Multi-stage compressor system for generating a compressed gas

Abstract

A compressor system for generating a compressed gas having at least one main compressor to which pre-compressed gas is fed using at least one pre-compressor, wherein the at least one main compressor is driven, as is also the at least one pre-compressor, by a common drive motor or separate drives, wherein at least one pre-compressor is assigned a hydrodynamic fluid coupling for changing the drive speed of the pre-compressor.

Description

FIELD

The invention relates to a compressor system for generating a compressed gas according to the preamble of claim 1.

BACKGROUND

The compressor systems which are used in industry and in the production industries consume a considerable proportion of electrical energy and the use of internal combustion engines means that in the case of mobile systems considerable quantities of the limited natural raw materials based on mineral oil or else renewable energy carriers, such as for example biofuels, are consumed. Owing to rising energy costs and environmental stress generated by the consumption of energy, more efforts are always having to be made to lower energy consumption and therefore at the same time lower the costs and the consumption of natural resources.

Compressor systems for generating compressed air are the most widespread systems in compressor systems which are currently in use. However, systems for generating process gases of any type or for feeding gases over relatively large distances are implemented as compressor systems whose drive power levels extend into the two-digit megawatt range. In industrial nations, it can therefore be assumed that approximately 10% of the entire amount of electrical energy consumed by all industrial companies is consumed by compressor systems. It is therefore not difficult to see that it is particularly important to improve efficiency in this area by new developments and innovations and therefore save energy.

SUMMARY

In contemporary compressor systems, the problem occurs that they can only be regulated insufficiently in terms of their delivery quantity, for example of compressed air. This gives rise, in particular in the partial load range, to insufficient efficiency and therefore to unnecessary consumption of energy.

Customary electric motors are only able to be operated at a rotational speed which depends on the power supply frequency, which cannot be influenced, of the available electric current. This type of compressor system therefore operates efficiently only when the compressor systems output precisely the air quantity which corresponds to their rotational speed. Remaining control alternatives are here only switching off the system or closing the intake regulator, which is equivalent to reducing the volume flow to virtually zero. However, both types of regulation have considerable disadvantages with respect to efficiency.

Closing the intake regulator leads to reduction in the volume flow to virtually zero, but the necessary drive power level continues to be very high and is as a rule more than 50% of the full load power. Lowering the vessel pressure does not lead to a significant improvement either, since said lowering does not take place until after a certain time delay and also requires a specific amount of time in order to reduce the vessel pressure. Even with this reduced vessel pressure, the power drain is still approximately 25% of the full load power and the vessel volume which is discharged is destroyed.

Although it is possible to switch off the drive motor immediately, the number of the renewed starts, in particular in the case of electric motors, is limited.

A considerable number of starts is therefore not possible.

A further possibility provides throttle regulation. Owing to its method of operation, of reducing the intake pressure, this type of regulation is also extremely inefficient.

An alternative for dealing with the problems of insufficient energetically appropriate regulation of the volume flow is provided by the developing technology in the form of frequency-regulated drives. In this context, the rotational speed of the drive motor can be regulated within technically defined limits. However, this type of regulation of the volume flow is relatively cost-intensive and also cannot be used in an optimum way with respect to efficiency. In particular in the partial load range, disadvantages occur owing to the relatively low efficiency of the frequency converter and of the electric motor.

In the frequently used screw compressors with or without fluid injection, when there is a combination of frequency-regulated drives there is an additional significant worsening of the efficiency since this type of compressor has efficiency which is different over the rotational speed range. In the case of a constant drive speed, it is possible to optimize the screw compressor for this drive speed. If this rotational speed range is undershot or exceeded, the efficiency decreases significantly. In the case of a low rotational speed, increased backflow of the already compressed gas occurs. In the case of a relatively high rotational speed, significant hydraulic losses occur, in particular with the screw compressors with fluid injection.

In compressor systems with a drive by means of an internal combustion engine, a certain amount of volume flow regulation is possible by means of a process of changing the rotational speed. However, this process is severely limited by the virtually constant torque demand of the fluid-injected screw compressor which is used almost exclusively. The regulation of the volume flow which is achieved on this basis is usually only 20 to 40% and is therefore also very uneconomic. In these compressor systems, the fuel consumption without compressed air being extracted is still 50% of the consumption at the full rotational speed and maximum operating pressure.

