts (22) are driven by means of a toothed belt (40). To be able to use a toothed belt for drive purposes, a rotational flank clearance between the two rotor elements (14) of greater than ±0.75°, in particular greater than ±1°, is provided.

Description

The invention relates to a dry-running vacuum pump, in particular a screw pump.

Dry-compressing vacuum pumps, such as e.g. screw pumps, comprise two rotor elements arranged in a suction chamber. With screw pumps, the rotor elements are formed as helical displacement elements. Each rotor element is supported by a rotor shaft. With a screw pump, both rotor elements are arranged in the suction chamber formed by the pump housing. Both rotor shafts extend through a housing wall that defines the suction chamber. Gears are connected with the two rotor shafts. With screw pumps, the two gears mesh with each other. Thereby, on the one hand, the two shafts rotating in opposite directions are synchronized and, on the other hand, the two shafts are driven. By providing the two meshing gears, only one of the two shafts has to be driven. For obtaining an efficient compressing process and a good volumetric efficiency narrow gaps are necessary between the rotors which require a very exact synchronizing. Typically, maximum synchronizing errors or circumferential backlashes between the rotors of ±0.25 are allowable. With dry-compressing vacuum pumps on the market, this can be achieved by providing meshing gears on the shaft ends. Due to the precision required and the small allowable tolerances, the costs are high.

Further, when providing meshing gears for synchronizing, it is necessary to provide for oil lubrication. As a result, a complex and complicated sealing is required in the housing wall through which the shaft ends extend.

An electron synchronizing of the two rotor shafts is also know. However, the same is also complex and costly. Usually, electric motors are provided as drive means for driving the rotor shaft. For increasing the rotary speed of the vacuum pump, these may be connected with a frequency converter. This also is a relatively costly component. Possibly, intermediate gearings are provided which in turn must also be oil-lubricated.

Further, it is known e.g. from DE 38 23 927 to synchronize two rotor shafts of a screw pump by means of a toothed belt. However, a corresponding product was not technically realized and offered on the market. For constructing screw pumps driven by means of a belt drive, with which pumps at least a vacuum of 200 mbar absolute pressure can be obtained when pumping against atmosphere, correspondingly small synchronizing errors have to be realized between the screw-shaped rotor elements. With the use of toothed belts, this would only be possible if the gears or toothed belt wheels driven by the toothed belts have a very small tooth gap clearance. According to ISO 13050 this is less than 0.1 mm or less than 0.2 mm, depending on the profile. Alternatively, with gears, a larger effective diameter of typically 0.2 to 0.4 mm would have to be provided. This leads to a forced tooth pitch error that compromises the synchronizing. Providing larger effective diameters, however, significantly shortens the service life of the toothed belts used. This is not accepted in dry-compressing vacuum pumps on the market.

It is an object of the present invention to provide a dry-compressing vacuum pump driven by a toothed belt and not having the above disadvantages.

The object is achieved, according to the invention, with the features of claim 1.

The dry-compressing vacuum pump comprises a suction chamber formed by a pump housing. Two rotor elements are arranged in the suction chamber, the vacuum pump particularly being a screw pump. Each rotor element is supported by a rotor shaft. The two rotor elements extend through a housing wall defining the suction chamber, so that one shaft end per rotor shaft extends from the suction chamber. One toothed belt wheel is arranged on the two shaft ends, respectively. Since the two rotor shafts are driven by means of a toothed belt, the two toothed belt wheels do not mesh with one another. Further, a drive means such as an electric motor is provided. Here, a belt pulley is arranged in particular on a drive shaft of the electric motor. The toothed belt is connected with the two toothed belt wheels and the drive means, in particular the belt pulley of the drive means. In order to provide a toothed belt for driving the two rotor shafts, a circumferential backlash between the two rotor elements of more than ±0.75°, in particular more than ±1° is provided according to the invention. The use of toothed belts is possible only because of the provision of such a large circumferential backlash.

In order to be able, despite the large circumferential backlash, to achieve a high vacuum of in particular less than 200 mbar absolute pressure when compressing against atmosphere, special designs of the compressing stages, i.e. of the displacer elements arranged on the rotor shaft, are preferred, wherein it is possible of course that the displacer elements are formed integrally with the rotor shaft.

