US20130183185A1

US20130183185A1 - Screw rotor for a screw type vacuum pump

Info

Publication number: US20130183185A1
Application number: US13/737,787
Authority: US
Inventors: Jürgen Dirscherl; Frank Gitmans; Gerhard Rüster; Markus Prasse
Original assignee: Vacuubrand GmbH and Co KG
Current assignee: Vacuubrand GmbH and Co KG
Priority date: 2012-01-12
Filing date: 2013-01-09
Publication date: 2013-07-18
Also published as: CN103256224A; EP2615307A1; EP2615307B1

Abstract

A screw rotor is for a screw type vacuum pump, preferably for a screw type vacuum pump having a pumping capacity less than 50 m³/h. The rotor has a rotor shaft, a rotor core which rests on the rotor shaft, and a rotor cover which rests on the rotor core and at least partially encloses the rotor core. The rotor core is made of a material having a thermal conductivity greater than 100 W/m·K, preferably a thermal conductivity greater than 200 W/m·K. A screw type vacuum pump has correspondingly designed rotors.

Description

CROSS-REFERENCE TO RELATED APPLICATION

This application claims priority of European Patent Application No. 12 000 151.6, filed Jul. 12, 2012, which application is incorporated herein by reference.

FIELD

The present disclosure relates to screw rotors for a screw type vacuum pumps and screw type vacuum pumps.

