EP3737863A1

EP3737863A1 - Compressor

Info

Publication number: EP3737863A1
Application number: EP19700332.0A
Authority: EP
Inventors: Thomas Dreifert; Kai NADLER; Bernhard Kliem; Roland Müller
Original assignee: Leybold GmbH
Current assignee: Leybold GmbH
Priority date: 2018-01-12
Filing date: 2019-01-04
Publication date: 2020-11-18
Also published as: WO2019137852A1; DE202018000178U1; CN111448392A; JP2021510404A; KR20200105817A; US20200362861A1; CN111448392B

Abstract

A dry-running rotary-screw compressor comprises two screw rotors in a housing (26) that forms a suction chamber. Preferably atmospheric pressure is applied to a compressor inlet (28) of the compressor and preferably a pressure of more than 2 bar (absolute) is applied to a compressor outlet (32) of the compressor. At least one displacement element (10, 12) is provided per screw rotor and has a helical recess with several turns. The at least one displacement element (10, 12) per screw rotor has an asymmetric profile with a single thread.

Description

compressor

The invention relates to a compressor, in particular a screw compressor.

For the compression of gases, in particular for the supply of compressed air, today primarily oil-injected screw compressors are used. With these, a compression of lbar (absolute) to 8.5 bar to 14 bar (absolute) can usually be carried out in a compressor stage. The subsidized intake volume flows are in the range of 30 to 5000 m ³ / h. Such screw compressors have two counter-rotating screw rotors. The screw rotors each have at least one screw-line-shaped depression, so that a displacement element is formed. The injection of oil into the pump chamber, in which the two screw rotors are arranged, serves to seal the gap between the rotors and the housing or the inner wall of the pump chamber. Only due to the provision of oil, sufficient tightness can be achieved in order to achieve high compression pressures, in particular up to 14bar in a compressor stage can. Furthermore, the oil serves to lubricate the rolling contacts between the two screw rotors. This makes it possible to dispense with a synchronization transmission for the two screw rotors. Furthermore, the oil serves to dissipate heat of compression. Only then can a low temperature be achieved with high efficiency. Lastly, that serves Oil also for damping mechanical noises. A major disadvantage of the use of oil is that the oil also reaches the gas to be delivered. The oil must be removed by multi-stage separators from the compressed air. This has the consequence that such compressors are expensive and require a lot of space. The use of oil-injected screw compressors is especially in areas where a high cleanliness of the compressed air is required, such as in areas of the pharmaceutical or food industry, not possible or only with extremely complex multi-stage oil separators.

To produce oil-free compressed air, it is known to use dry-compressing screw compressors. The two screw rotors are arranged without contact and synchronized with each other via an oil-lubricated gearbox. However, dry compressing screw compressors have the disadvantage that with a compressor stage only a compression to 4 to 5 bar (absolute) is possible. This is due in particular to the fact that high gaps occur through the gaps between the rotors and the housing. For example, to achieve pressures of 9 bar (absolute), two-stage screw compressors must be used. In addition to the two compressor stages, an intermediate cooling of the compressed air is also necessary, resulting in complex systems with many components and a large installation space.

Furthermore, so-called rotary-tooth compressors are known as dry-compressing compressors. These too have the disadvantage that they have to be designed to achieve high pressures of about 9 bar (absolute) in several stages.

Furthermore, dry compressing spindle compressors are known. These have a plurality of closed working chambers connected in series over a plurality of turns or wraps of a displacer. Theoretically, high compression pressures should also be achieved in one stage can be replaced so that spindle compressors multi-stage screw compressors or rotary tooth compressors could be replaced. However, spindle compressors have not been on the market so far, so that practical proof that high compression pressures can be achieved in one stage has not yet taken place. Spindle compressors are described, for example, in DE 10 2010 064 388, WO 2011/101064, DE 10 2012 202 712 and DE 10 2011 004 960.

The object of the invention is to provide a dry compressing compressor, with the one-stage high pressures of more than 5bar (absolute) can be achieved in particular.

The object is achieved according to the invention by a dry-compressing compressor according to claim 1.

