Improved micro gas turbine unit and operating method the same

TECHNICAL FIELD

The present invention relates to the field of micro gas turbine units.

This term means turbomachines usually comprising a compressor, a combustion chamber and a small turbine, typically working at a rotational speed even of 100,000 rpm.

Typically, not exclusively, such machines are made according to the standard ISO 3977-1:1997 and have dimensions compatible with powers ranging from 28 kW to about 100-200 kW.

PRIOR ART

In micro gas turbine units the rotor of the compressor and the rotor of the turbine are mounted on the same (single) shaft.

Below reference is made to two different types of layout both known in prior art which are denoted for convenience reasons as "conventional" and "modern" layouts: the conventional layout is based on high precision rolling bearings, while the modern layout uses aerodynamic bearings, of the air lubricated foil type.

In the conventional layout the shaft is held in position with respect to a frame, free to rotate, by means of rolling bearings, often angular contact bearings.

Due to the high number of revolutions in the operating phase of a micro turbine, such rolling bearings often work as much as they can as regards their mechanical properties: this derives from the fact that such bearings have certain limits (speed of use, operating temperatures, fatigue) that are not suitable for being used in the field of micro gas turbine units.

It has to be noted that under operating conditions the rotational speed of commercial micro turbines range from 43,000 to 116,000 rpm, while temperatures inside the micro turbine range from lOODEGC to lOOODEGC and they thermally stress bearings in a different manner depending on their positioning and on the cooling system, if any.

A particularly important drawback regards axial loads: fatigue-related limits in the life of the rolling bearings are not suitable for the application of interest .

It has to be noted, incidentally, that in a micro gas turbine unit the (axial) thrusts of the compressor and of the turbine are the main loads, since usually they are higher than radial loads (being the weight of the compressor and turbine rotors) by an order of magnitude. In order to better show this aspect, the table 1 below, shows the average approximate design data under nominal conditions of a micro gas turbine system.

table 1

the following table 2 on the contrary, shows the lifespans Li , Li o and Ls o (corresponding to 99%, 90% and 50% of reliability) for sets of known high precision angular contact rolling bearings, subjected to loads shown in table 1.

table 2

Such lifespans have to be ideally compared with the lifespan of a micro gas turbine unit, that currently is 60,000-80,000 hours.

As it is immediately clear, the time of use anyway is considerably lower than the life of the unit itself, resulting in replacing the bearings during the useful life of the unit, resulting in shutting down the machine, in manufacturing loss and in general losses as regards replacement time/costs.

In order to partially overcome such drawback the solutions related to the modern layout have been suggested wherein in micro gas turbine units the rolling bearings have been replaced by air foil bearings (known in itself in prior art) .

It has to be noted that by means of such solution some of the above problems are at least partially solved. In particular a more reliable, eco-friendly (oil-free) operation and a reduction of the problems related to high operating temperatures (design, refrigeration, reliability) are obtained. Although such solution has improved the general prior art, however it has still some drawbacks.

Such drawbacks are related mainly to the fact that air foil bearings have a life dependent from the number of starts and stops of the unit.

Such aspect is quite important in the field of micro gas turbine units, since due to their nature such units have relatively short operating cycles, with many starts and stops even in the same day.

The air foil bearing exerts its function when the rotational speed is such to generate an aerodynamic lift thereof .

If the rotational speed of the shaft is not enough to generate the complete aerodynamic lift of the bearing, a sliding contact is generated between the foils of the air foil bearing, which contact inevitably generates a given wear.

As it is easy to understand such not optimal operating condition is generated both when reducing the rotational speed of the shaft, getting near a stop, and in the event of a start, from the stop condition, of the unit.

That is to say the suggested solution, that provides air foil bearings, is optimal in a continuous operating condition, but it has still some drawbacks if it is employed on plants (such as micro gas turbine units) that are frequently started/stopped.

In order to reduce such drawback, therefore, it is necessary to lubricate (by high temperature solid lubrication) the foils and the pin of the air bearing. Considering that conventional solid lubricants (graphite and moly-disulfide) are limited to 150DEGC therefore they are not suitable for this aim and therefore innovative lubricants are used, such as polymer films and special coatings that are able to lubricate also at maximum temperatures of about 650DEGC.

The efficiency of such innovative lubricants however is optimal at a high temperature (start or stop with the machine in warmed up condition) , but it strongly decreases at ambient temperature (start at cold condition) .

Even if such solution has partially overcome the drawback, however the lifespan (particularly in the case of micro gas turbine units with frequent starts/stops in the same day) is limited on the whole.

Another drawback of foil bearings is their high costs. Still another drawback is the low load capacity (specific bearing load of 0,7 Mpa against 2Mpa of rolling bearings) and the impossibility of a retrofit (installation on already existing units) : the fact of selecting air foil bearings involves a design philosophy completely different from that of rolling bearings (for example shafts will have to have a higher diameter such to guarantee the sufficient peripheral speed allowing a suitable aerodynamic lift) .

Then it is necessary to select specific rapid starting procedures, to avoid unacceptable wear of the foil bearings .

Finally it has to be considered that foil bearings have to be suitably designed for each specific application; this makes their availability reduced on the whole.

