A rotor that rotates at high speed and a turbo compressor including the same

Abstract

The present invention describes a turbo compressor (1) comprising at least one rotor (10) whose shaft (11) is provided with at least one preloading element (60) that keeps the diameter of the shaft permanently larger than its characteristic diameter (D1). The present invention further comprises a high speed rotor and a method for its manufacture.

Description

Technical Field

[0001] The present invention relates in particular to a rotor for high-speed rotation, which is suitable for refrigeration equipment, such as a cooling device or a heat pump used in industrial processes, a turbine in an organic Rankine cycle, a fuel cell recycling device, an optical scanner, an inertial gyroscope, and any other application requiring a high-speed rotating rotor. The rotor is used in a radial turbo compressor lubricated by a gaseous working fluid.

Background Art

[0002] Cooling devices are used for cooling and refrigeration operations in many industrial and agricultural processes. Such compressors can also be used as heat pumps. Ammonia cooling devices are currently driven by large oil-lubricated piston compressors that are bulky, heavy and require regular maintenance. The oil used for lubrication further contaminates the ammonia and is a contaminant material that needs to be replaced and treated before disposal. From the perspective of efficiency, additional power is consumed due to the need to recycle and cool the lubricant, which is detrimental to the current trend of low energy consumption. Furthermore, the oil moving inside the cycle reduces the heat conduction in the heat exchanger.

[0003] Some turbo compressors have been disclosed in which the lubricant is the gaseous refrigerant itself, thus avoiding the use of oil as a lubricant. In that case, the rotor of the turbo compressor is supported by the gaseous refrigerant. An important aspect associated with such an arrangement is that the compressor can be miniaturized. However, the centrifugal effect induces an outward deformation of the rotating shaft, which can reduce the nominal bearing clearance and limit the ability of the bearings to support the shaft. This deformation has important implications for gas-lubricated bearings. In some applications, the shaft-bearing clearance can be of the same order as the centrifugal deformation of the shaft, leading to seizure.

[0004] Furthermore, unlike positive displacement machines, dynamic compressors are more complex to control because the rotor speed is limited by surge and choke, which affects both the mass flow rate and the achievable pressure ratio. The risk of damage to the machine is high if the turbocharger proceeds to surge, and it is almost impossible to stabilize the refrigeration cycle once it has entered a surge state, so it should be avoided. In many cases, the only option is to stop and restart the machine. Therefore, a cooling device driven by a turbocharger requires enhanced control to avoid dangerous conditions.

[0005] Therefore, there is room for improvement in this technical field, particularly in providing more compact and cost-effective compressors. An important factor is enabling a higher rotor speed despite the physical limitations described above. There is also a need for better control of the refrigeration cycle.

Summary of the Invention

Problems to be Solved by the Invention

[0006] Therefore, an object of the present invention is to provide a more environmentally friendly and compact turbocharger that does not require or substantially does not require lubricating oil, thus improving cycle efficiency.

[0007] A further object of the present invention is to provide a turbocharger with improved efficiency compared to known turbochargers. This includes lower power consumption and / or higher compression output.

[0008] A further object of the present invention is to provide a rotor with a minimal increase in diameter during rotation and adapted to high rotational speeds. A further object is to provide a rotor with minimal diameter variation from the rest position to functional rotation and vice versa.

[0009] A further object of the present invention is to provide a compact turbo compressor that enables oil-free operation, facilitates the operation of a multi-stage cycle, and reduces the losses of a heat exchanger.

[0010] Another object of the present invention is to provide an improved method for compressing a gaseous working fluid, particularly a gaseous refrigerant fluid, using a radial turbo compressor.

Means for Solving the Problems

[0011] According to the present invention, these objects are achieved by the subject matter of the independent claims and are further detailed through the dependent claims.

[0012] Compared to what is known in the art, the present invention provides an improved turbo compressor that is environmentally friendly and has improved efficiency.

[0013] Exemplary embodiments of the present invention are disclosed herein and illustrated by the following drawings.

Brief Description of the Drawings

[0014]

Figure 1

Figure 2

Figure 3a

Figure 3b

Figure 3c

Figure 3d

Mode for Carrying Out the Invention

[0015] The turbo compressor 1 described in this specification is particularly suitable for gaseous working fluids such as ammonia or other fluids with a low molecular weight, such as water or propane as an example. The rotational speed of the rotor is extremely high, typically, for a rotor having a diameter of about 20 mm to about 60 mm, higher than 100,000 revolutions per minute or 170,000 revolutions per minute or even 300,000 revolutions per minute, or even higher depending on the rotor output and its diameter. At such a high rotational speed, aerodynamic compressor design and thermal management still remain highly challenging parameters. The radial deformation of the rotor is also a major limiting constraint considering the thinness of the gap.

