WO2008064802A2 - Bearing and method for operating a magnetic bearing - Google Patents

Abstract

The invention relates to a bearing for the rotary mounting of a rotor on a stator, which has an idle position and a working position, comprising an actuator for moving the rotor or stator depending on a control signal and a control device to provide the control signal, wherein the control device provides the control signal such that the actuator accelerates the rotor or stator from the idle position in the direction of the working position during an acceleration phase and the actuator is then braked in a braking phase, the rotor or stator being released from the actuator in the braking phase and moves into the working position due to the acceleration imparted during the acceleration phase. A maintenance-free, non-contact, lubricant-free bearing can be achieved which is particularly efficiently operable by use of a permanent magnet bearing the stator section of which can be operated by a piezoelectric accelerator.

Description

 Warehouse and method for operating a warehouse

description

The present invention relates to bearings for rotatably supporting a rotor on a stator, and more particularly to maintenance-free and lubricant-free bearings, which can be used particularly well for supporting a radial gear in a turbomachine of a heat engine, such as a heat pump or a refrigerating machine.

13 shows a known heat pump, as described in "Technical Thermodynamics", Theoretical Background and Practical Applications, 14th revised edition, Hanser Verlag, 2005, pages 278 - 279. The heat pump comprises a closed circuit, By means of a first heat exchanger 80 and the evaporator, so much heat is extracted from the soil or groundwater that the working fluid evaporates Compressor 81 compresses it, which increases the pressure and temperature, and this compression is carried out by a reciprocating compressor.The compressed and high-temperature working fluid now enters the second heat exchanger 82, the condenser. or hot water circuit extracted so much heat that the refrigerant under h ohem pressure and high temperature is liquefied. In throttling or expansion member 83, the working fluid is expanded, ie the working fluid is released. Here, the pressure and temperature are reduced to such an extent that the working fluid is again able to re-absorb energy from the soil or groundwater in the evaporator. The cycle is now closed and starts again. As can be seen, the working medium serves as an energy transporter to pick up heat from the soil or groundwater and deliver it to the heating circuit in the liquefier. In this process, the second law of thermodynamics is fulfilled, which states that heat or energy from "self" can only be transferred from the higher temperature level to the lower temperature level, and vice versa only by external energy supply, here the drive work of the compressor, can happen.

Fig. 12 shows a typical h, log p diagram (h is the enthalpy, p is the pressure of a substance). Between point 4 and point 1 in the diagram of Fig. 12 an isobaric evaporation of the working fluid at low values for the pressure and the temperature (Ti pχ _f) takes place, here the heat is supplied Qβi.

Between point 1 and point 2, ideally, a reversible compression of the working medium vapor in an adiabatic compressor to the pressure p ₂ takes place. The temperature rises to T ₂ . It is here to perform a compression work.

Then, at high pressure P ₂ , an isobaric cooling of the working medium vapor is first carried out from 2 to 2 '. Overheating is reduced. Subsequently, a liquefaction of the working fluid takes place. Overall, the heat Q _{25 can} be dissipated.

In the throttle 83 is then the adiabatic throttling of the working fluid from the high pressure P ₂ to the low pressure pi instead. In the process, part of the liquid working fluid evaporates and the temperature decreases to the evaporation temperature Ti. In the h, log p diagram, the energies and indices of this process can be determined by means of enthalpies are calculated and illustrated as shown in FIG.

The working fluid of the heat pump thus takes in the evaporator heat from the environment, ie air, water, sewage or soil, on. The condenser serves as a heat exchanger for heating a heating medium. The temperature Ti is slightly below the ambient temperature, the temperature T ₂ considerably, the temperature T ₂ 'slightly above the required heating temperature. The higher the required temperature difference, the more work the compressor has to apply. It is therefore desirable to keep the temperature increase as small as possible.

With reference to FIG. 12, in the ideal case, a compression of the working substance vapors along the curve for the entropy s = constant up to point 2 is thus carried out. From here to point 3 the working fluid liquefies. The length of the section 2-3 represents the useful heat Q. From the point 3 to the point 4, the relaxation takes place and from point 4 to point 1, the evaporation of the working substance, the distance 4-1 reproduces the heat extracted from the heat source. In contrast to the T, s diagram, the amounts of heat and work can be taken as stretches in the h, log p diagram. Pressure losses in valves, the pressure and suction lines, the compressor, etc. deform the ideal course of the cycle in the h, log p diagram and reduce the effectiveness of the entire process.

In piston compressors, the aspirated working vapor initially has a lower temperature than the cylinder wall of the compressor and absorbs heat from it. As the compression progresses, the temperature of the working-material vapor finally increases above that of the cylinder wall, so that the working-substance vapor gives off heat to the cylinder wall. Then, when the piston sucks in steam again and compressed, the temperature of the piston wall is initially again fallen below and then exceeded, resulting in permanent leads to losses. Furthermore, overheating of the sucked Arbeitsstoffdarapfes will be necessary and necessary so that the compressor does not suck a liquid agent. Another disadvantage is the heat exchange with the oil circuit of the reciprocating compressor, which is indispensable for lubrication.

Occurring irreversibilities, such as heat losses during compression, pressure losses in the valves and flow losses in the pressure line for liquefying and in the condenser increase the entropy, ie the heat that can not be recovered. Furthermore, the temperature T _{2 is also} above the liquefaction temperature. Such an "overheat enthalpy" is undesirable, especially because the high temperatures involved accelerate the aging of the compressor and, in particular, the lubricating oil in a piston compressor, and also reduce the effectiveness of the process.

The liquefied low-temperature working fluid at the condenser outlet would have to be relieved by an ideal cycle through an engine, such as a turbine, to utilize the excess of energy that existed against the temperature and pre-compression conditions. For reasons of this required large expense, this measure is omitted and the pressure of the working fluid is abruptly reduced by the throttle 83 to the low pressure and the low temperature. The enthalpy of the working substance remains approximately the same. Due to the sudden reduction in pressure, the working fluid must partially evaporate in order to lower its temperature. The necessary heat of evaporation comes from the excess temperature located working fluid, so it is not removed from the heat source. The totality of the losses caused by the expansion in the throttle 83 (FIG. 13) is referred to as relaxation losses. These are exergy losses because heat of temperature T is converted into heat of temperature T ₀ . delt is. These losses can be reduced if the liquid agent can deliver its heat to a medium of a temperature less than T. This subcooling enthalpy can be exploited by an internal heat exchange, which, however, again requires additional equipment. In principle, the internal heat exchange is a limit, because in the compression of the vapors, the OberhitZungstemperatur T ₂ increases, whereby the profits achieved are partially compensated again, and machine and lubricating oil are subjected to thermal stress. Finally, the overheating increases the volume of the steam, which reduces the volumetric heat output. This heat is used only to preheat the vapors of the working fluid flowing to the compressor as far as necessary to ensure that all the droplets contained in the vapor of the working fluid have been converted to steam with certainty.

In general, the ratio of the enthalpy difference between point 1 and point 4 to the enthalpy difference between point 2 and point 1 of the h, log p diagram is a measure of the economics of the heat pump process ,

A currently popular tool is R134a, which has CF ₃ -CH ₂ F as its chemical formula. This is a working fluid that is no longer harmful to the ozone layer, but has a 1000 times greater effect than carbon dioxide in terms of the greenhouse effect. However, the Rl34a agent is popular because it has a relatively large enthalpy difference of about 150 kJ / kg.

Although this working fluid is no longer an "ozone killer", there are still considerable demands on the integrity of the heat pump cycle, such that no working fluid molecules escape from this closed circuit, since these would cause very considerable damage due to the greenhouse effect. This encapsulation causes Considerably additional costs when building a heat pump. In addition, chemical lubricants are required, which can also lead to an environmental impact.

Furthermore, it can be assumed that until the implementation of the next stage of the Kyoto Protocol due to the greenhouse effect by 2015, Rl34a will also be banned, which has already been done with much more harmful means.

A disadvantage of existing heat pumps, therefore, apart from the fact of the harmful lubricants and the harmful working fluid and the fact that due to the many losses in the heat pump cycle, the efficiency of the heat pump is typically not more than a factor of 3. In other words, about 2 times the energy used for the compressor can be taken from the heat source, such as groundwater or soil. Considering now heat pumps in which the compressor is driven by electric current, and considering at the same time that the efficiency in power generation is perhaps equal to 40%, it turns out that - in terms of the overall energy balance - a heat pump the benefit is doubtful. Based on the primary energy source, 120% - 3 "40% of heating energy is provided, and a conventional heating system with a burner achieves efficiencies of 90 - 95%, ie with a high technical and thus financial expense only an improvement of 25 - 30% achieved.

Better systems use primary energy to drive the compressor. So gas or oil is burnt to create the compressor power with the energy released by the combustion. An advantage of this solution is that the energy balance is actually more positive. This is due to the fact that, although only about 30% of the primary energy carrier can be obtained as drive energy, but for the waste heat of then about 70% can be used with the heating. The heating energy provided is then 160% = 3 ^• 30% + 70% of the primary energy source. A disadvantage of this solution, however, is that a household, although he has no more classical heating, still needs an internal combustion engine and a fuel storage. The cost of engine and fuel storage are still added to the cost of the heat pump, which is indeed a high-closed circuit due to the climate-damaging coolant.

All of these things have meant that heat pumps can only compete to a certain extent in competition with other types of heating.

Conventional heat pumps have the disadvantage that, when operating on the basis of a piston principle, they have inherent losses due to the fact that the piston has to be reciprocated. In addition, the piston must be lubricated, and it is difficult or impossible to prevent the lubricant from coming into contact with the working fluid.

It has been found that the replacement of a piston pump with a centrifugal compressor operating, for example, on the turbo principle has some advantages. A decisive advantage of such a principle is that it is possible to use water as the working medium, and this water can even come from an open circuit. Regardless of whether actually a completely closed or fully open circuit is used, no contamination of the water vapor or of the water with lubricants should take place in such a steam atmosphere. On the other hand, if a radial compressor is used in a heat engine, the efficiency should be as high as possible, so that for the storage of a radial wheel, the additional high speeds have as little energy as possible should be spent.