The dry compressing compressor according to the invention has a suction chamber formed by a housing. Two intermeshing screw rotors are arranged in the pump chamber. These are rotated in opposite directions to convey the gas. For this purpose, each screw rotor has at least one displacement element which has a helical recess for forming the turns. In particular, only one displacement element can be provided per screw rotor, which may optionally be formed integrally with a rotor shaft. Furthermore, the housing has a compressor inlet, to which preferably atmospheric pressure is applied. A pressure of more than 2 bar (absolute) is preferably applied to a compressor outlet, wherein it is particularly preferred for the compressor outlet to have a pressure of more than 5 bar (absolute).

With the dry-compressing compressor according to the invention, it is possible to achieve high levels of pressure since, according to the invention, the at least one displacement element per screw rotor is designed to be catchy and has an asymmetrical profile. In a particularly preferred embodiment that is asymmetric profile designed such that no or at most a small blow hole is formed. Since there is no through-blowing hole, a short-circuit is only provided between two adjacent chambers in the case of an asymmetrical profile which is preferably designed according to the invention. In a particularly preferred embodiment, the so-called Quimby profile is provided as the asymmetrical profile. Asymmetrical profiles have two different profile flanks. Although these are expensive to manufacture due to the required two separate steps, this can be achieved by a very dense working chamber.

Even the provision of catchy possibly also symmetrical rotor profiles already has the advantage that a higher density can be achieved. In the case of profiles having two or more passages of the displacement elements which mesh with each other, the column results in connections across a plurality of chambers, so that the leakage adversely affects the conveyed gas flow and the energy conversion quality.

In a further preferred embodiment of the dry compressing compressor according to the invention, the number of turns of the at least one displacement element or, in the case of a plurality of displacement elements, the sum of the turns of the displacement elements of a screw rotor is greater than the ratio of the pressure prevailing at the compressor outlet to the pressure prevailing at the compressor inlet. The number of turns n thus results

P pause

Pein where paus is the outlet pressure and p _{a is} the inlet pressure of the compressor. It is particularly preferred that the number of turns or loops is calculated according to n> Paus + 4.

anguish By such a large number of turns or wrappings per screw rotor is a continuous, but relatively slow compression of the gas. This makes it possible to dissipate the heat generated during the compression well.

Furthermore, it is preferable that the built-in volume ratio of the dry compressing screw compressor between the theoretical delivery volume at the inlet stage (Vein) and the theoretical delivery volume at the outlet stage (V _out ) to the pressure ratios at the inlet (Pein) and at the outlet (P _out ) is adjusted. Here, p _and p _from as de- finiert absolute pressures. Preferred is a volume ratio V of

where n has a value of k-0.3 to k + 0.3 and preferably a value between k-0.1 and k + 0.1 k is the isotropic exponent of the gas mixture to be delivered.

In a further preferred embodiment, the displacement elements have at least one region or section in which the chamber volume decreases from a volume Vein to an intermediate volume VVK.

In a further preferred or alternative embodiment, the reduction of the delivery volume of the stages (working chambers) from the large inlet volume (Vein) to the smaller outlet volume (V _out ) is divided into two areas. In this case, it is particularly preferred that, in a first region, the working chamber closed off towards the suction side is reduced within a small rotation angle range to a specific volume (volume of the pre-compression VVK). It is preferred that

VVK - X Vein where x = 0.1 to 0.5, in particular x = 0.2 to 0.4 and more preferably x = 0.3. The precompression raises the temperature of the gas by the compression work to a moderate value of 150 ° C - 200 ° C. In the second region of the compression, the working chamber volume decreases considerably less strongly as a function of the angle of rotation than in the first region. The angle of rotation and thus the number of stages is considerably higher in the second range than in the first range. Due to the moderate increase in temperature in the first region, the large housing surface in the second region and the relatively long residence time of the gas in the second region due to the larger angle of rotation, in the second region by heat transport into the housing, a further increase in temperature of the gas through the compression is largely avoided.

The compression of the gas taking place here is selected such that the resulting heat of compaction can be well dissipated via the side walls of the housing so that the temperature of the gas does not increase or increases only slightly. The maximum temperature change is preferably less than 50 ° C., and more preferably less than 30 ° C.

A particular advantage of the selected distribution of the volume reduction is that a largely homogeneous temperature distribution is achieved in the component. As a result, thermal peak loads and the associated strong component strains can be prevented.

The ratio between the inlet volume (Vein) and the volume of the precompression (transition from the first to the second range, VVK) can be related to the internal volume ratio v, of the compressor where j = 2 to 5, in particular j = 2.5 to 3.5 and particularly preferably j = 3. In a particularly preferred embodiment, the precompression takes place in the described first range within 1.5 to 3 rotor revolutions (turns).