From the above it is clear how such drawbacks reflect on the lack in the development of micro gas turbine technology, that on the contrary can be a valid solution for the autonomous production of localized energy.

OBJECTS AND SUMMARY OF THE INVENTION

It is the general object of the present invention to overcome the drawbacks of prior art.

In particular it is the object of the present invention to raise the life of bearings, while reducing the need of maintenance, such to contribute in spreading such machines .

Another object of the invention is to suitably distribute thrusts, such to suitably load the bearings.

Still another object of the invention is to distribute axial loads on bearings, preferably by reducing the load of the main thrust bearing.

It is a further object of the present invention to provide a micro gas turbine unit that is relatively simple to be made, strong, reliable and that needs a relatively limited maintenance.

Another object of the invention is to provide a solution able to release fluid dynamic design of micro gas turbine unit from design constraints usually present as regards axial thrusts of the turbine and compressor, such to allow designing intended to achieve better performances and efficiencies.

These and further objects of the present invention are achieved by a micro gas turbine unit embodying the characteristics of the annexed claims, which are an integral part of the present description.

The general idea at the base of the present invention provides to make a micro gas turbine unit comprising a frame, a compressor part provided with at least one compressor rotor, a turbine part provided with at least one turbine rotor, which compressor rotor and turbine rotor are connected by a shaft, wherein according to the invention in at least one operating condition of the unit the turbine rotor is coupled to the frame such to discharge thereon all or a part of an axial load.

According to a preferred embodiment the turbine rotor is coupled to a first end of the shaft by a first spline coupling, said first spline coupling being a helical spline, which helical spline ends at an annular shoulder of the shaft, and wherein the compressor rotor is coupled to a second and opposite end of the shaft by a second spline coupling, there being further provided:

- at least one main thrust air bearing comprising first and second aerodynamic elements intended to generate an aerodynamic lift of the first elements relative to the second elements above a rotational speed of relative lift between said first and second elements, said first elements being fastened to the turbine rotor and said second elements being fastened to the frame, such to discharge thereon an axial thrust generated by the turbine rotor in an operating condition wherein the turbine rotor rotates at a speed higher than or equal to said lift speed,

- an auxiliary radial/thrust bearing, interposed between the frame and the shaft near the spline coupling.

Thus, in practice, the main thrust air bearing upon the self-lift phase (once the lift speed is exceeded) discharges the thrust on the frame while under sliding contact conditions (at low speed) it discharges it on the shaft.

Thus the air bearing is not more affected by wear due to repeated start/stop cycles. The provision of the air bearing, together with the manner it is mounted and with the provision of the helical spline (wherein grooves form a helix around the shaft) in practice makes a kind of "load partitioner" . The aim of the "load partitioner" is to distribute the axial load between the auxiliary thrust bearing and the main air bearing.

In substance, at the start and up to the generation of the aerodynamic lift (that is below the lift speed) the auxiliary bearing supports the axial load while the main thrust bearing works as relieved and thus with a minimum or null wear. In such condition the thrust of the turbine is exerted by the turbine rotor on the shoulder of the shaft rather than on the air bearing. Through the shoulder and the shaft such thrust is then transferred to the support of the shaft, that is to the auxiliary bearing, as in a conventional bearings-rotor layout. As soon as the lift speed is reached or exceeded, on the contrary, the main air bearing starts to bear and it relieves the auxiliary bearing: the shoulder of the shaft does not receive any more the thrust since the turbine rotor, forced by the aerodynamic pressure generated in the film of the main bearing, moves on the spline coupling and moves away from the shoulder.

Thus the load is transferred between the two bearings such to prevent the main bearing from being worn during the start/stop phase of the machine and such to relieve the auxiliary bearing of the axial load under full operating condition.

In practice, and stated in other words, the splined profile on the turbine side acts as a load partitioner in two manners :

- transfer of load from start/stop conditions in full operating condition involving a micrometric movement of the rotor on the spline (for this function it does not need to be helical) ;

- generating an axial thrust due to the moment exerted by the fluid on the blades of the turbine (it requires the spline to be an helical spline) ; such thrust modifies the distribution of the load under full operating condition between the main and auxiliary bearing.

Actually the major function as load partitioner is the second one of the two manners just described that is the function carried out by a helical spline: theoretically speaking the first manner requires a micrometric axial movement (since the aerodynamic film is in the order of a micron) .

Under full operating conditions, if the spline coupling is not a helical spline (for example by employing a straight spline also on the turbine side) , the thrust of the turbine Ft would be completely supported by the main thrust bearing as well as the compressor thrust Fc would be supported by the auxiliary bearing.

On the contrary by using as the actuator the helical spline coupling (turbine side) according to the invention, it is possible to transfer a portion of the thrust of the turbine from the hub to the shaft. Thus any desired distribution of the axial load between the two thrust bearings (main and auxiliary bearings) is obtained depending on a single design parameter, namely the helix angle of the spline.

Therefore the helical spline acts as an actuator since the torque exerted by the working fluid on the blades of the rotor is translated into an axial thrust on the shaft due to the kinematic bond between rotation and axial movement characteristic of the helical pair, where such movement depends on the helix angle.