The Internet publication "Permanent Magnetic Bearings" by Johan K. Fremerey from November 2000 (which can be accessed via the link www.fz-juelich.de/zat/magnet/0b30.pdf) provides an overview of permanent magnetic bearings that are used for non-contact In particular, such bearings are wear and maintenance-free and such bearings are also free from friction energy losses.The complete elimination of the bearing friction losses occurs in fast moving systems only with largely exclusion of air friction, ie under vacuum conditions, wear However, even if the bearing is not operated in vacuum, the friction loss is much lower than with corresponding contact bearings, such as ball bearings, rolling bearings or other plain bearings.

In particular, permanent magnetic bearings are used with repulsive force effect at the bearing gap, as described in British Patent 642,353. Permanent magnetic bearings have both magnetic rotor-side and stator-side magnetic rings axially magnetized so that above the ring, e.g. is the North Pole and below the ring is the South Pole. Such rings are made of crystalline powder. The magnetic powder is pressed into a mold in the presence of a strong magnetic field. The crystals align with their preferred magnetization axis in the direction of the magnetic field. The pellets are then sintered. Suitable material is neodymium-iron-boron, (NdFeB), which has a high coercitive field strength. Such permanent magnet rings made of this material can be produced at comparatively low prices.

A repulsive bearing consists of several stacked, axially magnetized rings on the rotor and on the stator constructed with the magnetization direction of adjacent rings within each stack is opposite. The rotor and stator magnets facing each other at the same height at the position gap are magnetized in the same direction. As a result, surface flows with radially repelling effect are opposite to the bearing gap, so that the bearing is radially stable.

As long as the magnetic rings of the rotor and the stator are at the same level, the repulsive forces are balanced in the axial direction. When leaving this equilibrium position, however, the repulsive force becomes increasingly effective in the displacement direction. The bearing tends to shift from the unstable axial equilibrium position and the magnitude of the height of a single magnet ring in the axial direction. The deflected by a ring height position are oppositely magnetized rings on the bearing gap at the same height opposite. The repulsive bearing transforms into a radially active bearing with radial flow in the bearing gap and corresponding bearing instability. Such a breakout force results from a displacement of the rotor with respect to the stator For axial displacement, electrical stabilizing coils are used directly on the circumference of the stator magnets, whereby the coil current is controlled in such a way that it can move the rotor Surface current of the stator magnet surrounded by the coil and thus its tensile strength attenuates when the shaft magnet moves in the direction of this Statorπvagneten.On the opposite stator magnets mounted there coil simultaneously enhances the local surface current, so that its tensile force increases The current regulation is controlled by a sensor which continuously measures the axial position of the wave magnets r be used. she are charged for this purpose with a high-frequency current. Induced short-circuit currents at the ends of the waves influence the high-frequency current in the coils. As the shaft approaches, the current in the coil in question increases. The currents of both coils are compared in a differential circuit. The difference signal is rectified and serves as a sensor signal whose polarity and amplitude clearly reflect the direction and magnitude of the axial shaft deflection. Such permanent magnetic bearings are used for rotary anode storage in an X-ray tube, for a spinning centrifuge, for a gas friction vacuum gauge, for a disk chopper for neutron beams, for a turbomolecular pump, for an X-ray pulse selector, for a drum chopper for neutron beams, for a flywheel Energy storage or used within a crystal pulling system.

The problem with such bearings or with other frictionless bearings is that the bearing, without mechanical stress and thus the risk of damage to produce, must be returned from a rest position to a working position and. Friction-free or friction-reduced non-contact bearings are characterized namely by the fact that the bearing has a rest position in which the rotor and the stator touch, and in which the rotor is not or only slightly rotates with respect to the stator, and that they have a working position, in which the rotor and the stator do not touch, and in which the rotor can rotate relative to the stator without contact. When switching on such a bearing, it is therefore important that the rotor and the stator come into their respective working position, namely from a rest position, wherein when the rotor or the working device in which the rotor is arranged, switched off the rotor comes to a parking position or rest position without mechanical damage occurring during this position change. On the other hand, this position change may consume only little energy and must be initiated by simple measures so that the rotor can always be turned off when there is no need for the rotor rotation, as is the case in particular when the rotor for energy production or energy conversion is used for example in a heat engine.

The object of the present invention is to provide an improved bearing and concept for operating a bearing that is efficient and economical.

This object is achieved by a bearing according to claim 1 or a method according to claim 22.

The present invention is based on the finding that, for a bearing which has a rest position in which the rotor and the stator touch, and which has a working position in which the rotor and the stator do not touch, an actuator for Moving the rotor from the rest position to the working position in response to a control signal is used. This control signal is delivered such that actuation of the actuator with the control signal causes the actuator to first accelerate the rotor or stator from the rest position towards the working position in an acceleration phase and then decelerate the actuator in a deceleration phase. According to the invention, however, the rotor is not braked by the deceleration of the actuator, so that the rotor detaches due to its mass from contact with the actuator or the stator and continues its movement generated due to the acceleration. According to the invention, the control device is designed to provide a control signal such that the rotor or the stator are driven so strongly that the rotor or the stator are brought into their working position relative to one another. In other words, the rotor or the stator is thus "pushed" in order to cover a large part of the distance from the rest position to the working position, so to speak, in free flight According to the invention, the change from the rest position to the working position and also the reverse change becomes dynamic completed, the rotor or stator is thus not "driven" slowly into its working position, but is sufficiently accelerated to fly independently due to the applied acceleration in the working position.

A significant advantage of the present invention is that the actuator can thus be implemented much smaller and thus considerably less expensive with regard to its actuating stroke. The actuator does not have to have such a large stroke to overcome the entire distance between the rest position and the working position. Instead, the actuator only needs to have a small stroke, but must be able to achieve very high acceleration in the small stroke. In the case of actuators available on the market, however, the achievement of a high acceleration, in particular in the direction of propagation, is considerably less expensive than that of providing a large stroke length. Preference is given in particular to piezoelectric actuators which are distinguished by a high speed that can be generated, but which at the same time are distinguished by the fact that the price of such actuators increases considerably as the maximum distance increases.

In a preferred embodiment of the present invention, a bearing is used as the bearing, which is radially stable to accommodate large radial forces due to high rotational speeds of rotors, but which is axially metastable and requires an active control. In such a bearing with an axially considered metastable state, the actuator used according to the invention for "throwing" the rotor or stator can be used be used simultaneously for active control of the axial position of the rotor and / or stator, since the stroke, which is required for an active control, is typically much smaller than the distance that the rotor and / or stator must move relative to each other in such camps, to get from the rest position to the working position. In such a preferred embodiment with active axial control and for the use of the mechanical actuator, this actuator is thus both when starting, so then when it is transitioned from the rest position to the parking position, as well as in operation for active control. At the same time, when the bearing moves from the working position to the rest position, the actuator is reused to come into mechanical contact with the rotor after an acceleration phase either at the beginning or during the braking phase, ie to catch the rotor then, at the end of the deceleration phase, safely guide the rotor to the rest position.

The inventive dynamic bearing operation could thus be compared to a juggler throwing a plate into the air. When the plate is in the hands of the juggler, the plate is still in the resting position. If the plate is now thrown up, the plate starts from the rest position into the working position. If the juggler then holds the plate in the air with a pole, this corresponds to an active control of the bearing during operation. Then, at the end of the performance, the juggler will drop the plate and not just let it crash to the ground, but catch it with his hand to safely guide the plate to its resting position.

In a preferred embodiment of the present invention, the rotor does not move in the park position. The rotor rests on the stator. The rotor is therefore in a lossless rest or "parking position." However, in the working position, the rotor rotates, and if the working position is metastable, the remaining free space is lost. level is stabilized by mechanical displacement of the rotor relative to the stator or the stator relative to the rotor. In particular, if permanent magnetic bearings are used, only such a shift will cause losses, while the entire "bearing work" is provided lossless.

In a preferred embodiment of the present invention, one side of an axle is radially stably supported by a fixed permanent magnet (rotor and stator), and the opposite side of the axle is provided with a permanent magnet rotor or stator with the stator mechanically or axially displaced can be.

A particularly preferred location of use of the bearing according to the invention is within a turbocompressor arrangement of a heat engine, such as a heat pump or a refrigerating machine. Normal centrifugal compressors have z. B. ball bearings. As has been stated, electromagnetic bearings which must be continuously supplied with electric power in order to maintain the magnetic field are also known. In contrast to these bearings, the bearing of the invention requires, especially when it is designed as a permanent magnet bearing, no permanent electrical power. Even permanent magnets require permanent electromagnetic power when they are provided with coils for damping or amplifying circular currents. The mechanical actuator arrangement according to the invention, however, only consumes power, if necessary, for a change from the rest position to the working position and vice versa, if necessary for active regulation and only when a position change is carried out. On the other hand, if the bearing is in equilibrium, the operator will also be at rest, and the bearing will require no power during operation. Preferred embodiments of the present invention will be explained below in detail with reference to the accompanying drawings. Show it:

1a is a schematic representation of a bearing according to the invention according to an embodiment of the present invention, wherein the bearing is in its working position;

Fig. Ib is a representation of the bearing, wherein the bearing is in its rest position;

Fig. 2a is a time-lapse diagram of the distance traveled by the actuator or stator over time;

Fig. 2b is a velocity-time diagram of the movement of the stator or actuator corresponding to Fig. 2a;

Fig. 2c is an acceleration-time diagram of the acceleration exerted by the actuator;

Fig. 2d is a schematic representation of the distance covered by the rotor;

Fig. 3 is a schematic representation of a preferred embodiment of the present invention with permanent magnet bearing and piezo actuator;

Fig. 4 is a schematic representation of the active control performed during operation of the preferred bearing;

5 shows a schematic representation of the bearing according to the invention for supporting a radial wheel in a turbomachine of a heat pump; 6a is a schematic block diagram of the heat pump;

FIG. 6b shows a table for illustrating various pressures and the evaporation temperatures associated with these pressures; FIG.