Due to the large circumferential backlash allowed by the invention, a comparatively high return flow occurs in particular at the suction-side compressing elements, i.e. the compressing elements downstream of the pump inlet.

With screw-type rotors having a pitch variable in the delivery direction, the profile engagement gap in the inlet region is decisive for the maximum allowable synchronizing error, due to the large pitch of the winding of the displacer elements existing there. Even comparatively small angular deviations lead to undesired flank contacts in the suction-side displacer elements. To avoid this, a large circumferential backlash must be selected. For achieving a good volumetric efficiency of the pump despite the large gap created thereby, it is preferred to increase the number of windings having a large pitch and a large profile gap in the inlet region. In particular, two to three windings are preferably provided in this region. In addition or as an alternative, the number of windings in the outlet region, i.e. on the pressure side, may also be increased. This results in a lesser pressure gradient in the inlet region and thus also results in a reduce return flow. In the outlet region, the windings have a smaller pitch.

Further, in an alternative preferred embodiment, it is possible that the two screw rotors have a plurality of rotor or displacer elements or displacer stages. Preferably, at least two displacer elements or displacer stages are provided.

Such a vacuum pump screw-type rotor preferably comprises at least two helical displacer elements arranged on a rotor shaft. The at least two displacer elements preferably have different pitches, the pitch being constant for a respective displacer element. For example, the vacuum pump screw rotor comprises two displacer elements, wherein a first, suction-side displacer element has a larger constant pitch and a second, pressure-side displacer element has a smaller constant pitch. Due to the preferred provision of a plurality of displacer elements which each have a constant pitch, manufacture is significantly facilitated.

Preferably, each displacer element has at least one helical recess having the same contour all along its length. Preferably, the contours differ for each displacer element. Each individual displacer element preferably has a constant pitch and an unvarying contour. This facilitates the manufacture significantly so that the manufacturing costs can be reduced drastically.

For a further improvement of the suction capacity, the contour of the suction-side displacer element, i.e. in particular of the first displacer element seen in the pumping direction, is asymmetric in shape. Due to the asymmetric design of the contour or the profile, the flanks may be designed such that the leakage surfaces, the so-called blowholes, are preferably eliminated completely or at least have a reduced cross section. A particularly suitable asymmetric profile is the so-called “Quimby profile”. Such a profile may be relatively difficult to produce, but has the advantage that no continuous blowhole exists. A short circuit only exists between two adjacent chambers. Since the profile is an asymmetric profile with different profile flanks, at least two work steps are required for the manufacture, since, due to their asymmetry, the two flanks have to be made in different work steps.

The pressure-side displacer element, in particular the last displacer element in the pumping direction, is preferably provided with a symmetric contour. In particular, the symmetric contour has the advantage that the manufacture is simpler. Specifically, both flanks with a symmetric contour may be produced with a rotating end mill or a rotating disk mill in one work step. Such symmetric profiles have only small blowholes, which, however, are continuous, i.e. provided not only between two adjacent chambers. The size of the blowhole decreases when the pitch is reduced. In this respect, such symmetric profiles may be provided in particular with the pressure-side displacer element, since the same, in a preferred embodiment, has a smaller pitch than the suction-side displacer element and, preferably, also than the displacer element arranged between the suction-side and the pressure-side displacer element. Although such symmetric profiles are slightly less tight, these have the advantage that the manufacture is significantly simpler. In particular, it is possible to produce the symmetric profile in a single work step and preferably with a simple end mill or disc mill. This reduces the costs drastically. A particularly well suited symmetric profile is the so-called “cycloid profile”.

Providing at least two such displacer elements results in the corresponding screw vacuum pump to be able to generate low inlet pressures at low power consumption. Further, the thermal load is also low. Arranging at least two such displacer elements, designed in a preferred manner, with a constant pitch and an unvarying contour in a vacuum pump leads to essentially the same results as obtained with a vacuum pump having a varying pitch. In case of high built-in volume ratios, three or four displacer elements may be provided per rotor.

