MECHANICAL COMPONENT FOR THERMAL TURBO MACHINERY

BACKGROUND OF THE INVENTION

The present invention relates to the technology of thermal turbo machines. It refers to a mechanical component for thermal turbo machinery according to the preamble of claim 1.

PRIOR ART

The increasing use of renewable energy sources to produce electricity requires additional operational flexibility from fossil-fuel steam and gas turbines. To compensate for renewable energy fluctuations in the electrical grid, a gas turbine (GT) engine needs to be more flexible, operating in peaking and partial loading modes as well as the base-load operation mode. At the constant rotational speed Ω of a turbine train, these flexible engine operation conditions induce variations of cooling-flow/metal temperature, mass-flow and pressure. These changes may generate unexpected asynchronous excitations acting upon the rotating and non-rotating mechanical components.

In general, the GT power depends directly on the mass flow, which varies under variable flexible operation conditions. The mass flow is the function of flow velocity U and its density ρ. The flow velocity U has a direct impact on Reynolds number Re=(ρ U d)/μ_f, where μ_fdenotes the free stream dynamic viscosity of a fluid and d means the characteristic diameter of the streamlined object or component, which is streamlined by said flowing medium.

Based on the experimental data given in the literature, the fluid flow excites the streamlined component within the range of Reynolds number varying between 30 and 5000. Then, regular vortex shedding as an oscillating flow takes place downstream the component, which is stimulated with the excitation function f_e. This excitation frequency is determined from the dimensionless Strouhal number St with St=(d f_e)/U. Acting perpendicularly to the oncoming flow direction upon the streamlined component, the excitation force F(t) is then determined from

F(t)=½c_wρU²A sin(2πf_et)

Where t is time, c_wmeans dimensionless drag coefficient depending on the shape of a streamlined component as given in the handbooks, and A denotes the contour area of the streamlined object projected perpendicularly to the oncoming flow. In this equation, the term of (ρU²)/2 corresponds to the dynamic pressure, which also alternates under different operation modes of a GT engine.

Thus, for varying operation conditions of a GT engine at the constant rotational speed Ω, numerous asynchronous f_eand synchronous kΩ (where k=1, 2, 3, . . . , ∞) resonances of the component can occur that might be unknown in the ordinary design process focusing mainly on a so-called “on-design point” of the engine.

Mechanical components of a gas turbine are usually designed for the base load nominal operation condition concerning ISO temperature varying between −15° C. and 45° C. This is called “on-design mode operation”. Under the flexible operation condition from the base load to part load of the GT engine, component base-load temperature T_breduces by even up to 120 K what generally depends on the type of a gas turbine. This temperature variation δT changes material properties like Young's modulus, what has a direct impact on the variation of the natural frequencies of a GT component as expressed by

$\begin{matrix} ω_{T_{b} \pm δ T} = \sqrt{\frac{E_{T_{b} \pm δ T}}{E_{T_{b}}}} ω_{T_{b}} & (1) \end{matrix}$

Where ω_Tbdenotes the reference eigenfrequency of the component at the base-load operation temperature T_b, ω_Tb±δTis the component eigenfrequency depending on part-load operation which subjects to ‘changing temperature δT with respect to the base-load temperature T_b, E_Tbis Young's (elastic) modulus at temperature T_breferring to the base load of a GT engine, and E_Tb±δTis Young's modulus at temperature T_b±δT referring to the part load of a GT engine.

In the past GTSC (Gas Turbine Single Cycle) and GTCC (Gas Turbine Combined Cycle) installed base have been designed mainly for the base load engine operation at component temperature T_b, called frequently on-design point. In general, there is a technical risk that the part-load GT operation can result in an unexpected resonance ω_Tb±δTof the GT component leading towards HCF (High-Cycle Fatigue) damages. For the operational flexibility conditions, both phenomena of component frequency variation and asynchronous excitations must be considered in the design process of new and installed base engines.

For over 100 years, the Campbell diagram has been used as best engineering practice preventing the components from their resonances (see FIG. 1 and e.g. document US 2009/0301055 A1). This diagram controls the frequency changes of rotating blade “B” and standstill vane “V” in terms of the rotational speed Ω of a turbine train.