The inventively high number of turns in the second region can be achieved here in a preferred embodiment by a single displacement element per rotor. However, it is also possible to provide a corresponding number of turns in this pressure-side region, for example by two displacement elements. By providing a high number of windings in accordance with the invention in this region, in which, according to the invention, preferably only a relatively small compression of the medium to be conveyed takes place per turn, it is possible to dispense with rotor internal cooling. This is due in particular to the fact that, due to the relatively low compression in this area, the temperature increase of the displacement element caused by the compression is lower. Furthermore, due to the high density of the pumped medium through the medium itself, good heat dissipation from the displacement element into the compressor housing takes place in this area.

Preferably, the screw rotors and the at least one provided displacement element are designed such that at least 6, in particular at least 8 and particularly preferably at least 10 windings are provided between an area in which 5% -20% of the outlet pressure prevails and the pressure-side rotor end are. The pressure-side rotor end is in this case the area of the compressor outlet. The inventively high number of turns in this area can be provided here in a preferred embodiment in the case of a single pressure-side displacement element provided for each rotor. However, it is also possible to provide a corresponding number of turns in this pressure-side region, for example on two displacement elements. By providing a high number of turns in accordance with the invention in an area in which, according to the invention, then only a relatively small compression of the medium to be delivered takes place per turn, it is possible to dispense with a rotor internal cooling. This is due in particular to the fact that, due to the relatively low compression in this area, the temperature increase of the displacement element caused by the compression is lower. Furthermore, due to the relatively high density of the medium in this area, the conveyed medium itself also provides good heat removal from the displacement element into the compressor housing.

In addition, due to the preferably high number of turns, a large surface is available for heat exchange with the housing.

It is particularly preferred that the preferably at least 6, in particular at least 8 and particularly preferably at least 10 turns are provided in a pressure-side displacement element.

Furthermore, for the inventive design of screw rotors without internal cooling, it is preferred that the pressure-side displacement element has an average working pressure of more than 2 bar (absolute) on at least 6, in particular at least 8 and particularly preferably at least 10 turns. The aim is in particular a shallow pressure gradient within the compressor. Therefore, the pressure should rise slowly over many turns of, in particular, 6 to 10 turns.

According to the invention, it is thus possible, even in rotors without rotor cooling and with a housing made of aluminum or an aluminum alloy between the surface of the at least one displacement element and the inside of the pump chamber, in particular in the pressure-side region, a cold gap with a Height of 0.03 mm - 0.2 mm and in particular 0.05 mm - 0.1 mm. Such a relatively large gap height can be due to the inventive, described above Embodiment of the particular 6, preferably 8 and more preferably 10 last turns are provided.

In a further preferred embodiment of the invention, a comparison with the diameter relatively long screw rotor is selected. In particular, the at least one displacement element per screw rotor or in the case of a plurality of displacement elements per screw rotor together have a ratio of length L to diameter D, for which the following applies: and particularly

By providing a long rotor, in particular many chambers, the usable area for heat dissipation is increased. Due to the resulting good heat exchange, the gas temperatures of the compressed gas are relatively low. The provision of many chambers also has the advantage that the pressure differences between adjacent chambers are small and thereby a high density can be achieved. By reducing the flow rate per stage from the inlet to the outlet side, the compression process becomes thermodynamically particularly effective and the gas temperatures remain relatively low. It is particularly preferred here for the internal volume ratio to be adapted to the ratio of outlet pressure to inlet pressure such that neither over-compression nor compression due to re-aeration occurs.

The inner volume ratio can be achieved by varying the pitch of the turns. Preferably, a change in the pitch of the turns, in particular such that these from the compressor inlet to Compressor outlet decreases or becomes steeper. The change in slope can be continuous and / or in stages.

In addition to or instead of the variation of the pitch, a change in the head and foot diameters of the profile can also take place in steps or continuously. Again, a continuous change of the head and foot diameter is preferred, so that the rotor is conical, in particular in combination with a continuous change in the pitch.

In a particularly preferred embodiment, the pressure ratio between the outlet pressure and the inlet pressure is at least 5. In a particularly preferred embodiment, the outlet pressure is at least 2 bar (absolute), in particular at least 5 bar.