In a particularly preferred embodiment, a pressure-side, i.e. in particular a last displacer element in the pumping direction, has a large number of windings for reducing the obtainable inlet pressure and/or for reducing the power consumption and/or the thermal load. By a large number of windings, a larger gap between the screw-type rotor and the housing may be accepted, while the performance remains the same. Here, the gap may have a cold gap width of 0.05-0.3 mm. A large number of outlet windings or a large number of windings in the pressure-side displacer element may be produced in an economic manner, since this displacer element has a constant pitch and in particular also has a symmetric contour. This allows for a simple and economic manufacture, so that providing a larger number of windings is acceptable. Preferably, this pressure-side or last displacer element has more than 6, in particular more than 8 and particularly preferred more than 10 windings. In a particularly preferred embodiment, the use of symmetric profiles has the advantage that both flanks of the profile can be cut simultaneously using a mill. Here, the mill is also supported by the respective opposite flank, so that a deforming or bending of the mill during the milling operation and inaccuracies caused thereby are avoided.

For a further reduction of the manufacturing costs, it is particularly preferred to form the displacer elements and the rotor shaft as one piece.

In another preferred embodiment, the change in the pitch between adjacent displacer elements is discontinuous or erratic. Possibly, the two displacer elements are arranged at a distance from each other in the longitudinal direction, so that a circumferential, cylinder ring-shaped chamber is formed between two displace elements, which serves as a tool run-out. This is advantageous in particular with integrally formed rotors, since the tool producing the helical line can be guided out in a simple manner in this region. If the displacer elements are manufactured independently and are thereafter mounted on a shaft, providing a tool run-out, in particular such a circular cylindrical region, is not required.

In a preferred development of the invention, no tool run-out is provided at the change in pitch between two adjacent displacer elements. In the region of the change in pitch, preferably both flanks have a discontinuity or recess for guiding out the tool. Such a discontinuity has no decisive influence on the compression performance of the pump, since it is a discontinuity or recess that is very limited locally.

The vacuum pump screw rotor preferably comprises a plurality of displacer elements. These may each have the same or different diameters. It is preferred in this respect that the pressure-side displacer element has a smaller diameter than the suction-side displacer element.

With displacer elements manufactured independently of the rotor shaft, these are mounted on the shaft e.g. by press fitting. In this respect, it is preferred to provide elements such as dowel pins to fix the angular position of the displacer elements relative to each other.

In particular in case of the integral design of the screw-type rotor, but also in case of a multi-part design, it is preferred to manufacture the same from aluminum or from an aluminum alloy. It is particularly preferred to manufacture the rotor from aluminum or an aluminum alloy, in particular AlSi9Mg or AlSi17Cu4Mg. The alloy preferably has a high proportion of silicon of preferably more than 9%, in particular more than 15%, in order to reduce the expansion coefficient.

In a further preferred development of the invention, the aluminum used for the rotors has a low expansion coefficient. It is preferred for the material to have an expansion coefficient of less than 22·10⁻⁶/K, in particular less than 20·10⁻⁶/K. In another preferred embodiment, the surface of the displacer elements is coated, wherein in particular a coating against wear and/or corrosion is provided. In this respect it is preferred to provide an anodic coating or another suitable coating, depending on the application.

The vacuum pump has at least two compression stages.

Further, it is preferred for the dry-compressing vacuum pump of the present invention that the vacuum pump has a maximum volumetric efficiency of at least 75%, in particular at least 85%. The volumetric efficiency is the quotient of the real maximal obtained volume flow and the theoretically possible volume flow in a loss-free pump relative to the suction chamber geometry and the operating speed. The maximum volumetric efficiency is usually reached in a range between 1 and 10 mbar.

The toothed belt used preferably not only serves for driving, but also for synchronizing the rotor shafts. With screw pumps, the rotor shafts rotate in opposite directions. Thus, in a preferred embodiment, the toothed belt is designed as a double-sided toothed belt. In top plan view, the toothed belt thus preferably extends between the two toothed belt wheels connected with the shaft end.

In a preferred embodiment comprising the above described rotor, tooth gap clearances of the two toothed belt wheels of more than 0.10 mm can be accepted.

Here, the tooth gap clearance is defined by the combination of the tooth shape of the toothed belt wheels used and the tooth shape and size of the teeth of the toothed belt. Due to the relatively large tooth gap clearance, the service life of the toothed belts is significantly extended.

For achieving a further extension of the service life of the toothed belts, it is further preferred that the effective diameter is not enlarged and that thus no forced tooth pitch error occurs.