Fig. 7 is a block diagram of a preferred embodiment of the heat pump operated with groundwater, seawater, river water, seawater or brine;

Fig. 8a shows an alternative embodiment of the condenser of Fig. 7;

FIG. 8b shows an alternative embodiment of the condenser with reduced return in off mode; FIG.

Fig. 8c is a schematic representation of the condenser with a gas separator;

Fig. 9a shows a preferred implementation of the evaporator of Fig. 7;

9b shows an alternative embodiment of the evaporator with the use of the liquefier effluent as boiling aid;

FIG. 9c shows an alternative embodiment of the evaporator with a heat exchanger for using groundwater for boiling assistance; FIG.

Fig. 9d an alternative embodiment of the evaporator with feed from the side and drain in the middle;

9e is a schematic representation of the expander with indication of preferred dimensions; 10a shows an alternative implementation of the evaporator for reducing the height of the riser pipe;

10b shows an implementation of an alternative realization of a connection of a heating line to the condenser with a turbine / pump combination;

Fig. IIa is a schematic representation of the compressor executed by a plurality of sequentially arranged turbomachines;

FIG. IIb shows a schematic representation of the setting of the rotational speeds of two cascaded flow machines as a function of the setpoint temperature; FIG.

Fig. Llc is a schematic plan view of a radial wheel of a turbomachine according to a preferred embodiment of the present invention;

Fig. Hd is a schematic cross-sectional view with a merely schematic representation of the Radialrad blades to illustrate the different extension of the blades with respect to the radius of the Radialrads;

Fig. 12 is an exemplary h, log p diagram; and

Fig. 13 shows a known heat pump, which performs the left-handed circuit of FIG. 12.

1 a shows a bearing for rotatably supporting a rotor 500 on a stator 502, the bearing having a rest position, which is shown in FIG. 1 b, in which the rotor and the stator touch, and wherein the bearing also has a rest position Working position, as shown in Fig. Ia, in which the rotor and the stator do not touch. Instead, a gap 504 exists in this working position between rotor and stator, which is shown in Fig. Ia in the axial direction with respect to a rotation axis 506. In the exemplary embodiment shown in FIG. 1a, the pivot bearing gap 508 is arranged parallel to the axis of rotation 506 and, depending on the design of the rest position, can also exist in the rest position.

In a preferred embodiment, the rotor is located on the stator on a specially provided support surface 510, which may be specially designed to accommodate mechanical loads well. However, no special precautions are necessary for these bearing surfaces 510, since the inventive operation of the mechanical actuator 512 ensures in any case that the rotor and / stator are carefully "thrown off" and carefully "caught" again.

Although in Fig. Ia and Ib the difference between the working position and the rest position is drawn so that the rotor moves relative to the stator in the axial direction to come to the working position or to come to the rest position, the present invention equally applicable to a situation in which the stator is moved in the axial direction with respect to the rotor. Furthermore, there are embodiments in which both the rotor and the stator are at least partially moved in different directions, so that results in the working position of the gap 504 between the rotor and stator, which is not present in the rest position, since the rotor on the Stator rests. For the following description, however, it is preferable, on account of the simpler illustration, to represent the working position so that the rotor is moved to the working position.

The bearing according to the invention in Fig. Ia further comprises the mechanical actuator 512 for moving the rotor of the

Rest position in the working position in response to a

Control signal 514, by a controller 516 is delivered. In particular, the controller 516 is configured to provide a control signal to the actuator that causes the actuator to accelerate the rotor or stator from the rest position toward the working position in an acceleration phase, and further causes the actuator 512 to brake in phase is slowed down. By braking the actuator is achieved that the rotor or stator detaches due to its mass of a contact with the actuator or possibly the stator and continues its movement. In particular, the control device 516 is designed to operate the actuator so that the rotor or stator are driven so strongly that the rotor or stator reaches the working position.

For this purpose, there are various possibilities that can also be combined with each other. In the embodiment shown in FIG. 1a, the actuator is provided with two "outlets" 520 and 522 which are intended to schematically indicate that the mechanical actuator engages the rotor when the alternative of FIG. 520 occurs or engages the stator is provided if the alternative 522 is present, Alternatively, a mixed solution may be provided in which the actuator is in mechanical engagement with both the rotor and the stator.

However, in a preferred embodiment of the present invention, it is preferred that the mechanical actuator implement only the alternative indicated by the "exit" 522 such that the mechanical actuator is actuated by mechanical actuation of the stator and immediate transmission of force via the seating surface 510 (FIG. Fig. Ib) the rotor 500 quasi "accelerated" indirectly. Then, as the mechanical actuator 512 transitions to the deceleration phase, the stator 502, which is mechanically coupled to the mechanical actuator, is also decelerated while the rotor 500 now detaches from the support surface 510 of the stator 502 and continues to fly the tor by the mechanical actuator 512 braked again and returned to its rest position.

The text below discusses the movement or speed or acceleration of the actuator and possibly stator and the rotor, exemplary numerical values being mentioned. The actuator is supplied with a control signal such that it is accelerated from a start deflection value at 0 μm until it has a maximum velocity at a solution point 530, as shown in Fig. 2b. The speed shown by way of example is 250 μm / ms. Then, after the so-called "solution point", the actuator and the stator preferably connected to the actuator are decelerated, which is reflected in the inflection point 530 of the motion-time diagram of Fig. 2a When the maximum displacement of 20 μm of an actuator is reached, the speed decreases and the actuator "contracts again" until it has a maximum negative speed of, for example, -500 μm / ms , wherein this maximum speed is then brought back to 0, when the actuator is then again "retracted" completely after approximately 400 μs. In the acceleration-time diagram of Fig. 2c, an acceleration phase thus results initially with a positive acceleration of approximately 2000μm / ms ^2. In the demarcation phase 534 delimiting to the acceleration phase 532, a negative acceleration v on -10,000 μm / ms ² , for example, to achieve fast braking. In order for the actuator to end up safely in its rest position, a further smaller acceleration phase 536 takes place after the deceleration phase 534, which is designated "2a" in FIGURE 1a., However, the acceleration phase 536 can also be omitted. Then, however, the negative acceleration would have to be changed from the course in FIG. 2c, for example by transition to an acceleration course which is no longer constant over the Time is and varies only between different states.

It should be noted that the course shown in FIGS. 2a to 2c is the same as the course in FIG. 2d. However, it can be seen from the progressions that it is preferred to decelerate more than to accelerate. This is also readily possible since in the acceleration phase both the stator and the rotor must be accelerated, while in the braking phase only the stator or preferably only a small part of the stator, namely the bearing section of the stator and not the stator Rotor must be braked because the rotor has detached from the stator. It is preferred that more than 50% of the maximum stroke of the actuator be used for acceleration, with values in the range of 65 to 90% being particularly preferred. It also follows that the negative acceleration in the deceleration phase 534 is greater in magnitude than the positive acceleration in the acceleration phase 532. Relationships between positive acceleration in the acceleration phase and negative acceleration in the deceleration phase of 2 to 50 are preferred, wherein in Fig. 2c by way of example a ratio of 5 is drawn. In general, this ratio depends on the mass ratio of the mass to be accelerated to the mass to be braked.

Furthermore, it is preferable to operate the second acceleration phase 536 with an even higher positive acceleration than was used for the braking phase in absolute terms, so that the largest possible percentage of the actuator's stroke length can be utilized for acceleration, without the actuator in the deceleration phase overloaded. In particular, piezoelectric actuators are preferably designed so that in each case it is ensured that they do not suffer damage in the deceleration phase. On the other hand, piezoelectric actuators have the advantage that in the acceleration phase in the direction in which they can bear the highest load. In the acceleration phase, the acceleration is highest because the actuator must also accelerate the rotor weight. It is therefore preferred to carry out the acceleration phase with an actuator having a piezoelectric construction, by causing an extension of the piezoelectric actuator in the acceleration phase. The actuator is thus preferably claimed in the acceleration phase to pressure and not to train. Preferably, such mechanical actuators are used, which have an operating distance less than 100 microns. Furthermore, speeds are preferred which reach values of between 50 and 2000 μm / ms at the end of the acceleration phase and which are at the end of the deceleration phase at -200 to 4000 μm / ms.

FIG. 2 d shows the movement course of the rotor, which has to move over a width of the gap 504 which is> 300 μm. By comparison, the actuator deflection is indicated at 540. At 540, there is thus a compressed representation of the course of movement of FIG. 2a, wherein attention must be paid to the changed time scale.

At time 530, the rotor disengages from the stator and moves into free motion. Due to the acceleration originally imposed on the rotor, the rotor then continues to fly between time 530 and a time of about 3.5 ms until it is decelerated due to frictional forces or the inherent magnetic bearing force. Furthermore, in a preferred embodiment of the present invention, when the rotor enters the control region 542, the actuator is actively controlled by the actuator, with the duration of the active control shown at 544 in FIG. 2d. From the sketch in FIG. 2 d, it can be seen that the rotor travels most of the gap 504 in free movement, that is to say without contact with the stator or with the actuator. As will be described below with reference to FIG. 3, the amount of the gap, eg the region of 300 μm, is equal to the height of a magnet ring of a repulsive magnetic bearing, as shown in FIG. According to the invention it is preferred that the gap is greater than the maximum stroke of the actuator and in particular at least two to 20 times as large. Thus, due to the fact that the actuator does not have to have the entire gap width as the maximum stroke, a particularly inexpensive operator can be used.