Briefly by means of the spline coupling the transfer of the load between the two bearings is achieved such to guarantee an optimal life of the thrust /radial auxiliary bearings and to reduce loads on the air thrust bearing. According to an advantageous characteristic, dependent or independent from the previous ones, the second spline coupling is a straight tooth spline coupling.

According to a further advantageous characteristic, dependent or independent from the previous ones, the main thrust air bearing is an air foil bearing, wherein the first aerodynamic elements comprise a washer and the second aerodynamic elements comprise a plurality of corrugated foils.

According to a further advantageous characteristic, dependent or independent from the previous ones,

- with the turbine in the stopped condition or rotating at a speed lower than the lift speed, the turbine rotor is in contact with the annular shoulder of the shaft, such that an axial thrust of the turbine rotor is discharged on the shaft through the annular shoulder;

- with the turbine rotating at a speed equal or higher than the lift speed, the turbine rotor is moved away from the annular shoulder of the shaft, such that an axial thrust of the turbine rotor is discharged partially on the frame through the main thrust air bearing and partially on the shaft through the helical spline coupling . According to a further advantageous characteristic, dependent or independent from the previous ones, the unit further comprises a third radial bearing coupled between the frame and the shaft near the main thrust bearing, more preferably placed between the compressor and the turbine such to reduce their operating temperature .

The coupling of the third radial bearing with the frame is a slide coupling, where the third radial bearing is free to axially slide with respect to the frame, such to permit an axial expansion of the shaft.

According to a further advantageous characteristic, dependent or independent from the previous ones, at least one of the auxiliary radial/thrust bearing and the third radial bearing (if any) comprises a rolling bearing.

According to the above characteristic the rolling bearing is preferably a double direction bearing and more preferably it comprises a plurality of single-row angular contact bearings.

As an alternative or in combination at least one of the auxiliary radial/thrust bearing and the third radial bearing (if any) comprises a magnetic bearing.

According to a further advantageous characteristic the helix angle β of the helical spline coupling ranges from 45 DEG to 135 DEG.

Accordingly, another object of the invention is an operating method of a micro gas turbine unit, said unit comprising a frame, a compressor part provided with at least one compressor rotor, a turbine part provided with at least one turbine rotor, which compressor rotor and turbine rotor are connected by a shaft, wherein the turbine rotor is coupled to a first end of the shaft by a first spline coupling ending at an annular shoulder of the shaft

there being further provided at least:

- at least one main thrust air bearing comprising first and second aerodynamic elements intended to generate an aerodynamic lift of the first elements relative to the second elements above a rotational speed of relative lift between said first and second elements, said first elements being fastened to the turbine rotor and said second elements being fastened to the frame,

- an auxiliary radial/thrust bearing, interposed between the frame and the shaft near the compressor,

said operating method providing for at least

- a first phase, in which the shaft rotates at a rotational speed lower than said rotational speed of relative lift between said first and second elements, in which first phase the thrust of the turbine is exerted on the annular shoulder of the shaft, so that the auxiliary bearing supports the axial load,

a second phase, in which the shaft rotates at a rotational speed equal or higher than said rotational speed of relative lift between said first and second elements, in which second phase the thrust of the turbine is exerted on said second elements.

Additional (and optional) features of the method will became apparent from the following description.

With reference now to the difference between conventional and modern layouts mentioned above, other advantages deriving from the present invention are summarized below:

- the foil bearing in the solution of the invention has no limits on start-stop cycles; this is a considerable advantage compared to the modern layout;

- rolling bearings in the solution of the invention have a life and reliability considerably higher than those of a conventional layout;

- the support system in the solution of the invention has fewer temperature-related problems, compared to the conventional layout, the most stressed rolling bearing has lower loads and, since it is on the compressor side, it is subjected at a less extent to temperature-related problems downstream the combustion chamber and upstream the turbine; the main thrust foil bearing, placed in a hotter area, usually withstands temperatures much higher than rolling bearings (foil bearings withstand above 300DEG more than rolling bearings and such margin can be further increased considering that the innovation makes not important the role of the tribology coating in foil bearings) ;

- the unit (or machine) in the solution of the invention can work under optimal fluid dynamic conditions, since bearings are a less compelling limit for thrusts exerted by turbine and compressor;

- radial bearings in the solution of the invention are not of the sliding type (hydrodynamic or aerodynamic) , as in the modern rotor-bearing layout completely based on air foil bearings, accordingly, the dynamics of the rotor is not affected by instability problems that affect sliding bearings, known in literature with the name whirl, whip and pneumatic hammering, no further reference being made thereto here, referring to scientific texts for an appropriate analysis;

- compared to the usual interference fit of rotors of turbine and compressor, the spline couplings suggested in the solution of the invention allow the rotors and the bearings to be easily disassembled for overhaul and maintenance reasons; such characteristic is particularly interesting for a small machine as a micro gas turbine unit ;

- it is possible to perform retrofit of the conventional layout, that is different from the one of the solution of the invention due to the lack of spline couplings and of the foil thrust bearing;

- in the solution of the invention, in the design phase, the use of the helical spline coupling allows loads acting on the most stressed bearings to be suitably distributed by means of a single parameter (the helix angle) ; such degree of design freedom is not available in the known conventional layout;

- in the solution of the invention it is not necessary a device for the rapid start-up of the unit and materials for the solid lubrication to prevent the foil bearing from getting worn, which are necessary in the modern layout .