As well as the regulation of the volume flow, regulation of the operating pressure cannot be carried out in an optimum way in contemporary compressor systems either. A higher operating pressure with the same volume flow of the system results in a higher torque demand owing to a constant drive speed, and the drive motor cannot make available this additional torque. In practice this means a significant limitation in terms of the flexibility of the use of the compressor. Increasing the operating pressure is in practice only ever possible by modifying the system or by using a system with a relatively high operating pressure which is over-dimensioned for the regulating mode and can then be correspondingly reduced.

However, with such a procedure the investment in the compressor system is significantly higher and the flexibility is still considerably limited, since with the relatively low operating pressure the volume flow is not increased but instead there is merely a reduction in power. The system therefore cannot be used completely within the scope of its performance capability. This applies, in particular, to the screw compressors which are used.

A higher delivery quantity is critical, in particular, for screw compressors, since technical problems occur here both with respect to a limited rotational speed range and with respect to the manufacturability with limited rotor diameters. These limits can be reached starting from a drive power level of approximately 250 to 315 kW.

In particular, the economic limits should also be evaluated critically, since the manufacture is made more difficult as a result of relatively small manufacturing numbers and a considerable expenditure. Starting from the abovementioned power range, other compressor systems also come into the picture, for example multi-stage turbo compressors. These turbo compressors are distinguished by better efficiency compared to screw compressors. The specified power ranges do not constitute a fundamental limitation, but they can become an impassable limitation for smaller and medium-sized manufacturers.

A further significant point are the geographic differences and conditions of use which result from worldwide marketing of these systems. The configuration of a compressor system usually occurs according to the viewpoints of the application which is provided. It is customary in this context to put limitations on the conditions of use in order to avoid unnecessarily increasing manufacturing costs of such a compressor system. These conditions of use are limited by a maximum climatic temperature of use and a maximum altitude of use. An excessively high temperature of use may mean that the cooling of the compressor system is not sufficient. Both the drive motor and the actual compressor have their limitations in use with respect to the operating temperature, and in the event of inadequate cooling can be damaged or their service life can be significantly restricted if the operating temperature is very close to the critical limitation in use over a relatively long period of time.

The same applies correspondingly also to use at relatively high altitudes. Owing to the relatively low air pressure, internal combustion engines can experience overheating since a sufficient quantity of oxygen for combustion is not available. Sufficient cooling is also put at risk since given the same cooling air volume flow the mass flow is reduced as a result of the lower air density. The cooling is considerably reduced as a result. This applies both to the drive motor and to the actual compressor. At an altitude of 2000 m above sea level this can already lead to reduction in the cooling power of up to 20%. If compressor systems are to be constructed for such conditions of use, a larger cooling system must be installed and, if appropriate, a larger drive motor also becomes necessary in order to avoid overloading. Both these things entail significantly higher costs.

The publication DE 2848030 A1 discloses a method for the multi-stage compression of gases, in particular oil-free air, wherein here the entire compression process is divided among a plurality of stages in such a way that the individual compressors are operated at the best point, and at the same time achieve the lowest power drain, and that given a partial load only the first and the last compressor stages are operated under conditions which are changed to a greater extent in comparison to the rated load, while the pressure conditions, the volume flows and therefore the optimum operation remain the same in the middle stages and when there is a change in load the quantity adaptation takes place in the first compressor stage, and the adaptation of the pressure to the consumer takes place in the last stage.

However, it proves a disadvantage here that the multi-stage compressor for carrying out this method connects the flow compressor and expulsion compressor, in particular screw compressor, in series in hydraulic terms, by means of pipelines, and each compressor stage is operated with the rotational speed which is optimum for it with at least two output shafts via a power split transmission. The number of compressor stages is to be a maximum of four, in the first stage it is always a turbo compressor and in the last stage it is an expulsion compressor. The stages lying between them are turbo compressors or expulsion compressors. This makes the large structural expenditure with only restricted regulating capability clear.

A combined arrangement of turbo compressors and screw compressors is proposed in the disclosures in US-PS 36 40 646 and FR-PS 13 97 614, but these solutions also have disadvantages, since here the turbo compressors and screw compressors run on one shaft at the same rotational speed, as a result of which optimum operation with good efficiency levels is not possible for the volume range and pressure range relating to this invention, in the same way as with the multi-stage compressors of the same generic type.