In the Campbell Diagram of a rotating blade “B” and non-rotating vane “V” shown in FIG. 1, eigenfrequencies ω_B,N,Tband ω_V,N,Tbdepend on the stiffening effect of the centrifugal load growing from 0 to the nominal speed Ω_Nand the softening effect of temperature increasing from the ambient temperature T_ato base-load temperature T_b, where kΩ (where k=1, 2, . . . , ∞) denotes the harmonic excitation due to non-uniform circumferential pressure distribution of a flow medium within the turbine.

Because of non-homogeneous pressure distribution of the fluid medium along the circumferential direction of a turbine housing, the blade ω_Band vane ω_Vfrequencies can be stimulated by harmonic excitations determined with kΩ (where k=1, 2, . . . , ∞) and illustrated with dashed lines in FIG. 1. Therefore, especially at the nominal rotational speed Ω_N(see vertical dashed line in FIG. 1), the vane and blade frequencies must differ from kΩ_N. At the nominal rotational speed kΩ_N, these blade ω_B,N,Tband vane ω_V,N,Tbfrequencies correspond to the base-load temperature T_band mass flow of the on-design operation condition.

However, the base-load temperature T_bcannot be shown explicitly in the conventional Campbell diagram. Therefore, in case of the part load operation reducing the component temperature by δT, the resonance risk of the blades and vanes must be determined with Eq. (1) to show the shift of their frequencies along the vertical line of the nominal speed Ω_N(see FIG. 2). Since the blade ω_B,N,Tb-δTand vane ω_V,N,Tb-δTfrequencies coincide with the harmonic excitations kΩ_N, the bladed disc or vane assembly will experience HCF damages.

Concerning an approximate sense of Eq. (1), this adjustment of the conventional Campbell diagram to the part-load analysis of turbine blading seems to be not reliable enough. On the other hand, a new engineering procedure is required for determining safety regimes of part-load GT operation of the blades, vanes as well as other components in terms of temperature variation T_b±δT at the nominal speed Ω_Nof a turbine train.

For the part-load operation condition, which generally corresponds to the engine power reduction, temperature and mass flow become crucial engineering parameters in assessment of HCF risk. The conventional Campbell diagram shown in FIG. 2 does not provide enough technical details in the design process, because the metal temperature variation δT of a rotating blade or stationary vane is not shown explicitly.

Therefore, a Part-Load Resonance Diagram triggered by temperature variation δT is proposed as illustrated in FIG. 3 (right-hand side picture). The Part-Load Resonance Diagram of FIG. 3 shows for a rotating blade “B” and stationary vane “V” the eigenfrequency change ω(δT)_B,Nand ω(δT)_V,Nwith respect to temperature reduction δT from base-load temperature T_bunder the part-load GT operation condition at the nominal rotational speed Ω_Nof a turbine train, where a darker zone corresponds to the allowable temperature range of interest regarding needs of minimal T_TATfor GTCC operation.

Indeed, this diagram extends the conventional Campbell's diagram information in detail and measures the eigenfrequencies variation of components in terms of temperature variation δT of base-load temperature T_bat the constant rotational speed Ω_Nof a turbine train what corresponds to the part-load operation condition.

In the Part-Load Resonance Diagram, usually a particular range of temperature variation is of interest (see the darker zone in FIG. 3), which assures enough high T_TATfor a stable steam turbine operation in Combined Cycle plants. This diagram allows for risk check of asynchronous excitation triggered by reducing mass flow of fluid medium. Those asynchronous excitation frequencies can be either computed with time-marching CFD approach or measured in the engine. For practical engineering judgment, the well-known closed forms based on Karman vortex street and the Strouhal number St can be used for assessing these asynchronous excitations in FIG. 3.

Thus, eigenfrequency curves ω(δT) of each GT component must avoid coincident points with horizontal excitation lines representing harmonic and asynchronous excitation at the nominal rotational speed. A typical wobbling effect of ±5% of the nominal rotational speed Ω_Ndoes not have a significant impact on the change of the eigenfrequencies of rotating blades, and this phenomenon can be neglected in the analysis without reducing the reliability. In case of significant change of the rotational speed Ω, then an additional Part-Load Resonance Diagram must be created at this speed of interest as demonstrated for Ω_Nin FIG. 3. For the stationary components, the rotational speed has no impact on the analysis.

