Drive system for a multi-stage screw compressor

Abstract

The present invention relates to a drive system for a multi-stage screw compressor comprising a first compressor stage with a first main rotor mounted non-rotatably on a first shaft and a second compressor stage with a second main rotor mounted non-rotatably on a second shaft. The drive system includes a drive shaft for driving the shafts of both main rotors of the first and second compressor stages, wherein the drive of the shaft of the first main rotor is independent of the drive of the shaft of the second main rotor, so that the rotational speeds of the two main rotors can preferably be adjusted independently of each other.

Description

Application area

[0001] The present invention relates to improvements in the drive of a multi-stage screw compressor and in particular a two-stage screw compressor in which the main rotors of the stages are driven via a common drive shaft. Technical background

[0002] In a multi-stage screw compressor, a 3-wheel gearbox is conventionally used, in which a drive wheel arranged on the drive shaft engages with the two drive pinions arranged on the respective screw shaft of the first and second compressor stages to drive the two compressor stages.

[0003] This entails several disadvantages, which are described in detail below.

[0004] In a two-stage screw compressor, an intermediate pressure develops between the first and second stages, dependent on the final pressure, which essentially corresponds to the final pressure of the first stage. This intermediate pressure arises because the first stage delivers a higher volume flow than the second stage can dissipate. Consequently, the speed of the second stage must be set to a defined ratio relative to the speed of the first stage to achieve optimal intermediate pressure. The respective pressure ratios of the individual compressor stages, in turn, determine the distribution of the total output within the two-stage compressor.

[0005] Intermediate pressure has a significant impact on the overall efficiency of the compressor. Essentially, a compressor generates pressure by enclosing a certain volume of air and continuously reducing the volume of that enclosed air until a defined pressure is reached. The volume ratio—that is, the initial volume of the enclosed gas (when the compressor's pressure chamber is closing) to the final volume at the point of defined pressure (when the pressure chamber opens again)—is determined during the compressor's design and remains constant. Maximum efficiency is achieved when the compressor's pressure chamber opens and the pressure in the adjacent pressure chamber is the same.

[0006] In a two-stage compressor, this pressure chamber is, in the first stage, the space between the two compressor stages with the intermediate pressure described above. In the second stage of the compressor, this pressure chamber is the pressure vessel connected to the compressor or the consumer network, which maintains the constant final pressure.

[0007] If the pressure in the pressure chamber is higher than in the pressure chamber at the time of opening, this is called undercompression. Conversely, if the pressure in the pressure chamber is lower than in the pressure chamber at the time of opening, this is called overcompression. In both cases, significant efficiency losses occur.

[0008] In a two-stage compressor, the intermediate pressure corresponds to the intake pressure of the second compressor stage, where the intake pressure is the pressure present when the pressure chamber of the respective compressor stage is closed. Consequently, if the intermediate pressure deviates from the optimal setting, the optimal intake pressure is not generated during the compression process in the second stage, resulting in over- or under-compression with respect to the final pressure and thus a significant loss of efficiency in the two-stage compressor.

[0009] Therefore, if a two-stage compressor is operated with a volume ratio fixed by its design at different final pressures, the intermediate pressure must be adjusted to the respective final pressure by adjusting the speed of the two compressor stages in order to achieve maximum efficiency.

[0010] However, if a two-stage screw compressor is operated at different power levels, the rotational speed of the first stage is changed to influence the performance of the screw compressor, and the rotational speed of the second stage must be adjusted accordingly to achieve the optimal intermediate pressure that generates the highest efficiency.

[0011] Typically, however, the rotors of the two compressor stages are installed in a common rotor housing, so the position of the driven main rotors is fixed and the gearbox consequently has a constant center distance. With the conventional 3-wheel drive via a common drive wheel, the gear ratio between the first and second stages can only be changed to a very small extent by modifying the tooth geometry, for example by shifting the profile of the meshing gears.