The publication WO 2006 061 221 A1 discloses a method for regulating the maximum rotational speed of a compressor which is driven by means of a motor via a hydrodynamic coupling. The background here is that the compressor is used in a vehicle and is also driven by the drive motor thereof, with the result that the hydrodynamic coupling ensures that a defined maximum rotational speed of the compressor is not exceeded. The hydrodynamic coupling therefore serves only to a limited degree to regulate the compressor, which is additionally conceived of as a single-stage working machine and is to be kept only in the upper limit of its rotational speed.

Finally, WO 2009 090 075 A1 relates to a turbocharger-turbo compound system, in particular for a vehicle having an internal combustion engine with an exhaust gas power turbine which is arranged in the exhaust gas flow in order to convert exhaust gas energy into mechanical energy, and having a compressor which can be driven by means of the exhaust gas power turbine, wherein the compressor is connected in a second drive connection to the output shaft of the internal combustion engine or can be shifted into such a connection, and the turbo compound drive connection has at least two parallel power branches with different transmission ratios, in order to vary the rotational speed ratio between the turbocharger drive connection and the internal combustion engine output shaft. The invention is characterized in that a hydrodynamic coupling is provided in both power branches. This is intended to ensure that various transmission ratios can be set in order to change the rotational speed ratio between the drive connection between the exhaust gas power turbine and the compressor, and the internal combustion engine output shaft by switching the drive power flow from one power branch to the other power branch.

Since this is a specific solution for improving turbocharger-turbo compound systems in vehicles, the basic benefit of hydrodynamic couplings for adapting the rotational speed ratio with an effect on a compressor which is to be driven is known, but not for improving a compressor and the regulating capability thereof, but rather for the sake of better use of an exhaust gas turbocharger in a vehicle with an internal combustion engine in the different operating states which are present.

The system described below serves to generate compressed air or some other gas which is suitable for compression. In this context, the object of the present invention is to combine, in particular, the methods of operation and the different characteristics of different compressor types with one another in such a way that an optimum concept of such a novel compressor system is produced. The aimed-at improved properties relate here to a maximum efficiency level with low manufacturing costs, a good regulating capability and a high degree of flexibility.

The desired improvements are to be achieved in that in a compressor system with one or more main compressors, one or more pre-compressors are connected upstream. By means of this pre-compression, the entire compressor power is divided between at least two compressors. By means of suitable intermediate cooling, the efficiency level of such a system is significantly improved. Furthermore, with such a system it is possible to regulate the operating pressure, permitting the operator to set different operating pressures without modifying the system.

In order then to achieve a variable rotational speed of the pre-compressor according to the invention, a further element is necessary. According to the invention, a fluid coupling serves as the connecting element between the drive of the turbo compressor and the drive motor or the main compressor. This fluid coupling permits the rotational speed of the turbo compressor to be changed independently of the rotational speed of the drive motor or main compressor. This change ensures that there is an increase or a reduction in the volume flow or the admission pressure. The use of a turbo compressor in connection with a fluid coupling yields the optimum combination of the operating characteristics, since both are hydraulic machines by their nature. The adaptation of the torque and rotational speed is optimum with this combination and permits the maximum overall efficiency level.

The desired volume flow can therefore be regulated in an infinitely variable fashion by means of the quantity of the fluid circulating in the fluid coupling. The rotational speed of the secondary part of the coupling determines the rotational speed of the turbo compressor and therefore the volume flow and the admission pressure. The rotational speed of the secondary part in the fluid coupling is at maximum at the drive speed of the primary part of the fluid coupling minus a necessary difference in rotational speed for the transmission of power, and can be regulated in a downward direction by reducing the filling quantity in the fluid coupling.

This necessary slip at the same time also constitutes the efficiency loss. In order to keep the efficiency loss within limits, it may be significant to change the rotational speed of the turbo compressor by means of one or more transmission ratios. An excessively large difference in rotational speed of the turbo coupling between the primary and secondary rotational speed can be prevented with these fixed stages, which improves the efficiency level significantly even in the partial load range. Basically, it may be expedient for all designs of the inventive compressor system here that the transmissions which are used are embodied in a shiftable fashion.