In prior art, several proposals have been made to manipulate the vibration behavior of components in thermal turbo machinery.

Document U.S. Pat. No. 6,290,037 B1 discloses a vibration absorber in which an absorber end mass is coupled to a primary mass by means of a cantilevered beam, wherein at least a portion of the beam comprises a shape memory alloy (SMA). Preferably, the end mass is coupled to the primary mass with several discrete SMA wires which may be individually heated. When each of the SMA wires is heated above a predetermined temperature, the SMA material undergoes a phase change which results in a change in the stiffness of the SMA wire. Heating of the various wires in various combinations allows the operational frequency of the absorber to be actively tuned. The frequency of the absorber may therefore be tuned to closely match the current vibration frequency of the primary mass, thereby allowing the absorber to be adaptively tuned to the frequency of the primary mass in a simple and straightforward manner.

Document U.S. Pat. No. 6,796,408 B2 discloses a method for damping vibrations in a turbine. The method includes performing structural dynamics analysis on the turbine to determine at least one area of high vibration stress on the turbine, and performing thermal analysis of the turbine to determine at least an approximated maximum operating temperature at the area of high vibration stress. Additionally, the method includes utilizing hysteresis damping to dampen operational vibrations. The hysteresis damping includes selecting a shape memory alloy (SMA) having a martensitic-to-austenite transformation temperature substantially similar to the approximate maximum operating temperature of the component at the area of high vibration stress, and disposing the selected SMA on the turbine on the related area of high vibratory stress.

Document U.S. Pat. No. 7,300,256 B1 discloses a damping arrangement for a blade of an axial turbine, in particular a gas turbine, which includes a damping element which is arranged in a recess in the blade aerofoil of the blade and frictionally dampens the vibrations of the blade. In such a damping arrangement, simplified manufacture and assembly and a reliable and effective function are achieved by the recess being configured as a cavity extending in the radial direction through the inside of the blade aerofoil, the damping element being inserted in the radial direction into said cavity.

In document DE 10 2010 003 594 A1, turbine blades have a vibration damping element formed with a shape memory alloy (SMA) element. The damping element is coupled with a surrounding area such that heat transferred to the SMA element from hot fluid flowing around one blade is changed based on a vibration state of the blade. The SMA element is formed with a SMA wire. The SMA element is extended in end surfaces of covers or a supporting wing. The SMA element couples the blade with the surrounding area in transverse to a longitudinal axis of the blade.

Document US 2012/0183718 A1 discloses a part, which includes a structure and at least one shape memory alloy element that is pre-stressed and embedded at least in part within said structure. The shape memory alloy is suitable for dissipating the mechanical energy of said structure when it vibrates in a given frequency band.

However, the situation at part-load is neither discussed nor solved in any of these prior art references.

SUMMARY OF THE INVENTION

It is an object of the present invention to provide a mechanical component for thermal turbo machinery with enhanced protection against High-Cycle Fatigue (HCF), which takes into account the influences at part load operation.

This object is obtained by a mechanical component according to claim 1.

According to the invention, a mechanical component for thermal turbo machinery, especially a steam or gas turbine, comprises a part, especially base part, and at least one additional device being mechanically coupled to said part in order to influence the vibration characteristic of said part during operation of the turbo machine

It is characterized in that the mechanical coupling between said part and said at least one additional device changes with the temperature of said at least one additional device.

According to an embodiment of the invention said at least one additional device is a device, which changes with temperature its form and position relative to said part in order to establish an additional mechanical contact between said part and said at least one additional device within a predetermined temperature range.

Specifically, said at least one additional device is a bi-metallic device.

Alternatively, said at least one additional device is a shape-memory-alloy device.

According to another embodiment of the invention said additional mechanical contact is a stiffening contact, which mechanically stiffens said part.

Alternatively or additionally, said additional mechanical contact is a friction contact, which dampens vibrations in said part.