[0016] The rotational speed of the rotor can be defined by the number of revolutions per minute and thus indicates its angular velocity, but the rotational speed is better defined by its DN number which is related to its tangential speed. The tangential speed takes into account both the angular velocity of the rotor and its outer diameter and can be expressed by the following formula: DN is equal to or directly proportional to the angular velocity × the radius of the rotor.

[0017] Therefore, in this specification, the terms "high speed" or "speed of high-speed rotation" mean a DN number exceeding about 3 million and even exceeding 4 million, expressed in typical rpm×mm units. Considering the angular velocity value, when the terms "high speed" or "speed of high-speed rotation" are applied to a rotor having a diameter greater than about 20 mm or greater than about 50 mm, it more appropriately indicates an angular velocity exceeding 100,000 revolutions per minute, preferably exceeding 140,000 revolutions per minute, and more preferably about 300,000 revolutions per minute. A person skilled in the art will understand that, provided that the tangential velocity remains within the above-mentioned value range, even at a lower angular velocity, a larger diameter is also relevant. Therefore, the speed of high-speed rotation clearly represents the rotational speed at which the rotor is subjected to a force that will result in a significant increase in its diameter. In the framework of this specification, considering the gap around the rotor, an increase in the rotor diameter of 0.01% or more can already be considered significant. The speed of high-speed rotation is also related to the speed at which the rotor is functional, which means that this rotational speed can compress the gas fluid.

[0018] Referring to FIG. 1, the turbo compressor 1 according to the present disclosure includes a housing 40 in which a rotor 10 is disposed so as to be rotatable inside. Inside the housing 40, one or more bearings 20 are provided to maintain the rotor 10 in a predetermined position. Specifically, one or more bearings 20 surround the shaft 11 of the rotor 10 while maintaining a gap in relation to the outer surface of the shaft 11. The bearing 20 is preferably adapted to the gas fluid L. Therefore, the rotor 10 can rotate about its longitudinal axis A without contacting the bearing. The rotor 10 is rotationally driven by a motor (not shown). The motor can be any electric motor suitable for high-speed rotation, such as a rotational speed exceeding 100,000 revolutions per minute, or exceeding 150,000 or 170,000 revolutions per minute, and even exceeding 300,000 revolutions per minute.

[0019] Regarding the design of the motor, the windings are selected to avoid unbalanced attractive forces on the rotor that could lead to possible mechanical instabilities. Moreover, high speeds inevitably result in high iron losses due to a frequency of approximately 2800 Hz. The selected motor can exhibit the best compromise between iron losses and copper losses.

[0020] One end of the rotor 10 includes a compressor impeller 30 that deflects the gaseous working fluid L radially. For this reason, the compressor is a radial compressor. The deflected gaseous working fluid L is induced along the axis 11 of the rotor 10. The resulting layer of gaseous working fluid L centered on the axis 11 enables the rotor 10 to float and remain centered on its longitudinal axis A inside the bearing during rotation.

[0021] Referring to FIG. 2, the rotor 10 can be viewed in more detail. For example, the outer surface of the axis 11 may include one or more surface regions 13 having a specific treatment. For example, such surface regions 13 may be provided with grooves, such as V-shaped grooves that facilitate the formation of a gas film around the axis 11 during rotation. Such surface regions 13 may alternatively or additionally be provided with a coating or protective layer. Outside these surface regions, the outer surface of the axis 11 has no pattern or coating.

[0022] The rotor 10 may include one or more protrusions 12 that enable the rotor to be stabilized, for example, inside the housing 40. The rotor further includes, at one of its ends, a compressor impeller 30 having a compression vane 31 or an equivalent shape that can deflect the gaseous working fluid L. The housing 40 is provided with an inlet 50 for collecting the gaseous working fluid L. The inlet 50 is disposed near the compressor impeller 30. The inlet 50 may further have a conical shape in order to concentrate the flow of the gaseous working fluid L towards the compressor impeller 30.

