Acoustic resonance excited heat exchange

Abstract

New exemplary heat exchange configurations that incorporate internal or external surfaces equipped with perturbators, for changing the thermal behavior of the system, or for modulating the surface temperature distribution of the flow surfaces. This is achieved by applying an acoustic wave to the fluid flow in a heat exchange passage, and selecting the frequency of the acoustic exciting wave to be the same as the acoustic resonance frequency of the heat exchange passage itself. As the traveling waves interact with the boundaries confining the heat exchange passages, constructive interference of the incident and reflected waves give rise to a standing wave. Thus, the heat exchange passages act as a resonator, and by superimposing this standing wave on the separating and reattaching fluid flow, significant heat transfer improvement can be achieved. This is accomplished without the need to significantly increase the pressure required to achieve the desired through flow.

Claims

1. A heat transferring device, comprising: at least one internal passageway for fluid flow, said at least one internal passageway having at least one static element configured to generate turbulence in said fluid flow, and said at least one internal passageway having at least one acoustic resonance frequency when fluid flow is present in said at least one internal passageway; and a source for generating acoustic waves with a frequency at a harmonic of said at least one acoustic resonance frequency, said source being at least one of an externally powered device and a passive device, said source configured to apply said acoustic waves to a fluid passing through said at least one internal passageway, such that a standing wave is generated in said at least one internal passageway, wherein the at least one static element is such that the turbulence generated in said fluid flow comprises separating and reattaching flows which interact with said standing wave in said at least one internal passageway.

2. A heat transferring device according to claim 1, wherein said at least one static element which creates turbulence is said fluid flow comprises at least one of a rib, a pin, a fin, a dimple, a pin-fin, and a periodic array of any of the foregoing perturbation elements.

3. A heat transferring device according to claim 1, wherein said passive device is actuated by said fluid flow.

4. A heat transferring device according to claim 1 wherein said at least one acoustic resonance frequency is a plurality of acoustic resonance frequencies, said plurality of acoustic resonance frequencies being associated with different segments of said at least one internal passageway.

5. A heat transferring device, comprising: at least one internal passageway for passing fluid therethrough, having at least one perturbation element, said at least one perturbation element comprising at least one of a rib, a pin, a fin, a dimple, a pin-fin, and a periodic array of any of the foregoing perturbation elements; wherein said at least one internal passageway is constructed to have at least one acoustic resonance frequency when fluid flow is present therein, such that a harmonic of said resonance frequency matches an acoustic wave frequency derived from a source of pressure fluctuations, such that a standing wave is generated in said at least one internal passageway, wherein the at least one perturbation element is such that it generates in said fluid flow separating and reattaching flows which interact with said standing wave in said internal passageway.

6. A heat transferring device according to claim 5, wherein said acoustic source is either an externally powered device or is passively generated by said fluid flow.

7. A heat transferring device according to claim 5 wherein said at least one acoustic resonance frequency is a plurality of acoustic resonance frequencies, said plurality of acoustic resonance frequencies being associated with different segments of said at least one internal passageway.

8. A method of changing the thermal behavior of a heat transferring device, comprising: providing a heat transferring device with at least one internal passageway, said at least one internal passageway having at least one static element which creates turbulence in fluid flow through said at least one internal passageway, and said at least one internal passageway having at least one acoustic resonance frequency to fluid flow in said at least one internal passageway; generating acoustic waves with a frequency matching a harmonic of said at least one acoustic resonance frequency; applying said acoustic waves to fluid passing through said at least one internal passageway, such that a standing wave is generated in said at least one internal passageway, wherein said changing of said thermal behavior arises from the interaction of said standing wave with a separating and reattaching flow of said fluid passing through said at least one internal passageway.

9. A method according to claim 8 wherein said at least one internal passageway comprises at least one channel which enables at least partial through flow and is at least partially bounded by semi-permeable or solid walls.

10. A method according to claim 8 wherein said at least one acoustic resonance frequency of said at least one internal passageway is associated with either the entire extent or a portion of said at least one internal passageway.