According to a further embodiment of the invention said at least one additional device has the form of a longitudinal beam or curved plate, which is fixedly connected at both ends to said part, such that it establishes said additional mechanical contact in an area between both ends, when it changes with temperature its form and position relative to said part.

According to just another embodiment of the invention said at least one additional device has the form of a longitudinal cantilever or curved plate, which is fixedly connected at one end to said part, such that it establishes said additional mechanical contact with its other, free end, when it changes with temperature its form and position relative to said part.

According to a further embodiment of the invention additional sub-parts are provided on said at least one additional device in an area of said additional mechanical contact in order to influence the character of said additional mechanical contact.

According to another embodiment of the invention a heating or cooling means is provided for actively changing the temperature of said at least one additional device.

Specifically, said part is a blade or vane of a gas turbine.

Specifically, said part is an exhaust gas housing of a gas turbine.

It can as well be part of a combustor, compressor, or any other system whose operation temperature varies enough for changing remarkably Young's Modulus E as given in Eq. (1).

BRIEF DESCRIPTION OF THE DRAWINGS

The present invention is now to be explained more closely by means of different embodiments and with reference to the attached drawings.

FIG. 1 shows a Campbell Diagram of a rotating blade “B” and non-rotating vane “V”, whose eigenfrequencies ω_B,N,Tband ω_V,N,Tbdepend on the stiffening effect of the centrifugal load growing from 0 to the nominal speed Ω_Nand softening effect of temperature increasing from the ambient temperature T_ato base-load temperature T_b, where kΩ (where k=1, 2, . . . , ∞) denotes the harmonic excitation due to non-uniform circumferential pressure distribution of a flow medium within the turbine;

FIG. 2 shows a Conventional Campbell Diagram of a rotating blade “B” and stationary vane “V” concerning their eigenfrequency increase to values ω_B,N,Tb-δTand ω_V,N,Tb-δTdue to temperature reduction δT under the part-load GT operation condition at the unchanged nominal rotational speed Ω_Nof a turbine train;

FIG. 3 shows a part-load Resonance Diagram (right-hand side picture) for a rotating blade “B” and stationary vane “V” recording eigenfrequency change ω(δT)_B,Nand ω(δT)_V,Nwith respect temperature reduction δT from base-load temperature T_bunder the part-load GT operation condition at the nominal rotational speed Ω_Nof a turbine train, where a darker zone corresponds to the allowable temperature range of interest regarding needs of minimal T_TAT(turbine outlet temperature) for GTCC operation;

FIG. 4 illustrates four design strategies of a rotating shrouded turbine blade (as an arbitrary example) against HCF, namely (1) Mass Strategy (MS): Component Mass Variation, (2) Stiffness-Strategy (SS): Component Stiffness Increase, (3) Damping-Strategy (DS): Component Damping Enlargement, and (4) Mistuning-Strategy (MTS): Component mistuned in terms of Excitation;

FIG. 5 shows typical deformation curves of systems having a bi-metallic configuration (dashed curve) or made of a shape memory alloy (solid lines) with respect to temperature T, where T_a, T_TAT-min, T_bdenote the ambient, minimal Turbine Outlet Temperature for GTCC operation, and base-load temperature of a GT engine, respectively, and q_C,mindenotes the threshold deformation of the system for a contact with the GT component of interest above temperature T_TAT-min;

FIG. 6-8 show an embodiment of the invention by applying a bimetallic TMD to the stationary Exhaust Gas Housing of a GT for achieving a stiffening effect to shift the original system's eigenfrequency by δω due to the additional bending stiffness of the bi-metallic system, where T_Tmeans the threshold temperature of interest;

FIG. 9 shows TMD configuration for multi-contacts amplifying stiffening effect and/or frictional damping performance with respect to the mode shapes of the baseline part, with FIG. 9(a) showing examples of a thin-wall, thick-wall and solid sub-parts for generating different contact characteristics, and with FIG. 9(b) illustrating the results of the sub-parts arranged for stiffening and damping effects in the resonance response function;

FIG. 10 illustrates re-design degrees of freedom with TMD in mechanical contact with the base part, where “α” and “β” correspond to the stiffening or damping concept with one sub-part (see FIG. 9) depending on vibration magnitudes of the base part; and

FIG. 11 shows a part-load Resonance Diagram for a rotating blade “B” and stationary vane “V” equipped with the proposed TMD, which shifts the original eigenfrequency ω(δT)_B,Nand ω(δT)_V,Nby required values δω_B(δT) and δω_V(δT) (as illustrated with two long-dashed curves) to avoid the resonances that appear within the operation zone of the part-load condition, or which increases the damping performance of the overall system in terms of varying mass flow {dot over (m)}, which can generate asynchronous excitations of the baseline component.

