COMPRESSOR

Abstract

A dry-compressing compressor comprises two screw rotors in a housing defining a suction chamber. At a compressor inlet of the compressor preferably atmospheric pressure prevails and at a compressor outlet of the compressor preferably a pressure of more than 2 bars (absolute) prevails. For each screw rotor at least one displacement element including a helical recess defining a plurality of windings is provided. The at least one displacement element per screw rotor has a single-pass asymmetrical profile.

Description

BACKGROUND

1. Field of the Disclosure

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

2. Discussion of the Background Art

For compressing gases, in particular for providing compressed air, primarily oil-injected screw compressors are nowadays used. They can usually perform a compression from 1 bar (absolute) to 8.5 to 14 bars (absolute) in one compressor stage. Here, the delivered intake volume flows range from 30 to 5000 m³/h. Such screw compressors comprise two counter-rotating screw rotors. The screw rotors each comprise at least one helical deepened portion such that a displacement element is formed. The injection of oil into the suction chamber, where the two screw rotors are arranged, serves for sealing the gaps between the rotors and the housing and/or the inner wall of the suction chamber. By providing oil, a sufficient tightness can be attained for realizing high compression pressures of in particular up to 14 bars in one compressor stage. In addition, the oil serves for lubricating the rolling contacts between the two screw rotors. Therefore, a synchronization gear for the two screw rotors is not required. Further, the oil serves for discharging compression heat. Only in this manner, a low temperature can be attained at a high efficiency. Finally, the oil serves for damping mechanical noise. An essential disadvantage of the use of oil is that the oil enters the gas to be delivered. The oil must be removed from the compressed air with the aid of multi-stage separators. As a result, such compressors are complex and require a large installation space. The use of oil-injected screw compressors in particular in areas where a high purity of the compressed air is required, such as in the field of pharmaceutical or food industry, is not possible or possible only when using extremely complex multi-stage oil separators.

For generating oil-free compressed air, it is known to use dry-compressing screw compressors. Here, the two screw rotors are arranged in a contactless manner and synchronized to each other via an oil-lubricated gear. However, dry-compressing screw compressors have the drawback that one compressor stage only allows for a compression to 4 to 5 bars (absolute). The reason for this is in particular that large leakages occur through the gaps between the rotors and the housing. For reaching pressures of 9 bars (absolute), for example, two-stage screw compressors must therefore be used. Besides the two compressor stages, an intermediate cooling of the compressed air is necessary, which results in complex equipment comprising many components and requiring a large installation space.

In addition, dry-compressing compressors configured as so-called rotary tooth compressors are known. These, too, have the drawback that they must be of a multi-stage configuration for achieving high pressures of approximately 9 bars (absolute).

In addition, dry-compressing spindle compressors are known. These comprise a plurality of closed working chambers arranged one behind the other along a plurality of windings or loops of a displacer. Theoretically, high compression pressures are said to be achieved even with a one-stage design such that the spindle compressors can substitute multi-stage screw compressors or rotary tooth compressors. However, spindle compressors are so far not commercially available such that there is no evidence that high compression pressures can be reached with a one-stage design.

Spindle compressors are described in DE 10 2010 064 388, WO 2011/101064, DE 10 2012 202 712 and DE 10 2011 004 960, for example.

It is an object of the disclosure to provide a dry-compressing compressor with the aid of which high pressures of in particular more than 5 bars (absolute) can be reached even with a one-stage design.

SUMMARY

The dry-compressing compressor according to the disclosure comprises a suction chamber defined by a housing. In the suction chamber, two screw rotors engaging with each other are arranged. These are counter-rotated with respect to each other for delivering the gas. For this purpose, each screw compressor comprises at least one displacement element having a helical recess for defining the windings. In particular, for each screw rotor only one displacement element can be provided which can be integrally formed with a rotor shaft. Further, the housing comprises a compressor inlet where preferably atmospheric pressure prevails. At a compressor outlet preferably a pressure of more than 2 bars (absolute) prevails, wherein it is particularly preferred that a pressure of more than 5 bars (absolute) prevails at the compressor outlet.