In particular main compressors which are capable of achieving a final pressure which is independent of the admission pressure are suitable for a system according to the invention. These are, in particular, all compressors which operate according to the positive displacement principle, such as, for example screw compressors, piston compressors, scroll compressors or else rotary gear compressors.

The compressor type which is used most frequently at present is the screw compressor. It is used with or without fluid injection. In order to achieve an optimum efficiency level with this compressor type, it should be attempted to make the volume ratio or pressure ratio variable and therefore to set an optimum compression ratio which is dependent on the actual admission pressure and the actual final pressure. This type of setting of an optimum volume ratio or pressure ratio is already used today in compressors in the refrigerant industry. By means of a slider or another suitable adjustment instrument it is possible to set an optimum ratio during operation.

Such adjustability of the ratio is provided in a design of the present invention with a screw compressor as an expedient embodiment. Depending on the actual final pressure and the actual admission pressure, an optimum internal volume ratio of the screw compressor can be set in which, for example, a regulating slide and/or suitable valves are influenced in such a way that a variable connection can be produced between the compressor space of the screw compressor and the outlet region, and therefore the internal volume ratio can be changed in terms of its effectiveness.

When the optimum compression ratio is set, it is appropriate to use the ratio of the actual outlet pressure from the screw compressor to the inlet pressure into the screw compressor. This also applies correspondingly if the main compressor is composed of two or more screw compressors. The optimum compression ratio can also be set here by virtue of the fact that it is selected with the ratio of the actual outlet pressure to the actual inlet pressure of the respective screw compressor. In the other suitable compressor types it may also be appropriate to carry out suitable adaptation measures for optimizing the efficiency level.

Basically all types of compressors are suitable for pre-compression, but the turbo compressor is particularly well suited for pre-compression. By virtue of its design, it is capable of sucking in large volume flows and bringing them to a corresponding final pressure which depends, however, very strongly on the rotational speed. The characteristic of the turbo compressor therefore permits significant changes in the admission pressure and in the volume flow to be generated by changing the rotational speed.

Turbo compressors with a radial through-flow are particularly well suited to the requirements described here for the novel compressor system. These radial turbo compressors permit pressure ratios of up to four to be formed in one stage. This therefore produces an extremely compact overall size and a pre-compressor design which is best suited to the requirements.

The drive of the pre-compressor can be provided directly by means of the drive motor via a corresponding transmission or as part of the main compressor with the corresponding power splitting. The latter gives rise to a particularly compact and efficient unit which can be driven either by an internal combustion engine or by an electric motor.

For larger volume flows it is also possible to connect a plurality of turbo compressors in parallel or in series, in particular if extremely high admission pressures are required. Intermediate cooling also ensures an improved efficiency level here.

In addition, the regulating range of the turbo compressor can be extended by means of a suitable inlet directing apparatus. However, other regulating elements for extending the characteristic diagram of the turbo compressor are also possible. These regulating elements can be mounted upstream or downstream of the turbo compressor. Further regulating elements for the main compressor can be mounted upstream of the pre-compressor or downstream of the pre-compressor, i.e. simultaneous also upstream of the main compressor or else downstream of the main compressor. The arrangement is directed according to the requirements of the compressor concept composed of a pre-compressor and main compressor.

When a plurality of pre-compressors are used there is the possibility of driving them jointly by way of one fluid coupling or else of using a suitable fluid coupling for each pre-compressor. A combination also appears possible such that, for example, two pre-compressors are driven by one fluid coupling in the case of three or more pre-compressors. All that is decisive here is that the combination is technically and commercially appropriate for the respective application.

In a design with a fluid-injected compressor, the cooling and lubricating fluid of the compressor can also be used to operate the fluid coupling, thereby permitting a common circuit with an optimum ratio between costs and benefits. The fluid which is required for the fluid coupling is extracted here from the main flow of the cooling and lubricating fluid which has been cooled back. If the pre-compressor is integrated structurally into the main compressor, the liquid for operating the fluid coupling in the main compressor can be removed and also fed back again to the circuit there, provided that the necessary pressure differences are complied with.