Especially with piezoelectric actuators, the maximum deflection stroke is proportional to the height of the actuator. For piezoelectric actuators, the ratio between deflection and stroke is at e.g. 0.14 percent. This means that an actuator having a stroke of e.g. 300 microns, in order to overcome in Figure 2d without free movement of the gap 504, would have to have a height which is 215 mm. However, to cover only a height of 20 microns according to the present invention, an actuator with a height of 14 mm is sufficient. From a price point of view, a piezo actuator with a stroke of 20 μm would cost about 20 euros, while a piezo actuator for a stroke of 300 μm would cost well over 500 euros.

Referring now to Fig. 3, a preferred embodiment of the present invention will be described, in which a permanent magnet bearing with magnetized stacked rings is employed. Stator 502 has a number of magnetic rings 560a, 560b, 560c, 56Od, 56Oe and 56Of attached thereto. Also, magnetic rings 562a, 562b, ..., 562e are also attached to the rotor. Furthermore, the controller 516 is coupled on the input side to a control device 517, which is also shown in Fig. Ia and performs a detection of the gap 504 in any known manner. In particular, capacitive, inductive, optical and in any other way either directly or indirectly via a position determination of the rotor and / or the stator detects a gap length of the gap 504 and is used as the actual value in the gap detection device 517 or the regulator 517.

There, the actual value is compared with a desired value in order to provide, depending on the target-actual comparison, a control signal which is filtered by any known loop filter having a desired frequency characteristic. This control signal is fed to the controller 516, which then supplies the input signal or control signal 514 for the mechanical actuator 512. In response to the control signal 514, the actuator 512 thus causes a displacement of the stator or the magnetic rings attached to the stator or, in general, of the bearing section coupled to the stator such that the relative position between the bearing section of the stator and the bearing section of the stator Rotor is changed.

The controller 511 is for detecting an actual position of the rotor with respect to the stator and / or an actual position change of the rotor with respect to the stator to obtain an actual value. Further, the controller is for moving the rotor or stator so that the actual position of the rotor with respect to the stator approaches a target position of the rotor relative to the stator at the working position, when the actual position has been detected, or that Actual position change at least approaches zero or becomes zero when the actual position change is detected. Measured in this alternative is the change in direction or whether the change is positive, negative or zero, and the position bearing magnet stator to rotor is corrected until the change in direction is approximately or exactly equal to zero

This active control of the magnetic bearing takes place in the working position or in the operating mode of the rotor and will be explained with reference to FIG. 4. Fig. 4 shows a cross section through three stacked each other Magnetic rings whose magnetization is indicated by directional arrows 564, wherein the arrow direction of the arrow 564 indicates whether up or down is the north pole or south pole of the magnetised ring. After oppositely polarized rings attract, there is an upper force F ₀ and a lower attraction F _u , as indicated in FIG. Then, when an axial force is applied to the rotor, that is, when the rotor moves upward, for example, the force F ₀ becomes larger and the force F _u becomes smaller, so that the deflection is further enhanced by the inherent magnetic force inherent in the magnetic bearing why the bearing is axially metastable. To compensate for this situation, the stator is controlled so that it is also deflected upward, so that the force F _{0 is} again smaller and the force F _{u is} greater, so that again a balance of power is set. Corresponding variables and tendencies associated with each other are shown in the table drawn in FIG. It follows that the stator in the control region, which is schematically drawn in Fig. 2d at 544, is constantly moved slightly upwards or downwards to compensate for acting on the rotor axial forces or to keep the bearing stable.

Fig. 5 shows a preferred embodiment of the present invention for supporting a radial impeller for a turbocompressor, as e.g. in the compressor 16 of Fig. 7 or in a turbomachine 172 or 174 of Fig. IIA is used, and as described in more detail in Fig. HC and Fig. HD.

In particular, a rotor shaft 570 is arranged on the rotor 500, which extends through a bore 572 in the stator. The bearing again comprises a stator bearing section 560 as well as a rotor bearing section 562, with identically polarized magnet rings facing each other in the two bearing sections when the bearing is in the optimally adjusted bearing position or when no bearings are present. ne axial force on the rotor 500 acts. Also shown in FIG. 5, the actuator 512 is arranged to mechanically couple the stator section 560 to the rest of the stator. This has the advantage that the mass which the accelerator - apart from the rotor mass - has to accelerate and, in particular, also decelerate again, is as small as possible.

Further, in Fig. 5, the pressure distribution is indicated when the radial wheel in the heat pump of Fig. 6A operates. On the upper side is then a small vapor pressure, while on the lower side there is a large vapor pressure. This means that the radial wheel experiences an axial force acting upwards to the small pressure, so that due to the operation of the radial wheel, a rotor deflection takes place up here, which is compensated by a Statorregelung up.

The axial thrust is thus directed upwards. However, this axial thrust occurs relatively slowly, as a result of which a typical piezo ring 512 or a piezo actuator in the form of a plurality of radially distributed rods can easily keep up. The change in force thus has a lower slope than the Aktuaktor can endure by working on train or pressure. As the actuator, a piezoelectric actuator having a disk-like structure and which is typically supplied with a controlled DC voltage and having a maximum voltage of e.g. 2kV / mm can withstand.

As already stated, a return from the working position to the rest position takes place in that, in order to bring about a "downward movement" of the rotor, the stator is deflected to bring the bearing into imbalance, then the rotor moves down to a certain extent from right to left on the movement curve in FIG. 2d, whereby the magnetic bearing forces and, in particular, the ever-increasing F ₀ and the im- The decreasing F ₀ of FIG. 4 accelerates the movement. Then, just before a "collapsing" impact of the rotor on the support surface, the actuator is deflected in an acceleration phase to then be decelerated Preferably, after a lapse of a small period of time after the beginning of the braking phase, the rotor is then engaged with the stator at the support surface Touch are thus collected, this "hitting" is mechanically unproblematic, since at the time of impact, the rotor and the stator due to the drive by the actuator move almost equally fast or the stator moves only slightly slower than the rotor. The motion conditions and in particular the deceleration in the braking phase are adjusted so that the rotor is then safely landed safely, without destructive mechanical stresses, which could, however, impact when the rotor impinges unrestrained on the bearing surface of Fig. Ia would.

Hereinafter, a preferred use of the bearing according to the invention within a heat pump and in particular within a turbo compressor of a heat pump for storing a radial wheel, where the advantages of the lubricant-free bearing particularly come to light, since the preferred heat pump concept represents an open circuit in which the working fluid Water is taken from the environment and then released into the environment again.

FIG. 6 a shows a heat pump according to the invention which initially has a water evaporator 10 for evaporating water as the working fluid in order to produce a steam in a working steam line 12 on the output side. The evaporator comprises an evaporation space (not shown in FIG. 6a) and is designed to generate an evaporation pressure of less than 20 hPa in the evaporation space, so that the water evaporates at temperatures below 15 ^° C. in the evaporation space. The water is preferably groundwater water, in the ground free or in collector pipes circulating brine, so water with a certain salinity, river water, seawater or seawater. According to the invention, all types of water, ie, calcareous water, lime-free water, saline water, or salt-free water, are usefully used. This is because all types of water, that is all these "hydrogens", have the favorable water property, namely that water, also known as "R 718", is an enthalpy difference useful for the heat pump process Has ratio of 6, which is more than 2 times the typical usable enthalpy difference ratio of z. B. R134a corresponds.

The steam is fed through the suction line 12 to a compressor / condenser system 14, which is a turbomachine such. B. has a radial compressor, for example in the form of a turbocompressor, which is designated in Fig. 6a with 16. The turbomachine is designed to compress the working steam to a vapor pressure at least greater than 25 hPa. 25 hPa corresponds to a liquefaction temperature of about 22 ⁰ C, which may already be a sufficient heating flow temperature of a floor heating, at least on relatively warm days. In order to generate higher flow temperatures, pressures greater than 30 hPa can be generated with the turbomachine 16, wherein a pressure of 30 hPa has a liquefaction temperature of 24 ⁰ C, a pressure of 60 hPa has a liquefaction temperature of 36 ⁰ C, and a pressure of 100 hPa corresponds to a liquefaction temperature of 45 ° C. Underfloor heating systems are designed to heat sufficiently with a flow temperature of 45 ^° C, even on very cold days.

The turbomachine is coupled to a condenser 18, which is designed to liquefy the compressed working steam. By liquefying, the energy contained in the working steam is supplied to the condenser 18, in order then to be fed to a heating system via the supply line 20 a. to be led. The working fluid flows back into the condenser via the return line 20b.

According to the invention, it is preferable to extract from the high-energy water vapor directly through the colder heating water the heat (energy) which is taken up by the heating water so that it heats up. The steam is so much energy withdrawn that this is liquefied and also participates in the heating circuit.

Thus, a material entry into the condenser or the heating system takes place, which is regulated by a drain 22, such that the condenser in its condenser has a water level, which always remains below a maximum level despite the constant supply of water vapor and thus condensate ,

As has already been stated, it is preferred to take an open circuit, ie to evaporate the water, which is the heat source, directly without a heat exchanger. Alternatively, however, the water to be evaporated could first be heated by a heat exchanger from an external heat source. However, it should be remembered that this heat exchanger again means losses and equipment expense.

Moreover, in order to avoid losses for the second heat exchanger, which has hitherto necessarily been present on the condenser side, it is also preferred to use the medium directly there as well, that is, when thinking of a house with underfloor heating, the water which from the evaporator comes to circulate directly in the underfloor heating.

Alternatively, however, it is also possible to arrange a heat exchanger on the condenser side, which is supplied with the supply line 20a and which has the return line 20b, this heat exchanger dissipating the water present in the condenser. cools and thus heats a separate underfloor heating fluid, which will typically be water.

Due to the fact that water is used as the working medium and due to the fact that only the vaporized portion of the groundwater is fed into the turbomachine, the degree of purity of the water does not matter. The turbomachine, as well as the condenser and possibly directly coupled underfloor heating always supplied with distilled water, so that the system has a reduced maintenance compared to today's systems. In other words, the system is self-cleaning, since the system is always fed only distilled water and the water in the drain 22 is thus not polluted.

