Patent Publication Number: US-2015082794-A1

Title: Apparatus for acoustic damping and operational control of damping, cooling, and emissions in a gas turbine engine

Description:
FIELD OF THE INVENTION 
     The invention relates to an apparatus for acoustic damping of a range of acoustic frequencies, and for operational control that optimizes acoustic damping, cooling, and emission control in a combustion section of a gas turbine engine. 
     BACKGROUND OF THE INVENTION 
     Gas turbine engines often use a portion of air from the compressor for cooling and emissions control. Combustion gas temperatures can approach or exceed limits for structures in the working gas flow path. Therefore, cooling of surfaces adjacent the combustion gas (“hot surface”) may be implemented. Film/effusion cooling holes are provided through walls of flow-directing structures lining the working gas flow path so that a portion of the compressed air bypasses the combustor inlets and flows through these walls. This approach is used on such structures as the combustion chamber liner, transition ducts, transition exit pieces, and other components. The holes provide film cooling and effusion cooling of these components. 
     In conventional gas turbines a transition duct directs a flow of combustion gas traveling at about mach 0.1 to 0.3 between each combustor and the first row of turbine blades. Compressed air in a plenum surrounding this duct has higher static pressure than the combustion gas within the transition duct. This drives compressed air from the plenum through cooling holes in the duct walls. An emerging technology for can annular gas turbine engines provides transition duct structures that direct gas from the combustor to the first row of turbine blades along a mainly tangential and partly axial flow path at proper speed and orientation to drive the first row of rotating blades without an intervening row of stationary vanes. An assembly of such transition ducts is disclosed in U.S. Pat. No. 7,721,547 to Bancalari et al. issued May 25, 2010, incorporated in its entirety herein by reference. 
     In the emerging technology the transition ducts accelerate the combustion gas above mach 0.3 to about mach 0.8. This increased speed provides a decrease in static pressure in the newer transition duct design, so a greater pressure difference exists between the compressed air in the plenum and the combustion gas in the duct. This, pressure difference can provide more air than is needed for film cooling. It is so great that film cooling can overshoot and separate from the hot inner surface of the duct, reducing cooling effectiveness, unless the cooling holes are kept smaller than in prior designs and smaller than is ideal. Smaller holes clog with particles more quickly. 
     Acoustic damping resonators have been used in gas turbine engines to damp vibrations during operation. They may be called Helmholtz resonators or High Frequency Dynamics (HFD) damping resonators. Examples are disclosed in U.S. Pat. No. 6,530,221. Each resonator includes a chamber in an enclosure welded to a wall lining the working gas flow path. They are used on structures such as a combustion chamber liners and transition ducts. The resonator enclosure may have holes that admit cooling air to purge the resonator chamber. This prevents contamination from entering the chamber from the working gas, and cools the resonator walls and flow path wall. The cooling air passes through the resonator walls, impinges on the flow path wall, and then passes through effusion/film cooling holes in the flow path wall to form a cooling film on the hot inner surface of the flow path wall. These film cooling holes also act as Helmholtz resonation ports energized by the working gas flow. 
     The volume of a resonator is the main determinant of its resonant frequency. Space limitations can limit the size of damping resonators, thus limiting them to damping high frequencies only, such as over 1000 Hz. But unwanted intermediate frequencies between 50-1000 Hz are generated in gas turbine engines under some conditions. Damping resonators are often needed most at areas of high heat release. This exposes their enclosure attachment welds to high temperatures via heat conduction through the flow path wall, which can be a design-limiting factor. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The invention is explained in the following description in view of the drawings that show: 
         FIG. 1  is a front view of a prior art assembly of combustion gas flow directing structures called transition ducts and exit pieces in an emerging technology thereof. 
         FIG. 2  is a back perspective detailed view of a prior art exit piece of one of the flow directing structures in the emerging technology of  FIG. 1 . 
         FIG. 3  is a back sectional view of a prior art combustion chamber liner with a circular array of acoustic damping resonators on the outer surface thereof. 
         FIG. 4  is a side sectional view of a damping resonator taken on line  4 - 4  of  FIG. 3 . 
         FIG. 5  is a side sectional view of two damping resonators with different chamber volumes showing aspects of an embodiment of the invention. 
         FIG. 6  is a side sectional view of a larger damping resonator embodiment. 
         FIG. 7  is a side sectional view of a larger damping resonator embodiment with an intermediate reinforcing rib not extending to the cover plate. 
         FIG. 8  is a schematic view of four sets of damping resonators with four respectively different chamber volumes controlled in three groups by three air inflow throttle valves. 
         FIG. 9  is a frequency/airflow response chart for a resonator sized to damp a frequency or band of frequencies at about 700 Hz. 