In the case of a design of the pre-compressor without integration into the main compressor, the operating fluid of the fluid coupling can also be extracted from the main flow of the cooling and lubricating fluid and also fed back again into this circuit while complying with the necessary pressure differences. If these pressure differences are not present, it is possible to use a fluid pump to ensure that the necessary pressure difference is present. This applies both to the variant in which the pre-compressor is integrated into the main compressor, and also to the variants in which the pre-compressor is not integrated into the main compressor.

In compressors without fluid injection there is the possibility of integrating a corresponding circuit with a fluid for operating the fluid coupling or else of also using an often included oil circuit for lubrication of the bearings and cooling of the housing. Given a correspondingly sufficiently large quantity of oil and pump size it is also possible to form here a compact structural unit without a large amount of additional expenditure. It is also possible here to install this in an integrated fashion in the compressor housing of the main compressor or else to accommodate the pre-compressor in a separate housing and to extract the fluid from the oil circuit for lubrication of the bearings and cooling of the housing of the main compressor and also to feed said fluid back into said oil circuit. If the quantity of oil and/or the pump are not sufficient, they are to be newly dimensioned in accordance with the requirements of the fluid coupling.

A further possibility is to operate the fluid coupling in an independent oil circuit. This would be appropriate, in particular, when the fluid coupling was installed in a separate housing together with the transmission and the pre-compressor. However, such a solution requires a considerably higher structural expenditure, since both a separate oil pump and a separate oil cooler would be necessary. Whether this additional expenditure is economically justified must be tested in an individual case. In an arrangement of these structural components in a separate housing, it may be additionally appropriate as an alternative that the hydrodynamic fluid coupling has its own drive and is therefore not coupled in this respect to the main compressor.

For this purpose, it is also basically to be noted that the question of a separate drive for the fluid coupling is basically relevant and is to be answered as a function of the operating parameters of the compressor system. If basically a compact design of the pre-compressor and main compressor with all the structural components and drive in one housing is to be aimed at, it may nevertheless be appropriate in a specific application case, either technically or else economically, to separate the components structurally and drive them separately.

In order to regulate the filling level of the fluid coupling it is appropriate to compare the desired final pressure of the compressor with the actual operating pressure. If the actual operating pressure is lower than the desired operating pressure, the filling of the fluid coupling can be increased until the maximum rotational speed is reached. If the aimed-at operating pressure is exceeded, the operating fluid in the fluid coupling is reduced in an analogous fashion until the desired operating pressure is reached.

There is provision here to use suitable valves to influence either the inflow of the operating fluid or else the outflow of the operating fluid. A combination of valves in the region of the inflow as well as the outflow of the operating fluid is also a structural alternative. The continuous inflow and outflow of operating fluid simultaneously ensures corresponding cooling, since the waste heat of the fluid coupling must be carried away. A further alternative for monitoring and regulating the fluid coupling is a rotational speed monitoring means.

Monitoring the achieved admission pressure can also be used as an indication of the operating state of the fluid coupling.

The following advantages can now be achieved with the embodiments illustrated above:

1. Volume Flow Adaptations

The main compressor is kept at a constant rotational speed by a drive motor and therefore takes in a constant volume flow of the gas to be compressed. If an admission pressure is then built up by means of the pre-compressor, the volume flow of the entire compressor system increases. Given an admission pressure of 1 bar (psig), the volume flow is doubled given an identical temperature, tripled given 2 bar (psig) and even quadrupled given 3 bar (psig).

2. Relatively Low Manufacturing Costs

In order to implement a compressor system with a desired volume flow, given an admission pressure of 1 bar (psig) and an identical temperature, the main compressor only has to be given dimensions which are half as large as those for a compressor system without pre-compression. Given an admission pressure of 2 bar (psig), the overall size is even reduced to approximately a third, and analogously in the case of 3 bar (psig) it is reduced to approximately a quarter of the customary overall size.

In the case of compressor systems with drive by means of an electric motor, it is additionally possible to dispense with the use of a frequency converter, with the result that a significantly more cost-effective standard electric motor can also be used. The additional financial outlay on the pre-compressor, the fluid coupling and the additional intermediate cooler is considerably lower than the achievable cost savings, as a result of which the solution according to the invention is economically appropriate.