- the solution of the invention comprises variants based on a support of the shaft different from rolling bearings; for example "oil-free" implementations of the suggested solution are possible, based on the use of magnetic bearings (active ones or active-passive combined ones) for supporting the shaft, in addition to the main thrust foil bearing;

- the "oil-free" arrangement of the present invention has expected performances higher than the known layout, since it would not have a life limited by the number of start/stop cycles or any other wear problems and it would have even less frictions;

the "oil-free" arrangement of the solution of the invention has expected performances higher than a magnetic layout based, without distinction, on active or active-passive combined bearings; since the major portion of (axial) load is supported by the foil bearing, the power input is lower than the magnetic layout;

- performances of the solution of the invention are higher than known solutions and they fully justify possible higher costs;

- the costs expected in implementing the solution of the invention are lower than those of known solutions with air bearings in connection with comparable and better performances under the most common operating conditions; lifespans are higher if the unit is frequently started, they are lower (only in the event of using rolling bearings) if the operation is mainly under a full operating condition, where the first situation is the most common for a micro turbine.

the "oil-free" arrangement in the solution of the invention allows a system controlling the dynamics of the rotor to be incorporated by using active magnetic bearings for supporting the shaft, a thing that is not possible in the modern layout.

Further advantageous characteristics are the subject matter of the annexed claims, that are an integral part of the present description.

BRIEF DESCRIPTION OF THE DRAWINGS The invention will be described below with reference to some not limitative embodiments, provided by way of example and not as a limitation in the annexed drawings. These drawings show different aspects and embodiments of the present invention and, where appropriate, reference numerals showing like structures, components, materials and/or elements in different figures are denoted by like reference numerals.

In the annexed figures:

Figure 1 is a diagram of a micro gas turbine unit according to the invention;

Figure 2 is a graph about the selection of parameters for implementing the micro gas turbine unit according to the invention under a first design condition in which compressor and turbine thrusts are directed toward the inner side of the machine;

Figure 2A is a graph about the selection of parameters for implementing the micro gas turbine unit according to the invention under a second design condition in which compressor and turbine thrusts are directed toward the outer side of the machine;

Figure 3 is - perspective view - an assembly of compressor rotor / shaft / turbine rotor of the micro gas turbine unit according to the invention in a partially disassembled condition;

Figure 4 is -longitudinal section - the turbine rotor and part of the shaft of the assembly of fig 3;

Figure 5 is - in perspective view - a part of the shaft of fig.4 provided with a helical spline;

Figures 6 and 7 are - partial section - a part of the shaft and the turbine rotor of the previous figures in two conditions, with the turbine rotor in the stop condition and with the turbine rotor rotating at a full operating speed;

Figure 8 is - partial section - a part of the shaft and the compressor rotor of the previous figures;

Figures 9-11 are details of an air bearing of the previous figures;

Figure 12 is a perspective view in partial phantom of the turbine rotor comprising a third bearing,

Figure 13 is a section of another embodiment of the invention ;

Figure 14 is an enlarged sectional view of a particular of the embodiment of fig.13.

DETAILLED DESCRIPTION OF THE INVENTION

While the invention is susceptible of various modifications and alternative forms, some preferred embodiments are shown in the drawings and will be described below in details.

It should be understood, however, that there is no intention to limit the invention to the specific embodiment disclosed, but, on the contrary, the intention of the invention is to cover all modifications, alternative constructions and equivalents falling within the scope of the invention as defined in the claims.

The use of "for example", "etc", "or" indicates non ^¬ exclusive alternatives without limitation unless otherwise defined.

The use of "comprise" means "comprise, but not limited to, " unless otherwise defined. Terms as "vertical" and "horizontal", "upper" and "lower"

(with no other indications) have to be read with reference to the assembling (or operating) conditions and with reference to the standard terminology in use in common speech, where "vertical" means a direction substantially parallel to that of the vector of the force of gravity "g" and horizontal means a direction perpendicular thereto.

With reference to the annexed figures, they show a micro gas turbine unit 1 comprising a frame 2, a compressor part 3 provided with at least one compressor rotor 31, a turbine part 4 provided with at least one turbine rotor 41.

According to a feature of the invention, with the unit 1 in at least one operating condition, the turbine rotor 41 is directly coupled to the frame such to discharge thereon all or a part of an axial load.

The term "directly coupled" means that it is not coupled only, or solely, to the frame, by the shaft 5.

In particular according to the embodiment of the annexed figures, the compressor rotor 31 and the turbine rotor 41 are connected by a shaft 5: the turbine rotor 41 is coupled to a first end of the shaft 5 by a first spline coupling, that in the example is an helical one 51, ending at an annular shoulder 53 of the shaft 5.