[0012] In other words, with conventional 3-wheel transmissions, if the total output of the two-stage compressor changes due to a change in the gear ratio between the drive transmission and the first compressor stage, the gear ratio to the second compressor stage is no longer defined, but is instead determined by the center distance between the drive shaft and the second compressor stage. As a result, the second compressor stage can no longer transport the intended volume flow from the intermediate space, leading to an unfavorable intermediate pressure during this change in output, which in turn results in a significant loss of efficiency.

[0013] In summary, if the total power, final pressure or volume flow of a two-stage compressor is changed, a significant and undesirable loss of efficiency results when using the conventional 3-wheel gearbox, especially in the case of a fixed and unchangeable distance between the main rotor axes and the drive shaft. Description of the invention

[0014] The present invention aims to avoid the aforementioned problem and to enable changes in the overall power, final pressure, or volume flow rate of a multi-stage, and in particular two-stage, compressor, especially in the case of a fixed and unchangeable distance between the main rotor axes and the drive shaft, without significant efficiency losses. Furthermore, the present invention offers numerous other advantages, which will be discussed in detail below. Solution to the task

[0015] This problem is solved by a drive system for a multi-stage compressor according to claim 1. Further advantageous embodiments of the invention are set forth in the dependent claims.

[0016] A drive system according to the invention for a multi-stage screw compressor comprises a first compressor stage with a first main rotor mounted non-rotatably on a first shaft and a second compressor stage with a second main rotor mounted non-rotatably on a second shaft. The drive system includes a drive shaft for driving the shafts of both main rotors of the first and second compressor stages. According to the invention, the drive of the shaft of the first main rotor is independent of the drive of the shaft of the second main rotor.

[0017] Advantageously, in a preferred embodiment of the invention, the rotational speeds of the main rotors can be adjusted independently of each other due to the drive of the shaft of the first main rotor being independent of the drive of the shaft of the second main rotor.

[0018] In particular, the ratio of the rotational speeds to each other is preferably freely selectable, whereas the speed ratio of the two rotors in a conventional 3-wheel gearbox is fixed by the design.

[0019] In the context of the present invention, an independent drive of the two shafts of the main rotors is to be understood as meaning that the two shafts are driven by a common drive shaft, but, in contrast to the 3-wheel transmission of the prior art, not via the same drive wheel on the drive shaft.

[0020] In a preferred embodiment, the drive shaft can have two drive wheels arranged non-rotatably on the drive shaft. The first drive wheel of these two drive wheels engages with a first output pinion arranged non-rotatably on the shaft of the first main rotor, while the second drive wheel of the two drive wheels engages with a second output pinion arranged non-rotatably on the shaft of the second main rotor.

[0021] The 4-wheel gearbox according to the invention makes it very easy to adjust the intermediate pressure by freely selecting the gear ratios between the two compressor stages, and it also allows for coverage of ranges that are not achievable with a simple change in the tooth geometry of a 3-wheel gearbox. This leads to significant advantages in compressor operation when the total power output changes, the final pressure varies, and the volumetric flow rates are altered, as thermodynamically induced power losses due to over- or under-compression can be avoided.

[0022] Since, when using the 4-wheel transmission according to the invention, the intermediate pressure is no longer essentially determined by the predetermined center distance as described above for the 3-wheel transmission, the arrangement of the individual compressor stages can be freely selected. The resulting compact and weight-saving design, as well as the possibility of greater integration of functions into the compressor housing, result in significant cost advantages.

[0023] Since the 4-wheel drive according to the invention allows for the free arrangement of all compressor stages, it also offers the possibility, unlike 3-wheel drives, of freely designing the inlets and outlets, as well as the spaces within the compressor housing. This enables the implementation of a fluid-mechanically optimal arrangement by reducing deflections and adapting the arrangement of the internal flow channels and the inlets and outlets. Consequently, in addition to the increased compactness, this results in a further significant efficiency gain.