[0023] Here, it is specified that the rotor 10 must be very well balanced so as to be able to obtain a complete rotational movement about the longitudinal axis A, including in the case of very high rotational speeds as mentioned above. The rotor 10 is further preferably lightweight in order to avoid inertia. At such high rotational speeds, weight can be a negative factor that increases potential defects and can result in speed limitations, vibrations, and rotational mechanical instabilities.

[0024] In one embodiment, the material of the rotor 10 or at least the shaft 11 is selected to be lightweight and at the same time have durability.

[0025] In one embodiment, the material of the rotor 10 or at least its shaft 11 is selected to exhibit minimal deformation under load. Specifically, the material is selected such that at the speed of high-speed rotation, its diameter shows only a minimal increase under centrifugal force. Specifically, the increase in its diameter is less than 0.1% of its original diameter or less than about 20 - 30 micrometers.

[0026] In one embodiment, the rotor 10 or at least its shaft 11 has a Young's modulus of more than 400 GPa, preferably 500 - 800 GPa. It can have a Vickers number of more than 2000, preferably about 2600 or more. Preferably, the material of the rotor 10 or at least its shaft 11 has a specific elastic modulus of 0.01 - 0.5 GPa / kg / m 3 and a specific strength of 0.01 - 0.9 MPa / kg / m 3 and is selected in such a way, where the specific elastic modulus means the ratio between the Young's modulus and the material density, and the specific strength means the ratio between the yield strength and the material density.

[0027] Therefore, any material having a suitable density and a suitable Young's modulus corresponding to the above values can be used, including ceramic materials such as steel or metal alloys or carbide materials.

[0028] The rotor shaft 11 can be hollow or at least partially hollow so as to define an internal space 14 (Figs. 3a, 3b). Preferably, the internal space extends over the entire length of the shaft 11. The rotor shaft 11 has a natural diameter D1 corresponding to its outer diameter in the absence of any external constraints. Specifically, the natural diameter D1 corresponds to the outer diameter of the shaft 11 in the absence of rotation at a temperature of about 20°C. The natural diameter D1 can be, for example, 30 mm. However, the diameter of the shaft 11 can vary as necessary, as evaluated by those skilled in the art. The natural diameter D1 can typically be between 20 mm and 100 mm.

[0029] Depending on the applied constraints, the outer diameter of the shaft 11 can increase and correspond to a load diameter D3 that is larger than the natural diameter D1 of the shaft 11. The load diameter D3 can be due to the centrifugal force under the speed of high-speed rotation. Alternatively, the load diameter D3 can be due to a temperature increase. The load diameter D3 can also be due to a combination of parameters such as the speed of high-speed rotation and a temperature increase. Depending on the intensity and nature of the load or stress applied to the shaft 11, the increase in diameter from its natural diameter D1 can be about 0.01% - 0.2%, typically about 0.05% - 0.1%. This results in a significant variation in diameter during rest compared to during the operating operation of the rotor. Such variations can be larger than the gap between the shaft 11 and surrounding fixed components such as the bearing 20.

[0030] Typically, the gap between the outer surface of the shaft 11 and the closest surrounding components such as the bearing 20 is less than 30 micrometers, or less than about 20 micrometers, and even less than about 10 micrometers, depending on the overall dimensions of the turbocompressor. If the natural diameter D1 allows such a gap, the increase in diameter under stress is considered to fill the gap and further cause direct contact with the surrounding elements, preventing the rotation of the rotor 10. Alternatively, if the natural diameter D1 is small enough to sufficiently predict such an increase under stress, the initial gap is too large to enable the support of the rotor 10 by the gaseous working fluid L.

[0031] To limit the diameter variation of the shaft 11, at least one preloading element 60 (Figs. 3c, 3d) makes it possible to artificially increase the natural diameter D1 of the shaft 11 at a value close to or equal to the loading diameter D3 corresponding to the preloading diameter D2. Thus, the preloading element 60 makes it possible to permanently maintain the outer diameter of the shaft 11 at the preloading value D2 even in the absence of any external stress. In other words, at least one preloading element 60 prevents the shaft 11 from taking its natural diameter D1 in the rest state. That is to say, it can be said that the diameter variation of the shaft 11 is restricted between its rest position and its operating rotation by at least one preloading element 60.

[0032] According to one aspect, the diameter of the shaft 11 including at least one preloading element 60 varies by less than 0.5%, preferably less than 0.2%, or about less than 0.05% when the rotational speed of the shaft increases from the rest condition to the high-speed rotation speed condition.