BACKGROUND

Numerous processes in research and industry require a vacuum in the range of 10²Pa to 10⁻²Pa (fine vacuum range), wherein, frequently, condensing and/or aggressive vapors or gases must also be conveyed. To generate a negative pressure in this range, liquid-sealed or -lubricated vacuum pumps, such as oil-sealed rotary vane pumps, are used. The use of such pumps, in which the pumped medium comes into contact with oil or other liquids, has many disadvantages. The pumped media may contaminate the lubricant or react with it, which reduces the lubricating and sealing effect. Backflow of gaseous components or decomposition products of the lubricant into the process unit may significantly interfere with the processes therein.
For this reason, for quite some time there have been efforts to develop so-called “dry” vacuum pumps, i.e., pumps in which the pumped media do not come into contact with a liquid.
At higher pressures, i.e., in the range of 10⁵Pa to 10²Pa, diaphragm vacuum pumps are advantageous, since the pump chamber is hermetically separated from the drive area by the diaphragm, which is clamped in a gas-tight manner. However, due to the limited compression ratio and the valves which are usually activated only by the gas flow, pressures below 50 Pa are difficult to achieve.
In addition to fine vacuum pumps such as reciprocating pumps, scroll pumps, claw pumps, and Roots pumps, screw type vacuum pumps are known.
In screw type vacuum pumps (screw pumps for short), two helical rotors intermesh with one another in a contactless manner in a suitably shaped screw pump stator, so that due to their counterdirectional rotation, gas is conveyed from an inlet to an outlet. All statements in this regard, including in the discussion below, relate to oil-free screw pumps having contactless compression.
One advantage of the screw pumps is the high potential compression, since screw pumps may intrinsically have a multistage design in which each screw thread acts as a stage. Thus, screw pumps provide the possibility of achieving a good ultimate vacuum in the range of <1 Pa, using only one pair of rotors.
A so-called cantilevered bearing of this rotor pair is possible in screw pumps. In a cantilevered bearing, the bearing is provided from only one side of the rotor pair. No bearing is present on the other side of the rotor pair. Thus, the screw pump stator may be designed without a bearing unit. This allows simple disassembly of the screw pump stator for maintenance and cleaning, for example.
A general problem with screw pumps is the high heat release, in particular in the area of the compression on the atmosphere side. At low intake pressures, only a small quantity of gas is conveyed from the suction side to the atmosphere side. Thus, there is very little gas exchange inside the pump. In addition, a negative pressure prevails up to the last atmosphere-side screw thread in the pump chambers, which are formed by the intermeshing screw threads.
When the pump chamber is open at the last screw thread on the atmosphere side in the course of rotation of the rotor, gas flows from the outlet back into this pump chamber. The inflowing gas together with the gas that is conveyed here from the suction port is expelled in the course of rotation of the rotors. This pulsing of the gas at the outlet results in a high drive power requirement, and releases large quantities of heat in a relatively small volume.
The backflow of the gas may be reduced by end plates situated tightly against the screw rotor and having openings at appropriate positions. However, since these end plates at the same time hinder the gas discharge, little improvement is gained by this configuration.
Another approach for reducing the backflow is to provide check valves at such end plates. However, these check valves must open and close at the rotational frequency of the rotors. However, the frequency of typically 6000-25,000 min⁻¹is usually too high for this purpose; i.e., check valves having a sufficient size respond too slowly.
To reduce the temperature and power problem, screw rotors having a pump chamber volume which decreases toward the outlet are commonly used. This may be achieved, for example, by a reduced screw pitch or a reduced screw radius toward the outlet side. This results in an internal compression of typically 2 to 10. The power requirements of the pump and the heat release at the end of the screw on the atmosphere side, based on the pumping capacity of the pump, may thus be reduced almost by this compression factor.
A disadvantage of this method is that the manufacture of the rotors becomes much more difficult due to the continuous or also discontinuous change in the screw profile. Another disadvantage is that the internal compression at high intake pressures may result in internal overpressure. This may overload the drive motor and cause damage to the pump. Complicated pressure relief valves in the pump chamber stator in the area of the internal compression are therefore often necessary. When noncompressible liquids are conveyed, whether they are drawn in by suction or formed by condensation in the interior, hydrostatic blockages may develop, with the result that the pump abruptly stops due to overload. This may result in costly consequential damage to the unit and the drive.