DETAILED DESCRIPTION OF DIFFERENT EMBODIMENTS OF THE INVENTION

An overall idea of the present invention is to introduce an additional device into a new or existing design of a baseline component or part of a thermal turbo machine, especially a gas or steam turbine, which by mechanical coupling with the component passively changes the mechanical properties of the baseline component in terms of variation of the operation temperature of the engine.

This additional device, from now on called Thermal Memory Device (TMD), increases the reference stiffness of the baseline component and also enlarges the frictional damping onto mechanical contacts between the baseline component and the additional device. These additionally created mechanical properties of the baseline component with a TMD protect the engine from High-Cycle Fatigue under high temperature operations like in gas turbines. Through the created mechanical contacts onto the baseline component, the TMD does not cause any thermal stress during rapid variation of the thermal boundary conditions because the baseline component and additional device can slide relatively to each other without generating any thermal stress concentration during their thermal expansions. The aerodynamic performance of the engine is not impacted by applying the TMD inside the baseline component like for instance a cooled turbine blade or vane.

At the on-design point, each component of a GTCC system must be free of resonance in accordance with Campbell diagram (see FIGS. 1 and 2). The part load conditions of e.g. a GT engine generate two effects of (1) Metal Temperature Reduction of components, and/or (2) Generation of unexpected asynchronous excitations that are usually known from the test engine or field experiences.

In case of the unexpected resonance under the part-load operation, there are 4 standard resonance-mitigation strategies, such as Mass-Strategy (MS), Stiffness-Strategy (SS), Damping-Strategy (DS), and Mistuning-Strategy (MTS), as illustrated in FIG. 4.

According to the Mass-Strategy (MS), the mass of vibrating areas of the large component, e.g. an Exhaust Gas Housing, is locally changed. This is not an effective solution because frequency shifts of 2-3 Hz require a significant modification of the geometry of the large baseline component.

The Damping-Strategy (DS) is based on the friction or impact dissipation mechanism and does not relate to a straightforward engineering solution. Also, the Mistuning-Strategy (MTS) is an out-of-the-box solution of the engineering practice, which usually corresponds with too high costs for its validation.

Therefore, the Stiffness-Strategy (SS), which increases the overall stiffness of the component, is applied as the most simple and efficient mitigation. Often, an additional coupling like e.g. a bolt or stab is welded between the components or component parts, which increases the system's frequency of interest. However, this stiffening solution placed in the flow channel of a turbine generates aerodynamic losses or can easily lead toward new TMF (Thermal-Mechanical Fatigue) damages. For the components operating above evaluated temperature, an additional stiffness caused by the bolt does not allow for the thermal expansion of the overall system and TMF cracks can appear on the zones of thermal stresses driven by variable part-load operation conditions.

In GT technology, the thermally loaded components are usually designed for internal cooling and comprise thin-shell structures to avoid too high thermal stress concentrations during fast start-ups or shut-downs of an engine. In other words, a typical GT vane comprises a hollow space for internal cooling, which can be used for introducing an additional structural element which stiffens the baseline component for shifting its eigenfrequency above the resonance of interest.

To control this stiffening process in terms of temperature, the internal (additional) component or element is made of bimetallic material (BM) or shape memory alloy (SMA) whose characteristics are shown in FIG. 5. Deformations of the bi-metallic system (BM) are a substantially linear function in terms of temperature T. The shape memory alloy (SMA) demonstrates “binary” behavior of the deformation with the typical pseudo elastic-plastic hysteresis, as illustrated in FIG. 5.