[0033] According to another aspect, the diameter of the shaft 11 including at least one preloading element 60 varies non-linearly when the rotational speed of the shaft 11 increases from the rest condition to the high-speed rotation speed condition. Specifically, the diameter of the shaft 11 remains or substantially remains at its preloading value D2 from the rest condition to a predetermined high speed value, and the diameter increases when exceeding such a predetermined high speed value. Alternatively, the diameter of the shaft 11 may continuously increase with the increase of the rotational speed at a rate lower than that which would be considered in the absence of one or more preloading elements 60. Thus, the preloading diameter D2 may correspond to the diameter that the shaft 11 is considered to naturally take under the functional high-speed rotation speed.

[0034] The preloading diameter D2 may be locally provided at a specific location of the shaft 11 such as at or near the end portion. Alternatively, the preloading diameter D2 may be provided along the entire length of the shaft 11, resulting in a homogeneous cylindrical preloaded shaft. Alternatively, the preloading diameter D2 may be provided along some portions of the shaft 11.

[0035] The preloading element 60 according to the present disclosure can be any insertable element that provides a mechanical force from its internal space 14 outwardly against the wall of the shaft 11 to urge the wall of the shaft 11 to expand mechanically. The preloading element 60 can take the form of a ring having a diameter D2' larger than the inner diameter of the shaft 11 (Figs. 3c, 3d). Alternatively, the preloading element 60 can be a cylinder inserted into the internal space 14 of the shaft 11 and having a diameter D2' larger than the inner diameter of the shaft 11. The preloading element 60 can have a conical outer shape or a truncated shape or a shape equivalent thereto that facilitates its insertion into the shaft 11. A given preloading element 60 can include several reinforcing elements 61 that enable it to be kept lightweight while maintaining its strength. The preloading element 60 may or may not be hollow. Any other arrangement can be used, provided that the preloading element 60 provides mechanical interference with the shaft 11. The preloading element 60 must be very well balanced in order to avoid or limit any vibrations or failures during operation. Considering the high rotational speed, the shape of the preloading element 60 preferably excludes any flat or sharp corners in order to avoid mechanical vulnerability. Specifically, the edges of the preloading element 60 are carefully rounded to reduce or avoid stress concentration at the boundary of the preloading interface.

[0036] According to one aspect, the diameter D2' of the preloading element 60 is about 0.05% to 0.1%, for example about 0.10% or 0.20% larger than the inner diameter of the shaft 11. According to another aspect, the diameter D2' of the preloading element 60 is about 10 micrometers to about 90 micrometers, for example about 40 or 50 micrometers larger than the inner diameter of the shaft 11. Those skilled in the art will understand that the mechanical interference can be appropriately designed according to specific requirements.

[0037] The preloading element 60 is preferably selected to have a low sensitivity to centrifugal force at high speeds, such that its expansion during operating operation is limited to remain low. However, it is worth noting that due to mechanical interference between the preloading element 60 and the shaft 11, the preloading element 60 itself is constrained, and thus its sensitivity to deformation conditions is reduced.

[0038] Preferred preloading elements 60 have low thermal expansivity. These preloading elements may include or be made of a ceramic material or other equivalent material having a low coefficient of thermal expansion. The preloading element 60 may have a coefficient of thermal expansion lower than or equal to that of the shaft 11 so as to reduce the thermal expansion of the assembly of the shaft and the preloading element 60 compared to the case of the shaft 11 alone, at least as long as the preloading element 60 remains in contact with the shaft 11.

[0039] Similarly, with respect to the mechanical expansion of the diameter at the rotational speed, the preloading element 60 is preferably defined to remain in direct contact with the wall of the shaft 11 over the entire range of high functional rotational speeds. This is particularly suitable when the preloading element 60 is forced inside the shaft while remaining in a state where it has no other fixed points other than direct contact with the wall of the shaft 11. However, it is possible to combine the preloading element 60 with a fixing element integrated with the shaft 11 so as to be maintained even when the diameter of the shaft 11 becomes larger than the preloading element 60. Such a fixing element may be a central passage inside the shaft that can fix the preloading element upward. Alternatively, the fixing element may be one or more recesses or protrusions on the inner wall of the shaft 11.

[0040] According to a preferred arrangement, the preloading element 60 may include or be made of, for example, silicon nitride SiN or Si3N4 or an equivalent material. In another arrangement, the preloading element may be made of the same material as the shaft so that there is no difference in thermal expansion.