Another approach is to use two separate screw pumps, having different pumping capacities, connected in series, each on its own having no internal compression (see EP 0 811 766 B1), whereby a pressure relief valve may be connected between the pumps (see WO 2007/088989 A1). However, these approaches as well result in a high level of design complexity (two pump units).
In known screw pumps of fairly large size, liquid cooling of the pump housing is often used to control the thermal conditions. For larger pumps, internal liquid cooling of the rotors is also used, but this involves a high expenditure of effort.
In addition, it is not uncommon for gas from the outside to also be admitted into the pump chamber in the area of the last screw thread on the atmosphere side. The purge gas cools this area and transports heated gas away from the last screw threads. The high level of complexity as well as the unavoidable deterioration of the ultimate vacuum of the pump are disadvantageous.
For compact screw pumps having typical rotor spacings of 20 to 100 mm and a pumping capacity <50 m³/h, internal liquid cooling of the rotors cannot be used due to space and cost reasons. In addition, liquid cooling of the housing would be disadvantageous in such equipment, which should have flexible use in research laboratories, for example, while the customary pumps, which are much larger, for weight reasons are usually stationarily installed in industrial facilities. Compact screw type vacuum pumps thus require novel approaches to control the difficult thermal situation at the end of the rotors on the atmosphere side.
Another aspect for compact screw type vacuum pumps is the selection of the material of the rotors. Such screw rotors, which are often designed in one piece with the rotor shafts, are usually made of cast iron alloys or steel alloys, since these materials have high rigidity (modulus of elasticity) and good machinability. The thermal conductivity of this material class is only moderate, but is generally adequate in conjunction with external water cooling and optional internal oil cooling. In addition, rotor temperatures of >150° C. at the surface are still acceptable for such materials.
A disadvantage of conventional steels and also cast iron alloys is the limited chemical resistance. Aggressive chemicals must be kept away from such pumps by means of cold traps or the like. Furthermore, complicated operations using purge gas are frequently involved. Nevertheless, when aggressive media are conveyed, such pumps often have only short service lives.
Steel alloys having high chemical resistance, such as Hastelloy, are usually difficult to machine, so that production of the screw profiles, which often have complicated shapes and narrow tolerances, is complex and expensive.
Another disadvantage of steel or cast rotors is their high weight, which has an adverse effect on the required drive power during acceleration, as well as the imbalance of the rotors. Approaches for avoiding this problem with the aid of a rotor made of aluminum on a steel shaft are known (DE 100 39 006 A1).
For applications using chemically aggressive substances, rotors made of chemically resistant plastics would be advantageous. Due to the limited rigidity (low modulus of elasticity) of plastics, a shaft made of a more rigid material is generally necessary inside the rotor. Such a system, composed of a rotor made of plastic with a steel rotor shaft in the interior, is known (WO 2010/061939 A1).
A disadvantage of the latter-mentioned system is that practically all usable plastics have a low thermal conductivity. Even with a high filler content of carbon fiber, for example, it is difficult to achieve a thermal conductivity greater than 1 W/m·K. For use in the screw pump, this means that with a high rate of heat release at the end of the rotor on the atmosphere side, the heat is not adequately dissipated, and the plastic material at that location quickly heats to high temperatures. This may result in excessive thermal expansion or even thermal damage (decomposition, melting) of the material. The high rate of thermal expansion is disadvantageous, since the rotors, which run by one another rapidly (typically >6000 min⁻¹) at a small clearance (typically <0.1 mm), may then contact one another, which may result in significant consequential damage.
The above-described problems with screw rotors for a screw type vacuum pump have previously been addressed in the prior art (GB 2 243 189 A). In the cited document, for an application in conjunction with chemically aggressive substances in a screw type vacuum pump, two rotors which run in mutual engagement are provided which are made of cast iron, but provided with a thin coating composed of protective materials, in particular plastic. The problem of the high heat release and the dissipation of the heat is not addressed therein. In fact, in the cited document, due to the design of the rotor cores of the rotors made of cast iron, the thermal conductivity is not high enough to actually protect the plastic material from destructive heating. Since in this case the rotor shaft is separate from the rotor core, i.e., the rotor core is wedged onto the rotor shaft, here as well no design is disclosed which represents an optimum solution for the heat dissipation.