FIGS. 6-8 show an example of how a standard baseline component can be equipped with an internal system made of conventional bi-metallic system, according to an embodiment of the present invention. Baseline component in this case is a stationary exhaust gas housing 10 of a gas turbine (see for example document U.S. Pat. No. 8,915,707 B2). The exhaust gas housing 10 comprises two concentric rings, namely an outer ring 11 and an inner ring 12. Both rings 11 and 12 are connected by a plurality of radial struts 13. Each strut 13 has a wing-like aerodynamic cross-sectional profile and a hollow interior 14 (FIGS. 7, 8).

As can be seen in FIGS. 7, 8, which show a cross-section in the plane A-A, a bi-metallic thermal memory device (TMD) 15 is arranged within a strut parallel to the longitudinal axis 21 of said strut. Thermal memory device 15 is positioned near the wall of strut 13, extends through hollow interior 14 of strut 13, and is rigidly fixed at both ends to the outer ring 11 and inner ring 12 by means of suitable fixations 16a and 16b. Thermal memory device 15 itself is divided in longitudinal direction into two bounded metal parts or beams 15a and 15b, which consist of metals with different thermal behavior to establish the necessary bi-metallic effect.

For temperatures below a threshold value T_T, there is no mechanical contact between the inner surface of the baseline component 10 and the external surface of the bi-metallic system 15, as illustrated in FIG. 7. Above the threshold temperature T_Tof interest, the bi-metallic component 15 comes in contact with the baseline part 10 (contact area 17 in FIG. 8), what increases the overall eigenfrequency of the coupled internally system 11, 12, 13, 15 by the required frequency range δω. The frequency shift δω can be enforced by applying additional components with thermal memory made of bi-metallic or shape memory materials. Also, instead of one simple beam 15 clamped at its both ends, two cantilevers 22, 23 (FIG. 8) with one free end each can be used for getting two contacts with the baseline part 10.

Cantilever beams 22, 23 of different lengths could be considered for arranging contacts at different locations with respect to vibration nodes and antinodes of mode shapes of the baseline part 10 (see FIG. 8).

Furthermore, as shown in FIG. 9(a), the external surface of the thermal memory device 15 can be equipped with additional sub-parts 18, 19, 20, whose shapes better match with the internal contour of the baseline part (in the example strut 13). Additionally, these shapes can be arranged for creating the best friction damping performance during vibrations of the entire system. In particular, sub-part 18 is a hollow thick-wall part, sub-part 19 is a hollow thin-wall part, and sub-part 20 is a fully solid part. These different sub-parts 18, 19 and 20 each generate a different contact normal and tangential stiffness.

Thus, the thermal memory device TMD being in contact with the baseline part has two functions:

- 1) Stiffening effect for shifting the eigenfrequency of interest, and
- 2) Damping of the forced vibration through the frictional dissipation onto the contact.

Accordingly, two S-Stiffness and D-Damping Design Strategies SS and DS are thus realized in the structure of FIG. 9(a), as illustrated in FIG. 9(b). As explained above, each sub-part can be designed as a thin-wall, thick-wall or/and solid structure for reaching the damping performance on the contacts at different radii r1, r2 and r3 in accordance with an elastic-friction dissipation mechanism or other approaches known in the open literature. Because the thermal memory device 15 always presses sub-parts 18, 19 and 20 against the baseline part 13, contact wear would not have an impact on the damping performance and the entire free-failure operation. Anyway, the wear at said contacts can be minimized with a specific coating.

For the stationary baseline components, even of large dimensions like an EGH (Exhaust Gas Housing), the stiffening effect or/and the damping performance could be validated in a typical annealing oven by using a standard system for measuring vibrations in evaluated temperatures.

Depending on needs of the design protection, either stiffening or damping performance of the system can be enforced as schematically illustrated in FIG. 10. The vibrations of the baseline component, known from the measurements or/and computations, are considered as the kick-off point of the re-design.

At the region of interest of the baseline component, the thermal memory device TMD comes in the required technical (flat) or Hertz contact within the baseline component. In terms of the cross-section of the device, the overall baseline component stiffness can be increased or reduced after being in contact at the operation temperature of interest. Then, the overall stiffness of the entire system increases by ratio “α” from the reference stiffness of the baseline as illustrated with a dark region (“stiffness increase”) in FIG. 10.