Following what disclosed above, limiting the advantages of the invention and versatility thereof, instead of the helical spline it is also possible to use a first straight spline coupling. On the contrary the compressor rotor 31 is coupled to a second - and opposite - end of the shaft 5 by a second spline coupling 52, in this non limitative example a straight tooth spline.

The spline couplings 51 and 52 are free, not forced, that is they allow the parts to carry out a relative movement along the grooves.

Both of them can be made with profiles both of the straight-sided type and involute type.

To this end the diameters are selected such not to cause interference and in particular, with reference to DIN 5480 standard currently in force, they are preferably of the H/e, H/f, H/h type.

As regards the spline on the turbine side, then, there is preferably provided the possibility of performing an axial micrometric movement by providing a suitable axial clearance to allow the rotor and the washer integral thereto to be detached from the annular shoulder when passing from the start conditions to full operating conditions.

As it can be noted in the annexed figures there is further provided at least one main thrust air bearing 61.

Such bearing 61 comprises first 611 and second 612 aerodynamic elements intended to generate an aerodynamic lift of the first elements 611 relative to the second elements 612 (or vice versa, in the same manner) upon reaching or exceeding a rotational speed of relative lift between the first and second elements 611, 612. The first elements 611 are fastened to the turbine rotor 41 while the second elements 612 are fastened to the frame 2, such to discharge thereon an axial thrust generated by the turbine rotor 41 in an operating condition where the turbine rotor rotates at a speed higher than or equal to said lift speed.

The unit 1 then comprises the auxiliary radial/thrust bearing 62, interposed between the frame 2 and the shaft 5 near the spline coupling 52.

Generally speaking, therefore, the radial/thrust bearing 62 (preferably angular contact ball bearing) is placed on the compressor side, as an auxiliary axial and radial support, while the thrust air bearing 61, preferably a foil bearing, is placed on the turbine side, as the main axial support.

The connections with the shaft 5 of rotors 31, 41 of the compressor and of the turbine are a conventional spline coupling 52 (grooves with straight generatrices) and a helical spline 51 (grooves with helical generatrices) respectively. The latter, as it will be disclosed in details below, acts as a "thrust partitioner" between the main thrust bearing 61 and the radial/thrust one or auxiliary one 62.

The bearing 61 has pads or foils 612 fastened to the frame and the washer 611 integral with the turbine rotor 41; their relative position and the preloading of the foils can be adjusted by means of spacers.

Such main thrust bearing 61 in full operating condition discharges the thrust (preferably the most considerable portion) directly on the frame 2 (and not on the shaft 5) .

Thus the average life of bearings 62, limited due to fatigue reasons almost only by the axial load (that with the machine in optimal performance conditions is an order of magnitude higher than the radial one) , is much higher than the life of the machine, since under full operating conditions the major portion of the axial load (coming from the turbine) is supported by the air thrust bearing 61 and directly discharged on the frame 2.

Such portion is variable and it is defined in design phase by selecting the helix angle of the spline 51. Moreover the life of such main thrust bearing 61, limited only by the number of starts/stops of the machine, is practically infinite, since the load is completely borne by the angular contact bearing 62 till the generation of the speed forming the air film dynamized by virtue of the positioning of the foils or pads 612 on the frame 2. When the foil bearing 61 has no aerodynamic lift, the thrust of the turbine rotor 41 acts on the shoulder 53 of the shaft 5.

After the generation of the air film, on the contrary, its micrometric thickness causes the washer 611 to move by the same amount such that the contact on the shoulder 53 does not occur anymore and the only transfer of axial load to the shaft 5 occurs through the surfaces of the helical spline 51, depending on the design helix angle.

Now with reference - more analytically - to the description of the forces acting in the various operating steps of the unit 1 just described (even if generally speaking) the positioning of the main thrust bearing 61 is such to discharge the thrust of the turbine:

On the frame 2 through the aerodynamic bearing 61 upon reaching and/or exceeding the lift speed

On the shaft 5 (and later on the frame through the angular contact bearing 62) at rotational speed of the turbine lower than the lift one (that is at low speed, during the start and stop steps) .

Thus the air bearing 61 is not more affected by wear determined in the start/stop steps, as on the contrary in known solutions.

In the preferred, but not limitative, case the main thrust air bearing 61 is an air foil bearing, wherein the first aerodynamic elements 611 comprise a washer and the second aerodynamic elements 612 comprise a plurality of pads composed of a plurality of corrugated foils 614 and a smooth top foil 613, such as shown in figs. 9-11. The second aerodynamic elements 612 are fastened to the frame such as shown in fig.4, while the washer is placed on the rear part of the turbine rotor 41, such that the opposite surfaces of the sliding kinematic pair (that is the top foils of the pads and the surface of the washer) , are separated by a given clearance ca .

In a preferred embodiment there is also provided a spacer 21, seen in fig.4, allowing such clearance ca to be regulated such to guarantee the optimal operation of the foil bearing.

In particular, to this end, the two surfaces of the pair of the air foil bearing (the smooth top foils 613 of the pads and the washer 611) are preferably in contact when the rotational speed of the turbine is zero (unit 1 in stop condition) , such that the elasticity of the corrugated foils exerts a given (slight) preloading. The clearance ca therefore is provided (slightly) negative (the structure composed of the foils 612 is elastic) .