SUMMARY

On the basis of the latter-mentioned prior art, the teaching of the present invention is based on an object of providing a screw rotor for a screw type vacuum pump in which use in the laboratory under chemically aggressive conditions is in any case possible from a design standpoint, and the above-described thermal problems are likewise solved.
With reference to a screw type vacuum pump as a whole, having two helical rotors in mutual contactless engagement with one another in a pump chamber of a screw pump stator which is shaped to fit, the above-described object can be achieved by the use of appropriately configured rotors.
In summary, the following advantages result for the rotors for compact screw pumps for use in research and industry, in particular using chemically aggressive substances:

- Effective cooling of the rotor is possible.
- The rotor has low thermal expansion.
- The rotor shaft has sufficiently high rigidity (modulus of elasticity).
- When a suitable material is used for the rotor cover, the surface of the rotor may have high chemical resistance and contact tolerance, i.e., without a tendency toward scoring upon contact with the opposite screw rotor.
- The rotor may be quite light in weight in order to reduce potential imbalances.
- When suitable materials are used for the rotor cover, the production of the screw profile, which is often very demanding and associated with narrow tolerances, is simplified by good machinability of the rotor material.

The present inventors have realized that the cooling of the screw rotor for screw pumps having a compact design must be carried out primarily by discharging heat from the pump chamber via the rotor and the rotor shaft. For this purpose it is provided to construct the rotor with a rotor core made of a highly thermally conductive material which is surrounded by a rotor cover, preferably made of a chemically resistant plastic. Materials having a thermal conductivity greater than 100 W/m·K, preferably greater than 200 W/m·K, for example aluminum or copper and some alloys, are used as highly thermally conductive material. Plastics and iron alloys (steel, cast iron) do not achieve these values, and do not have sufficient heat dissipation for effective cooling of the rotor toward the interior or via the rotor shaft.
In one design, the rotor shaft is also made of a highly thermally conductive material, so that the heat is transported from the pump chamber by solid-state heat conduction from the rotor via the rotor shaft. It is particularly advantageous if the rotor shaft is designed in one piece with the rotor core, since in that case the solid-state heat conduction occurs inside the rotor core, without interfering interfaces, into the rotor shaft and to the outside.
In an alternative design, instead of a solid, highly thermally conductive material a hollow shaft is used for the rotor shaft, through which a cooling gas such as air is conveyed, and which is drawn in by the rotation of the shaft itself, for example by a type of blower on a free end of the shaft. The cooling gas is led through the rotor shaft to the area of the highest heat release, and at that location cools the highly thermally conductive rotor core from the inside. The heated gas is delivered into the outlet of the pump, for example, where it may be used as a purge gas, or is recirculated back through the rotor shaft. Thus, in this design as well, the rotor is made of a highly thermally conductive rotor core which is surrounded by a rotor cover, the rotor core being in contact with the hollow rotor shaft or in direct contact with the cooling gas. The material of the hollow rotor shaft may also be highly thermally conductive.
The thickness of the casing material, i.e., the rotor cover, results from the need, on the one hand, for the layer to be diffusion-proof and mechanically stable, and on the other hand, for the thermal conduction through the layer to the core material to still be great enough to avoid overheating at the surface. This means that in one design the highly thermally conductive core material extends into the screw threads, and is not present only as an essentially cylindrical part. This means that the highly thermally conductive material, at least in this section, has the screw profile (reduced by the casing wall thickness), the casing material in these areas preferably having a thickness of 0.1 to 10 mm.
Although the rotor cover often has only a comparatively low specific thermal conductivity of usually <5 W/m·K (typical for plastics, for example), sufficient heat dissipation through this layer to the rotor core is achieved on account of the small thickness of the rotor cover. As a result of the application, the highly thermally conductive rotor core extends outwardly into the screw threads preferably in the area of the highest heat release, i.e., at the end of the rotor on the atmosphere side, in which a high level of heat dissipation through the rotor is necessary.
However, in one embodiment the rotor has one or more sections in which the highly thermally conductive rotor core does not extend, or does not extend completely, outwardly into the screw threads. Possible materials for the rotor core and the rotor cover are considered for purposes of explanation.