With the aid of the dry-compressing compressor according to the disclosure high pressures can be reached with a one-stage design since, according to the disclosure, the at least one displacement element per screw rotor is of a single-pass configuration and has an asymmetrical profile. According to a particularly preferred embodiment, the asymmetrical profile is configured such that no or merely a small blowhole occurs. Since no continuous blowhole exists, in a profile which is preferably asymmetrical according to the disclosure a short-circuit merely occurs between two adjacent chambers. According to a particularly preferred embodiment, the so-called Quimby profile is provided as the asymmetrical profile. Asymmetrical profiles have two different profile edges. Although the manufacture is complex due to the required two separate operating steps, an extremely tight working chamber can be realized.

Providing single-pass, possibly even symmetrical rotor profiles offers the advantage that a larger tightness can be achieved. In the case of profiles having two more passes of the respective meshing displacement elements, connections across several chambers are formed through the gaps such that the leakage affects the delivered gas flow and the energy conversion quality.

According to another preferred embodiment of the dry-compressing compressor according to the disclosure, the number of windings of the at least one displacement element or, in the case of a plurality of displacement elements the sum of the windings of the displacement elements of a screw rotor is larger than the ratio of the pressure prevailing at the compressor outlet to the pressure prevailing at the compressor inlet. The number of windings thus results from

$n>\frac{\mathrm{Pout}}{\mathrm{Pin}}$

wherein p_out is the outlet pressure and p_in is the inlet pressure of the compressor. It is particularly preferred that the number of windings or loops is calculated as follows

$n>\frac{\mathrm{Pout}}{\mathrm{Pin}}+4.$

Due to such a large number of windings or loops per screw rotor, a continuous but relatively slow compression of the gas is achieved. Thereby, it is possible to easily discharge heat produced during the compression.

In addition, it is preferred that the installed volume ratio of the dry-compressing screw compressor between the theoretical delivery volume at the inlet stage (V_in) and the theoretical delivery volume at the outlet stage (V_out) is adapted to the pressure ratios at the inlet (p_in) and the outlet (P_out). Here, p_in and p_out are defined as absolute pressures. Preferred is a volume ratio V_i of

$V_i=\frac{\mathrm{Vin}}{\mathrm{Vout}}={\left(\frac{\mathrm{Pout}}{\mathrm{Pin}}\right)}^{1\text{/}k}$

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

According to another preferred embodiment, the displacement elements comprise at least one area or portion where the chamber volume Vin de- creases to an intermediate volume V_VK.

According to another preferred or alternative embodiment, the decrease of the delivery volume of the stages (working chambers) from the large inlet volume (V_in) to the smaller outlet volume (V_out) is divided into two areas. Here, it is particularly preferred that in the first area the working chamber closed towards the suction side is reduced to a specific volume (volume of the precompression V_VK) within a small rotation angle range. Here it is preferred that

V_VK=x V_in

wherein x=0.1 to 0.5, in particular x=0.2 to 0.4, and particularly preferred x=0.3. Due to the compression operation, the precompression raises the temperature of the gas to a moderate value of 150° C.-200° C. In the second area of the compression, depending on the rotation angle, the working chamber volume decreases to a considerably smaller extent than in the first area. The rotation angle and thus the number of stages in the second area is considerably larger than in the first area. Due to the moderate temperature rise in the first area, the large housing surface in the second area and the relatively long dwell time of the gas in the second area due to the larger rotation angle, in the second area another temperature rise of the gas due compression can be avoided to a large extent by heat transport into the housing.

The compression of the gas is selected such that the produced compression heat can be easily discharged via the side walls of the housing such that the temperature of the gas does not rise or rises only to a small extent. Here, the maximum temperature change is preferably less than 50° C., and particularly preferably less than 30° C.

A particular advantage of the selected division of the volume decrease is that a largely homogeneous temperature distribution in the component is achieved. Thereby, thermal peak loads and the associated large component expansions can be avoided.

The ratio between the inlet volume (V_in) and the volume of the precompression (transition from the first to the second area V_VK) can be related to the internal volume ratio v, of the compressor

$v_{\mathrm{VK}}=\frac{\mathrm{Vin}}{\mathrm{Vout}}={\left(v_i\right)}^{1\text{/}j}$

wherein j=2 to 5, in particular j=2.5 to 3.5, and particularly preferred j=3.