[0024] Furthermore, the easily achievable adjustment of the best possible intermediate pressure according to the invention, through the free choice of the transmission ratios between the two compressor stages, also enables the optimization of an oil-flooded compressor operated at constant speed, since in this case the injection temperature of the oil, which also has a strong influence on the intermediate pressure and the efficiency, can be chosen thermodynamically favorably just above the precipitation point of condensation.

[0025] In this sense, a preferred embodiment of the present invention relates to an oil-injected compressor in which the drive system according to the invention is used.

[0026] Ultimately, the 4-wheel gear arrangement according to the invention also has advantageous effects on the design and service life of the bearings used in the system for the rotor shafts, as will be shown below.

[0027] A screw compressor typically comprises one or more pairs of main and secondary rotors that mesh together in a helical configuration. The screws of the two rotors each have different pitches and numbers of teeth, so that during rotation, a chamber forms between each tooth gap of the rotors. This chamber continuously shrinks, thus generating the desired pressure. Consequently, different pressures arise at different axial points within the screw pair. This pressure gradient creates axial and radial forces on the rotors, which are absorbed by the rotor bearings. The forces induced and generated by the rotor geometry depend on the intermediate and final pressure, but not on the compressor's power output. The resulting radial force is absorbed by the radial bearings, and the axial force by the axial bearings.Due to the different geometries of the main and secondary rotors, the resulting forces have different effects in the radial and axial directions. The main rotor typically experiences the greatest axial force.

[0028] Since, as described above, the forces acting on the rotor bearings are caused by the pressure gradient generated in the rotor helix, and this pressure gradient depends on the respective intake and discharge pressures of the individual stages, the intermediate pressure, which represents the discharge pressure of the first stage as well as the intake pressure of the second stage, also influences the bearing forces of the individual compressor stages. If the intermediate pressure is too high, the bearings of the first stage are subjected to greater stress; if it is too low, the bearings of the second stage are subjected to greater stress. As described above, the intermediate pressure in conventional three-wheel drives cannot be kept constant when the total power output changes, when different discharge pressures are used, when volume flows are altered, and also when different injection temperatures occur in the case of an oil-flooded compressor.Therefore, all these cases must be taken into account in the bearing design, which inevitably leads to a higher bearing load and larger, more expensive bearings or reduced bearing lifespan.

[0029] Since it is possible to keep the intermediate pressure constant in all the cases mentioned with the 4-wheel gearbox according to the invention, an exact, cost-effective bearing with a long service life can be used.

[0030] Advantageously, in a preferred embodiment of the drive system according to the invention, the two drive wheels on the drive shaft can be positioned in contact with each other, which can lead to increased stability of the drive wheels against axially acting forces.

[0031] Furthermore, in a preferred embodiment, the transmission between the second drive wheel and the second output pinion can have a different or the same gear ratio as the transmission between the first drive wheel and the first output pinion, in order to ensure the higher speed of the second compressor stage that is usually necessary in operation.

[0032] Advantageously, the toothing between at least one of the drive gears and the corresponding output pinion can be inclined relative to the axial extent of the drive shaft. Preferably, the toothing between both drive gears and their respective output pinions can also be inclined relative to the axial extent of the drive shaft, and the angles formed by the toothing of each drive gear with the output pinion relative to the axial extent of the drive shaft can be different. In particular, the angles formed by the toothing of each drive gear with the output pinion relative to the axial extent of the drive shaft can have opposite signs.

[0033] Helical gearing of the drive wheels and the respective output pinions can be advantageously used to control axial forces acting on the rotors and their shafts during operation, as described in detail below.

[0034] As shown above, the pressure gradient in the screw pairs of the individual compressor stages leads to radially and axially acting forces on the rotors and their shafts.