In a further particularly preferred embodiment, the dry-compressing compressor at the compressor inlet and / or at the compressor outlet in each case preferably within the compressor housing to a gas collecting space.

Incidentally, it is preferable that the dry compressing screw compressor is a two-shaft compressor. These are preferably mounted on both sides, so that narrow gaps can be realized both between the displacement elements and between the displacement elements and the inner wall of the suction chamber. Preferably, a synchronization of the two rotor shafts by a preferably arranged outside of the pump chamber synchronization gear. The bearing lubrication can be done with grease and / or oil. Likewise, the gearbox lubrication can take place via grease and / or oil. This is possible insofar as both the bearings and the synchronization gear are preferably arranged outside the pump chamber and thus is further avoided that the gas to be delivered is contaminated with oil. Preferably, the housing is made of aluminum or an aluminum alloy. Particularly preferred here as the aluminum alloy for the housing AISi7Mg or AIMgO, 75Si. In particular, the coefficient of thermal expansion (coefficient of expansion) of the material of the screw rotors is less than the expansion coefficient of the material of the housing. It is particularly preferred that the coefficient of expansion of the screw rotors is smaller than 12 * 10 ⁶ 1 / K. This can be achieved by rotors made of iron or steel materials.

The two screw rotors arranged in the pump chamber have at least one displacement element which has a helical recess. The helical recesses form several turns. According to the invention, the at least one displacement element is made of a steel or an iron alloy. It is thus particularly preferred that the screw rotors, including the displacement elements, are made of a steel or iron alloy. The housing is likewise produced from a steel or iron alloy or from aluminum or an aluminum alloy.

Each displacement element preferably has at least one helical recess which has the same contour over its entire length. The contours are preferably different for each displacement element. The single displacement element thus preferably has a constant pitch and a constant contour. This simplifies the production considerably, so that the production costs can be greatly reduced.

To further improve the suction power, the contour of the suction-side displacement element, that is to say in particular of the first displacement element in the pumping direction, is preferably asymmetrical. As a result of the asymmetrical design of the contour or of the profile, the flanks can be designed such that the leakage surfaces, the so-called blow-off more preferably completely disappear or at least have a small cross-section. A particularly suitable asymmetric profile is the so-called "Quimby-ProfM". Although such a profile is relatively difficult to produce, it has the advantage that there is no continuous blow hole. A short circuit exists only between two adjacent chambers. Since this is an asymmetrical profile with different profile flanks, at least two working steps are required for the production because the two flanks must be produced in different work steps due to their asymmetry.

The pressure-side displacement element, in particular the last displacement element in the pumping direction, is preferably provided with a symmetrical contour. The symmetrical contour has the particular advantage that the production is easier. In particular, both flanks with a symmetrical contour can be produced by a rotating end mill or by a rotating side milling cutter in one work step. Although such symmetrical profiles have blow holes, they are continuous, ie not only provided between two adjacent chambers. The size of the blow hole decreases as the slope decreases. In this respect, such symmetrical profiles can be provided in particular in the case of the pressure-side displacement element, since in a preferred embodiment it has a smaller pitch than the suction-side displacement element and preferably also as the displacement elements arranged between the suction-side and the pressure-side displacement element. Although the density of such symmetrical profiles is slightly lower, they have the advantage that the production is much easier. In particular, it is possible to produce the symmetrical profile in a single work step and preferably with a simple end mill or side milling cutter. This reduces the costs considerably. A particularly suitable symmetrical profile is the so-called "cycloid profile". The provision of at least two such displacement elements means that the corresponding screw compressor can generate high outlet pressures at low power consumption. The thermal load is also low. Arranging at least two displacement elements having a constant pitch and a constant contour in a compressor designed according to the invention leads to substantially the same results as in a compressor having a displacement element with a changing pitch. At high built volume ratios can be provided per rotor three or four displacement elements.