The distilled water discharged through the drain can thus - if no other regulations stand in the way - be easily returned to the groundwater. Alternatively, however, it can also be used here, for example. B. in the garden or in an open space to be seeped, or it can be supplied via the channel, if regulations dictate - a sewage treatment plant.

The combination according to the invention of water as a working medium with the 2 times better usable enthalpy difference ratio compared to R134a and because of the reduced requirements for the closed system (rather an open system is preferred), and because of the use the turbomachine, through which the required compression factors are achieved efficiently and without purity impairments, an efficient and environmentally neutral heat pump process is created, which becomes even more efficient when the liquefier in the liquefier directly liquefied, since then no heat exchanger is needed throughout the heat pump process , In addition, all associated with the piston compression losses fall away. In addition, the very low water losses which would otherwise be incurred in the throttling can be used to improve the evaporation process, since the effluent water at the effluent temperature, which will typically be higher than the groundwater temperature, will be used to advantage Evaporator by means of structuring 206 of a drain pipe 204, as will be explained in Fig. 9a, to trigger a bubble evaporation, so that the evaporation efficiency is increased.

Hereinafter, referring to FIG. 7, a preferred embodiment of the present invention will be explained in detail. The water evaporator comprises an evaporation chamber 100 and a riser 102, in which groundwater from a groundwater reservoir 104 in the direction of arrow 106 moves up into the evaporation chamber 100. The riser 102 opens into an expander 108, which is designed to widen the relatively narrow tube cross section in order to create the largest possible evaporation surface. The expander 108 will be funnel-shaped, that is to say in the form of a paraboloid of revolution of any shape. It can have round or angular transitions. All that matters is that the diameter facing the evaporation chamber 100 or the area facing the evaporation chamber 100 is greater than the cross-sectional area of the riser pipe in order to improve the evaporation process. Assuming that about 1 1 per second flows through the riser up into the evaporation chamber, about 4 ml per second are evaporated in the evaporator at a heating power of about 10 kW. The remainder, cooled by about 2.5 ° C., passes over the expander 108 and lands in a collection sump 110 in the evaporation chamber. The collecting sump 110 has a drain 112, in which the amount of 1 1 per second less the evaporated 4 ml per second is discharged again, preferably back into the groundwater reservoir 104. For this purpose, preferably a pump 114 and a valve for overflow control intended. It should be noted that nothing has to be pumped actively here, since due to gravity, when the pump or the valve 114 is opened, water flows from the evaporator catch basin 110 via a return pipe 113 down into the groundwater reservoir. The pump or the valve 114 thus ensure that the water level in the catch basin does not rise too high or that no water vapor penetrates into the drain pipe 112 or that the evaporation chamber is reliably decoupled from the situation at the "lower" end of the return pipe 113.

The riser is arranged in a riser 116, which is filled by a preferably provided pump 118 with water. The levels in 116 and 108 are interconnected according to the communicating tube principle, with gravity and the different pressures in 116 and 108 providing water transport from 116 to 108. The water level in the riser tank 116 is preferably arranged so that even at different air pressures, the level never falls below the inlet of the riser 102, so that the ingress of air is avoided.

Preferably, the evaporator 10 includes a gas separator configured to receive at least a portion, e.g. For example, at least 50% of a gas dissolved in the water to be evaporated is to be removed from the water to be evaporated, so that the removed part of the gas is not sucked from the compressor via the evaporation space. Preferably, the gas separator is arranged to supply the remainder of the gas to a non-evaporated water so that the gas is carried away from the non-evaporated water. Dissolved gases may include oxygen, carbon dioxide, nitrogen, etc. These gases usually evaporate at a higher pressure than water, so that the gas separator can be arranged below the expander 108, so that oxygen vaporized in the gas separator etc. escapes from the water which is not yet evaporating and is preferably fed into the return line 113. The Feed is preferably at the point of the return line 113, at which the pressure is so low that the gas is taken back by the returning water into the groundwater. Alternatively, however, the separated gas can also be collected and disposed of at certain intervals or vented on an ongoing basis, ie discharged to the atmosphere.

Typically, the groundwater, seawater, river water, seawater, brine or any other naturally occurring aqueous solution will have a temperature between 8 ⁰ C and 12 ⁰ C. By lowering the temperature of 1 1 of water to I ⁰ C, a power of 4.2 kW can be generated. If the water is cooled by 2.5 ⁰ C, then a power of 10.5 kW is generated. Preferably, the riser is flowed through by a stream of water with a current in dependence on the heating power, in the example one liter per second.

When the heat pump is operating at a relatively high load, the evaporator will vaporize about 6 ml per second, which corresponds to a vapor volume of about 1.2 cubic meters per second. Depending on the required heating water temperature, the turbomachine is controlled with regard to its compaction performance. If a heating flow temperature of 45 ° C is desired, which is by far sufficient even for extremely cold days, then the turbomachine must increase the steam produced at perhaps 10 hPa to a pressure of 100 hPa. In contrast, a flow temperature of z. B. 25 ° for underfloor heating, so only by a factor of 3 must be compressed by the turbomachine.

The power generated is therefore determined by the compressor power, so on the one hand by the compression factor, ie how much compressed the compressor, and on the other by the volume flow generated by the compressor. If the volume flow increases, the evaporator must evaporate more, with the pump 118 adding more groundwater to the riser pipe. transported basin 116 so that the evaporation chamber is fed more groundwater. If, on the other hand, the turbomachine delivers a lower compression factor, less groundwater flows from the bottom to the top.

It should be noted, however, that it is preferred to control the flow of groundwater through the pump 118. According to the principle of communicating tubes, the level in the container 116 or the delivery rate of the pump 118 determines the flow through the riser. Thus, an increase in efficiency of the system can be achieved because the control of the flow is decoupled from the suction power of the turbomachine.

No pump is needed to pump the groundwater from below into the vaporization chamber 100. Instead, this is done by "self." This automatic upgrade to the evacuated vaporization chamber also helps to easily reach the 20 hPa negative pressure, without the need for evacuation pumps or anything else, and instead just a riser with a height greater than 9 The necessary negative pressure can, however, also be produced with a significantly shorter riser, if, for example, the implementation of Fig. 10a is used. The implementation of the high pressure on the negative pressure is effected via a turbine 150, wherein the turbine thereby extracts energy from the working medium.At the same time, the negative pressure on the return side is brought back into the high pressure, with the energy required for this by a Pump 152. Pump 152 and turbine 150 are via a force coupling g 154 coupled so that the turbine drives the pump, with the energy that has removed the turbine from the medium. A motor 156 is still needed only to compensate for the losses that the system has of course, and to achieve the circulation, so a system from its rest position to the dynamic mode shown in Fig. 10a.

In the preferred embodiment, the fluid machine is configured as a rotary compressor with a rotatable wheel, where the wheel may be a low-speed radial, medium-radial, semi-axial, or propeller, as known in the art. Radial compressors are described in "Turbomachines", C. Pfleiderer, H. Petermann, Springer-Verlag, 2005, pages 82 and 83. Such radial compressors thus include as a rotatable wheel the so-called. Center runner whose form depends on the individual requirements. In general, any turbomachines can be used, as they are known as turbine compressors, fans, blowers or turbocompressors.

In the preferred embodiment of the present invention, the radial compressor 16 is designed as a plurality of independent turbomachines, which can be controlled independently of each other at least in terms of their speed, so that two turbomachines can have different speeds. Such an implementation is shown in Fig. IIa, in which the compressor is designed as a cascade of n fluid flow machines. At any point after the first turbomachine one or more heat exchangers, for example, for hot water heating, which are denoted by 170, is preferably provided. These heat exchangers are designed to cool the gas which has been heated (and compressed) by a preceding turbomachine 172. Here, the superheating enthalpy is usefully used to increase the efficiency of the entire compaction process. The cooled gas is then further compressed with one or more downstream compressors or fed directly to the condenser. It is removed heat from the compressed water vapor, so that z. B. service water to higher temperatures than z. B. 4O ⁰ C to heat. However, this reduces not the overall efficiency of the heat pump, but even increases it, since two sequential turbomachines with interposed gas cooling with a longer life due to the reduced thermal stress and less energy to achieve the required gas pressure in the condenser, as if a single turbomachine without gas cooling would be present ,

The cascaded independently operated turbomachines are preferably controlled by a controller 250, the input side receives a desired temperature in the heating circuit and possibly also an actual temperature in the heating circuit. Depending on the desired oil temperature, the rotational speed of a turbomachine arranged earlier in the cascade, which is denoted by ni by way of example, and the rotational speed n2 of a turbomachine arranged later in the cascade are changed, as illustrated with reference to FIG. If a higher oil temperature is input to the controller 250, both speeds are increased. However, the speed of the previously arranged turbomachine, which is denoted by ni in Fig. IIb, raised with a smaller gradient than the rotational speed n2 of a later arranged in the cascade turbomachine. As a result, when the ratio n ₂ / ni of the two rotational speeds is plotted, a straight line with a positive slope results in the diagram of FIG. IIb.

The intersection between the individually plotted rotational speeds ni and n ₂ can take place at any desired location, that is to say at any desired temperature and, if appropriate, can not take place. In general, however, it is preferable to lift a turbomachine arranged closer to the condenser in the cascade more strongly in terms of its rotational speed than a turbomachine arranged earlier in the cascade, if a higher desired temperature is desired. The reason for this is that the turbomachine, which is later arranged in the cascade, must process already compressed gas that has been compressed by a turbomachine that was previously arranged in the cascade. Furthermore, this ensures that the blade angle of blades of a radial wheel, as also explained with reference to FIGS. 11c and 6d, is always as good as possible with respect to the speed of the gas to be compressed. Thus, the adjustment of the blade angle consists only in optimizing a low-turbulence compression of the incoming gas. The other parameters of the angle adjustment such as gas flow rate and compression ratio, which would otherwise have made a technical compromise in the choice of blade angle and thus only at a desired temperature optimum efficiency are brought according to the invention by the independent speed control to its optimum operating point and therefore have no influence more on the choice of the blade angle, so there is always an optimal efficiency despite a fixed blade angle.