According to a particularly preferred embodiment, the precompression is performed in the described first area at 1.5 to 3 rotor revolutions (windings).

According to a preferred embodiment, the inventive large number of windings in the second area can be realized by a single displacement element for each rotor. However, it is also possible to provide a corresponding number of windings in this discharge-side area by two displacement elements, for example. By providing an inventive large number of windings in this area, where, according to the disclosure, preferably the medium to be delivered is only compressed to a small extent per winding, it is possible to do without internal cooling of the rotors. The reason for this is in particular that due to the relatively small extent of compression in this area the temperature increase of the displacement element caused by compression is small. In addition, in this area, due to the high density of the delivered medium, a good heat dissipation from the displacement element into the compressor housing via the medium is realized.

Preferably, the screw rotors and the at least one provided displacement element are configured such that between an area where 5%-20% of the outlet pressure prevails, and the discharge-side rotor end at least 6, in particular at least 8, and particularly preferably at least 10 windings are provided. Here, the discharge-side rotor end is the area of the compressor outlet. Here, according to a preferred embodiment, the inventive large number of windings in this area can be provided at a single discharge-side displacement element provided per rotor. However, it is also possible to provide a corresponding number of windings in this discharge-side area at two displacement elements, for example. By providing an inventive large number of windings in an area where, according to the disclosure, the medium to be delivered is only compressed to a relatively small extent, it is possible to do without internal cooling of the rotors. The reason for this is in particular that due to the relatively small extent of compression in this area the temperature increase of the displacement element caused by compression is smaller. In addition, in this area, due to the high density of the delivered medium, good heat dissipation from the displacement element into the compressor housing via the medium is realized.

Moreover, due to the preferably large number of winding, a large surface area for heat exchange to the housing is available.

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

In addition, for configuring screw rotors without internal cooling according to the disclosure, it is preferred that the discharge-side displacement element has a mean working pressure of more than 2 bars (absolute) at at least 6, in particular at least 8, and particularly preferably at least 10 windings. In particular, it is intended to realize a flat pressure gradient inside the compressor. Therefore, the pressure should slowly rise across many windings, in particular 6 to 10 windings.

According to the disclosure, it is thus preferably possible to provide a cold gap having a height of 0.03 mm-0.2 mm, and in particular 0.05 mm-0.1 mm between the surface of the at least one displacement element and the inside of the section chamber, in particular in the discharge-side area, even in the case of rotors without internal cooling of the rotors or a housing of aluminum or an aluminum alloy. Such a relatively large gap height can be provided due to the inventive configuration of the particularly 6, preferably 8, and particularly preferably 10 last windings, as described above.

According to another preferred embodiment of the disclosure, a relative long screw rotor relative to the diameter 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, said plurality of displacement elements jointly have a ratio of length L to diameter D where the following applies:

$\frac{L}{D}>\frac{\mathrm{Pout}}{2\mathrm{Pin}}-2$ $\mathrm{and}\mathrm{in}\mathrm{particular}$ $\frac{L}{D}>\frac{\mathrm{Pout}}{2\mathrm{Pin}}-1$

By providing a long rotor having in particular many chambers, the area usable for heat dissipation is increased. Due to the resulting good heat exchange, the gas temperatures of the compressed gas are relatively low. Providing many chambers further offers the advantage that the pressure differences between adjacent chambers are small and thus a large tightness can be achieved. Due to such a reduction of the delivery volume per stage from the inlet to the outlet side, the compression process becomes particularly effective in terms of thermodynamics and the gas temperatures remain relatively low. Here, it is particularly preferred that the internal volume ratio is adapted to the ratio of outlet to inlet pressure such that neither overcompression or compression by re-aeration occurs.

The internal volume ratio can be attained by varying the pitch of the windings. Preferably, the pitch of the windings is in particular changed such that it is decreased and/or becomes steeper from the compressor inlet to the compressor outlet. The pitch can be changed continuously and/or stepwise.

In addition to or instead of the variation of the pitch, the head or foot diameter of the profile can be changed continuously or stepwise. Again, a continuous change of the head or foot diameter is particularly preferred such that the rotor has a conical configuration, in particular in combination with a continuous change of the pitch.