In order to increase the achievable outlet pressure and / or to reduce the power consumption and / or the thermal load, in a particularly preferred embodiment, a pressure-side displacement element, that is, in particular in the pumping direction last displacement element, has a large number of windings. Due to the high number of turns, a larger gap between the screw rotor and the housing can be accepted with consistent performance. The gap can have a cold gap width of 0.05-0.3 mm. A large number of outlet turns or number of turns in the pressure-side displacement element is inexpensive to produce, since according to the invention this displacement element can have a constant pitch and preferably also a symmetrical contour. On the outlet side, an asymmetric profile can also be used. As a result, a simple and inexpensive production is possible, so that the provision of a larger number of turns is acceptable. This pressure-side or last displacement element preferably has more than 6, in particular more than 8, and particularly preferably more than 10 turns. The use of symmetrical profiles in a particularly preferred embodiment has the advantage that both flanks of the profile can be cut simultaneously with a milling cutter. In this case, the milling cutter is additionally supported by the respectively opposite flank, so that deformation or bending of the milling cutter during the milling process and thus caused inaccuracies are avoided. To further reduce the manufacturing cost, it is particularly preferred to form the displacement elements and the rotor shaft in one piece.

In a further preferred embodiment, the pitch change between adjacent displacement elements is discontinuous or erratic. Optionally, the two displacement elements are arranged in the longitudinal direction at a distance from each other, so that between two displacement elements a circumferential cylindrical chamber is formed, which serves as a tool outlet. This is particularly advantageous for the production of integrally formed rotors, since the helix producing tool can be brought out in this area in a simple manner. If the displacement elements are produced independently of each other and then mounted on a shaft, it is not necessary to provide a tool outlet, in particular such a ring-cylindrical area.

In a preferred embodiment of the invention, no tool outlet is provided between two adjacent displacement elements on the change of pitch. In the area of the pitch change, preferably, both flanks have a defect or recess in order to be able to lead out the tool. Such a defect has no appreciable influence on the compression capacity of the compressor, since it is a locally highly limited defect or recess.

The compressor screw rotor according to the invention has in particular a plurality of displacement elements. These can each have the same or different diameters. In this case, it is preferred that the pressure-side displacement element has a smaller diameter than the suction-side displacement element. In the case of displacement elements which are produced independently of the rotor shaft, they are mounted on the shaft, for example by press fits. In this case, it is preferable to provide elements such as dowel pins for fixing the angular position of the displacement elements to one another.

It is particularly preferred that the screw rotor is made in one piece, in particular one made of steel or an iron alloy. The screw rotor can also have a rotor shaft that carries the at least one displacement element. This has the advantage, in particular when providing a plurality of displacement elements, that they can be produced independently of one another and are then connected to the rotor shaft in particular by being pressed on or shrink-fitted. In this case, it is possible to provide feathering keys or the like for defining the angular position of the individual displacement elements.

In the preferred provision of several displacement elements per screw rotor, it is possible to form the displacement elements in one piece.

According to the invention, it is preferred that the screw rotors have no rotor inner cooling. In this respect, it is particularly preferred that the screw rotors do not have any channels through which liquid coolant flows in particular. However, the screw rotors can have bores or channels, for example for weight reduction, for balancing or the like. It is preferred in particular that the screw rotors are solid.

Furthermore, it is preferred that the housing has an average heat flow density in the region of the displacement elements which is less than 80,000 W / m ² , preferably less than 60,000 W / m ² and in particular less than 40,000 W / m ² . The mean heat flux density is the ratio of compaction power to the wall area of the compression area. In the case of the dry-compressing screw compressor according to the invention, a gas aftercooler and / or a condensate separator for separating off the condensate and / or a silencer resulting from the compression may optionally additionally be provided on the compressor outlet. It is also possible to provide an intake air filter or an intake silencer at the compressor inlet.

Particularly preferably, with the compressor according to the invention a delivery level of at least 70 percent, preferably at least 85 percent, can be achieved for at least one operating point of the compressor. The decisive factor is the ratio of theoretically possible and practically achieved volume flow. The achievable by the compressor according to the invention high degree of delivery is a measure of the good tightness of the compressor.

Furthermore, the compressor according to the invention preferably has a high isothermal grade of at least 45 percent, preferably at least 60 percent. The isothermal grade is the ratio of ideal isothermal compaction performance and real compaction performance. The isothermal grade in turn represents a measure of good tightness and good cooling of the compressor.