11. A method according to claim 8 wherein said at least one acoustic resonance frequency is a plurality of acoustic resonance frequencies, said plurality of acoustic resonance frequencies being associated with different segments of said at least one internal passageway.

12. A method according to claim 8, wherein said at least one perturbation element comprises at least one of a rib, a pin, a fin, a dimple, a pin-fin, a periodic array of any of the foregoing perturbation elements, a surface protrusion and a surface indentation that causes said fluid flow to be locally separated and reattached.

13. A method according to claim 8, wherein said generated acoustic wave is derived from a source of pressure fluctuations and matches said harmonic of said resonance frequency.

14. A method according to claim 8 wherein said method modulates the surface temperature distribution of said at least one internal passageway.

15. A method according to claim 8 wherein said acoustic wave frequency is a harmonic of any of a longitudinal, transverse, lateral, radial, or mixed mode(s) of said acoustic resonance.

16. A method according to claim 8, wherein said at least one acoustic resonance frequency is in the audible, the inaudible, the infrasound or the ultrasound frequency ranges.

17. A method according to claim 8 wherein said acoustic wave is generated either by at least one externally powered source, or is passively generated by said fluid flow.

18. A method according to claim 17 wherein said at least one externally powered source produces temporal acoustic pressure fluctuations through vibroacoustics or thermoacoustics.

19. A method according to claim 8 wherein said acoustic wave is generated by any combination of at least one of a fluid-dynamic, fluid-resonant, or fluid-elastic generator.

20. A method of changing the thermal behavior of a heat transferring device according to claim 8, wherein said acoustic waves are generating by at least one of an externally powered device and a passive device actuated by said fluid flow.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

(1) The present invention will be understood and appreciated more fully from the following detailed description, taken in conjunction with the drawings in which:

(2) FIG. 1A to FIG. 1I illustrate several different configurations of rib perturbations used in the flow path of the plate cooling elements of a compact heat exchanger;

(3) FIG. 2 is a presentation graph illustrating the performance of a large number of prior art perturbator technologies, typically employed in commercial compact heat exchangers;

(4) FIG. 3 illustrates schematically an exemplary heat exchanger showing additional acoustic elements suitable for implementing the methods of the present disclosure for increasing the thermal efficiency of the heat exchange;

(5) FIGS. 4A to 4D show some sample cavities, which can be used for passive excitation or self-excitation of the fluid input to the heat exchanger of FIG. 3;

(6) FIG. 5 illustrates schematically a standard solar plate heat exchanger with passive excitation inputs, to achieve a higher heat transfer efficiency;

(7) FIGS. 6A to 6D illustrate experimental wind tunnel results on a rib turbulator section of a heat exchange passage, showing the enhancement of the thermal efficiency when longitudinal acoustic excitation is appliedFIG. 6A shows the channel geometry, FIGS. 6B and 6C show the estimated Nusselt numbers along the flow path, with and without acoustic resonance excitation, and FIG. 6D shows the heat transfer enhancement effect associated with acoustic resonances;

(8) FIG. 7 is a graph showing the longitudinal Nu variation at the centerline position of the channel shown in FIGS. 6B and 6C, in the presence and absence of the acoustic resonance excitation;

(9) FIG. 8 shows graphically the static wall pressure data for the unexcited and the acoustic resonance excited cases of FIGS. 6B and 6C; and

(10) FIGS. 9A and 9B illustrate schematically two views of the internal cooling channels of a typical high pressure turbine blade.

DETAILED DESCRIPTION

(11) Reference is now made to FIG. 3, which illustrates schematically an exemplary heat exchanger 30 showing additional acoustic elements suitable for implementing the methods of increasing the efficiency of a heat exchanger by using periodic acoustic excitation with acoustic waves tuned to the acoustic resonance frequency of the heat exchanger passageways. The exemplary heat exchanger shown in FIG. 3 is a roughened plate heat exchanger 30, which is shown on the right-hand side of FIG. 3 with its plates disassembled to show the passage of the fluid through the heat exchanger and across the roughened plates 31. An acoustic source 32 is used to inject acoustic waves 33 into the input manifolds of the heat exchanger, either by passing the flow through the acoustic source 32, or by directing the acoustic wave to impinge on the fluid, such that the sound pressure changes are imposed upon the flow. A loudspeaker can be used for such acoustic wave generation. The frequency of the acoustic excitation source is tuned to be the same as the acoustic resonance frequency of the heat exchanger passage or passages. Therefore the traveling wave 33 generated by the excitation source 32, on entry into the heat exchanger 30, establishes a standing wave pressure pattern, and such a standing wave pattern can have pressure amplitudes substantially larger than that of the input wave itself, in accordance with the acoustic Q-factor of the resonant cavity of the heat exchanger passageways.