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

According to another particularly preferred embodiment, the dry-compressing compressor comprises at the compressor inlet and/or at the compressor outlet a respective gas collection chamber preferably inside the compressor housing.

Moreover, it is preferred that the dry-compressing compressor is a compressor having two shafts. The latter are preferably supported on both sides such 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, the two rotor shafts are synchronized by a synchronization gear preferably arranged outside the suction chamber. The bearings can be lubricated by grease and/or oil. Likewise, the gear can be lubricated by grease and/or oil. This is possible since both the bearings and the synchronization gear are preferably arranged outside the suction chamber and it is thus avoided that the gas to be delivered is contaminated by oil.

Preferably, the housing is made from aluminum or an aluminum alloy. Here, an aluminum alloy AlSi7Mg or AlMg07,5Si for the housing is particularly preferred. In particular, the heat expansion coefficient (expansion coefficient) of the material of the screw rotors is smaller than the expansion coefficient of the material of the housing. It is particularly preferred that the expansion coefficient of the screw rotors is smaller than 12*10⁻⁶1/K. This can be achieved with rotors made from iron or steel materials.

The two screw rotors arranged in the suction chamber comprise at least one displacement element having a helical recess. The helical recesses define several windings. According to the disclosure, the at least one displacement element is made from a steel or iron alloy. It is thus particularly preferred that the screw rotors including the displacement elements are made from a steel or iron alloy. The housing is also made from a steel or iron alloy or from aluminum or an aluminum alloy.

Preferably, each displacement element comprises at least one helical recess having the same contour along its overall length. Preferably, the contours are different for each displacement element. Thus the individual displacement element preferably has a constant pitch and an unvarying contour. Thereby, the manufacture is considerably simplified such that the manufacturing costs can be largely reduced.

For further improving the suction capacity, the contour of the suction-side displacement element, that is in particular the first displacement element as seen in the pumping direction, is preferably of an asymmetrical configuration. Due to the asymmetrical configuration of the contour and/or the profile, the edges can be configured such that the leakage areas, the so-called blowholes, can in particular completely disappear or have at least a smaller cross-section. A particularly suitable asymmetrical profile is the so-called “Quimby” profile. Although such a profile is relatively difficult to produce, it offers the advantage that no continuous blowhole exists. A short-circuit occurs only between two adjacent chambers. Since this is an asymmetrical profile having different profile edges, at least to working steps are required for the production since the two edges have to be produced in two different working steps due to their asymmetry.

The discharge-side displacement element, in particular the last displacement element as seen in the pumping direction, preferably has a symmetrical contour. The symmetrical contour in particular offers the advantage that it is easier to produce. In particular, the two edges having a symmetrical contour can be produced in one working step using a rotating end milling cutter or a rotating side milling cutter. Although such symmetrical profiles have blowholes, these are continuous, i.e. do not only exist between two adjacent chambers. The size of the blowhole decreases with decreasing pitch. Thus, such symmetrical profiles can in particular be provided for the discharge-side displacement element since, according to a preferred embodiment, it has a smaller pitch than the suction-side displacement element and preferably also than the displacement element arranged between the suction-side and the discharge-side displacement element. Even though the tightness of such symmetrical profiles is somewhat smaller, they offer the advantage that they are considerably easier to produce. In particular, it is possible to produce the symmetrical profile in a single working step and preferably using a simple end milling cutter or side milling cutter. Thereby, the costs are considerably reduced. A particularly suitable symmetrical profile is the so-called “cycloidal profile”.

Providing at least two such displacement elements results in the corresponding screw compressor being capable of generating high outlet pressures at a low power consumption. Further, the thermal load is small. Arranging at least two displacement elements having the configuration according to the disclosure with a constant pitch and an unvarying contour in a compressor leads to essentially the same results as with a compressor having a displacement element with a varying pitch. At high installed volume ratios three or four displacement elements per rotor can be provided.