The rotor core may be encased in various ways. If a very thin coating (<0.1 mm) is applied, under some circumstances mechanical refinishing of the layer may be dispensed with. However, such layers are often not completely diffusion-proof, so that the layer may be infiltrated by the pumped media and then chip off under vacuum. For thicker coatings, the shape of the screw profile must be laboriously refinished. Thicker coatings are usually fused on after the application (for example, by electrostatic powder coating). This often leads to rounding of the edges, resulting in defects at the outer edges after the final machining.
A distinction is made between these coating processes and the extrusion coating of a rotor core with a thermoplastics plastic. In this process, the layer thickness may be selected practically as desired (i.e., so that it is also diffusion-proof), and the edges are precisely formed. At the same time, this process also allows filling of fairly large plastic volumes.
A comparison of the mechanical and thermal parameters of various materials shows that of the materials having a very high thermal conductivity of >100 W/m·K, copper and some copper alloys appear to be very suitable. The reason is the high thermal conductivity, the still acceptable thermal expansion, and the still acceptable modulus of elasticity. Aluminum and its alloys show much less favorable values for all three parameters, but are lighter in weight. Due to the much lower modulus of elasticity, aluminum is not very suitable as a rotor shaft material, but may be used as a rotor core material, in which case the rotor shaft would have to be made of a different material such as copper, or be composed of a hollow shaft having internal gas cooling. For protection of corrosion-sensitive materials such as copper, a coating with Ni, Cr, Ag, or Au, for example, may be provided.
Other metals having high thermal conductivity, such as gold, silver, alkali metals and alkaline earth metals, zinc, molybdenum, or tungsten and their alloys are ruled out because of excessive material costs, poor machinability, reactivity, or low modulus of elasticity. Novel materials such as CFRP often have anisotropic properties which are difficult to control, in particular in the shaping of solid bodies. Furthermore, the manufacture is often expensive and complicated. In addition, specialized ceramics such as AlN have interesting material properties, but are difficult to machine. Nevertheless, these materials are of interest in the future for the rotor core or parts of the rotor core of rotors according to the invention.
Disadvantages of copper are its high specific weight and comparatively high material cost. Therefore, the second aspect of the invention, namely, that the highly thermally conductive core material extends outwardly to just under the plastic surface only where this is thermally necessary, i.e., in particular in the area of the compression on the atmosphere side, is particularly advantageous. In other areas of the rotor which are subject to less thermal stress, the formation of the core material until it reaches into the screw threads may be dispensed with, and at such location the rotor may be composed of a relatively small cylindrical rotor core that is surrounded by plastic as the rotor cover.
Chemically highly resistant plastics such as PPS, PEEK, or fluoroplastics, which preferably are reinforced with fillers such as carbon fiber or glass fiber, are preferably used as casing material. For example, the density of PEEK with carbon fiber reinforcement is only approximately 16% that of copper. Thus, by means of the system in which the highly thermally conductive core material extends outwardly to just under the plastic surface only where this is thermally necessary, i.e., in particular in the area of the compression on the atmosphere side, the weight of the screw rotor and thus, potential imbalances, may be significantly reduced. Profile-related imbalances may be largely compensated for at the rotor core itself, so that only minor corrections to the rotor cover are necessary at the completed rotor, and large balancing rings or boreholes may be dispensed with.
Further advantages of materials such as PPS, PEEK, or fluoroplastics are the good machinability and the contact tolerance, i.e., a low tendency toward scoring. The machining of such plastics is much simpler and quicker, and thus more cost-effective, than highly corrosion-resistant stainless steels, for example.
The system thus results in a rotor having a chemically highly resistant and diffusion-proof surface, and at the same time, very high thermal conductivity of the overall system, at least in the area of high heat release during operation, with surprisingly favorable manufacturing costs. The latter results due to the fact that materials such as copper or aluminum are used only where necessary, material-conserving manufacturing processes such as injection molding are used, and the materials are easily machinable.