Providing a toothed belt for driving and synchronizing the two rotor shafts in particular has the advantage that no oil lubrication has to be provided. This has the particular advantage that the sealing of the shaft ends with respect to the suction chamber may be designed in a significantly more economic manner. Moreover, it is possible to use grease-lubricated roller bearings. In particular, the two shafts are supported in the housing wall through which the shaft ends are passed, wherein these bearings may be grease-lubricated bearings. The opposite shaft ends supported in the region of the inlet side are preferably supported in grease-lubricated bearings, but oil-lubricated bearings may also be used.

Further, a belt tensioning means may be provided to constantly keep the belt taut. Preferably, this is an automatic tensioning means in which the tension is generated e.g. by a spring or the like or a fixed bias is applied during assembly. Likewise, it is possible to tension the belt by configuring the drive motor such that it is displaceable.

Another advantage of the toothed-belt drive according to the invention is that a variation of the vacuum pump speed is possible in a simple manner. For that purpose, it is merely necessary to exchange the toothed belt pulley connected with the drive means. When the toothed belt pulley is exchanged, the toothed belt must be exchanged as well, as needed.

The invention will be explained hereinafter in detail with reference to a preferred embodiment and the accompanying drawings.

In the Figures:

FIG. 1 is a schematic longitudinal section through a screw-type vacuum pump,

FIG. 2 is a schematic illustration of the drive of the vacuum pump,

FIG. 3 is a schematic illustration of a combination of a toothed belt and a toothed belt disc with a tooth gap,

FIG. 4 is a schematic illustration of a combination of a toothed belt and a toothed belt disc without a tooth gap,

FIG. 5 is a schematic top plan view of a first preferred embodiment of a vacuum pump screw-type rotor,

FIG. 6 is a schematic top plan view of a second preferred embodiment of a vacuum pump screw-type rotor,

FIG. 7 is a schematic sectional view of displacer elements with an asymmetric profile, and

FIG. 8 is a schematic sectional view of displacer elements with an asymmetric profile.

FIG. 1 is a greatly simplified schematic illustration of a pump housing 10. A suction chamber is formed inside the pump housing 10, in which chamber two rotor elements 14 are arranged. In the embodiment illustrated the rotor elements 14 are screw-type rotors. The screw-type rotors 14 have helical compression elements that mesh with each other. The two screw-type rotors 14 are driven in opposite directions. In the embodiment illustrated the two screw-type rotors 14 have two pump stages 16, 18.

The two rotor elements are respectively arranged on a rotor shaft 22. On the suction side, the two rotor shafts 22 are supported in a housing cover 24 via bearing elements 26. On the opposite side, shaft ends 28 extend through a housing wall 30. The two rotor shafts 22 are supported in the housing wall 30 by grease-shaped bearings 32.

The dry-compressing vacuum pump convey a medium through an inlet 34 to an outlet 36.

For driving the two rotor elements 14, the two shaft ends 28 are each connected with a respective toothed belt wheel 38, wherein the two toothed belt wheels 38 do not mesh with each other. Synchronizing is effected via a toothed belt 40 (FIG. 2) not illustrated in FIG. 1. The toothed belt is designed as a double-sided toothed belt and, for synchronizing the two toothed belt wheels 38 or the two shaft ends 28 connected with the toothed belt wheels, is passed between these. Further, a drive means 42 is provided whose drive shaft 44 is connected with a toothed belt disc 46.

FIG. 3 schematically illustrates teeth of a toothed belt disc 38 or 46 in connection with a toothed belt 40. A tooth 48 of the toothed belt 40 is designed such that gap, illustrated in hatched lines, is formed opposite a tooth interstice 50 of two adjacent teeth 52 of the toothed belt wheel 38. Thereby, a certain play exists between the toothed belt 40 and the toothed belt wheel 38. The synchronizing of the two rotor shafts 22 may be somewhat compromised thereby, but the service life of the toothed belt 48 is extended.

As an alternative, a toothed belt may be provided, as schematically illustrated in FIG. 4. The same shows no distances between the tooth interstice 50 and the tooth 48 of the belt 40, which is referred to as a zero gap.