3. Improvement of the Efficiency Level

The overall efficiency level given good intermediate cooling can be improved by more than 15% by means of an optimum selection of the compression ratio between the pre-compressor and the main compressor. This gives rise to considerable reductions in the energy consumption and therefore to significant cost savings during the operation of the compressor system.

4. More Flexible Use of the Compressor System

If a higher operating pressure is desired by the operator of the system, said operating pressure can be formed by lowering the admission pressure and therefore reducing the volume flow. The main compressor requires a relatively high torque for the relatively high pressure, and the pre-compressor can reduce its torque demand by means of a lower admission pressure until equilibrium is established again between the available torque of the drive motor on the one hand and the required torque of the pre-compressor and main compressor on the other.

Adapting to the ambient conditions with respect to the height above sea level and the ambient temperature is therefore also possible. The utilization factor of the drive motor can also be taken into account by suitable monitoring. This is possible, for example in the case of electric motors, by monitoring the winding temperature. Therefore, overloading of the motor is prevented, and at the same time the compressor system can always be used until the limiting values for use are achieved. This applies accordingly also to a combination with an internal combustion engine as the drive. The system is therefore able to react in a flexible way to different conditions of use, as a result of which a significantly extended range of use of the compressor system according to the invention is possible without structural adaptations.

5. Extension of the Range of Use of the Compressor System

In compressor systems which are driven by an electric motor, it is possible according to the invention to adapt the utilization factor of the motor in an optimum way by monitoring various parameters, for example the power drain and the current voltage or else by monitoring the winding temperature.

Other parameters for monitoring the electric motor are also possible. Extreme ambient conditions such as a low air pressure due to being located at a high altitude or extremely high ambient temperatures can be compensated by varying the admission pressure and therefore changing the power drain of the pre-compressor and the main compressor. Overloading of the motor can thereby be prevented and at the same time utilization factor of the motor to its limits of use can be made possible without damaging the motor or shortening its service life.

This also applies accordingly in the use of an internal combustion engine. In this context, other parameters are then monitored, such as, for example, the exhaust gas temperature, the actual torque as a function of the fuel injection quantity or else the temperature of the cooling water, and are used as an assessment of the utilization factor.

In the case of screw compressors without fluid injection, the compressor system according to the invention is also advantageous. There is the problem here that the maximum final pressure is limited as a result of the two-stage embodiment. Said maximum final pressure is in the range from 10 to 11 bar (psig) (in special cases also up to 13 bar (psig) and is conditioned by the maximum permissible outlet temperature downstream of the second stage. It is basically the case that the higher the operating pressure the higher the outlet temperature.

If the admission pressure upstream of the first stage is then kept at a level above 0.5 bar (psig) by the pre-compression according to the invention, the final compression pressure downstream of the second stage can be raised to 15 bar (psig) or higher without exceeding the critical outlet temperatures. Furthermore, the regulating range of the intake volume flow can also be significantly extended.

In the case of relatively high operating pressures, there is the problem in conventional rotational speed-regulated drives that at relatively low rotational speeds so much compressed gas flows back that there can as a result already be a high outlet temperature. The second problem is the torque demand. This is known to rise with the operating pressure and can already exceed the limiting values for the maximum permissible torque as a result of the flowing-back of the compressed gas at low rotational speeds. Therefore, the compressor system must be kept at a minimum rotational speed which is also associated with the inadequate cooling of the electric motor at low rotational speeds.

The regulating range of the volume flow is therefore very restricted in conventional systems, and at 10 bar (psig) as the maximum operating pressure can already lead to a situation where the motor speed can be lowered only by approximately 40%. The volume flow regulating range is therefore only up to 40%. A compressor system according to the invention with comparable performance can implement a volume flow regulating range up to 70% to 80% with the same operating pressure, for example 10 bar (psig).

Even in compressor systems with fluid injection, the operating pressure can be significantly increased with an efficiency level which is still good and a volume flow regulating range which is still high if the admission pressure is kept at a minimum level. The admission pressure ensures that a relatively high operating pressure can still be formed with a technically optimum compression ratio of the main compressor. The standard of atmosphere at 10 bar (psig) signifies a compression ratio of 11, wherein an admission pressure of 0.5 bar (psig) given compression to 15.5 bar (psig) also signifies a compression ratio of the main compressor of 11.