[0041] As an example, the rotor of the present disclosure made of tungsten carbide and having a nominal diameter of 30 mm and an inner diameter of 25 mm has an increase in nominal diameter of 25 micrometers at 140,000 revolutions per minute. To stably support the rotor 10, a gap of 11 micrometers is required. By including a preloading element 60 made of silicon nitride and providing a mechanical interference of about 50 micrometers, it is possible to limit the increase in the diameter of the shaft 11 to about 3 micrometers.

[0042] The gas working fluid L can theoretically be any gaseous fluid, specifically hydrofluorocarbons (HFCs), hydrochlorofluorocarbons (HCFCs), CO2, ammonia, etc., or combinations thereof if possible, which are refrigerant gases used in industrial cooling devices. Ammonia offers the advantage of a large latent heat. Therefore, compared to conventional refrigerant fluids, the mass flow rate required to provide the same cooling capacity is less. When the mass flow rate decreases and the speed of sound increases, there is a tendency to direct the ideal rotor speed towards a higher value in order to maximize efficiency for a given load. The rotors and turbo compressors described herein are particularly suitable for using ammonia as the gas working fluid L.

[0043] According to one aspect of the present disclosure, two radial compressor stages are provided to achieve an overall pressure ratio of 6.5. This arrangement makes it possible not to exceed the mechanically limited impeller tip speed, which can be approximately 500 ms -1 ~600 ms -1 For a cooling device with a cooling capacity of 250 kW, an ideal rotor speed of 130 krpm can be provided for each stage with an impeller tip diameter of approximately 50 mm. Due to rotational mechanical constraints, the two compressor stages are preferably not assembled on the same rotor, and thus the cooling device has to be driven by two individually driven compressors. This additional cost provides greater flexibility and improved off-design performance. The compressors according to the present disclosure can of course include only one stage or three or more stages. Since each rotor can be driven individually, all rotors can be the same or different.

[0044] The present disclosure also encompasses a method of manufacturing a rotor 10 combined with one or more preloading elements 60. Specifically, the preloading element 60 can be forced into the internal space 14 of the shaft 11 by applying a high mechanical pressure to urge the preloading element 60 to fit inside the shaft 11. Such a press-fitting process can be performed at room temperature, such as about 20°C. Alternatively, the preloading element 60 can be press-fitted at a higher temperature, such as above about 100°C or above about 200°C, to thermally provide an increase in the diameter of the shaft 11. Alternatively, liquid pressure can be provided within the internal space 14 of the shaft 11 to mechanically increase its diameter, and the preloading element 60 can be inserted into the internal space 14 either under pressure or under ambient pressure conditions. The external geometry of the preloading element 60 can be adapted to facilitate its insertion into the shaft 11. For this purpose, it can have a conical or truncated shape. According to one aspect of the present disclosure, the method includes the step of preloading the shaft 11 using either thermal or mechanical conditions to increase its intrinsic diameter D1. The method further includes the step of inserting one or more preloading elements 60 into the internal space 14, with or without pressure. Preferably, the insertion step is performed while the preloading condition is applied to the shaft 11. The method also includes the step of removing the preloading after one or more preloading elements 60 have been inserted. After this step, the diameter of the shaft 11 is maintained at its preloading value D2 through the inserted preloading element 60.

[0045] The present disclosure also encompasses a method for cooling operations, such as cooling or refrigerating an industrial room, using a high-speed rotating rotor, such as the rotor described herein.

[0046] This disclosure also encompasses a method for managing and controlling the stability of a rotor at a rotational speed exceeding 170,000 revolutions per minute. In fact, the reduction of clearance distortion caused by centrifugal load and thermal load is significantly reduced by this concept, and thus the machine becomes much more robust. By minimizing the clearance distortion, a gas bearing-supported rotor, which is prone to becoming rotationally dynamically unstable, becomes more stable.

[0047] In this disclosure, the term "load" has the same meaning as "stress" and refers to any external restraint that results in an increase in the diameter of the shaft of the rotor 10. Such restraints include centrifugal force, temperature rise, and combinations of both.

[0048] In this disclosure, the expression "operating operation" at least preferably means the rotational movement of the rotor 10 at its operating speed or functional speed, which is preferably the speed at which the rotor provides the required compression output. This expression also involves the flow of the gaseous working fluid L around the shaft 11 to maintain it centered on the shaft 11 during rotation.