Accordingly no considerable wear will be present upon starting and stopping the machine, since in such steps the thrust acting on the main thrust bearing is very light, that is no external load is added to the preloading of the foils 612.

When the foil bearing 61 has no aerodynamic load capacity (below lift speed) , the thrust of the turbine rotor 41 acts on the shoulder 53 of the shaft 5. Differently after forming the air film and after generating the relevant aerodynamic pressure, its micrometric thickness causes the washer 611 to be moved by the same amount and the turbine rotor 41 to be moved in the axial direction (with reference to the shaft) such that the contact between the hub of the turbine rotor 41 and the shoulder 53 does not occur anymore.

Accordingly under nominal conditions and starting from a rotational speed sufficient for the aerodynamic self- lift (higher than the lift speed) , the only transfer of the axial load from the turbine rotor 41 to the shaft 5 can take place through the helical spline 51, depending on the design helix angle.

Such helix spline 51 acts as a "thrust partitioner" between the main thrust bearing 61 and the auxiliary bearing 62.

The principle of the operation of the "thrust partitioner" that is the load distribution law, is shown below .

The load distribution law has been defined which is followed under full operating condition by the pair of helical splines used as mechanical actuator besides as a simple coupling member. With reference to the diagram in fig.l it schematically shows the forces acting under nominal conditions on the components of the rotor according to the changes given by the innovation.

Vectors Ft and Fc are the thrust Ft of the turbine and the thrust Fc of the compressor, respectively, generated by the pressure of the fluid on the relevant blades. The force Ft-R acts on the bearing 61.

Vectors R show the actions and (equal) reactions that the turbine rotor exerts on the shaft through the helical spline coupling 51.

The axial force Fc-R is the thrust to which the shaft is subjected and that will be supported by the auxiliary thrust bearing 62.

Finally the moment Mt is the resisting torque of the turbine, due to the action of the pressures on the blades thereof .

Inventors have analysed the different scenarios as the helix angle β of the spline 51 changes from 45DEG to 135DEG, where a 90DEG value corresponds to a spline with straight generatrices (straight teeth) .

Firstly it has to be noted that the transfer of the load R from the hub to the shaft through the splined surfaces due to the axial thrust Ft has a small amount (in the order of 6.7% of the thrust for to which the maximum transfer corresponds) and it is due to the deformation of the members of the pair.

That is to say, if the hub and shaft were perfectly rigid, the thrust of the turbine would not be transferred at all to the shaft through the splined surfaces, but it would be discharged completely on the main thrust bearing 61.

Secondly the transfer of the load R from the hub to the shaft through the splined surfaces due to the moment Mt is regulated by the following law valid for perfectly rigid members of the pair

Af, _r

H = ¹

tan wherein rp is the radius of the pitch circle in the event of involute spline profiles or the inner radius for straight-sided profiles.

In the event of deformable members, the law is met with a slight error (in the order to 0.6% of the transferred load for β =135DEG, to which the maximum load corresponds) .

By using the results from FEM simulations and assuming for simplicity reasons the members of the spline coupling to be rigid (assumption that does not result in very considerable errors, as just mentioned) , it is possible to establish the load distribution under full operating condition between the two thrust bearings by suitably selecting the helix angle.

To this end, figure 2 shows as a function of the helix angle β, for the data supposed above in table 1, the trend of the load transfer R (continuous thick curve) in line with the above equation.

On the basis of such trend, also the axial thrust Tt of the turbine on the main bearing 61 is drawn (dashed curve) and the thrust Ts of the shaft on the auxiliary bearing 62 is drawn, dash-point curve, with positive values for forces directed from the compressor to the turbine .

Such two thrusts are calculated by relations Tt= Ft-R and Ts=Fc-R respectively, that can be deduced from the analysis of figure 1, wherein the thrusts of turbine and compressor, Ft and Fc respectively, are obviously constant under full operating condition, denoted in table 1.

The continuous thin straight line on the contrary is the value of the reference thrust Tref=Ft-Fc, to which the only thrust bearing of a conventional plant is subjected. As it is clearly visible, by employing a straight tooth spline (β=90ϋΕΟ , the load on the main thrust bearing 61 is higher than the one related to a conventional plant, since it is equal to the whole thrust Ft of the turbine, while the auxiliary radial/thrust bearing 62 bears the thrust Fc of the compressor, directed from the compressor to the turbine (positive) .

Starting to decrease the helix angle β the loads on both the bearings 61, 62 are reduced, since the helix is oriented such to generate on the shaft a thrust R opposite to that of the compressor Fc .

By selecting β equal about to 82DEG, the main thrust bearing 61 operates under reference conditions (Tt=Tref ) , while the auxiliary one 62 is axially relieved and therefore, since it is subjected only to the radial load, it will have an optimal lifespan (in the order of one million of hours with reliability of 90%, as deduced from the calculation of the lifespan Li o for angular contact ball bearings subjected only to the radial load specified in Table 1) .