(12) In the plate-type heat exchanger 30 shown in FIG. 3 it is possible to position an acoustic excitation source in either the hot flow stream or the cold flow stream or both, and to analyze the acoustic resonance of each stream independently. The resonance frequency depends on the working fluid and the temperature of the fluid within the flow path, since the speed of sound differs with the temperature of the medium through which the sound is passing. The source 32 generates a traveling propagating wave 33 outside of the heat exchanger cavity, having a frequency calculated to be the same as that of the resonance of the heat exchanger path, but on entering the heat exchanger cavity, generates a standing wave in accordance with the geometry, length, and form of the passageways of each section of the heat exchanger.

(13) The system of FIG. 3 shows an active excitation source. However, a more convenient way of achieving traveling pressure waves may be by using a passive excitation source rather than an active excitation source. The passive excitation source will have an output tone which can itself be tuned to the resonance frequency of the heat exchanger cavity, and can be self excited at the correct frequency without any input tone being applied. In other words, the steady fluid flow through the passive excitation source generates an acoustic tone at the same frequency as the self-resonance of the heat exchanger cavity. Thereafter, the acoustic wave is carried through the flow into the cavity of the heat exchanger, such that the produced standing wave is self-maintaining without any external input to generate the wave form.

(14) Reference is now made to FIGS. 4A to 4D, which show some sample means of creating acoustic waves which can be used for passively exciting the fluid in the heat exchanger. These drawings show respectively a simple cavity, a shallow depth cavity and a deep cavity, and a cavity with a vibrating component that produce the acoustic source excitation by means of fluid-dynamic, fluid-resonant, and fluid-elastic interactions.

(15) Reference is now made to FIG. 5, which illustrates schematically a standard solar plate heat exchanger 50, with its plates disassembled in order to show the separate hot and cold fluid flow paths, adapted according to the methods of the present disclosure to achieve a higher heat transfer efficiency than a conventional design. In the entry flow paths 51,52, of the heat exchanger, there are positioned passive acoustic excitation sources, shown in FIG. 5 as acoustic horns 53. An enlarged cross-sectional view of such a horn resonator 53 is shown in the blown up section of the drawing. The flow 51 through the horn 53 generates an acoustic wave because of the resonance in the vibrating flap cavity 54, as is known in a conventional horn. By this means, the acoustic wave is generated in the horn 53 by the flow 51 itself, and impressed upon the fluid before entry into the heat exchanger 50. The resonance frequency of the horn is adapted to be equal to a resonance frequency of one or more passageways in the heat exchanger, such that the impressed acoustic wave generates a standing wave within the aforesaid passageways. The resonance frequency can either be selected to be that of the entire passageway of the heat exchanger, or it can be tuned to be that of a certain section of the passageway of the heat exchanger, where for instance more efficient heat transfer is required because of a localized heating problem inherent to the design. In addition, based on the internal features of the heat exchanger, the generated excitation frequency can be designed to be at any overtone (higher harmonic) of the fundamental (base resonance) frequency of the longitudinal, transverse, lateral, radial (but not limited to) and mixed modes of resonance. As previously stated, the passive resonator, such as the horn 53, can be positioned in either the hot or the cold fluid flows, or in both as is shown in FIG. 5.

(16) The configuration shown in FIG. 5 is particularly convenient, since it enables the improvement of existing heat exchangers to be achieved by the simple addition of such passive resonators at the input or output manifold(s) of the heat exchanger, without the need to modify the heat exchanger elements themselves. Moreover, in the presence of a change in inlet or outlet fluid temperature supply, the tonal sound generation will be self-adapted to match with the changed acoustic resonance frequency of the heat exchanger passages. Furthermore, the robust design of such a passive resonator is such that it can withstand the hot gas environments present, which may be destructive to other methods of generating an acoustic wave. In addition such a passive resonator does not present any significant obstruction to the fluid flow, such that the kinetic efficiency of the heat exchanger is not affected.