In view of this, it is further preferred that a turbomachine arranged more in the cascade in the direction of the condenser has a direction of rotation of the radial wheel which is opposite to the direction of rotation of a radial wheel previously arranged in the cascade. Thus, a nearly optimal entry angle of the blades of both axial wheels can be achieved in the gas flow, such that a favorable efficiency of the turbomachine cascade occurs not only in a small target temperature range, but in a much larger target temperature range between 20 and 50 degrees, which is an optimal area for typical heating applications. The speed control according to the invention and, if appropriate, the use of counter-rotating axial gears thus provides optimum matching between the variable gas flow with changing desired temperature on the one hand and the fixed blade angles of the axial wheels on the other hand. In preferred embodiments of the present invention, at least one or preferably all axial wheels of all flow machines are made of plastic with a tensile strength above 80 MPa. A preferred plastic for this is polyamide 6.6 with inserted carbon fibers. This plastic has the advantage of tensile strength, so that Axialräder the fault machines can be made of this plastic and yet can be operated at high speeds.

Axial wheels are preferably used according to the invention, as shown for example in Fig. 11c at reference numeral 260. Fig. 11c shows a schematic plan view of such a radial wheel, wherein Fig. Hd shows a schematic cross-sectional view of such a radial wheel. A radial wheel, as is known in the art, includes a plurality of internally outwardly extending vanes 262. The vanes extend from a center axis 264 distance, designated r _w , entirely outward of the axis 264 of FIG radial wheel. In particular, the radial wheel comprises a base 266 and a lid 268 which is directed to the intake manifold or to an earlier stage compressor. The radial wheel includes an intake port, designated ri, for sucking gas, which gas is then discharged laterally from the radial wheel, as indicated at 270 in FIG. Hd.

When Hc is considered, the gas is at a relatively higher speed in the direction of rotation in front of the blade 262, while at a reduced speed behind the blade 262. However, for high efficiency and high efficiency, it is preferred that the gas be ejected from the radial wheel at all sides with as uniform a speed as possible, ie at 270 in FIG. Hd. For this purpose, there is a desire to mount the blades 262 as close as possible. However, for technical reasons, an arbitrarily tight attachment of from the inside, ie from the radius r _w outwardly extending blades is not possible because then the suction port with the radius ri is more and more blocked.

In accordance with the invention, it is therefore preferred to provide blades 272, 274, 276, respectively, that extend less than the blade 262. In particular, the vanes 272 do not extend from r _w to the outside, but extend outward from Ri with respect to the radial wheel, where Ri is greater than rw. Similarly, as shown by way of example in Fig. 11c, vanes 274 extend outwardly from R ₂ only, while vanes 276 extend outwardly only from R ₃ , where R _{2 is} greater than Ri and where R _{3 is} greater than R ₂ is.

These relationships are shown diagrammatically in FIG. Hd, with a double hatching, for example in the region 278 in FIG. Hd, indicating that there are two blades in this region which overlap and are therefore characterized by the double-hatched region. Example, the hatching from bottom left to top right in the area 278 denotes a vane 262 _w extends from r to by the very outside, while extending from top left to bottom right of the area 278 hatching indicates a vane 272 which merely to extends from Ri to the outside with respect to the radial wheel.

Preferably, at least one blade, which does not extend so far inwardly, is thus arranged between two blades extending deeper inwardly. This means that the intake area is not clogged or areas with a smaller radius are not covered too much with blades, while areas with a larger radius are more densely covered with blades, so that at the outlet of the radial wheel, ie where the compressed Gas leaves the radial wheel, as homogeneous a velocity distribution of the exiting gas exists. The GE- Speed distribution of the exiting gas is particularly homogeneous in the preferred radial wheel according to the invention in Fig. 11c on the outer circumference, because the distance of blades accelerating the gas, and due to the "stacked" arrangement of the blades is substantially smaller than in one case in which, for example, only the blades 262 are present, which extend from the very inside to the very outside and thus inevitably at the outer end of the radial wheel have a very large distance, which is substantially larger than the radial wheel according to the invention, as shown in Fig. llc is.

At this point it should be noted that the relatively complex and complicated shape of the radial wheel in Fig. Llc can be made particularly favorable with plastic injection, in particular, can be easily achieved that all blades, including the blades that are not completely extend inside to the very outside, so the blades 272, 274, 276 are firmly anchored, since they are both connected to the lid 268 so on the base 266 of FIG. Hd. The use of plastic in particular with the plastic injection molding technology makes it possible to produce any shape accurately and inexpensively, which is not readily possible or only very consuming or possibly even impossible with axial metal wheels.

It should be noted at this point that very high rotational speeds of the radial wheel are preferred, so that the accelerations acting on the blades assume very considerable values. For this reason, it is preferred that, in particular, the shorter blades 272, 274, 276 are fixedly connected not only to the base but also to the cover, such that the radial wheel can readily withstand the accelerations that occur.

In this context, it should also be noted that the use of plastic is also favorable due to the superior impact resistance of plastic. So is not always exclude that ice crystals or water droplets hit the radial wheel at least the first stage compressor. Due to the high accelerations, very high impact forces arise which are easily withstood by plastics with sufficient impact resistance. Furthermore, the liquefaction in the liquefier preferably takes place on the basis of the cavitation principle. Here steam bubbles fall due to this principle in a volume of water in itself. Considerably microscopically, these also give rise to considerable speeds and forces which, viewed over the long term, can lead to material fatigue which, however, is easily manageable if a plastic with sufficient impact resistance is used.

The compressed gas discharged from the last compressor 174, that is to say the compressed steam, is then supplied to the condenser 18, which may be designed as shown in FIG. 7, but which is preferably configured as shown in FIG. 8a is shown. The condenser 18 contains a volume of water 180, and preferably an arbitrarily small volume of steam 182. The condenser 18 is designed to feed the compressed vapor into the water of the volume of water 180, so that where the vapor enters the liquid immediately results in condensation as schematically indicated at 184. For this purpose, it is preferred that the gas supply has a widening region 186, such that the gas is distributed as extensively as possible in the condenser water volume 180. Typically, due to the temperature layers in a water tank above, the highest temperature and below the coolest temperature. Therefore, via a float 188, the heating flow is arranged as close as possible to the surface of the water volume 180 in order to always remove the warmest water from the condenser water volume 180. The heating return is fed to the bottom of the condenser, so that the steam to be liquefied always comes in contact with the coolest possible water that is due to the circulation using a heating circulation pump 312 again moved from below in the direction of the steam-water boundary of the expander 186.

The embodiment in Fig. 7, in which only a single circulating pump 312 exists, is sufficient if the condenser is arranged in a building so that the areas to be heated are below the condenser, so that due to gravity in all heating pipes a larger Pressure as in the liquefier.

10b, on the other hand, shows an implementation of a connection of a heating line to the condenser with a turbine / pump combination if the condenser is to be arranged at a lower height than the heating line or if a conventional heating system requiring a higher pressure is to be connected , If the condenser is therefore to be arranged lower, that is to say below a surface to be heated or the heating line 300, the pump 312 is designed as a driven pump, as shown at 312 in FIG. 10b. Further, a turbine 310 is provided in the heater return 20b for driving the pump 312, which is connected via a power coupling 314 to the pump 312. The high pressure then prevails in the heater and the low pressure prevails in the condenser.

As a result of the steam continuously introduced into the condenser, the water level in the condenser would increase more and more, the drain 22 is provided, above which, so that the water level in the condenser does not change substantially, also z. B. have to drain about 4 ml per second. For this purpose, a drain pump or a drain valve 192 is provided for pressure control, such that without pressure loss, the required amount of z. B. 4 ml per second, ie the amount that is supplied to the steam liquefier with the compressor running, is discharged again. Depending on the implementation of the process into the riser, as shown at 194. After all the pressures between a bar and the present in the evaporation space pressure along the riser 102, it is preferred to feed the outlet 22 at the point 194 in the riser at which approximately the same pressure exists as he after the pump 192 and the Valve 192 is present. Then no work must be done to supply the drain water to the riser again.

In the embodiment shown in Fig. 7 is carried out completely without a heat exchanger. The groundwater is thus evaporated, the steam is then liquefied in the condenser, and the liquefied steam is finally pumped through the heating and returned to the riser. However, since not all of the water flowing through the riser amount of water is evaporated, but always only a (very small) share, thus, water that has flowed through the underfloor heating, fed to the groundwater. If such a thing is prohibited by local regulations, although the present invention does not involve any contamination, the process may also be designed to deliver the amount of 4 ml per second corresponding to about 345 liters per day to the channel. This would ensure that no medium that was in a heating system of a building is directly fed back into the groundwater.

However, the return 112 from the evaporator can be fed without problems into the groundwater, since the water returning there was in contact only with the riser and the return line, but did not exceed the "evaporation limit" between the evaporator expander 108 and the outlet to the turbomachine Has.

It should be noted that the evaporation space in the exemplary embodiment shown in FIG. 7 as well as the condenser or the steam space 182 of the condenser are removed. must be sealed. As soon as the pressure in the evaporation chamber rises above the level required for the water pumped through the riser pipe to evaporate, the heat pump process "stops".

Reference is now made to FIG. 8a, which illustrates a preferred embodiment for the condenser 18. FIG. The compressed steam supply line 198 is placed in the condenser so that the steam just below the surface of the condenser water volume 180 can escape into this volume of water. For this purpose, the end of the steam supply line around the circumference of the tube arranged around nozzles, through which the steam can escape into the water. To ensure the best possible mixing occurs, so that the steam comes into contact with the coldest possible water in order to liquify as quickly and efficiently as possible, an expander 200 is provided. This expansion is located in the condenser water volume 180. It has at its narrow point a circulation pump 202, which is designed to suck cold water at the bottom of the condenser and to put through the expander in an upward broadening flow. In this way, the largest possible amounts of the steam entering the liquefier water 180 should be brought into contact with the coldest possible water which is supplied by the circulation pump 202.