In the first preferred embodiment (FIG. 5) of the vacuum pump screw-type rotor, the rotor has two displacer elements 110, 112 forming the two pump stages 16, 18. A first, suction-side displacer element 110 has a large pitch of about 50-150 mm/rotation. The pitch is constant throughout the displacer element 110. The contour of the helical recess is constant as well. The second pressure-side displacer element 112 also has a constant pitch and a constant contour of the recess over its length. The pitch of the pressure-side displacer element 112 is preferably in the range of 10-30 mm/rotation. An annular cylindrical recess 114 is provided between the two displacer elements. The same serves to realize a tool run-out, due to the integral design of the screw-type rotor illustrated in FIG. 5.

Further, the integrally formed screw-type rotor has two bearing seats 116 and a shaft end 118. For example, a gear is connected with the shaft end 118 for driving.

In the second preferred embodiment illustrated in FIG. 6, the two displacer elements 110, 112 are manufactured separately and are then fixed on a rotor shaft 120, e.g. by pressing. This way of manufacturing may be somewhat more complex, but the cylindrical distance 114 between two adjacent displacer elements 110, 112 is not required as a tool run-out. The bearing seats 116 and the shaft ends 118 may be an integral part of the shaft 120. A continuous shaft 120 may also be made of another material different from that of the displacer elements 110, 112.

FIG. 7 illustrates a schematic sectional view of an asymmetric profile (e.g. a Quimby profile). The asymmetric profile illustrated is a so-called “Quimby profile”. The sectional view shows two screw-type rotors meshing with each other, their longitudinal direction being perpendicular to the drawing plane. The oppositely directed rotation of the rotors is indicated by the two arrows 115. With reference to a plane 117 perpendicular to the longitudinal axis of the displacer elements, the profiles of the flanks 119 and 121 are designed differently per rotor. The opposing flanks 119, 121 thus have to be manufactured independently. The manufacture which is therefore somewhat more complex and complicated, has the advantage, however, that no continuous blowhole exists and a short circuit exists merely between two adjacent chambers.

Such an asymmetric profile is preferably provided in the suction-side displacer element 110.

The schematic sectional view in FIG. 8 again shows a cross section through two displacer elements or two screw-type rotors which again rotate in opposite directions (arrows 115). With reference to the axis of symmetry 117, the flanks 123 of each displacer element are symmetrically designed. The preferred embodiment of a symmetrically designed contour illustrated in FIG. 8 is a cycloid profile.

A symmetric profile, as illustrated in FIG. 8, is preferably provided in the pressure-side displacer elements 112.

Further, it is possible that more than two displacer elements are provided. These may possibly also have different head diameters and corresponding base diameters. In this respect it is preferred that a displacer element with a larger head diameter is arranged at the inlet, i.e. at the suction side, so as to realize a higher suction capacity in this region and/or to increase the built-in volume ratio. Further, combinations of the above described embodiments are possible. For example, one or a plurality of displacer elements may be manufactured integrally with the shaft or an additional displacer element may be manufactured independently of the shaft and may then be mounted on the shaft.

Claims

1. A dry-compressing vacuum pump comprising

two rotor elements which are arranged in a suction chamber,

two rotor shafts, each supporting a rotor element,

two toothed belt wheels respectively arranged on one shaft end extending from the suction chamber,

a drive means driving the rotor shafts, and

a toothed belt connected with the drive means and the toothed belt wheels,

wherein

a circumferential backlash between the two rotor elements of more than ±0.75° is provided.

2. The dry-compressing vacuum pump of claim 1, wherein the maximum volumetric efficiency of the vacuum pump at an operating point in particular between 1 and 10 mbar is at least 75%.

3. The dry-compressing vacuum pump of claim 1, wherein, for synchronizing rotor shafts rotating in opposite directions, the toothed belt is designed as a double-sided toothed belt.

4. The dry-compressing vacuum pump of claim 3, wherein the toothed belt extends between the two toothed belt wheels.

5. The dry-compressing vacuum pump of claim 1, wherein the tooth gap clearance of the two toothed belt wheels is larger than 0.15 mm.

6. The dry-compressing vacuum pump of claim 1, wherein the rotor shafts are supported by grease-lubricated bearings, one bearing being provided per rotor shaft in a housing wall through which the shaft ends are passed.

7. The dry-compressing vacuum pump of claim 1, wherein the vacuum pump compresses against atmosphere and generates a vacuum of at least 200 mbar absolute.

8. The dry-compressing vacuum pump of claim 1, furthere comprising a belt tensioning means provided on the housing wall.