Furthermore, it is preferable for the dry-compressing compressor according to the invention to be operated with a medium-speed motor. In particular,

1

In particular, the speed is more than 3000 "*> and particularly preferred

1

more than 4,000 "i". On the other hand, the speed is preferably lower than

1

10,000 min. _B

At relatively low speeds of, for example, conventional asynchronous

1

For engines in the range of 3000 »i«, large rotor diameters must be used. This leads to unfavorable ratios of promoted Gas volume and leakage areas. This is approximately proportional to

Rotor diameter. On the other hand, very high speeds result in more than

1

10.000 _™ to very high demands on the balancing of the rotors and the displacement elements. This is very difficult to achieve with catchy threads. Furthermore, with increasing power density due to high speeds, the cooling of the compressor becomes difficult. Another disadvantage of very high speeds with very narrow tooth gaps is the high gas friction in the gas passages. This reduces energy efficiency. At medium speeds according to the invention, a good compromise between tightness, balancing properties, gas friction and heat transfer or temperature level can be achieved.

Preferably, an intensive cooling of the housing to keep the gas and components cold. If appropriate, this can also be done without rotor internal cooling in the embodiment of the compressor according to the invention. Low gas temperatures cause a reduction in compression work and thus have a positive effect on the power consumption of the compressor.

In a preferred development of the invention, the rotors or the displacement elements can be coated with inlet layers, for example based on PTFE or molybdenum sulfide, in order to reduce the gap heights without impairing operational safety.

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

Show it :

1 is a schematic plan view of a preferred embodiment of a screw rotor of the screw compressor according to the invention, 2 is a schematic sectional view of displacement elements with asymmetric profile,

Fig. 3 is a schematic sectional view of displacement elements with symmetrical profile, and

Fig. 4 is a schematic sectional view of a screw compressor.

The screw rotors shown in FIGS. 1 to 3 can be used in a screw compressor according to the invention, as shown in FIG. 4.

In a preferred embodiment of the screw compressor, the rotor has a compression direction, i. H. in Figure 1 from left to right, changing or variable slope. In a first suction-side region 10, which forms a first displacement element, a large pitch of approximately 50-150 mm / revolution is provided. The slope changes in the range 10, d. H. in the precompression range, at 55-65% of the inlet slope, i. approx. 30 - 100 mm / revolution. In a second pressure-side region 12, which corresponds to a second displacement element 12, the pitch is significantly lower. In this range, the slope is in the range of 10 - 30 mm / revolution. In the illustrated embodiment, the at least one displacement element per screw rotor thus formed by a screw rotor with variable preferably continuously changing pitch. This corresponds to a multiplicity of displacement elements arranged one behind the other in the conveying direction.

Both in the inlet region and in the outlet region, in each case a gas collection chamber 14 is provided in the illustrated preferred embodiment. Furthermore, the integrally formed screw rotor has two bearing seats 16 and a shaft end 18. With the shaft end 18, for example, a gear is connected to the drive.

Likewise, it is possible for the individual displacement elements 10, 12 to be produced separately and if necessary also to be fixed separately on the rotor shaft, for example by being pressed on. The bearing seats 16 and the shaft ends

18 may be an integral part of the shaft 20 here. The continuous shaft 20 may also be made of another material different from the displacement elements 10, 12.

Furthermore, conical rotors can also be provided. These in turn have according to the invention a plurality of displacement elements. Again, it is particularly preferred that the plurality of displacement elements are realized by a variable pitch. Conical rotors are also designed to be catchy.

Fig. 2 is a schematic sectional view of an asymmetric profile (e.g., a Quimby profile). The illustrated asymmetrical profile is a so-called "Quimby profile". The sectional view shows two screw rotors which mesh with each other and whose longitudinal direction is perpendicular to the plane of the drawing. The opposite rotation of the rotors is indicated by the two arrows 15. Relative to a plane 17 running perpendicular to the longitudinal axis of the displacement elements are the profiles of the flanks

19 and 21 designed differently per rotor. The mutually opposite flanks 19, 21 must therefore be produced independently of each other. However, therefore, although somewhat more complex and difficult production has the advantage that no continuous blow hole is present, but only between two adjacent chambers is a short circuit.

Such an asymmetric profile is preferably provided in the suction-side displacement element 10. The schematic sectional view in Fig. 3 again shows a cross section of two displacement elements or two screw rotors, which in turn rotate in opposite directions (arrows 15). Relative to the axis of symmetry 17, the flanks 23 are designed to be symmetrical per displacement element. The preferred exemplary embodiment shown in FIG. 4, a symmetrically designed contour, is a cycloid profile.