[0035] The use of helical gearing on the drive gear and output pinion results in additional forces within the system. In particular, the helical gearing generates axial forces, dependent on the helix angle, which act on the drive shaft and the driven rotor. Unlike the forces generated by the gas forces, the forces generated by the gears depend on the compressor's power output and the selected gear ratio. Depending on the helix angle of the gearbox, the axial forces generated by the helical gearing can either load or relieve the axial rotor bearing of the driven main rotor. This may allow for the use of smaller axial bearings on the rotors, or conversely, the application of a defined force to the axial rotor bearing.

[0036] In the case of a 4-wheel transmission according to the invention, with an arrangement of two drive wheels, it is possible to control the axial force by changing the helix angle of the gear teeth on each of the two drive wheels. If helix angles are used that have opposite signs when viewed from the axial extent of the drive shaft, the forces generated by the helical gearing can at least partially cancel each other out, allowing the use of smaller bearings on the drive shaft. Furthermore, by using different helix angles, the direction and magnitude of the force can be matched to the existing bearings in all power ranges and gear ratios. This matching of the bearing forces has the advantage that smaller bearings can be used, resulting in cost savings and a longer bearing service life.

[0037] In addition to controlling the axial forces acting on the drive shaft, the gear forces acting on the rotor bearings can also be controlled or reduced by appropriately selected and matched helical gearing of the individual drive gears. If a high axial force is generated at the driven rotor of a screw compressor due to the geometry of the rotors and the resulting pressure differential, the axial force generated by the pressure differential can be counteracted or increased individually at each compressor stage using the 4-gear transmission according to the invention and corresponding helical gearing between the respective drive gear and associated output pinion.

[0038] This also applies to more than the described two-stage compressors (e.g., three-stage and multi-stage compressors) and can be adapted to all performance ranges and gear ratios. This adjustment of bearing forces has the advantage that smaller bearings can be used, resulting in cost savings, and a longer bearing service life can be achieved.

[0039] According to a further aspect of the present invention, the drive system described in detail above is used in a two-stage compressor, wherein in a preferred embodiment it is an oil-injected compressor in which the advantages of the invention described in detail above can be used to a particularly beneficial effect. Examples of implementation

[0040] The present invention is explained in more detail below with reference to various exemplary embodiments and the drawings. The drawings show: Fig. 1 : a compressor with a drive system according to the invention in perspective view, Fig. 2 : a view of the compressor from Figure 1 in frontal view, Fig. 3 : a perspective view of the two compressor stages of the compressor from Figure 1 , Fig. 4 : a frontal view of the in Fig. 3 shown compressor stages, Fig. 5 : a side view of the in Fig. 3 shown compressor stages, Fig. 6a / b: Side view of the gearbox of the compressor shown in the previous figures, Fig. 7a-f Schematic frontal views of various drive systems according to the invention

[0041] Figure 1Figure 1 shows a perspective view of a compressor 1 with a compressor housing 2 in which an embodiment of the drive system according to the invention is housed. The compressor shown in the figures is an oil-injected compressor, but the drive system according to the invention, as described above, can also be used in compressors that do not have oil injection.

[0042] The compressor has an inlet 3 and an outlet 4 through which the medium to be compressed is supplied and discharged, as indicated by the arrows. The drive shaft 5, which protrudes from the housing 2, is driven rotationally in the direction indicated by arrow R by means of a motor (not shown).

[0043] The in Figure 1 The compressor 1 shown is in Figure 2 shown again in a frontal view.

[0044] At the in Figure 1 and Figure 2The compressor shown (1) is a two-stage compressor, and the two compressor stages 10 and 20 are located in the Figures 3-5 depicted.

[0045] The first compressor stage 10 comprises a main rotor 11 and a secondary rotor 14, which rotate and interlock to compress the medium fed into the housing 2 via the inlet 3. The two rotors 11 and 14 are mounted on their respective shafts 12 and 15 to prevent rotation.

[0046] The second compressor stage 20 also comprises a main rotor 21 and a secondary rotor 24, which rotate and interlock to further compress the medium delivered by the first compressor stage 10. The two rotors 21 and 24 of the second compressor stage 20 are also mounted to their respective shafts 22 and 25 in a rotationally fixed manner.