According to a particularly preferred embodiment, for increasing the attainable outlet pressure and/or for reducing the power consumption and/or the thermal load, a discharge-side displacement element, that is in particular the last displacement element as seen in the pumping direction, comprises a large number of windings. A large number of windings allows for accepting a larger gap between the screw rotor and the housing at constant performance. Here, the gap can have a cold-gap width of 0.05-0.3 mm. A large number of outlet windings or of windings of the discharge-side displacement element is inexpensive to produce since, according to the disclosure, this displacement element can have a constant pitch and preferably also a symmetrical contour. On the outlet side an asymmetrical profile can be used. This allows for an easy and inexpensive production such that it is acceptable to provide a larger number or windings. Preferably, this discharge-side or last displacement element has more than 6, in particular more than 8, and particularly preferably more than 10 windings. According to a particularly preferred embodiment, the use of symmetrical profiles offers the advantage that both edges of the profile can be simultaneously cut with a milling cutter. Here, the milling cutter is supported by the respective opposite edge such that deformation or distortion of the milling cutter during the milling operation and resultant inaccuracies are avoided.

For further reducing the manufacturing costs it is particularly preferred to integrally form the displacement elements and the rotor shaft.

According to another preferred embodiment, the change of pitch between adjacent displacement elements is inconsistent or erratic. Possibly, the two displacement elements are arranged at a distance to each other in the longitudinal direction such that between two displacement elements a circular cylindrical chamber is defined which serves as a tool outlet. This is in particular advantageous for manufacturing integrally formed rotors since the tool producing the helical line can be easily removed in this area. If the displacement elements are manufactured separately from each other and are then mounted to a shaft, a tool outlet, in particular such an annularly cylindrical area need not be provided.

According to a preferred aspect of the disclosure, no tool outlet is provided between two adjacent displacement elements at the location where the pitch changes. In the area of the change of pitch both edges preferably have a discontinuity or recess for removing the tool. Such a discontinuity has no notable influence on the compression capacity of the compressor since it is a localized discontinuity or recess.

The compressor screw rotor according to the disclosure in particular comprises a plurality of displacement elements. These may have the same or a different diameter. Here, it is preferred that the discharge-side displacement element has a smaller diameter than the suction-side displacement element.

In the case of displacement elements manufactured separately from the rotor shaft, the former are mounted to the shaft by press-fitting. Here, it is preferred to provide elements, such as locating pins, for defining the angular position of the displacement elements relative to each other.

It is particularly preferred that the screw rotor is formed integrally in particular from a steel or an iron alloy. The screw rotor can also comprise a rotor shaft which supports the at least one displacement element. In particular when providing a plurality of displacement elements, this offers the advantage that these can be manufactured separately from each other and then be connected to the rotor shaft in particular by press-fitting or shrink-fitting. Here, it is possible to provide fitting keys or the like for defining the angular position of the individual displacement elements.

If a plurality of displacement elements per screw rotor are provided, it is possible to integrally form the displacement elements.

According to the disclosure, it is preferred that the screw rotors have no internal cooling. Hence it is particularly preferred that the screw rotors do not have any ducts through which in particular liquid coolant flows. However, the screw rotors can comprise boreholes or ducts for the purpose of weight reduction, for balancing or the like, for example. It is particularly preferred that the screw rotors are of a solid configuration.

In addition, it is preferred that the housing has a mean heat flow density in the area of the displacement elements of less than 80,000 W/m², preferably less than 60,000 W/m², and in particular less than 40,000 W/m².

The mean heat flow density is the ratio of the compression capacity to the wall surface of the compression area.

In the dry-compressing screw compressor according to the disclosure a gas aftercooler and/or a condensate separator for separating the condensate produced by compression and/or a silencer may additionally be provided at the compressor outlet. Further, it is possible to provide an inlet air filter or an inlet silencer at the compressor inlet.

Particularly preferably, with the aid of the compressor according to the disclosure a volumetric efficiency of at least 70 percent, preferably at least 85 percent for at least one operating point of the compressor can be achieved. A decisive factor is the ratio of theoretically possible and practically achieved volume flow. The high volumetric efficiency adapted to be achieved by the compressor according to the disclosure is an indication of the good tightness of the compressor.