Further, it is preferred to provide around the condenser a sound insulation 208 which may be actively or passively formed. Passive sound insulation, like thermal insulation, will insulate the frequencies of the sound produced by the liquefaction process as well as possible. Likewise, it is also preferred to silencing the other components of the system.

The sound insulation may alternatively be actively formed, in which case z. B. a microphone for sound measurement would have and in response to a sound Counteracting effect, such as an in-vibration displacement of an outer condenser wall, etc. with z. B. piezoelectric means.

The embodiment shown in Fig. 8a is somewhat problematic in that, when the heat pump is taken out of service, the liquid 180 in the condenser penetrates into the tube 198, in which otherwise a compressed vapor is present. In one implementation, a return valve in line 198, e.g. be arranged in the vicinity of the output of the line from the condenser. Alternatively, line 198 may be directed up to the point where no liquid is returned to the compressor when the compressor is switched off. When the compressor is then put back into operation, so is first pressed by the compressed steam, the water from the steam line 198 in the condenser.

Only then is a vapor in the condenser brought to condensation when a sufficient proportion of the water has been removed from the line 198. Such a kind of embodiment thus has a certain delay time, which is needed until the volume of water 180 is reheated by the compressed steam. Further, the work required to remove the water that has entered the conduit 198 from the conduit 198 is no longer recoverable and thus "lost" in terms of heating, such that lower efficiency losses are tolerated have to.

An alternative embodiment which overcomes this problem is shown in FIG. 8b. In contrast to Fig. 8a, the compressed vapor is now not supplied within a tube below the water level in the condenser. Instead, the vapor is, so to speak, "pumped" into the liquid in the condenser from the surface, for which purpose the condenser comprises a nozzle plate 210, which has nozzles 212 protruding with respect to the plane of the nozzle plate 210. The nozzles 212 extend below the water level of the water volume 180 in the condenser. The recessed portions between two nozzles, shown at 214 in Fig. 8b, on the other hand extend above the water level of the water volume 180 in the condenser, so that always between two nozzles is the water surface of the condenser water, which is interrupted by a nozzle. The nozzle 212 has nozzle openings through which the compressed vapor that propagates from the conduit 198 within the vapor volume 182 may enter the condenser water, as shown schematically by arrows 216.

If the compressor is taken out of operation in the implementation of Fig. 8b, this leads to the fact that the liquid penetrates only a little bit into the nozzles 212 of the nozzle plate 210, so that only very little work has to be spent in order to Re-commissioning the heat pump to push out the water from the nozzles again. In any event, the expander 200 ensures that, due to the guide through the expander, the liquid being pumped upwardly by the pump 202 is always as cold as possible and comes in contact with the warm vapor. The warm water then either enters immediately into the supply line 20a or spreads over the Aufweiter edge in the water volume, as shown by an arrow 218, so that in the liquefier outside of the expander occurs a temperature stratification, in particular due to the up - further form is disturbed as little as possible.

The flow velocity at the edge of the expander, that is, where the arrow 218 is indicated, is substantially lower than in the middle. It is preferred to operate the condenser as a temperature-layer store, such that the heat pump and in particular the compressor do not have to run continuously, but only have to run when may exist, as it is for normal heating systems, for example, work with an oil burner, also the case.

Fig. 8c shows a further preferred implementation of the condenser in schematic form. In particular, the condenser comprises a gas separator 220, which is coupled to the gas volume 182 in the condenser. Gas generated in the condenser, such as oxygen or another gas which may outgas in the condenser, accumulates in the gas separator vessel 220. This gas can then be pumped to the atmosphere by operating a pump 222, preferably at certain intervals, since continuous gas pumping is not necessary due to the small amount of gas produced. Alternatively, the gas can also be docked in the return 112 or 113 of FIG. 7, so that the gas is returned together with the returning groundwater back into the groundwater reservoir, where it is then dissolved again in the groundwater, or when it enters the groundwater reservoir, where it goes into the atmosphere.

After the system of the invention works with water, even with strong outgassing no gases that were not previously solved in the groundwater, so that the separated gas has no environmental problems in itself. Again, it is emphasized that due to the turbomachinery compaction according to the invention and the use of water as the working fluid at any point contamination or contamination by a synthetic refrigerant or by an oil due to an oil circuit occur. The system according to the invention has at each point as a working medium water or steam, which is at least as clean as the original groundwater, or even cleaner because of the evaporation in the evaporator than the groundwater, since it is distilled water when the compressed steam in Condenser has been re-liquefied. Referring now to Figure 9a, there is shown a preferred embodiment of the evaporator for advantageously utilizing the condenser flow to accelerate the vaporization process. For this purpose, the effluent, which is at a heating-return temperature, that is to say a much higher temperature than the groundwater extracted from the earth, is passed through the expander 108 of the evaporator, so that the wall of the drainage tube 204 acts as a germ acts for a bubble boiling. This is much more efficient steam generated by the evaporator than if no such germination is provided. Furthermore, it is preferred to make the wall of the drainage pipe 204 as structured as possible, at least in the expander, as shown at 206, in order to further improve nucleation for the bubble boiling. The rougher the surface of the downcomer 204 is, the more nuclei are created for the bubbling. It should be noted that the passage through the drainpipe 22 is only very small, as it is only 4 ml per second added to the condenser in one operating mode. Nevertheless, even with this small amount and due to the relatively high temperature compared to the groundwater, the much more efficient bubble boiling can be brought about in order to reduce the size of the evaporator with the same efficiency of the heat pump.

To accelerate the evaporation process, alternatively or additionally, a region of the evaporator on which water to be evaporated is located, that is to say the surface of the expander or a part thereof, may be made of a rough material in order to supply nuclei for nucleate boiling. Alternatively or additionally, a rough grid (near) can be arranged below the water surface of the water to be evaporated.

Fig. 9b shows an alternative implementation of the evaporator. While the process in FIG. 9a is merely a "through" support for bubble formation for efficient 9b is used and, as shown in Fig. 9a on the left in the picture, then, when it has passed through the evaporator, is discharged, the process in Fig. 9b itself is used to enhance the formation of bubbles. For this purpose, the condenser outlet 22 of FIG. 7 is optionally connected via a pump 192 or, if conditions permit, without a pump, to a nozzle tube 230 having an end 232 at one end and nozzle openings 234. The warm condenser water, which is discharged from the condenser via the drain 22 at a rate of, for example, 4 ml per second, is now fed to the evaporator. On its way to a nozzle opening 234 in the nozzle tube 230 or directly at the exit from a nozzle, it will evaporate to a certain extent below the water surface of the evaporator water due to the pressure which is too low for the temperature of the discharge water.

The resulting vapor bubbles are directly as boiling nuclei for the evaporator water, which is conveyed via the inlet 102, act. Thus, efficient bubble boiling in the evaporator can be triggered without major additional measures, such triggering being similar to FIG. 9a due to the fact that the temperature near the rough area 206 in FIG. 9a or in the vicinity of a nozzle opening 234 already is so high that immediately takes place at the present pressure evaporation. This evaporation forces the creation of a vapor bubble, which, if the conditions are favorably chosen, has a very high probability that it will not collapse again, but that it will develop into a surface-going vapor bubble which, as soon as it does has entered the vapor volume in the evaporation chamber is sucked through the suction pipe 12 from the compressor.

The exemplary embodiment shown in FIG. 9b requires that condenser water enter the groundwater circuit. is brought, since the emerging from the nozzle tube 230 medium finally enters the overflow of the evaporator in the return 112 and is thus brought into contact with the groundwater.

If water-legal conditions or other reasons exist that this is not permissible, then the exemplary embodiment shown in FIG. 9c can be used. Here, the warm condenser water supplied from the condenser outlet 22 is introduced, for example, at a rate of 4 ml per second into a heat exchanger 236 to deliver its heat to a groundwater coming from the main groundwater flow in the line 102 via a branch line 238 and a branch pump 240 has been branched off. The branched groundwater then substantially decreases the heat of the condenser outlet within the heat exchanger 236, so that preheated groundwater, for example, at a temperature of 33 ⁰ C is introduced into the nozzle tube 230 in order by the high temperature compared to the groundwater Bubble boiling in the evaporator to effectively trigger or support. By contrast, the heat exchanger delivers via a drain line 238 relatively strongly cooled waste water, which is then fed via a drain pump 240 of the sewer. Due to the combination of branch line 238 and branch pump 240 and heat exchanger 236, only groundwater is used or introduced in the evaporator without it being in contact with another medium. A relevance under water law thus does not exist in the exemplary embodiment shown in FIG. 9c.

Figure 9d shows an alternative implementation of the edge feed vaporizer. Here, in contrast to FIG. 7, the expander 200 of the evaporator is arranged below the water level 110 in the evaporator. This results in water flowing "from the outside" into the center of the expander, to then be returned to a central conduit 112. While the central conduit in Fig. 7 has served to feed the evaporator, it now serves in Fig. 9d - direct the unevaporated groundwater. In contrast, the conduit 112 shown in Fig. 7 has been used to remove unevaporated groundwater. In contrast, in FIG. 9d, this line at the edge acts as a groundwater supply.

Fig. 9e shows a preferred implementation of the expander 200 as it can be used in the evaporator or the expander as e.g. can also be used in the condenser, and as shown for example in Fig. 7 or Fig. 8a and 3b. The expander is preferably designed such that its small diameter preferably enters the expander at the center of the "large" expander surface. This diameter of this inlet (in Fig. 9d) is preferably between 3 and 10 cm and especially preferred embodiments between 4 and 6 cm.