BRIEF DESCRIPTION OF THE DRAWINGS

In the section which follows, the basic concept of this compressor system will be explained in more detail with reference to two schematic drawings, of which:

FIG. 1 shows the schematic design of a possible compressor system with a pre-compressor and a main compressor without fluid injection, and

FIG. 2 shows the schematic design of a possible compressor system with a pre-compressor and a main compressor with fluid injection and use of the fluid to operate the fluid coupling.

DETAILED DESCRIPTION

In the compressor system according to FIG. 1, the intake air is purified by means of an air filter 10 and then fed to the pre-compressor 20 via a connecting line 15. The pre-compressed intake air is fed to an intermediate cooler 25 via a further connecting line 16 and cooled as much as possible. The cooled and pre-compressed air is then fed to the main compressor 30.

This main compressor may basically be any type of compressor, even multi-stage compressors are possible, with or without intermediate cooling between the individual compressor stages of the main compressor. The main compressor 30 is driven by a drive motor 40. This drive motor 40 may have any possible design for driving a compressor. This may be, for example, electric motors, internal combustion engines or else turbines of any type.

In the main compressor 30, the compressed air is compressed to the desired operating pressure and fed to a pressure vessel 60 via a connecting line 17. In order to maintain a minimum operating pressure, a minimum pressure valve 65 is mounted at the outlet of the pressure vessel 60. The compressed air flows from this minimum pressure valve 65 via an extraction line 18 to the extraction point 70.

The fluid coupling 55 which is driven via the main compressor 30, serves, together with the transmission ratio 50, composed of the fluid coupling 55 and the transmission 56, to drive the pre-compressor 20. This is an exemplary illustration with a pre-compressor 20 and a flow coupling 55 which is assigned thereto.

However, for the comprehension of the invention it is essential that these are components which, depending on the compressor system, are to be used multiply and in the necessary configuration. It has thus already been stated that a plurality of pre-compressors 20 with a plurality of fluid couplings 55 can interact and they can also supply an admission pressure for a main compressor 30 which is configured differently. The transmission 56 which is illustrated here is therefore also only exemplary. A plurality of transmissions may also be necessary, such as also a separate drive of the pre-compressor and a separate fluid circuit. All these further combination options are not shown separately in the drawings but are instead also to be included in the technical teaching.

In the compressor system according to FIG. 2, the intake air is purified by means of an air filter 10 and then fed to the pre-compressor 20 via a connecting line 15. The pre-compressed intake air is then fed to an intermediate cooler 25 via a further connecting line 16 and cooled as much as possible.

The cooled and pre-compressed air is then fed to the main compressor 30. This main compressor may be in this design a compressor with fluid injection. The main compressor 30 is driven by a drive motor 40. This drive motor 40 may also have any possible design for driving a compressor. This can be, as stated above, for example, electric motors, internal combustion engines or else turbines of any type.

In the main compressor 30, the compressed air is compressed again to the desired operating pressure and fed to a pressure vessel 60 via a connecting line 17. In order to maintain a minimum operating temperature, a minimum pressure valve 65 is mounted at the outlet of the pressure vessel 60. The compressed air flows from this minimum pressure valve 65 via an extraction line 18 to the extraction point 70.

The fluid which is separated off from the compressed air in the pressure vessel 60 is fed again via a line 31 to a filter 34, a line 32 to a fluid cooler 35, and a line 33 to the main compressor 30. The fluid coupling which is driven by means of the main compressor 30, including the transmission ratio 50, serves to drive the pre-compressor 20. A partial flow of the cooling and lubricating fluid of the main compressor 30 serves as an operating fluid for the fluid coupling 55. This partial flow is extracted via a line 51 and then fed to the fluid coupling 55.

The operating fluid which emerges again from the fluid coupling 55 is fed again to the main compressor 30 via the line 52.

Claims

1. A compressor system for generating a compressed gas having at least one main compressor (30) to which pre-compressed gas is to be fed using at least one pre-compressor (20),

wherein

the at least one main compressor (30) is driven, as is also the at least one pre-compressor (20), by a common drive motor (40) or separate drives,

wherein at least one pre-compressor (20) is assigned a hydrodynamic fluid coupling (55) for changing the drive speed of the pre-compressor (20).