[0047] All shafts are mounted within the housing in a conventional manner, although the details of the mounting are not shown for the sake of clarity.

[0048] The drive shaft 5 has a first drive wheel 51 and a second drive wheel 52 for independently driving the two compressor stages 10, 20, each of which is mounted on the drive shaft 5 in a rotationally fixed manner.

[0049] The first drive wheel 51 has teeth on its circumferential surface that engage with corresponding teeth on a first output pinion 13, which is mounted non-rotatably on the shaft 12 of the first main rotor 11. Thus, rotation of the drive shaft 5 and the first drive wheel 51 results in a corresponding rotation of the shaft 12 of the first main rotor 11. In the case of the oil-injected compressor shown in the figures, the first auxiliary rotor 14 is driven by the first main rotor 11 via the helical gearing with a transmission ratio determined by the respective number of teeth on the two rotors. The rotational speed of the first main rotor 11 is determined by the transmission ratio between the first drive wheel 51 and the first output pinion 13.

[0050] In the case of a non-oil-injected compressor, the first main rotor and the first auxiliary rotor can also be rotaryally connected to each other via a further gearbox sufficiently known to those skilled in the art, so that the two rotors rotate synchronously in an interlocking manner during operation.

[0051] In the illustrated embodiment, the toothing 16 between the first drive gear 51 and the first output pinion 13 is inclined with respect to the axis 5A of the shaft 5, as will be described later in connection with the Figures 6a and 6b will be described in more detail.

[0052] The second drive wheel 52 also has teeth on its circumferential surface, which, however, engage with corresponding teeth on a second output pinion 23 that is rotationally fixed to the shaft 22 of the second main rotor 21. Thus, rotation of the drive shaft 5 and the second drive wheel 52 results in a corresponding rotation of the shaft 22 of the first main rotor 21. Here, too, the second auxiliary rotor 24 is driven by the second main rotor 21 via the helical gearing with a transmission ratio determined by the respective number of teeth on the two rotors. The rotational speed of the second main rotor 21 is determined by the transmission ratio between the second drive wheel 52 and the second output pinion 23 and can be adjusted independently of the rotational speed of the first main rotor 11 by appropriately adjusting the radial dimensions of the two elements.

[0053] Here again, it applies that in the case of a non-oil-injected compressor, the second main rotor and the second auxiliary rotor can be rotaryally connected to each other via a further gearbox that is sufficiently known to those skilled in the art, so that the two rotors rotate synchronously and interlock during operation.

[0054] In the illustrated embodiment, the toothing 26 between the second drive gear 52 and the second output pinion 23 is also inclined with respect to the axis 5A of the shaft 5, as will be described later in connection with the Figures 6a and 6b will be described in more detail.

[0055] In the Figures 6a and 6b The gearbox of the embodiment of the drive system according to the invention shown in the preceding figures is shown in a side view.

[0056] As already mentioned, in the illustrated embodiment the teeth arranged and engaged on the circumferential sides of the drive wheels 51 and 52 and the output pinions 13 and 23 are arranged at an angle with respect to the axis 5A of the drive shaft.

[0057] Here, the teeth arranged on the first drive wheel 51 form an angle α with the axis 5A, while the teeth arranged on the second drive wheel 52 form an angle β with the axis 5A, as shown in Figure 6a is shown schematically.

[0058] Due to this helical gearing 16, 26, when the drive shaft 5 is rotated in the direction of rotation R, the following results are obtained in the Figures 6a and 6b Axial forces F shown, acting on the drive shaft 5.

[0059] By appropriately selecting the angles α and β and the transmission ratios between the drive wheels 51, 52 and the associated output pinions 13, 23, the axial forces on the drive shaft 5 caused by the rotation can be controlled.

[0060] In the Figures 6a and 6b In the exemplary embodiment shown, the angles α and β with respect to the axis 5A of the drive shaft have opposite signs, so that the resulting forces are opposite to each other.