On the contrary with an angle β lower than 82DEG but higher than 76DEG (design value range), the main bearing 61 becomes more relieved compared to the reference condition, to the detrimental of the lifespan of the auxiliary bearing 62, that will receive from the shaft a negative thrust (from the turbine to the compressor) . For the thrust on the main thrust bearing 61 is cancelled and accordingly the thrust of the shaft on the auxiliary bearing 62 takes its maximum value, that is Ts=-Tref, resulting in the typical lifespan of a conventional layout.

In the design range of the helix angles, the maximum lifespan/reliability of the bearing 62 is achieved for the minimum one typical of the conventional layout for the most hard load on the main thrust bearing 61, equal to that of a conventional layout, occurs at the minimum (null) one for

The helix angle β is clearly shown in fig.5 and it corresponds to the acute angle measured between the tangent to one of the helixes drawn by the side of a tooth on the pitch cylinder and a plane perpendicular to the axis of the cylinder.

In more general terms, the graphical construction shown in fig. 2, aimed to find the possible design range for the helix angle β, consists of finding the zeros of the thrust functions Tt and Ts, i.e. the axial loads of the bearings (61) and (62), respectively, as a function of β. Such zeros locate the limits of β.

Let the positive direction of the axial forces be from turbine to compressor. The conceptual design of the innovative support system presented hitherto implicitly assumes that turbine and compressor thrust are positive and negative, respectively, as shown in fig. 1. In such case, only the axial forces exerted by the working fluid on compressor and turbine blades are considered and the total thrust load, computed as the algebraic sum of compressor and turbine axial forces, is positive. Nevertheless, compressor and turbine thrusts receive other contributions. Indeed, both such forces are the resultant of the axial thrusts exerted by the working fluid on both the blades and the backside of the impellers .

Consequently, together with the area of the impeller backside, the pressure on the clearances between casing cover and impeller back shroud (backside pressure) plays an important role in determining the nominal axial thrusts. For different impeller geometries, which may yield different pressure drops between compressor delivery and clearances, compressor, turbine as well as total thrust directions can vary noticeably in intensity and reverse. An axial force reversal can also occur during transient operation, e.g. the start/stop phase of the unit.

Therefore, the range of design helix angles β, which has been previously determined by means of the equation of the load transfer R and the graphical construction in fig. 2, may considerably change with the impeller design. Particularly, the upper and lower limits of the design angles, which are respectively 82 and 76 DEG in fig. 2, can remarkably change when turbine, compressor and resultant thrusts (Ft, Fc and Tref ) are different from the values reported in tab. 1.

As a rule, such design limits of β must be included within the range of helix angle feasible for helical spline couplings and gears, i.e. β must be in the range

45-90 DEG when both compressor and turbine thrusts are directed toward the inner side of the machine and the resultant thrust is positive (fig. 1) .

Differently, when the resultant thrust is negative (Tref<0), according to the proposed graphical construction the trend of the load transfer R must reach negative values in order to make the thrust functions (Ts and Tt) zero, as shown in fig. 2A.

According to the equation of load transfer, this is possible if the helix angle β is greater than 90 DEG, i.e. by means of a reversal of the helix handedness. Again, the helix angle cannot exceed 135 DEG, which is a safe operation limit for the actuator, like in helical gears .

In wider terms, the helix angle of the spline coupling (51) can be chosen with reference to the nominal working conditions on the basis of the target life of the bearings within a suitable range, which limits are assessed by means of the proposed graphical construction. The resulting choice of β must finally fall within the admissible range 45-135 DEG.

With reference now again to the non-limitative solution shown in the annexed drawings, it has to be noted that the second spline coupling 52 is a straight tooth spline coupling; however - as an alternative - it could be also another type of spline coupling, or, even, an interference fit.

As regards the main thrust air bearing 61 it is shown in a preferred embodiment in figures 9-11, showing the air foil bearing 61, the first aerodynamic elements 611 comprising a washer and the second aerodynamic elements 612 comprising a plurality of corrugated foils 614 and one smooth top foil 613.

As set forth above, the positioning of the parts is such that

- with the turbine in the stopped condition (as in fig.6) or rotating at a speed lower than the lift speed, the turbine rotor 41 is in contact with the annular shoulder 53 of the shaft 5, such that an axial thrust of the turbine rotor is discharged on the shaft 5 through the annular shoulder 53;

- with the turbine rotating at a speed equal or higher than the lift speed (as in fig.7), the turbine rotor 41 is moved away from the annular shoulder 53 of the shaft 5, such that an axial thrust of the turbine rotor 41 is discharged partially directly on the frame 2 through the main thrust air bearing 61 and partially on the shaft 5 through the helical spline coupling 51.

In preferred embodiments there is further provided a third radial bearing 63 visible in fig. 12 and coupled between the frame 2 and the shaft 5 near the main thrust bearing 61, the coupling of the third radial bearing 63 with the frame 5 being a slide coupling, where the third radial bearing 63 is free to axially slide with respect to the frame, such to permit an axial expansion of the shaft 5.

The use of the bearing 63 is suitable such not to cause the shaft 5 to work in a projecting manner.