Further, the compressor according to the disclosure preferably has a high isothermal efficiency factor of at least 45 percent, preferably at least 60 percent. The isothermal efficiency factor is the ratio of ideal isothermal compression capacity and real compression capacity. The isothermal efficiency factor is also an indication of good tightness and good cooling of the compressor.

In addition, it is preferred that the dry-compressing compressor is operated by a motor at a mean speed. In particular, the speed is higher than 3,000 1/min, and particularly preferred more than 4,000 1/min. On the other hand, the speed is preferably lower than 10,000 1/min.

At relatively low speeds in the range of 3,000 1/min of conventional asynchronous motors, for example, large rotor diameters must be used. This results in unfavorable ratios of delivered gas volume and leakage areas. This is approximately proportional to the rotor diameter. On the other hand, very high speeds of more than 10,0000 1/min entail very high demands on the balancing of the rotors or the displacement elements. This is difficult to achieve in the case of single-pass screw threads. In addition, with increasing power density due to high speeds, it becomes more and more difficult to cool the compressor. Another drawback of very high speeds with very small tooth gaps is the high gas friction in the gas paths. Thereby, the energy efficiency decreases. At mean speeds according to the disclosure a good compromise between tightness, balancing, gas friction and heat transfer or temperature level can be achieved.

Preferably, the housing is intensively cooled for keeping the gas and the components cool. In the embodiment of the compressor according to the disclosure, this can possibly also be achieved without internal cooling of the rotors. Low gas temperatures cause a reduction of the compression operation and thus have a positive effect on the power consumption of the compressor.

According to a preferred aspect of the disclosure, the rotors and/or the displacement elements can be coated with layers on the basis of PTFE or molybdenum sulfide, for example, in order to decrease the gap heights without affecting the operational safety.

BRIEF DESCRIPTION OF THE DRAWINGS

Hereunder the disclosure will be explained in detail on the basis of a preferred embodiment with reference to the accompanying drawings in which:

FIG. 1 shows a schematic top view of a preferred embodiment of a screw rotor of the screw compressor according to the disclosure,

FIG. 2 shows a schematic sectional view of displacement elements having an asymmetrical profile,

FIG. 3 shows a schematic sectional view of displacement elements having a symmetrical profile, and

FIG. 4 shows a schematic sectional view of a screw compressor.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

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

According to a preferred embodiment of the screw compressor, the rotor has a pitch changing and/or variable in the direction of compression, i.e. from left to right in FIG. 1. In a first suction-side area 10 defining a first displacement element a large pitch of approximately 50-150 mm/revolution is provided. Here, the pitch changes in the area 10, i.e. in the precompression area, to 55-65% of the inlet pitch, i.e. approximately 30-100 mm/revolution. In a second discharge-side area 12 corresponding to a second displacement element 12 the pitch is considerably smaller. In this area the pitch is in the range of 10-30 mm/revolution. In the illustrated embodiment, the at least one displacement element per screw rotor is thus defined by a screw rotor having a variable, preferably continuously changing pitch. This corresponds to a plurality of displacement elements arranged one behind the other as seen in the direction of delivery.

In the illustrated preferred embodiment, both in the inlet area and the outlet area a gas collection chamber 14 each is provided.

Further, the integral screw rotor comprises two bearing seats 16 and a shaft end 18. The shaft end 18 has connected thereto a gearwheel for driving purposes, for example.

Likewise, it is possible that the individual displacement elements 10, 12 are manufactured separately from each other and are separately affixed to the rotor shaft by pressing, for example. Here, the bearing seats 16 and the shaft ends 18 can be integral components of the shaft 20. Here, the continuous shaft 20 can be made from a material differing from that of the displacement elements 10, 12.

In addition, conical rotors can be provided. According to the disclosure, they comprise a plurality of displacement elements. Here, too, it is particularly preferred that the plurality of displacement elements are realized by a variable pitch. Conical rotors, too, are of a single-pass configuration.

FIG. 2 shows a schematic sectional view of an asymmetrical profile (e.g. a Qumiby 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 drawing plane. The counter-rotation of the rotors is indicated by two arrows 15. Relating to a plane 17 extending perpendicularly to the longitudinal axis of the displacement elements, the profiles of the edges 19 and 21 are of different configuration for each rotor. The opposing edges 19, 21 must thus be manufactured separately from each other. However, this somewhat more complex and difficult manufacture offers the advantage that no continuous blowhole exists but a short-circuit occurs merely between two adjacent chambers.