[0061] However, it would also be conceivable that the angles α and β with respect to axis 5A have the same sign, so that the resulting axial forces on the drive shaft 5 are in the same direction.

[0062] In the Figures 7a - 7eExamples of various arrangements of two compressor stages with a drive system according to the invention are shown. Due to the free choice of the center distances from the drive shaft 5 to the shaft 12 of the first main rotor 11 and to the shaft 22 of the second main rotor 21, indicated schematically by corresponding arrows in the figures by the use of a second drive wheel 52 according to the invention, there is a great deal of flexibility in how the two compressor stages are arranged relative to each other in the housing.

[0063] Figure 7fFigure 1 shows an example of a three-stage compressor which uses a drive system according to the invention, in which an additional third drive wheel 53 is mounted non-rotatably on the drive shaft 5 and engages in a third output pinion 33 which is mounted non-rotatably on the shaft 32 of a main rotor of the third compressor stage, in order to realize a drive of a third compressor stage which is conditioned via the common drive shaft 5 but is independent of the drive of the first two compressor stages.

[0064] Embodiment 1: Drive system for a multi-stage screw compressor (1) comprising a first compressor stage (10) with a first main rotor (11) mounted non-rotatably on a first shaft (12), a second compressor stage (20) with a second main rotor (21) mounted non-rotatably on a second shaft (22), wherein the drive system has a drive shaft (5) for driving the shafts (12, 22) of both main rotors (11, 21) of the first and second compressor stages (10, 20), wherein the drive of the shaft (12) of the first main rotor (11) is independent of the drive of the shaft (22) of the second main rotor (21).

[0065] Embodiment 2: Drive system according to embodiment 1, wherein, due to the drive of the shaft (12) of the first main rotor (11) independently of the drive of the shaft (22) of the second main rotor (21), the rotational speeds of the two main rotors (11, 21) can be adjusted independently of each other.

[0066] Embodiment 3: Drive system according to embodiment 2, wherein, due to the drive of the shaft (12) of the first main rotor (11) independently of the drive of the shaft (22) of the second main rotor (21), the ratio of the rotational speeds of the two main rotors (11, 21) to each other is freely selectable.

[0067] Embodiment 4: Drive system according to one of embodiments 1 - 3, wherein the drive shaft (5) has two drive wheels (51, 52) arranged non-rotatably on the drive shaft, of which a first drive wheel (51) engages with a first output pinion (13) arranged non-rotatably on the shaft (12) of the first main rotor (11) and of which a second drive wheel (52) engages with a second output pinion (23) arranged non-rotatably on the shaft (22) of the second main rotor (21).

[0068] Embodiment 5: Drive system according to one of the preceding embodiments, wherein the transmission between the second drive wheel (52) and the second output pinion (23) has a different gear ratio than the transmission between the first drive wheel (51) and the first output pinion (13).

[0069] Embodiment 6: Drive system according to one of the preceding embodiments, wherein the transmission between the second drive wheel (52) and the second output pinion (23) has the same gear ratio as the transmission between the first drive wheel (51) and the first output pinion (13).

[0070] Embodiment 7: Drive system according to one of the preceding embodiments, wherein the toothing (16, 26) between at least one of the drive wheels (51, 52) and the corresponding output pinion (13, 23) is inclined to the axial extension (5A) of the drive shaft (5).

[0071] Embodiment 8: Drive system according to embodiment 7, wherein the toothing (16, 26) between the two drive wheels (51, 52) and the respective output pinions (13, 23) is inclined to the axial extent (5A) of the drive shaft (5) and wherein the angles (α, β) formed by the toothing (16, 26) of the respective drive wheel (51, 52) with the output pinion (13, 23) to the axial extent (5A) of the drive shaft (5) are different.

[0072] Embodiment 9: Drive system according to embodiment 8, wherein the angles (α, β) formed by the toothing (16, 26) of the respective drive wheel (51, 52) with the output pinion (13, 23) to the axial extent (5A) of the drive shaft (5) have different signs.