As regards the auxiliary radial/thrust bearing 62 and the third radial bearing 63, if any, in one embodiment at least one of them comprises a rolling bearing, preferably a double direction bearing and still more preferably it comprises a plurality of single-row angular contact bearing.

The angular contact rolling bearing 62 shown is able to support both radial and axial loads, which are due to the weight of the rotor or a part thereof and to the action of the compressor and of the turbine respectively. Since the axial load is directed from the compressor 3 to the turbine 4 at the start and in an opposite direction under full operating condition, the bearing 62 is a double direction bearing; as an alternative, in the same manner, the same bearing 62 could be replaced by two single direction bearings put side by side or still, more generally, by a plurality of bearings exerting a function equivalent to that of a double direction bearing .

Preferably the bearing 62 is composed of two (or more) single-row angular contact bearings coupled with each other .

Figures show, for simplicity reasons, a solution where the bearing 62 is a four-point contact ball bearing. Such bearing is mounted by inserting the outer ring into a suitable cavity in the frame of the machine and by placing the inner ring on the relevant seat on the shaft 5, with suitable tolerances.

In particular the outer ring is axially constrained to the frame 2 such to react to axial actions.

In a "oil-free" variant the auxiliary radial/thrust bearing 62 and the third radial bearing 63, if any, comprise a magnetic bearing, of the active or passive type .

With reference now to fig. 13 and 14, another embodiment of the invention is schematically shown, in which with the same numeral references are indicated the same part as the first embodiment.

In the case of negative resultant thrusts, the layout shown in fig. 3 can be suitably modified.

Namely, with reference to fig. 1, in this case the thrust Ft reverses and the constraint 61 should be on the opposite part (the low-pressure side) of the turbine 4. To this purpose, the runner 611, which is mounted on the high-pressure side of the impeller 41 in fig. 4, might be moved on its opposite side and the location of bearing pads 612 might be modified accordingly.

Nevertheless, such solution can be considered not optimal by the point of view of the machine layout.

A more straightforward solution of the present invention consists in manufacturing a thrust collar/runner 611 that is either integral part of the turbine impeller or rigidly fastened to it, so that the bearing pads 612 can be located on both the sides of the runner regardless of the thrust direction.

Accordingly, the assembly drawing of the double-effect version of the invention, i.e. suited to both positive and negative nominal thrusts (with a consistent choice of β in agreement with the graphical construction) as well as transient loading conditions, is reported in Figure 13.

In this case, the main thrust bearing 61 is a double effect foil air bearing instead of a single effect one. In other words, the pads 612 are located on both sides of the thrust collar/runner 611.

The group of pads on the side that carries the thrust load in nominal working conditions is termed the "loaded" or active bearing, while the other group, on the opposite side of the thrust collar, is called the "slack" side or inactive bearing.

With reference to fig. 13, if the total thrust load in nominal conditions is positive (Tref>0), the active bearing is located on the right side of the runner and the helix angle β of the spline coupling 51 is lower than 90 DEG.

Conversely, if the resultant nominal thrust is negative (Tref<0), the active bearing is on the left side of the runner and β >90 DEG. In fig. 13 and 14 the second set of (non-locating) angular contact bearings 63 in back-to-back arrangement is added to the assembly near the turbine.

In this case, at low speeds the annular shoulder 53 does not receive the thrust load directly from the turbine impeller 41. Indeed, it exerts the thrust on the inner ring side 54 of the bearing 63.

Anyway, the behaviour of the invention does not change, as the bearing 63 is not constrained in the axial direction and, therefore, no axial load is transferred to the frame 2.

The flat washer 23 is fastened to the shaft 5 with a locking ring nut 24.

The spacer 22 is mounted between the washer 23 and the impeller 41 in order to adjust the axial gap of the turbine hub-shaft coupling.

Similarly, two spacers 21 are employed to adjust the axial clearance of the air bearings (the active as well as the inactive one) .

Particularly, as shown in fig. 14, the total axial clearance est of the turbine hub-shaft coupling is the sum of the clearances csl and cs2 (est = csl+cs2) .

In the following, the gap between pads and runner of the air bearings is the desired operating clearance and it is referred to as "hot" clearance.

The gap csl is the clearance between the spacer 22 and the turbine impeller 41, while cs2 is the gap between the turbine impeller 41 and the inner ring side 54 of the bearing 63.

Similarly, the total hot clearance cat of the double effect air bearing 61 is given by the sum of two contributions cal and ca2, which are the hot clearances between runner 611 and pads 612 of the active and the inactive bearing (cat=cal+ca2 ) .

In order to avoid wear of the double-effect air bearing 61 due to dry contact between runner and pads at low speed, the total hot clearance cat must be higher than the clearance est (cat>cst) .

Indeed, the axial clearance est of the coupling 51 must be very little (in the micron-length scale) .

Anyway, est must be greater than the equivalent RMS roughness of the two contact surfaces at the impeller 41/inner ring side 54 interface (or impeller/annular shoulder 53, if the bearing 63 is not employed) , in order to provide the relief of the secondary axial bearing 62 over the speed at which the runner 611 becomes airborne. Thus the above objects are achieved.

Obviously many variants to what described up to now are possible .