EP4330552B1 - Drive system for a multistage screw compressor - Google Patents

Description

Application area

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

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.

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

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.

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.

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 one connected to the Compressor connected pressure vessels or the consumer network, which has the constant final pressure.

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.

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.

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.

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.

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.

In other words, with conventional 3-wheel transmissions, if the total power of the two-stage compressor is changed by changing the gear ratio between the drive transmission and the first compressor stage, the gear ratio to the The distance between the drive shaft and the second compressor stage is no longer defined but is necessarily determined by the center distance. 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 performance, which in turn results in a significant loss of efficiency. In summary, any change in the total power, final pressure, or volume flow of a two-stage compressor using a conventional 3-wheel drive, especially with a fixed and unchangeable distance between the main rotor axes and the drive shaft, results in a significant and undesirable loss of efficiency.

The documents US 2017/292525 A1 , US 4 524 638 A , US 5 378 128 A , DE 100 47 940 A1 and JP 2020 037922 A reveal various drive systems suitable for single- and multi-stage screw compressors. JP 2020 037922 A This discloses a drive system for a multi-stage screw compressor with the features of the preamble of claim 1, as well as a corresponding screw compressor. Helical toothed input and output pinions and their technical effects and advantages are also known, but not disclosed in the prior art in an obvious manner with the other features of claim 1.

Description of the invention

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

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.

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.

Advantageously, in a preferred embodiment of the invention, the drive of the shaft of the first main rotor is independent of the drive of the shaft of the second. The speeds of the main rotors can be adjusted independently of each other.

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.

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.

The drive shaft has two drive wheels mounted on it that are fixed against rotation. The first of these two drive wheels engages with a first output pinion that is fixed against rotation on the shaft of the first main rotor, while the second of the two drive wheels engages with a second output pinion that is fixed against rotation on the shaft of the second main rotor.

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.

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.

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.

Furthermore, the invention allows for the simple adjustment of the best possible intermediate pressure by freely choosing the translation ratios between The two compressor stages also allow for 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 efficiency, can be chosen thermodynamically favorably just above the point of condensation.

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.

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.

A screw compressor typically comprises one or more pairs of main and secondary rotors that mesh together in a helical shape. 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 of the screw pair. This pressure gradient generates axial and radial forces on the rotors, which are absorbed by the bearings of the two rotors. The forces caused 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 contributions of the resulting forces differ in the radial and axial directions. The greatest axial force typically occurs in the main rotor.

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.

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.

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.

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.

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. The toothing between both drive gears and their respective output pinions is inclined relative to the axial extent of the drive shaft, with the angles formed by the toothing of each drive gear with the output pinion relative to the axial extent of the drive shaft being different. The angles formed by the toothing of each drive gear with the output pinion relative to the axial extent of the drive shaft have opposite signs.

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.

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.

The use of helical gearing on the drive gear and the driven 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. In contrast to 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. The forces resulting from the helical gearing Depending on the orientation of the gearbox's helix angle, the generated axial forces 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.

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.

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.

This also applies to more than the two-stage compressors described (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 lifespan can be achieved.

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

• Fig. 1: einen Verdichter mit einem erfindungsgemäßen Antriebssystem in perspektivischer Ansicht,
• Fig. 2: eine Ansicht des Verdichters aus Figur 1 in Frontalansicht,
• Fig. 3: eine perspektivische Ansicht der beiden Verdichterstufen des Verdichters aus Figur 1,
• Fig. 4: eine Frontalansicht der in Fig. 3 gezeigten Verdichterstufen,
• Fig. 5: eine Seitenansicht der in Fig. 3 gezeigten Verdichterstufen,
• Fig. 6a/b: Seitenansicht des Getriebes des in den vorherigen Figuren gezeigten Verdichters,
• Fig. 7a-f: Schematische Frontalansichten verschiedener erfindungsgemäßer Antriebssysteme

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

Figure 1 Figure 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.

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).

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

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

The first compressor stage 10 comprises a main rotor 11 and a secondary rotor 14, which rotate to compress the medium led into the housing 2 via the inlet 3. interlock. The two rotors 11 and 14 are mounted on their respective shafts 12 and 15 in a rotationally fixed manner.

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.

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.

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.

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.

In the case of a non-oil-injected compressor, the first main rotor and the first secondary rotor can also 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 in an interlocking manner during operation.

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.

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 on the shaft 22 of the second main rotor 21, so that a 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 teeth with a gear ratio that is The rotational speed of the second main rotor 21 is determined by the gear ratio between the second drive gear 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 size of the two elements.

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.

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.

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.

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.

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.

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.

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.

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.

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.

In the Figures 7a - 7e Examples 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.

Figure 7f Figure 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.

Reference symbol list

1

compressor

2

Compressor housing

3

inlet

4

Outlet

5

drive shaft

5A

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 secondary rotor

15

Wave of the first auxiliary rotor

16

Gearing between first drive gear and first output pinion

α

Angle between first tooth and axis of the drive shaft

20

second compressor stage

21

second main rotor

22

wave of the second main rotor

23

second output pinion

24

second auxiliary rotor

25

Wave of the second auxiliary rotor

26

Gearing between second drive wheel and second output pinion

β

Angle between second tooth and axis of the drive shaft

32

Wave of a third main rotor

33

third output pinion

F

Direction of action of axial forces

R

Direction of rotation of the drive shaft

Claims (6)

1. A drive system for a multistage screw compressor (1) comprising

   a first compressor stage (10) with a first main rotor (11) mounted on a first shaft (12) in a torque-proof manner,

   a second compressor stage (20) with a second main rotor (21) mounted on a second shaft (22) in a torque-proof manner,

   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 driving of the shaft (12) of the first main rotor (11) takes place independently of the driving of the shaft (22) of the second main rotor (21);

   wherein an independent driving of both shafts of the main rotors should be understood such that the drive shaft (5) has two drive wheels (51, 52) which are arranged on the drive shaft in a torque-proof manner, a first drive wheel (51) of which is in engagement with a first output pinion (13), which is arranged on the shaft (12) of the first main rotor (11) in a torque-proof manner, and a second drive wheel (52) of which is in engagement with a second output pinion (23), which is arranged on the shaft (22) of the second main rotor (21) in a torque-proof manner;

   wherein the gearing mechanism has a different transmission between the second drive wheel (52) and the second output pinion (23) to the gearing mechanism between the first drive wheel (51) and the first output pinion (13);

   characterised in that

   the toothing (16, 26) between the two drive wheels (51, 52) and the respective output pinions (13, 23) is formed at an incline 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. The drive system according to claim 1,
   wherein, due to the driving of the shaft (12) of the first main rotor (11) independently of the driving of the shaft (22) of the second main rotor (21), the rotational speeds of the two main rotors (11, 21) can be adapted independently of one another.
3. The drive system according to claim 2,
   wherein, due to the driving of the shaft (12) of the first main rotor (11) independently of the driving of the shaft (22) of the second main rotor (21), the ratio of the two rotational speeds of the two main rotors (11, 21) relative to one another can be freely selected.
4. A screw compressor (1) with a drive system according to any one of claims 1 to 3.
5. The screw compressor (1) according to claim 4,
   wherein the screw compressor (1) has precisely two compressor stages (10, 20).
6. The screw compressor (1) according to claim 4 or claim 5,
   wherein the screw compressor (1) is an oil-injected compressor.