2. The compressor system as claimed in claim 1,

wherein

both the drive speed of the fluid coupling (55) and the drive speed of the pre-compressor (20) are adapted by at least one suitable transmission (56).

3. The compressor system as claimed in claim 1,

wherein

the pre-compressor (20) is composed of at least one turbo compressor which is embodied with a radial design.

4. The compressor system as claimed in claim 1,

wherein

at least two turbo compressors are arranged as pre-compressors (20) with parallel intake in order to increase the volume flow, said turbo compressors being driven by one or more fluid couplings (55).

5. The compressor system as claimed in claim 1,

wherein

at least two turbo compressors are arranged as pre-compressors (20) with serial intake in order to increase the admission pressure, said turbo compressors being driven by one or more fluid couplings (55).

6. The compressor system as claimed in claim 5,

wherein

at least one intermediate cooler is arranged between the at least two turbo compressors in order to improve the efficiency.

7. The compressor system as claimed in claim 1,

wherein

in order to regulate the fluid coupling (55), regulating elements for the infinitely variable adaptation of the rotational speed of the fluid coupling (55) are provided in the inflow and/or in the outflow of the operating fluid of the fluid coupling (55).

8. The compressor system as claimed in claim 1,

wherein

the mechanical drive of the pre-compressor (20) is integrated together with the fluid coupling (55) into a housing of the main compressor (30).

9. The compressor system as claimed in claim 1,

wherein

an intermediate cooler (25) is arranged between the pre-compressor (20) and the main compressor (30) in order to improve the efficiency.

10. The compressor system as claimed in claim 1,

wherein

the at least one main compressor (30) is a fluid-injected compressor,

wherein the fluid of the main compressor (30) is also to be fed as working medium to the fluid coupling (55) of the pre-compressor (20).

11. The compressor system as claimed in claim 1,

wherein

the drive of the fluid coupling (55) is provided directly or by way of a transmission from the main compressor (30) or from the drive motor (40).

12. The compressor system as claimed in claim 1,

wherein

the compressor system comprises one or more screw compressors with or without fluid injection and at least one radial turbo compressor for pre-compressing the gas.

13. The compressor system as claimed in claim 1,

wherein

the fluid of the lubrication circuit and/or cooling circuit for the bearings, the drive transmission and/or the housing cooling of the main compressor (30) is at least partially also the working medium of the fluid coupling (55) of the pre-compressor (20).

14. The compressor system as claimed in claim 1,

wherein

the pre-compressor (20), the fluid coupling (55) and all the transmission ratios (56) which are arranged for the purpose of reaching the desired rotational speed are integrated into a screw compressor.

15. The compressor system as claimed in claim 2,

wherein

the transmission ratios are embodied in a shiftable fashion in order to regulate the rotational speed.

16. The compressor system as claimed in claim 12,

wherein

an optimum internal volume ratio of the screw compressor is set as a function of the actual final pressure and of the actual admission pressure by mechanical adjustment devices and/or valves.

17. A method for regulating a compressor system as claimed in claim 1,

wherein

in the case of a drive (40) of the compressor system by an electric motor, the winding temperature thereof and/or the power drain and/or the current voltage are detected as parameters of the utilization factor of the electric motor and used as a regulated variable, in order to regulate, through changes in the operating parameters of the fluid coupling (55), the admission pressure which is generated by the pre-compressor (20) in such a way that the electric motor is loaded or relieved of loading to a greater degree.

18. The method for regulating a compressor system as claimed in claim 1,

wherein

in the case of a drive (40) of the compressor system by an internal combustion engine, the output gas temperature and/or coolant water temperature and/or actual torque thereof are detected as a function of the fuel injection quantity and the rotational speed as parameters of the utilization factor of the internal combustion engine and used as a regulated variable, in order to regulate, through changes in the operating parameters of the fluid coupling (55), the admission pressure generated by the pre-compressor (20) in such a way that the internal combustion engine is loaded or relieved of loading to a greater degree.

19. The compressor system as claimed in claim 11,

wherein

the transmission ratios are embodied in a shiftable fashion in order to regulate the rotational speed.

20. The compressor system as claimed in claim 14,

wherein

an optimum internal volume ratio of the screw compressor is set as a function of the actual final pressure and of the actual admission pressure by mechanical adjustment devices and/or valves.