A symmetrical profile, as shown in FIG. 3, is preferably provided in the pressure-side displacement elements 12.

Furthermore, it is possible that more than two displacement elements are provided. These may possibly also have different head diameters and corresponding foot diameters. In this case, it is preferred that a displacement element with a larger head diameter at the inlet, i. is arranged on the suction side in order to realize a greater pumping speed in this area and / or to increase the built-in volume ratio. Furthermore, combinations of the embodiments described above are possible. For example, one or more displacement elements may be made integral with the shaft or an additional displacement element independent of the shaft and then mounted on the shaft.

4, two screw rotors, as shown in FIG. The compressor housing 26 has an inlet 28 through which gas is sucked in the direction of an arrow 30. Furthermore, the compressor housing 26 has a pressure-side outlet 32, through which the gas is ejected in the direction of an arrow 38. Preferably, the screw compressor according to the invention compresses air into a compressed air space. Between the surfaces 42 of the two displacement elements 12 and an inner surface 44 of a pump chamber 26 formed by the compressor housing 26, a gap is formed whose height preferably in the range of 0.03 mm - 0.2 mm and in particular in the range of 0.05 mm - 0.1 mm.

The gap between the flanks of the displacement elements preferably has a gap height of 0.1-0.3 mm.

The compressor housing 26 is closed in the illustrated embodiment with two housing covers 47. The left in Fig. 4 housing cover 47 has two bearing receptacles, in each of which a ball bearing 48 is arranged for mounting the two rotor shafts. On the right side in FIG. 4, the pins 50 of the two screw rotor shafts protrude through the covers 47. On the outside, a gear wheel 52 is arranged on each of the two shaft journals 50. In the exemplary embodiment shown, the two toothed wheels 52 mesh with one another in order to synchronize the two screw rotors with each other. Furthermore, two bearings 48 for supporting the screw rotors are also arranged in the right-hand cover 47 in FIG. In the housing walls 47, a seal, not shown, is provided in addition to the bearings 48.

In the lower in Fig. 4 shaft is the drive shaft which is connected to a drive motor, not shown.

Claims

claims

1. A dry-compressing compressor having a pump chamber forming a housing (26) with a Compressoreinlass (28) to which preferably atmospheric pressure and a compressor outlet (32), preferably at a pressure of at least 2 bar (absolute), preferably at least 5 bar (absolute) prevails, two screw rotors arranged in the pump chamber each having at least one displacement element (10, 12) having a helical recess for forming a plurality of turns, wherein at least one displacement element (10, 12) per screw rotor has a catchy asymmetrical profile, the screw rotors have no rotor internal cooling and the housing (26) in the region of the displacement elements (10, 12) has a mean heat flow density of less than 80,000 W / m ² , preferably less than 60,000 W / m ² and in particular less than 40,000 W / m ² having.

2. Dry-compressing compressor according to claim 1, characterized in that the profiles are formed such that no blow hole is formed.

3. Dry-compressing compressor according to claim 1 or 2, character- ized in that the profiles of the at least one displacement element (10, 12) per screw rotor are designed as Quimby profile.

4. Dry-compressing compressor according to one of claims 1 to 3, characterized in that a displacement element arranged near the outlet of the vacuum pump has a symmetrical profile.

5. Dry-compressing compressor according to one of claims 1 to 4, characterized in that at least one displacement element (10, 12) per screw rotor or in the case of several displacement elements (10, 12) per screw rotor together has at least a number (n) of turns, which is greater than the ratio of outlet pressure (Paus) to inlet pressure (p _e m), so that ΐh pauses

Pein 'preferably n> Paus + 4.

Sorry.

6. A dry compressing compressor according to any one of claims 1 to 5, characterized in that the built-in volume ratio between the delivery rate of the inlet stage (Vein) and the outlet stage (Vaus) to the pressure ratio between inlet pressure (pein) and outlet pressure (paus) is adjusted so that applies

wherein n has a value of k-0.3 to k + 0.3 and preferably between k-0.1 to k + 0.1 and k is the isotropic exponent of the gas mixture to be delivered.

A dry compressing compressor according to any one of claims 1 to 6, characterized in that the displacer elements have at least one region in which the volume of the inlet stage (Vein) decreases within a small rotation angle range to a pre-compression volume (VVK), the ratio between Inlet volume (Vein) and the volume of precompression (VVK) in relation to the internal volume ratio (v,) of the compressor is where j = 2 to 5, in particular j = 2.5 to 3.5 and particularly preferably j = 3.