(17) The adapted heat exchanger of FIG. 5, or similar systems constructed according to the methods of the present disclosure, thus provides enhancement of the thermal performance and efficiency of the heat exchanger. Alternatively, for the same thermal efficiency as a prior art heat exchanger of the same type, use of these methods would correspond to a reduction in the form factor and size of the heat exchanger. Use of the methods and systems of the present disclosure therefore enables the performance characteristic band of the heat exchangers shown in the graph of FIG. 2 to be moved upwards, indicating an increase in the thermal efficiency compared to prior art heat exchangers. An increase of up to 30% in the thermal efficiency of such heat exchangers has been shown to be readily achieved, even using unoptimized designs generated during feasibility tests of these systems.

(18) Reference is now made to FIGS. 6A to 6D, which illustrate some experimental wind tunnel results, showing how it is believed that the enhancement in heat exchanger thermal efficiency is achieved, when the methods of the present disclosure are applied to a heat exchange passage equipped with a rib turbulator. FIG. 6A shows the basic topology in a cross-sectional view of a ribbed plate rectangular section of such a heat exchanger. The channel wall plate 60 of the heat exchanger has a number of rib perturbators 61 having a predetermined height H and pitch according to the design of the channel. As is known in the art from aerodynamic analysis of ribbed channels, the fluid flow 62 over such ribs 61 is highly turbulent 62, and the ribs generate flow separation in the region 66, with flow reversal 63 occurring in the rib wake region, and reattachment 64 at a distance from the rib; further downstream, the turbulence generating process is repeated as the next rib is encountered. The standing waves of the methods and systems of the present disclosure can be adjusted to generate increased turbulence, or coherent structures, or recirculation, at the optimal points relative to the rib, resulting in enhanced heat transfer because of the acoustically improved mixing generated by the perturbator.

(19) FIGS. 6B and 6C are graphic representations of the heat transfer distributions in the region 65 of the rib configuration of FIG. 6A, as obtained in wind tunnel experiments, where the rectangular heat exchange channel is subjected to Reynolds number of Re.sub.D=134,000 and Re.sub.H=10,050, based on channel hydraulic diameter and rib height respectively. The heat transfer is represented by the Nusselt numbers Nu, measured along the ribbed plate channel of the heat exchanger. The representations are of a plan view of the channels, looking down onto the ribs, such that y is the position across the width of the channel, and x is the distance down the channel, both being normalized to H, the rib height. FIG. 6B shows the values obtained in a prior art, unexcited heat transfer configuration, used as the baseline measurement, while FIG. 6C shows the values obtained in a heat transfer configuration according to the present disclosure, in this case implementing longitudinal acoustic resonance excitation. An analysis of these two representations is useful in explaining what is believed to be the reason for the enhanced thermal transfer achieved by the systems of the present disclosure. However it is to be understood that the methods and systems of the present disclosure are not intended to be dependent on the proposed mechanisms responsible for the achievements of the improvements described herewithin, and that other explanations may be the reason for the results obtained.

(20) An analysis is now made of the baseline results of FIG. 6B. As a preliminary note, it is observed that no heat transfer data is available on the top and directly upstream of the rib, in the region 2.33<x/H<0, due to the blocked camera observation path associated with the thermal measurement. The local streamwise maximum in heat transfer, x.sub.max, is indicated by the thin black line crossing the channel in the region of x/H=9.

(21) The upstream region 19<x/H<2.33 is characterized by the unperturbed boundary layer development over a flat plate 60, prior to influence due to the presence of the rib obstacle 61. Associated with boundary layer thickening at increasing development length from the inlet, an overall gradual decrease in heat transfer is observed. Towards the lateral walls, higher levels of heat transfer are a result of the corner wall vortices associated with the rectangular channel flow geometry.