With respect to magnitudes of the relative contact vibration of the baseline component, the damping performance can grow or reduce. These damping performances can be also influenced in the re-design by applying of the particular contact form, contact stress magnitude or contact area as well as with specific coating increasing or decreasing friction coefficient. The designer has an option of adding additional contact areas as explained schematically with solid or hollow sub-parts in FIG. 9(a). Then, in terms of the vibration of the baseline component, the damping performance of the overall system in contact can be controlled with rate “β” by using one or more hollow and solid sub-parts (bright upper region “damping increase” in FIG. 10).

The final outcomes of the rotating blade or stationary vane equipped by the component with thermal memory are illustrated with long-dashed curves in FIG. 11 (upper right part of the Figure). The original eigenfrequencies of the blade ω(δT)_B,Nand the vane ω(δT)_V,Nare shifted by the required values δω_B(δT) and δω_V(δT) to avoid the resonances that appear within the operation zone of the part-load condition (see the two long-dashed curves in FIG. 11). By using the component with thermal memory as explained above in connection with FIGS. 6-9, there are options for softening the effect such that for instance the new eigenfrequencies of the part are lower than the original ones. This mechanism is within the scope of the present invention, too.

The technology of the adding and mounting a thermal memory device (TMD) for part-load operation, as described above, can be applied to rotating components or stationary parts of different dimensions. Thus, a complete exhaust gas housing as shown in FIG. 6 or single blades or vanes can be equipped with a suitable TMD.

The proposed bi-metallic systems (15 or 22, 23 in FIG. 8) may be made of arbitrary metals that are available on market or can be developed according to particular design reasons. Also, arbitrary shape memory alloys (SMAs) may be made of known elements or may be developed for achieving the desired purpose of the design. In other words, all known or newly developed bi-metallic or/and shape memory alloys of various shapes and fixations are part of the present invention. This applies also to arbitrary forms of the TMDs. Additionally, sub-parts of the TMD device (as shown for example in FIG. 9) for frictional damping can be made of different materials or are designed as brushes or others for arranging soft- or hard-contact stiffness.

Several TMDs can be arranged in series or parallel connection for weakening or enforcing overall stiffness and/or damping results of the entire system. Also, the bi-metallic and shape memory alloy can be combined together for defining bi-linear stiffness effect as the result of the linear and binary deformation, respectively. In addition, to vary the stiffness result continually or temporarily, local or overall cooling or heating effects of this system can be considered that can be arranged through different sources like electrical heaters (24 in FIG. 8), and others.

In general, the thermal memory device TMD can be also designed within the meaning of increasing stiffness of the baseline component, whose original frequency begins to get larger above the threshold temperature of interest. For this general design purpose, the generated mechanical contact between the device and baseline component does not generate any thermal stress concentration which appears in every conventional joining techniques of welding, brazing, mechanical joining and others in the operation of thermal engines. This type of the application corresponds mainly to the design concept based on Stiffness-Strategy (SS) as illustrated in FIG. 4 and can be used for arbitrary operation condition of a thermal engine. In other words, this invention is not only limited to flexible operation of the engine.

The present invention is descried with respect to needs of GTSC and GTCC systems. Indeed, the scope of this innovation can be applied to other engines and machines that are designed for the on-design point but need to operate additional under various part-load operation conditions. The TMD can be triggered by thermal and mechanical loading change or can be driven with an active-control system (e.g. a heater 24, as shown in FIG. 8). The invention can be used for engines operating with constant and variable rotational speeds.

LIST OF REFERENCE NUMERALS

10 exhaust gas housing (stationary)

11 outer ring

12 inner ring

13 strut

14 hollow interior

15 thermal memory device (TMD)

15
a,b metal part

16
a,b fixation

17 contact area

18,19,20 sub-part

21 longitudinal axis (strut)

22,23 cantilever

24 heater (e.g. electrical)

MECHANICAL COMPONENT FOR THERMAL TURBO MACHINERY

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