Preferably, such an asymmetrical profile is provided for the suction-side displacement element 10.

The schematic sectional view in FIG. 3 shows a cross-section of two displacement elements and/or two screw rotors which are again counter-rotating (arrows 15). Relating to the symmetry axis 17, the edges 23 of each displacement element are of a symmetrical configuration. The preferred exemplary embodiment of a symmetrical contour illustrated in FIG. 4 is a cycloid profile.

A symmetrical profile, as illustrated in FIG. 3, is preferably provided for the discharge-side displacement elements 12.

Further, it is possible that more than two displacement elements are provided. They can possibly have different head diameters and corresponding foot diameters. Here, it is preferred that a displacement element having a larger head diameter is arranged at the inlet, i.e. on the suction side, for realizing a larger suction capacity in this area and/or increasing the installed volume ratio. Further, combinations of the embodiments described above are possible. For example, one or a plurality of displacement elements can be integrally formed with the shaft, or an additional displacement element can be separately manufactured and then mounted to the shaft.

In the schematic view of a preferred embodiment of a screw compressor according to the disclosure illustrated in FIG. 4, two screw rotors, as illustrated in FIG. 1, are arranged in a housing 26. The compressor housing 26 comprises an inlet 28 through which gas is taken in in the direction indicated by an arrow 30. Further, the compressor housing 26 comprises a discharge-side outlet 32 through which the gas is discharged in the direction indicated by an arrow 38. Preferably, the screw compressor according to the disclosure compresses air in a compressed air chamber.

Between upper surfaces 42 of the two displacement elements 12 and an inner surface 44 of a suction chamber 46 defined by the compressor housing 26, a gap is formed whose height preferably lies in the range of 0.03 mm-0.2 mm and in particular in the range from 0.05 mm-0.1 mm.

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

In the illustrated exemplary embodiment, the compressor housing 26 is closed by two housing covers 47. The left housing cover 47 in FIG. 4 comprises two bearing supports where a ball bearing 48 each for supporting the two rotor shafts is arranged. On the right side in FIG. 4, journals 50 of the two screw rotor shafts protrude through the covers 47. On the outside a respective gearwheel 52 is arranged on the two shaft journals 50. In the illustrated exemplary embodiment, the two gearwheels 52 mesh with each other for synchronizing the two screw rotors with each other. Further, in the right cover 47 in FIG. 4, two bearings 48 for supporting the screw rotors are arranged. In the housing walls 47 a seal not illustrated is provided in addition to the bearings 48.

The lower shaft in FIG. 4 is a drive shaft connected to a drive motor not illustrated.

Claims

1. A dry-compressing compressor comprising

a housing defining a suction chamber and having a compressor inlet where preferably atmospheric pressure prevails and a compressor outlet where preferably a pressure of at least 2 bars (absolute), preferably at least 5 bars (absolute) prevails,

two screw rotors arranged in the suction chamber and each having at least one displacement element including a helical recess for defining a plurality of windings,

wherein at least one displacement element per screw rotor has a single-pass asymmetrical profile,

the screw rotors have no internal cooling of the rotors, and

the housing has a mean heat flow density of less than 80000 W/m2 in the area of the displacement elements.

2. The dry-compressing compressor according to claim 1, wherein the profiles are configured such that not blowhole is formed.

3. The dry-compressing compressor according to claim 1, wherein the profiles of the at least one displacement element of each screw rotor are configured a Quimby profile.

4. The dry-compressing compressor according to claim 1, wherein a displacement element arranged near the outlet of the vacuum pump has symmetrical profile.

5. The dry-compressing compressor according to claim 1, wherein at least one displacement element per screw rotor and/or in the case of a plurality of displacement elements per screw rotor said displacement elements jointly comprise a number (n) of windings which is larger than the ratio of outlet pressure (pout) to inlet pressure (pin) such that n > Pout Pin preferably n > Pout Pin + 4. applies.