(22) As the flow approaches the rib, 2.33<x/H<0, it undergoes a deviation imposed by the obstacle. Passing over the rib, the flow is locally accelerated and subsequently experiences an abrupt step change at the backward face of the rib. Forming an elongated recirculation bubble 63, and confined by the flow reattachment line, the separated flow region occupies a distance of approximately 8-10H, as shown in FIG. 6B. As the most prominent flow feature, this exerts large variations in heat transfer.

(23) Forming a low momentum zone, the rib wake separation bubble imparts a local minimum in Nusselt number Nu.sub.D=370 at the immediate vicinity of the rib, x/H=0. This is evident across the entire passage width (x/H). Further downstream of the rib from x/H1.5, the Nusselt number begins to increase monotonously as cooler flow is progressively entrained from the mainstreama consequence of the diminishing wake effects. At an increased axial position, this steep rise eventually reaches a global maximum (Nu580) in the vicinity of the reattachment point, where the strong impingement of the separated free shear layer on the bounding wall subjects the heated surface to cool high-momentum mainstream fluid. Although the aerodynamic reattachment point (x.sub.R) and streamwise maximum in heat transfer (x.sub.max) do not universally coincide for all separated flows, x.sub.max is considered to be a relevant indicator of the skin friction reversal point. Towards the side walls, the local heat transfer maxima levels increase, the locations of which are observed slightly further upstream. This curved spanwise distribution and laterally increasing heat transfer are attributed to the aerodynamic wall effects and rolled up corner vortices, being advected over the rib from the upstream separation point. Beyond the reattachment point, x/H>10, the heat transfer decreases monotonically in the streamwise direction with the redeveloping thermal boundary layer and eventually approaches its initial unperturbed boundary layer state, at approximately x/H>27.

(24) In order to investigate the heat transfer implications of the acoustic resonance excitation of the present application, in FIG. 6C there is shown the flow over a rib 61, subjected to harmonic 120 Hz acoustic forcing. This frequency corresponds to the acoustic resonance frequency of the channel shown in FIG. 6A on which the results of FIGS. 6B to 6D were obtained. The resulting distributions of Nusselt number are displayed in FIG. 6C. Compared with the baseline unexcited case in FIG. 6B, significant changes associated with the acoustic perturbations are visible in FIG. 6C. To highlight this, the longitudinal Nu variation at the centerline position of the channel is portrayed in the presence and absence of 120 Hz resonance excitation, in FIG. 7, hereinbelow. The points of local maximum heat transfer for both cases are indicated in FIG. 7 by a triangle.

(25) Regions upstream of the rib, x/H<2.33, feature flat plate boundary layer development, and appear to be impervious to acoustic excitation, as shown by the co-incident traces in the graph of FIG. 7. As opposed to this, downstream of the step disturbance and around the reattachment region, 0<x/H<15, as is seen in FIG. 6C, the sound forcing excitation has a significant effect on the local flow field and the associated heat transfer distribution. While the local minimum in heat transfer remains fixed at the rib back face, x/H=0, the steep streamwise gradient in the separation region appears to be augmented as a consequence of the sound excitation, as is seen in the steeper and spatially earlier rise of the Nusselt number for the excited curve of FIG. 7. An increase in maximum heat transfer level is observed. Reaching an earlier reattachment, the extent of the bubble is significantly reduced under the influence of acoustic excitation, shifting the location of centerline maximum heat transfer from x.sub.max/H=9.5 to x.sub.max/H=6. Towards the lateral walls at y/H3.33, a similar observation can be made, moving x.sub.max/H from 8.5 to 6, as seen in FIG. 6C. Therefore, while the line of unexcited flow reattachment shown in FIG. 6B exhibits a curved shape in the spanwise direction, it becomes flattened in the presence of sound excitation.

(26) It appears that the 120 Hz acoustic resonance excitation exerts attenuating influence on the extent of the rib wake separation, notably reducing the size of this prevalent flow structure. Therefore, together with the characteristic flow topology, the associated heat transfer pattern is shifted towards the rib and compressed in the streamwise direction. Further downstream, as the excited thermal boundary layer starts to develop at an earlier position, the local heat transfer level at the re-attached flow condition appears to be slightly lower with respect to the unexcited case of FIG. 6B.