For reliably connecting the rotor cover to the rotor core, a positive-fit connection, preferably having undercuts for interlocking, is necessary if adequate material adhesion is not achievable. For this purpose, grooves, holes, or channels, for example, may be introduced into the rotor core. A rough surface of the rotor core, achieved by sandblasting, for example, is also helpful.
The advantages of the rotor are particularly apparent in the preferred configuration of the screw pump having cantilevered rotors. For cantilevered rotors, the bearing and drive area is preferably under ambient air pressure, and is not in contact with the pumped media. To keep this bearing and drive area from having to be sealed off from the pump chamber by a shaft sealing ring or the like, the pressure side of the pump unit is usually situated on the drive side.
This area is doubly subjected to thermal stress: on the one hand from the motor, and on the other hand from the heat of compression at the end of the screw rotor on the atmosphere side. However, when high-efficiency synchronous motors or a gearing is/are used, and when the drive area is efficiently cooled using a blower, for example, the drive area may be kept at low operating temperatures fairly easily.
The waste heat from the compression may be several times that of the motor waste heat. The rotor design now allows very effective dissipation of the heat of compression from the pump chamber in the direction of the well-cooled drive area, with the aid of the highly thermally conductive rotor shaft made of a solid material.
In one design, a means for delivering this heat to the ambient air is situated in the drive area, on the rotor shaft. This may be, for example, a co-rotating fan impeller or disks made of copper or aluminum, for example. These elements deliver the heat from the rotor shaft to the air very effectively due to the rapid rotation. The heated air may be discharged via an externally applied cooling air flow. An air flow generated by a co-rotating fan impeller may also be used for cooling the motor.
In particular for cantilevered rotors, it is important that the lowest possible weight is present at the end of the rotors remote from the bearings. Motions of large masses at a large distance from the bearing, even with slight imbalance, may result in great deflections, and thus, rotor collisions.
In this case a second aspect becomes particularly important, according to which the highly thermally conductive rotor core (often having a high density) outside the highly thermally stressed part of the rotor, is not drawn to just under the plastic surface, but, rather, is reduced to the greatest extent possible. This significantly reduces the moved mass specifically at the end of the rotor remote from the bearing.
For the dimensionally accurate machining of the cantilevered rotors (in the pump), depending on the production method it is also necessary to hold the rotor in the machining tool on the side facing away from the bearing. If the rotor cover is not suited for this purpose, the highly thermally conductive rotor core may also be guided outwardly on the end-face side, on the side facing away from the bearing. If necessary, this area must subsequently be protected from corrosion attack, for example by covering with a plug made of PTFE, for example. Alternatively, the holding at the end face may also be carried out using a highly corrosion-resistant metal such as Hastelloy which is fixedly joined to the core material.
In an alternative form of the rotor for a cantilevered bearing, the highly thermally conductive rotor core and/or the rotor shaft is/are not present with the complete cross section over the entire length of the screw rotor, or is/are hollow or absent altogether. The part of the rotor facing away from the bearing may then be made of solid casing material or have a recess. All of these characteristics result in a marked reduction of the moved masses in the area of the rotor remote from the bearing.
The highly thermally conductive rotor core, for example made of copper or aluminum or an alloy thereof, may be manufactured from the solid material, or preferably by affixing a hollow screw to a shaft or by joining a solid screw having a short shank, both of which reduce the material expenditure for the manufacture. In one design with even less material usage, the rotor core as a whole or the hollow screw is pre-cast, or the latter is made of an appropriately curved sheet metal part.
In another embodiment, additional functional elements of the rotor are integrated into the rotor cover. These may be, for example, balancing weights on one or both sides of the screw, or also purge gas fans as disclosed in DE 10 2010 055 798 A1.
In another embodiment, the drive of the screw pump is achieved by a dual-shaft synchronous drive composed of magnetized cylinders on each of the two rotor shafts, which due to their mutual magnetic interaction synchronize the rotors in opposite directions. The two magnetized cylinders are enclosed by one or more windings, which when suitably energized generate migrating magnetic fields so that the two magnetized cylinders, and thus the rotor shafts, rotate synchronously in opposite directions.