The large diameter d _{2 of} the expander is in preferred embodiments between 15 and 100 cm and is smaller than 25 cm in particularly preferred embodiments. The small version of the evaporator is possible when efficient measures for triggering and supporting the bubble boiling are used, as explained above. Between the small radius di and the large radius d _Σ there is a curvature region of the expander, which is preferably designed to give a laminar flow in this region which is of a fast flow rate, preferably in the range of 7 to 40 cm per Second is lowered to a relatively small flow rate at the edge of the expander. Strong flow rate discontinuities, such as eddies in the area of the line of curvature or "bubbling" above the inlet when viewed from the top of the expander, are preferably avoided as they may be detrimental to efficiency.

In particularly preferred embodiments, the expander has a shape which results in the height of the water is above the Aufweiter surface smaller than 15 mm, and preferably between 1 and 5 mm. It is therefore preferred to use an expander 200 which is designed such that in more than 50% of the area of the expander, when viewed from above, there is a water level which is less than 15 mm. This ensures efficient evaporation over the entire area, which, in view of its efficiency, is particularly enhanced when measures are taken to trigger BIAS boiling.

The heat pump according to the invention thus serves for efficient heat supply of buildings and no longer requires working equipment that has a global climate-damaging influence. According to the invention, water is evaporated under very low pressure, compressed by one or more turbomachines arranged one behind the other and liquefied again into water. The transporting energy is used for heating. According to the invention, a heat pump is used, which is preferably an open system. Open system here means that groundwater or other available thermal energy-carrying aqueous medium is evaporated, compacted and liquefied under low pressure. The water is used directly as a working medium. The contained energy is therefore not transferred to a closed system. The liquefied water is preferably used directly in the heating system and then fed back to the groundwater. To decouple the heating system capacitively, it can also be completed via a heat exchanger.

The efficiency and usefulness of the present invention will be illustrated by way of a numerical example. If an annual heat requirement of 30,000 kWh is assumed, according to the invention about a maximum of 3750 kWh of electric current must be expended for the operation of the turbomachine since the turbomachine only has to supply about one-eighth of the total heat requirement. The eighth comes from the fact that only in the most extreme cold a sixth must be spent, and z. B. at transition temperatures such as in March or the end of October, the efficiency can rise to a value greater than 12, so that a maximum of one-eighth must be spent on average over the year.

With electricity costs of about 10 cents per kWh, which can be achieved for electricity when electricity is purchased, for which the power plant does not have to guarantee freedom of interruption, this corresponds to an annual cost of about 375 euros. If you want to produce 30,000 kWh with oil, you would need about 4000 1, which would correspond to a current cost of oil, which is unlikely to fall in the future, a price of 2800 €. According to the invention, one can therefore save 2425 euros per year! It should also be noted that compared to the combustion of oil or gas for purposes of heating by the inventive concept, up to 70% of the amount of released CO _{2 is} saved.

In order to reduce the production costs and also to reduce the maintenance and assembly costs, it is preferred to design the housings of the evaporator, the compressor and / or the condenser and also, in particular, the radial wheel of the flow machine made of plastic and in particular of injection-molded plastic. Plastic is well suited, since plastic is corrosion-resistant with respect to water and, according to the invention, advantageously the maximum temperatures are significantly below the deformation temperatures of usable plastics in comparison with conventional heaters. Furthermore, the assembly is particularly simple, since there is negative pressure in the system of evaporator, compressor and condenser. This means that there are far fewer requirements for the seals, since the entire atmospheric pressure helps to keep the housings tight. Plastic is also particularly well, since at no point high temperatures occur in the system according to the invention, the use of expensive special plastics, metal or ceramic would require. By plastic injection molding, the shape of the radial wheel can be arbitrarily optimized and yet easily and inexpensively manufactured in spite of complicated shape.

Depending on the circumstances, the method according to the invention can be implemented in hardware or in software. The implementation can be carried out on a digital storage medium, in particular a floppy disk or CD, with electronically readable control signals which can interact with a programmable computer system in such a way that the corresponding method is carried out. In general, the invention thus also consists in a computer program product with program code stored on a machine-readable carrier for carrying out the method according to the invention, when the computer program product runs on a computer. In other words, the invention can thus be realized as a computer program with a program code for carrying out the method when the computer program runs on a computer.

Claims

claims

A bearing for rotatably supporting a rotor (500) on a stator (502), the bearing having a rest position in which the rotor and the stator are in contact, and a working position in which the rotor and the stator are not in contact, with the following features:

an actuator (512) for moving the rotor or stator from the rest position to the working position in response to a control signal (514); and

control means (516) for providing the control signal (514), the control means being adapted to supply the control signal (514) to the actuator which causes

in an acceleration phase (532) the actuator (512) accelerates the rotor or stator from the rest position towards the working position, and

that the actuator (512) is decelerated in a braking phase (534), wherein the rotor or stator detaches from contact with the actuator (512), the rotor or the stator due to its mass, and

wherein the control means (516) is further adapted to provide such a control signal to the actuator that the rotor or stator is accelerated so much that the working position is reached.

A bearing according to claim 1, wherein the rest position is stable and the working position is metastable, the control means (516) further comprising a controller (517) for detecting an actual position of the rotor with respect to the stator or an actual position change of the rotor with respect to the stator to obtain an actual value and for moving the rotor or stator such that the actual position of the rotor with respect to the stator approaches a target position of the rotor with respect to the stator at the working position or that the actual position change approaches a zero value at least.

3. A bearing according to claim 2, wherein the bearing is formed so that it is stable in the working position radially with respect to a rotational axis (506) of the rotor (500), and in the working position axially with respect to the axis of rotation (506) is metastable.

4. Bearing according to one of the preceding claims, wherein the rotor (500) in the rest position at at least one contact point (510) on the stator (502) is applied, and wherein the rotor (500) in the working position through a gap (504) is separated from the stator (502), wherein the actuator (512) is mechanically fixed to the stator (502, 560) and is not mechanically fixed to the rotor (500).

5. Bearing according to one of the preceding claims, wherein the rotor (500) has a bearing portion (562), and wherein the stator (502) has a bearing portion (560), wherein a mass of the bearing portion of the stator is smaller than a mass of the Rotor (500) is.

6. Bearing according to one of claims 1 or 2, wherein the actuator (512) with the stator (502) is mechanically connected to accelerate the stator in the acceleration phase and decelerate the braking phase, so that the rotor in the Braking phase releases from the stator.

7. Bearing according to claim 2,

in which the actuator (512) is mechanically connected to the stator (502), and

wherein the controller (517) is adapted to move the stator (502) to approximate the actual position of the rotor with respect to the stator and the desired position of the rotor with respect to the stator.

8. Bearing according to one of the preceding claims, which is designed as a magnetic bearing having a rotor bearing portion (560) with a number of axially disposed magnetic rings (560a, 560b, 560c, 56Od) and a stator bearing portion (560) having a number of axially disposed magnetic rings (560a, 560b, 560c, 56Od), the number being greater than or equal to 1 for both the rotor bearing portion and the stator bearing portion, the bearing portions (560, 562) being so arranged to each other that in the working position opposite rings repel magnetically stronger than in the rest position, and wherein in the rest position magnetically opposite rings attract more than in the working position.

9. Bearing according to claim 8, wherein the magnetic rings are permanent magnetic.

10. Bearing according to claim 4, wherein the gap (504) in the working position 0.05 to 3.0 mm.

A bearing according to any one of the preceding claims, wherein the actuator (512) has a maximum stroke length which is less than a distance to effect relative movement between the rotor and the stator is, so that the bearing is moved from the rest position to the working position.

The bearing of claim 11, wherein the actuator (512) has a maximum stroke length that is less than one-half

Track is.

13. A bearing as claimed in any one of the preceding claims, wherein the actuator (512) is a piezoelectric linear actuator having an actuating direction, the actuator being connected to the stator (502) which moves by actuation along the actuating direction of the stator or the rotor is parallel to a rotational axis of the rotor achievable.

14. Bearing according to one of the preceding claims, wherein the actuator (512) is designed to achieve a maximum acceleration, which is greater than 150 microns / ms depending on the control signal.

A bearing according to any one of the preceding claims, wherein the acceleration phase and the braking phase are part of a total beta-tiger activity duration, the braking phase being shorter than the acceleration phase.

16. A bearing according to claim 15, wherein a portion of the acceleration phase in the total actuator activity duration is greater than or equal to 60%.

17. A bearing according to claim 16, wherein the rotor (500) and the actuator (512) are dimensioned so that a period of time in which the rotor is not in contact with the stator, at least twice as large as the total actuator -Aktivitätsdauer.

18. Bearing according to claim 17, wherein the control means (516) is adapted to the actuator (512) in to accelerate an acceleration phase (536) subsequent to the brake phase to cause the actuator to reach a home position with decreasing speed in which the operator was before a start of the acceleration phase (532).

19. Bearing according to one of the preceding claims, wherein the control means (516) is adapted to drive a motor for driving the rotor so that no rotation occurs before the beginning of the acceleration phase (532), and that a rotation is only started, when the working position is reached.

A bearing according to any one of the preceding claims, wherein the control means (516) is arranged to perform an acceleration phase and a subsequent braking phase (534) when the rotor is to go to rest, the actuator (512) being in the Acceleration phase is accelerated and braked in the braking phase, with a timing of the acceleration phase and the braking phase are carried out so that the rotor touches the stator in the braking phase, and that the rotor, while it touches the stator, brought into the rest position becomes.

21 turbomachine with a radial wheel, wherein the radial wheel is supported by a bearing according to one of claims 1 to 20.

22. A method of operating a bearing for rotatably supporting a rotor (500) on a stator (502), the bearing having a rest position in which the rotor and the stator are in contact, and a working position in which the rotor and do not touch the stator, with the following steps: Moving (512) of the rotor or stator from the rest position to the working position, wherein in an acceleration phase (532) the rotor or stator is accelerated from the rest position towards the working position, and wherein in a braking phase (534) an actuator or the Stator is braked, wherein the rotor or stator detaches due to its mass from contact with the actuator (512), the rotor or the stator, wherein the rotor or stator are accelerated in the acceleration phase so strong that the working position is reached ,

23. Computer program with a program code for performing the method according to claim 22, when the program runs on a computer.