8. Dry-compressing compressor according to claim 7, characterized in that the compression from the inlet volume (Vein) to the pre-compression volume (VVK) takes place within one and a half to three rotor revolutions (turns).

9. Dry-compressing compressor according to one of claims 1 to 8, characterized in that at least one displacement element (10, 12) per screw rotor or at several displacement elements (10, 12) per screw rotor together this a ratio of length (L) to diameter ( D) for which applies

L Paus 2

D 2 pe

and particularly

10. Dry-compressing compressor according to one of claims 1 to 9, characterized in that the pitch of the turns of the displacement elements (10, 12) varies, preferably varies and particularly preferably from the compressor inlet (28) to the compressor outlet (32) decreases ,

11. Dry-compressing compressor according to one of claims 1 to 10, characterized in that the head and foot diameter of the rotor preferably changes continuously, wherein the rotor is in particular conical.

12. Dry-compressing compressor according to one of claims 1 to 11,

characterized in that the pressure ratio p _en between outlet and inlet pressure is at least 5.

13. Dry-compressing compressor according to one of claims 1 to 12, characterized in that two screw rotors are provided with parallel axes.

14. Dry-compressing compressor according to one of claims 1 to 13, characterized in that the compressor inlet (28), in particular within the housing (26), a gas collecting space (14) is provided.

15. Dry-compressing compressor according to one of claims 1 to 14, characterized in that at the compressor outlet (32) a particular inside the housing (26) arranged gas collecting space (14) is provided.

16. Dry-compressing compressor according to one of claims 1 to 15, characterized in that in the housing (26) roller bearings (48) and vor- Preferably seals are arranged on both sides of the two screw rotors.

17. Dry-compressing compressor according to one of claims 1 to 16, characterized in that for synchronizing the two screw benrotoren a synchronization gear (52) is provided.

18. Dry-compressing compressor according to one of claims 1 to 17, characterized in that the rotational speed of the screw rotors larger

1 1

ßer 3000 ^ and in particular greater 4.000 ^, wherein the speed

1

preferably less than 10,000 "*>.

19. Dry-compressing compressor according to one of claims 1 to 18, characterized in that the one displacement element is designed as a pressure-side displacement element (12) and per screw rotor at least one further displacement element (10) is provided.

20. A dry-compressing compressor according to any one of claims 1 to 19, characterized in that between a surface (42) of the displacement member (12) and an inner surface (44) of the suction chamber (46) has a gap with a height of 0.03 mm to 0.2 mm, in particular 0.05 mm to 0.1 mm is formed.

21. Dry-compressing compressor according to one of claims 1 to 20, characterized in that the pressure-side displacement elements (12) over their entire length have a constant slope.

22. Dry-compressing compressor according to one of claims 1 to 21, characterized in that each screw rotor has a at least one displacement element (10, 12) supporting the rotor shaft.

23. Dry-compressing compressor according to one of claims 1 to 22, characterized in that the displacement elements (10, 12) of a screw rotor are integrally formed.

24. Dry-compressing compressor according to one of claims 1 to 23, characterized in that the screw rotors and in particular the at least one displacement element (10, 12) per screw rotor has a lower coefficient of expansion than the housing (26) up, wherein the coefficient of expansion of the housing ( 26) is in particular at least greater than that of the screw rotors or of the at least one displacement element (10, 12).

25. Dry-compressing compressor according to one of claims 1 to 24, characterized in that the screw rotors do not have flowed through in particular liquid coolant channels.

26. Dry-compressing compressor according to one of claims 1 to 25, characterized in that the screw rotors are solid.

27. Dry-compressing compressor according to one of claims 1 to 26, characterized in that a temperature difference in the region of the pressure-side displacement elements (12) between these and the housing (26) in normal operation is less than 50K, in particular less than 20K.

28. Dry-compressing compressor according to one of claims 1 to 27, characterized in that the distance between the region in which prevail 5% to 20% of the outlet pressure, until the last turn of the pressure-side displacement element (12) at least 20% to 30 % of the rotor length is.

29. Dry-compressing compressor according to one of claims 1 to 28, characterized in that a gap between the flanks at least one of the displacement elements, preferably a gap height of 0.1 to 0.3 mm.