The present disclosure is explained in greater detail below with reference to drawings which illustrate exemplary embodiments strictly by way of example. The drawings show the following:

FIG. 1 shows a screw rotor for a screw type vacuum pump, in cross section; and

FIG. 2 shows a screw type vacuum pump having two rotors, in cross section.

FIG. 1 shows a screw rotor 1 in cross section. The rotor 1 is intended for use in a screw type vacuum pump, for example in a screw type vacuum pump having a pumping capacity less than 50 m³/h.
The rotor 1 schematically illustrated in cross section in FIG. 1 basically consists of a rotor shaft 2, a rotor core 3 which rests on the rotor shaft 2, and a rotor cover 4 which rests on the rotor core 3. As illustrated, the rotor shaft 2 is separate from the rotor core 3. In principle, it is also possible for the rotor shaft 2 and the rotor core 3 to be designed as one piece.
The rotor cover 4 at least partially encloses the rotor core 3. In the exemplary embodiment illustrated in FIG. 1, the rotor cover 4 encloses the rotor core 3 on the rotor shaft 2 at all outer surfaces, i.e., at all surfaces which do not abut against the rotor shaft 2.
In the illustrated exemplary embodiment, the rotor core 3 is made of a material having a high thermal conductivity greater than 100 W/m·K, preferably a thermal conductivity greater than 200 W/m·K. In addition, the rotor shaft 2 is preferably made of a material having a high thermal conductivity, in the present case, preferably a thermal conductivity greater than 100 W/m·K.
Alternatively or additionally, the rotor shaft 2 may have one or more channels extending parallel to the axis of the rotor shaft for supplying gas in the direction of the rotor core 3, so that the rotor 1 as a whole is cooled from the inside.
The rotor core 3 in individual sections of the rotor 1 may extend into the screw threads thereof, as illustrated in region 5 in FIG. 1. Thus, the rotor core 3 then has practically the same outer dimensions as the rotor 1 as a whole, with only a thin layer which forms the rotor cover 4. In this area, thicknesses of the rotor cover 4 between 0.1 mm and 10 mm are conceivable. This design is implemented in particular at locations where significant heat develops in a screw type vacuum pump during operation of the rotor 1, thus, in particular where the compression occurs at atmospheric pressure, near the outlet of a pump chamber of a screw type vacuum pump.
In less stressed areas, the rotor core 3 may be entirely absent, so that the rotor cover 4 may form the complete rotor 1 outside the rotor shaft 2. This is apparent in region 7 at the top in FIG. 1.
With regard to the rotor cover 4, it can be made of a material which has a low thermal conductivity compared to the thermal conductivity of the rotor core 3 and of the rotor shaft 2, for example a thermal conductivity less than 5 W/m·K. In particular, it is recommended here that the rotor cover 4 is made of plastic, in particular a thermoplastic plastic. In appropriate uses for chemical applications, the selection of a chemically resistant plastic such as PPS, PEEK, or fluoroplastic is recommended. The strength of the plastic of the rotor cover 4 may be increased using fillers such as glass fibers or carbon fibers.
The rotor cover 4 is joined to the rotor core 3, i.e., mounted thereon, in an injection molding process. Copper or aluminum or alloys of these materials are recommended as materials for the rotor core 3 or parts thereof, and/or for the rotor shaft 2.
FIG. 1 shows the rotor shaft 2 of the rotor 1 protruding at both ends, i.e., projecting significantly with respect to the rotor core 3 and the rotor cover 4. The rotor 1 is supported at both ends.
In contrast, the rotors 1, 1′, which are installed in the screw type vacuum pump as illustrated in FIG. 2, are configured for a one-sided bearing at one end. In this case, the rotor shaft 2 has a significant projection with respect to the rotor core 3 and the rotor cover 4 only at its end used as the bearing, namely, protrudes into a bearing area.
In the rotors 1, 1′ in FIG. 2, it is also apparent that in the area of the end of the rotor 1, 1′ facing away from the end used for the bearing, the rotor shaft 2 and/or the rotor core 3, depending on the distance from the end used for the bearing, has/have a reduced cross section, a recess, or is/are missing completely. The volume missing compared to the complete outer dimensions of the rotor 1, 1′ is filled by the rotor cover 4.
FIG. 2 shows a schematic sectional view of a screw type vacuum pump having helical rotors 1, 1′ in mutual contactless engagement with one another, inserted therein. The screw type vacuum pump in FIG. 2 has, first of all, a screw pump stator 8 which essentially forms the housing of the screw type vacuum pump. The screw pump stator 8 contains a pump chamber 9, shaped to fit the rotors 1, 1′, which has at least one inlet 10 and one outlet 11. The gaseous medium is conveyed from the inlet 10 to the outlet 11 by the contactless rolling off of the two counter-rotating rotors 1, 1′ in the appropriately shaped pump chamber 9. The rotors 1, 1′ together with the rotor shaft 2, rotor core 3, and rotor cover 4 are configured in the same way as described in detail for the rotor 1 illustrated in FIG. 1.
The rotors 1, 1′ in the exemplary embodiment in FIG. 2 differ from the rotor 1 in FIG. 1 in that the rotors 1, 1′ are cantilevered, i.e., supported only on one side. There is no bearing on the opposite end of the rotors 1, 1′, i.e., at the top in FIG. 2.
In FIG. 2, a bearing and drive area in which the rotor shafts 2 of the rotors 1, 1′ are supported is situated beneath the pump chamber 9 in the screw pump stator 8. It is apparent that the outlet 11 of the pump chamber 9 is situated at the end of the pump chamber 9 facing the supported ends of the rotors 1, 1′.
The bearing and drive area is preferably under ambient air pressure. This area contains means 12, 12′; 13, 13′ for bearing the rotors 1, 1′, and means for synchronizing and/or for driving the rotors 1, 1′. The means 12, 12′; 13, 13′ can be for example roller bearings or ball bearings. FIG. 2 depicts roller bearings. In the example illustrated here, the latter means are composed of appropriately magnetized cylinders 14, 14′ which due to their mutual magnetic interaction synchronize the rotors 1, 1′ in opposite directions. The two magnetized cylinders 14, 14′ are surrounded by one or more windings 15, 15′, which when suitably energized generate migrating magnetic fields so that the two magnetized cylinders 14, 14′, and thus the rotor shafts 2 of the rotors 1, 1′, rotate synchronously in opposite directions. Thus, in the present case the drive of the screw type vacuum pump is configured as a dual- shaft synchronous drive 14, 14′; 15, 15′. These types of designs are known per se from the prior art.
Illustrated on the shaft in the drive area are heat transfer means 16, 16′ for discharging heat, which has been conducted here via the rotor shafts 2, to the ambient air. These may be co-rotating fan impellers or disks, for example. The heated air may be discharged via an externally applied cooling air flow (not illustrated). The air flow generated by the co-rotating heat transfer means 16, 16′ may also be used for cooling the drive 14, 14′; 15, 15′.
In addition, further functional elements 17, 17′ are indicated in FIG. 2 which may be used for balancing, for example. Functional elements 17, 17′ attached to the rotor cover and can be balancing weights which function to counter-balance the rotor itself, fan wheels, and/or the like. In one example, these may be purge gas fans for drawing in purge gas from the bearing area and thus flushing the bearings.
In the end region 18 illustrated in FIG. 2, the rotor cover 4 of the respective rotor 1, 1′ has axially inwardly extending recesses. Beneath the recesses, the rotor cover 4 of both rotors 1, 1′ in each case extends over the complete cross section of the rotor 1, 1′, transversely with respect to the axis of the rotor 1, 1′, since the respective rotor shaft 2 terminates just below this area.