(27) Reference is now made to FIG. 6D, which shows the acoustically-associated enhancement factor, EF, contrasting the excited case of FIG. 6C to the unexcited case of FIG. 6B. Upstream of the rib, indicated by uniform EF1, effects of acoustical excitation are notably absent. In the immediate wake of the fence (x/H=0-8), due to upstream shift of the bell shaped Nusselt curve, as shown in the graphs of FIG. 7, pronounced heat transfer augmentation is observed over the entire span of the channel. With a maximum around the centerline at x/H1.5 and |y/H|<2.5, the enhancement factor exhibits an increase up to 25%. This alteration is not as prominent towards the lateral side walls. Further downstream, the heat transfer enhancement decreases gradually. Decaying to zero slightly upstream of the unexcited reattachment point (x/H=8), it reaches a global minimum at x/H=12.5. Far away from the obstacle, the EF gradually re-approaches unity, and therefore the sound induced effects are observed to diminish.

(28) Reference is now made to FIG. 8, which, in order to assess the local pressure drop implications of acoustic forcing on the rib wake, shows graphically the static wall pressure data for the unexcited and excited cases of FIGS. 6B and 6C. FIG. 8 shows the pressure development along the channel centerline acquired over a distance of 30 rib heights. As an indicator of aerodynamic losses, the static wall pressure acquisition enables representation of aerodynamic flow separation and reattachment. To this end, normalized by the dynamic pressure head, the values are reported in reference to the pressure port upstream of the rib at x/H=10.

(29) The static pressure ahead of the fence is seen to exhibit a development which is typical for the mean flow topology in the presence of an obstacle. As the flow encounters the perturbator, 8>x/H>1, the initially streamwise constant static pressure rises due to the potential blockage effect. Consistent with the results shown in FIG. 6C, sound excitation is seen as not imposing any changes to the local pressure field, supporting the observation that the oncoming attached upstream boundary layer is unaffected by the acoustic forcing.

(30) In contrast, notable excitation effects are apparent in the fence downstream region, 0<x/H<13.5. Absent of forcing, the wall pressure in the separation zone initially reduces in the streamwise direction and reaches a global minimum at x/H=3. Thereafter, wall pressure exhibits a gradual rise with constant slope until shortly after the reattachment point, x/H=12.5. Further downstream, the curve maintains a relatively constant level in the redeveloping flat plate boundary layer. The integral pressure drop penalty incurred over the fence obstacle is characterized by a D'Arcy friction factor of around f=0.14.

(31) In the presence of the excitation, although the general trends are retained, there seems to be a greater initial drop in pressure in the immediate vicinity of the rib, followed by an earlier minimum in pressure at x/H=1.5. In comparison to the unexcited case, the initial rise is observed to be steeper than the prior observed linear trend. Downstream of the excited maximum heat transfer point (x.sub.max/H=6.5), the pressure gradient drops gradually. Remarkably, the identical plateau of downstream static pressure level is reached at around the same location, x/H=12.5. Therefore, the total pressure loss associated with flow over the fence is inferred to remain constant. This is an important result, since it means that the increased thermal transfer efficiency of the systems of the present disclosure is achieved essentially without any additional pressure penalty on the flow driving system.

(32) Regarding the invariance of static pressure downstream of the separation region both absent and present of forcing, it can be deduced that the associated aerodynamic loss mechanism is unaffected despite prominent excitation induced changes in the reattachment region. As it is primarily the recirculation bubble, which causes the pressure drop, the associated recirculation (integral vorticity) can be hypothesized to remain constant. For the conducive excitation, the slightly lower initial pressure at the rib back face, along with the earlier recovery, could be indicative of a smaller vortex of greater vorticity, rotating at a higher rate, immediately downstream of the rib.