6. The dry-compressing compressor according to claim 1, wherein the installed volume ratio between the delivery volume of the inlet stage (Vin) and the outlet stage (Vout) is adapted to the pressure ratio between inlet pressure (pin) and outlet pressure (pout) such that the following applies: V i = Vin Vout = ( Pout Pin ) 1  /  k

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

7. The dry-compressing compressor according to claim 1, wherein the displacement elements include at least one area where the volume of the inlet stage (Vin) decreases to a precompression volume (VVK) in a small rotation angle range, wherein the ratio between inlet volume (Vin) and the volume of the precompression (VVK) is related to the internal volume ratio (vi) of the compressor v VK = Vin Vout = ( v i ) 1  /  j wherein j=2 to 5.

8. The dry-compressing compressor according to claim 7, wherein the compression from the inlet volume (Vin) to the precompression volume (VVK) takes place during one and a half to three rotor revolutions (windings).

9. The dry-compressing compressor according to claim 1, wherein at least one displacement element per screw rotor and/or in the case of a plurality of displacement elements per screw rotor said displacement elements jointly have a ratio of length (L) to diameter (D) for which the following applies L D > Pout 2  Pin - 2 and   in   particular L D > Pout 2  Pin - 1

10. The dry-compressing compressor according to claim 1, wherein the pitch of the windings of the displacement elements varies, preferably changes and particularly preferably decreases from the compressor inlet to the compressor outlet.

11. The dry-compressing compressor according to claim 1, wherein the head and the foot diameter of the rotor preferably continuously changes, wherein the rotor is in particular of a conical configuration.

12. The dry-compressing compressor according to claim 1, wherein the pressure ratio Pout Pin between outlet and inlet pressure is at least 5.

13. The dry-compressing compressor according to claim 1, wherein two screw rotors with parallel axes are provided.

14. The dry-compressing compressor according to claim 1, wherein at the compressor inlet in particular inside the housing a gas collection chamber is provided.

15. The dry-compressing compressor according to claim 1, wherein at the compressor outlet a gas collection chamber is provided in particular inside the housing.

16. The dry-compressing compressor according to claim 1, wherein in the housing roller bearings and preferably seals are arranged on both sides of the two screw rotors.

17. The dry-compressing compressor according to claim 1, wherein for synchronizing the two screw rotors a synchronization gear is provided.

18. The dry-compressing compressor according to claim 1, wherein the speed of the screw rotors is higher than 3 ,  000   1 min, 1 min, 1 min.

19. The dry-compressing compressor according to claim 1, wherein the one displacement element is configured as a discharge-side displacement element and for each screw rotor at least one further displacement element is provided.

20. The dry-compressing compressor according to claim 1, wherein between an upper surface of the displacement element and an inner surface of the suction chamber a gap with a height of 0.03 mm to 0.2 mm is formed.

21. The dry-compressing compressor according to claim 1, wherein the suction-side displacement elements have a constant pitch along their overall length.

22. The dry-compressing compressor according to claim 1, wherein each screw rotor comprises a rotor shaft supporting the at least one displacement element.

23. The dry-compressing compressor according to claim 1, wherein the displacement elements of a screw rotor are of an integral configuration.

24. The dry-compressing compressor according to claim 1, wherein the screw rotors and in particular the at least one displacement element per screw rotor have a smaller expansion coefficient that the housing, wherein the expansion coefficient of the housing is in particular at least larger than that of the screw rotors and/or the at least one displacement element.

25. The dry-compressing compressor according to claim 1, wherein the screw rotors do not comprise any ducts through which in particular a liquid coolant flows.

26. The dry-compressing compressor according to claim 1, wherein the screw rotors are of a solid configuration.

27. The dry-compressing compressor according to claim 1, wherein a temperature difference in the area of the discharge-side displacement elements between the latter and the housing during normal operation is smaller than 50 K.

28. The dry-compressing compressor according to claim 1, wherein the distance between the area where 5 ° A) to 20% of the outlet pressure prevails and the last winding of the discharge-side displacement element is at least 20% to 30% of the rotor length.

29. The dry-compressing compressor according to claim 1, wherein a gap between the edges of at least one of the displacement elements preferably has a gap height of 0.1 to 0.3 mm.