Claims

1. A screw rotor for a screw type vacuum pump, the screw rotor comprising:

a rotor shaft, a rotor core which rests on the rotor shaft, and a rotor cover which rests on the rotor core and at least partially encloses the rotor core,

wherein the rotor core is made of a material having a thermal conductivity greater than 100 W/m·K.

2. The rotor according to claim 1, wherein

the rotor shaft is made of a material having a thermal conductivity greater than 100 W/m·K.

3. The rotor according to claim 2, wherein

the rotor shaft and the rotor core are designed as one piece.

4. The rotor according to claim 1, wherein

the rotor core extends into the screw threads of the rotor.

5. The rotor according to claim 4, wherein

the rotor core extends into the screw threads of the rotor only in an area of the rotor which during operation faces an outlet of a pump chamber.

6. The rotor according to claim 1, wherein

the rotor cover is made of a material which has a low thermal conductivity compared to the thermal conductivity of the rotor core and of the rotor shaft.

7. The rotor according to claim 6, wherein

the rotor cover is made of plastic.

8. The rotor according to claim 7, wherein

at least one of the rotor core, parts thereof, and the rotor shaft is made of copper, aluminum, or alloys of these materials.

9. The rotor according to claim 1, wherein

the rotor is configured for a one-sided bearing at only one end of the rotor shaft.

10. The rotor according to claim 9, wherein

in an area of the end of the rotor facing away from the one end used for the bearing, the rotor shaft has a reduced cross section, a recess, or is missing completely, and a volume that is missing compared to an otherwise complete outer dimension of the rotor, is filled by the rotor cover.

11. The rotor according to claim 1, wherein

a heat transfer means for delivering heat to the ambient atmosphere is situated on the rotor shaft at a distance from the rotor core.

12. A rotor according to claim 1, wherein the rotor core is made of a material having a thermal conductivity greater than 200 W/m×K.

13. A rotor according to claim 4, wherein the rotor core, at a location where the rotor core extends into the screw threads of the rotor, has a reduced thickness.

14. The rotor according to claim 13, wherein the rotor cover has a thickness of 0.1 mm to 10 mm in the location where the rotor core extends into the screw threads of the rotor.

15. The rotor according to claim 6, wherein the rotor cover is made of a material which has a thermal conductivity less than 5 W/m.

16. The rotor according to claim 7, wherein the plastic is thermoplastic.

17. The rotor according to claim 7, wherein the plastic is a chemically resistant plastic selected from the group consisting of PPS, PEEK, and fluoroplastic.

18. The rotor according to claim 7, wherein the plastic is reinforced with a filler selected from the group consisting of carbon fibers and glass fibers.

19. The rotor according to claim 1, wherein the screw type vacuum pump has a pumping capacity less than 50 m³/h.

20. A screw type vacuum pump comprising:

a screw pump stator with at least one inlet and one outlet, and

two helical rotors which rotate in mutual contactless engagement with one another in a fittingly shaped pump chamber of the screw pump stator, and thus convey a gaseous medium from the inlet to the outlet,

wherein the rotors each comprises:

21. The screw type vacuum pump according to claim 20, wherein

the outlet of the pump chamber is situated at an end of the pump chamber facing the supported ends of the rotors.

22. The screw type vacuum pump according to claim 20, comprising

a dual-shaft synchronous drive for driving the rotors.

23. The screw type vacuum pump according to claim 20, wherein the screw type vacuum pump has a pumping capacity less than 50 m³/h.