(33) At the edge of the rib separation, the mixing layer dynamics are assumed to be governed by shear-induced generation of vorticity and turbulence in the velocity gradient region. The instability may roll up into vortices and could give rise to ensuing development of large coherent structures via sequential vortex pairing and amalgamation. After initial laminar formation, the increasing scales of these quasi-deterministic building blocks could determine the entrainment of momentum into the shear layer and thus the extent of turbulent mixing. Therefore, the downstream thickening or spreading rate of the mixing layer may be related to the vortex pairing mechanism and could be associated with the growth rate of large spanwise-correlated vortical structures. The spatially stationary periodic fluctuation of a standing wave could either directly interact with the pre-existing coherent flow feature, or form a new structure to dominate the reattaching flow field. Through the use of pressure nodes and velocity antinodes, standing waves can therefore be considered an effective way of delivering the necessary perturbation in the desired location and direction; and thereby influencing the heat exchange mechanisms on the channel surface.

(34) Reference is now made to FIGS. 9A and 9B, which illustrate schematically two views of the internal cooling channels of a high pressure turbine blade, based on the geometry of the NASA/GE Energy Efficient Engine (E.sup.3) engine. In such blades, in order to keep the turbine vane and blade metal temperatures below allowable limits, internal cooling techniques route the compressor air, introduced from the blade roots 91,92, through intricate serpentine passages, the front serpentine 96, and the aft serpentine 97, inside the airfoil. Extracting heat from the rectilinear internal channel walls, the gas is eventually discharged into the main stream from the leading edge film cooling holes 93 located in the leading edge showerhead 98, from the trailing edge film cooling slots 94, and the blade tip. The trailing edge impingement cavities 95 also assist in heat dissipation.

(35) To promote heat exchange, the passage walls are lined with repeated geometrical disturbance elements, which yield improved mixing with the free stream and induce high levels of turbulence to the core flow. This approach is effective in raising the heat transfer to considerably higher levels, at the expense of an inevitably enlarged pressure drop penalty. The common types of such protrusions include a sequence of rib-shaped turbulators which induce periodic tripping of the boundary layer, unbounded shear layer formation and consecutive separation, followed by an eventual flow reattachment and wall-bounded shear layer development. This geometry thus provides an example of the methods of the present disclosure for increasing the heat transfer effectiveness.

(36) The rotor-stator interaction of high-speed turbines represents a prominent mechanism of unsteady aerodynamic forcing. Associated frequency spectra feature characteristic peaks, which indicate the blade passing event, and its higher harmonic multiples or overtones. For an engine incorporating a blade such as that illustrated in FIGS. 9A and 9B, strong periodic excitation, associated with the rotor-stator interaction, is produced in frequency ranges from few kHz up to around 25 kHz. In the article entitled Exploitation of Acoustic Effects in Film Cooling GT2014-26318 by M Collins and T. Povey, presented at the ASME Turbo Expo 2014Turbine Technical Conference and Exposition, acoustic characteristics of travelling pressure waves in film holes is discussed, and the effects of periodic wake passing and unsteadiness effects on coolant mass flow fluctuations showed the importance of matching hole geometry to external source frequencies.

(37) In this light, it should therefore be possible to apply the methods and results of the present disclosure also towards practical turbomachinery applications, such as that shown in FIGS. 9A and 9B. For the closely confined internal air flow inside the highly branched turbine blade cooling system, the acoustically coupled resonance behavior of interconnected passages and cavities is expected to exert a strong influence on the internal convection heat transfer. The acoustic resonance behavior of a characteristic high pressure (HP) turbine blade can be analyzed numerically. Typically having a length of less than 10 cm in span, the natural frequencies will be in the same frequency range (from few kHz up to around 25 kHz) as the strong periodic excitation produced by the rotor stator interaction. The internal design of the turbine blade, which defines the serpentine cooling passages, can be tailored to benefit from the aero-thermal impact of standing waves on roughened surfaces. Use of these resonance frequencies thus enables the heat transfer from the blade to the cooling air to be increased, according to the methods described in this disclosure, leading to a higher performance blade configuration.

(38) It is appreciated by persons skilled in the art that the present invention is not limited by what has been particularly shown and described hereinabove. Rather the scope of the present invention includes both combinations and subcombinations of various features described hereinabove as well as variations and modifications thereto which would occur to a person of skill in the art upon reading the above description and which are not in the prior art.