[0073] Embodiment 10: Screw compressor (1) with a drive system according to one of embodiments 1 - 9.

[0074] Embodiment 11: Screw compressor (1) according to embodiment 10, wherein the screw compressor (1) has exactly two compressor stages (10, 20).

[0075] Embodiment 12: Screw compressor (1) according to embodiment 10 or embodiment 11, wherein the screw compressor (1) is an oil-injected compressor. Reference symbol list

[0076] 1 Compressor 2 Compressor housing 3 Inlet 4 Outlet 5 Drive shaft 5 Axial axis of the drive shaft 51 First drive wheel 52 Second drive wheel 53 Third drive wheel 10. First compressor stage 11. First main rotor 12. Shaft of the first main rotor 13. First output pinion 14. First auxiliary rotor 15. Shaft of the first auxiliary rotor 16. Gearing between first drive wheel and first output pinion α. ​​Angle between first gearing and axis of the drive shaft 20 Second compressor stage 21 Second main rotor 22 Shaft of the second main rotor 23 Second output pinion 24 Second auxiliary rotor 25 Shaft of the second auxiliary rotor 26 Gearing between second drive wheel and second output pinion β Angle between second gearing and axis of the drive shaft 32 Shaft of a third main rotor 33 Third output pinion Direction of action of axial forces; direction of rotation of the drive shaft

Claims

1. Drive system for a multi-stage screw compressor (1) comprising a first compressor stage (10) with a first main rotor (11) mounted non-rotatably on a first shaft (12), a second compressor stage (20) with a second main rotor (21) mounted non-rotatably on a second shaft (22), wherein the drive system has a drive shaft (5) for driving the shafts (12, 22) of both main rotors (11, 21) of the first and second compressor stages (10, 20), wherein the drive of the shaft (12) of the first main rotor (11) is independent of the drive of the shaft (22) of the second main rotor (21);wherein an independent drive of the two shafts (12, 22) of the main rotors (11, 21) is to be understood as such that the drive shaft (5) has two drive wheels (51, 52) arranged non-rotatably on the drive shaft, of which a first drive wheel (51) is engaged with a first output pinion (13) arranged non-rotatably on the shaft (12) of the first main rotor (11), and of which a second drive wheel (52) is engaged with a second output pinion (23) arranged non-rotatably on the shaft (22) of the second main rotor (21); wherein the transmission between the second drive wheel (52) and the second output pinion (23) has the same transmission ratio as the transmission between the first drive wheel (51) and the first output pinion (13); characterized by the fact thatthe toothing (16, 26) between the two drive wheels (51, 52) and the respective output pinions (13, 23) is inclined to the axial extent (5A) of the drive shaft (5) and wherein the angles (α, β) formed by the toothing (16, 26) of the respective drive wheel (51, 52) with the output pinion (13, 23) to the axial extent (5A) of the drive shaft (5) are different; and wherein the angles (α, β) formed by the toothing (16, 26) of the respective drive wheel (51, 52) with the output pinion (13, 23) to the axial extent (5A) of the drive shaft (5) have different signs.

2. Drive system according to claim 1, wherein, due to the drive of the shaft (12) of the first main rotor (11) independently of the drive of the shaft (22) of the second main rotor (21), the rotational speeds of the two main rotors (11, 21) can be adjusted independently of each other.

3. Drive system according to claim 2, wherein, due to the drive of the shaft (12) of the first main rotor (11) independently of the drive of the shaft (22) of the second main rotor (21), the ratio of the rotational speeds of the two main rotors (11, 21) to each other is freely selectable.

4. Screw compressor (1) with a drive system according to one of claims 1 to 3.

5. Screw compressor (1) according to claim 4, wherein the screw compressor (1) has exactly two compressor stages (10, 20).

6. Screw compressor (1) according to claim 4 or claim 5, wherein the screw compressor (1) is an oil-injected compressor.

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