[0049] In this specification, the expression "preloading element" can be used in the singular or plural. It is understood that the turbo compressor can include one or more such preloading elements, for example, 2, 3, 4, or 5 or more. Similarly, if there are multiple preloading elements 60 in the turbo compressor, they can all be the same or, on the contrary, different from each other. The difference between two preloading elements can relate to their shape including their dimensions, and mechanical interference with the shaft 11, their composition, or properties defined by any of the corresponding Young's modulus, or density, yield stress, thermal expansion.

Explanation of Reference Signs

[0050] 1 Turbo Compressor 10 Rotor 11 Shaft 12 Protrusion 13 Surface Area 14 Internal Space 20 Bearing 30 Compressor impeller 31 Compression blade 40 Housing 50 Inlet 60 Preloading element A Longitudinal axis D1 Natural diameter D2 Preloading diameter L Gas working fluid

Claims

**Claim 1**: A high-speed rotor (10) having a hollow shaft (11) and an impeller (30) that defines an internal space, the high-speed rotor being such that the hollow shaft has a specific diameter (D1) and a preload diameter (D2) that is larger than the specific diameter, and including at least one preload element (60) that maintains the preload diameter in a state where there is no external restraint, the preload element being an insertable element that provides a mechanical force from the internal space to the outside to the wall of the shaft. A high-speed rotor characterized by this. **Claim 2** The high-speed rotor according to claim 1, wherein the preload diameter (D2) corresponds to the diameter that the shaft (11) naturally assumes under high-speed rotation around its longitudinal axis (A) or at high temperature. **Claim 3** The high-speed rotor according to claim 2, wherein the high-speed rotation speed has a DN value of 3 million or more. **Claim 4** The high-speed rotor according to any one of claims 1 to 3, wherein the shaft (11) is hollow so as to define an internal space (14), and the at least one preload element (60) is inserted into the internal space. **Claim 5** The high-speed rotor according to claim 4, wherein the preload element is a ring or a cylinder having a diameter (D2') larger than the specific internal diameter of the shaft (11). **Claim 6** The high-speed rotor according to claim 4 or 5, wherein the preload element (60) is maintained in the shaft (11) using only direct contact with the wall of the shaft (11). **Claim 7** One or more shafts (11) and preload elements (60) have a specific modulus of elasticity of 0.01 to 0.5 GPa / kg / m ３ and a specific strength of 0.01 to 0.9 MPa / kg / m ３ The high-speed rotor according to any one of claims 1 to 6, characterized in that the specific modulus of elasticity represents the ratio between Young's modulus and the material density, and the specific strength represents the ratio between the yield strength and the material density. **Claim 8**: The high-speed rotor according to any one of claims 1 to 7, wherein the diameter of the hollow shaft varies by less than 0.5% when the rotational speed of the shaft increases from the rest condition to the speed condition of functional high-speed rotation. **Claim 9** The high-speed rotor according to any one of claims 1 to 8, wherein the edge of the preloading element (60) is rounded. **Claim 10** In a method for combining a rotor (10) with at least one preloading element (60), the rotor (10) has a hollow shaft (11) defining an internal space having an inner diameter, the preloading element (60) has a diameter (D2') larger than the inner diameter, the step of preloading the shaft to increase its natural diameter, and the step of inserting the at least one preloading element (60) into the internal space (14) under preloading conditions. **Claim 11** The method according to claim 10, wherein the preloading conditions include or define either a thermal condition or a fluid pressure condition that can temporarily increase the natural diameter (D2) of the shaft. **Claim 12** The method according to any one of claims 10 or 11, wherein the preloading conditions include or define the mechanical pressure applied on the preloading element to press-fit the preloading element (60) into the internal space (14) of the shaft (11). **Claim 13** In a device including at least one high-speed rotor according to any one of claims 1 to 9, the at least one high-speed rotor is integrated inside a housing (40) including an inlet (50) such that the compressor impeller (30) faces the inlet (50), and is positioned on the bearing in such a way that it rotates freely without contacting the bearing (20) under the flow of the gaseous working fluid (L). **Claim 14** The device according to claim 13, wherein the gap between the shaft (11) of the at least one rotor (10) and the surrounding bearing (20) is less than 30 micrometers. **Claim 15** The device according to claim 13 or 14, selected from a turbo compressor, a turbine, a fuel cell recirculation device, an optical scanner, and an inertial gyroscope.

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