         FIG. 10  is a frequency/airflow response chart for a resonator sized to damp a frequency or band of frequencies at about 1100 Hz. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
       FIG. 1  shows an assembly  10  with a plurality of flow directing structures  12 . Each flow directing structure  12  may include a cone  14  and an associated exit piece  16 . Each cone  14  receives combustion gas  18  from a respective combustor (not shown), and begins accelerating the combustion gas  18  to a speed appropriate for delivery onto the first row of turbine blades (not shown). Acceleration of the combustion gas  18  is accomplished by an acceleration geometry  20  using flow cross section constriction based on the Bernoulli principle and the Venturi effect. In the example shown, this is embodied as a cone-shaped duct  14  with a centerline  25  oriented mainly tangentially and partly axially relative to the turbine axis  26 . The cone  14  may abut the exit piece  16  at a joint  22 . Adjacent exit pieces abut each other at joints  24 . The exit pieces form an annular chamber immediately upstream of the first row of turbine blades (not shown). Combustion gas  18  enters each cone  14  and travels along a flow path within the cone  14 . The acceleration geometry  20  accelerates the combustion gas  18  to more than mach 0.3, and especially to between 0.3-0.8 as appropriate for direct delivery onto the first row of turbine blades. Upon entering the exit piece  16  the combustion gas may continue to accelerate to the final speed and may morph from a circular cross section to a non circular cross section. The combustion gas then enters an annular chamber formed by the plurality of exit pieces and forms a helical flow about the turbine longitudinal axis  26  for a short time prior to reaching the first row of turbine blades. Other embodiments may vary the specific shape of the flow directing structure  12  and the acceleration geometry  20 , and these various configurations are considered within the scope of the disclosure. 
       FIG. 2  shows a grid of ribs  30  on an outer surface of an exit piece  16  of the flow directing structures  12  of  FIG. 1 . The ribs structurally reinforce the exit piece to withstand the pressure difference created by the acceleration geometry  20  ( FIG. 1 ). Pockets  32  are defined between the ribs on the outer surface of the exit piece  16 . These ribs and pockets may be present on the combustor, the cone  14 , the exit piece  16 , or anywhere needed. 
       FIG. 3  is a sectional view of a prior art combustion chamber liner  28  surrounding a combustion chamber  34 , which may be cylindrical about an axis  36 . This view is taken on a section plane normal to the axis  36  through a circular array of acoustic damping resonators  33 , looking upstream relative to the combustion gas flow  18 . Each resonator  33  has a top wall  38  with coolant inlet holes  40 , a bottom wall  42  with holes  44  for coolant exit from the resonating chamber  48 , and side walls  45  between the top and bottom walls. The bottom wall  42  is formed by the combustor liner  28  bounding the combustion gas flow  18 , which flows generally axially, although it is shown here in a swirl for clarity. The coolant exit holes  42  may serve three functions: 1) as Helmholtz resonation ports that energize resonant vibrations in the chamber  48 ; 2) as coolant exits; and 3) for effusion/film cooling of the liner  28 . The air plenum  37  receives compressed air from the turbine compressor (not shown). Some of this air  50  enters the coolant inlet holes  40  in the top wall  38  of each resonator, and then escapes  52  into the combustion chamber  34 , providing effusion/film cooling of the inner surface of the liner  28 . 
       FIG. 4  is a side sectional view of a resonator  33  taken along line  4 - 4  of  FIG. 3 . Acoustic vibrations occur in each chamber  48  when the working gas  18  flows past the holes  44  in the liner  28 . These vibrations are excited by fluid dynamics mechanisms such as Helmholtz resonance. Such resonators have been used on the combustion chamber liner and the transition piece in prior art. However, the welds  56 ,  58  can reach thermal limits during operation—especially the upstream welds  56 , which are not cooled by film cooling  52  exiting the resonator. This can be a design-limiting factor. 
       FIG. 5  is a side sectional view of two acoustic dampers  60 ,  62  exemplifying an embodiment of the invention formed in respective pockets  32  between reinforcing ribs  30  on the outer surface of a flow directing structure  12  such as the exit piece of  FIG. 2 . Two resonating chambers  48 A,  48 B are covered by two respective perforated resonator plates  64 A,  64 B enclosing the two respective pockets  32  by spanning between the ribs  30  surrounding each pocket. The two plates are attached at two different heights H1, H2, thus providing two different damping frequencies. An airflow control manifold  66  may cover and enclose the resonators. Compressed air  50  from the air plenum  37  surrounding the flow directing structure  12  may be metered into the manifold  66  by fixed inlet hole(s)  69  and/or by an air inlet throttle valve  68 . The throttle may be a type that never fully closes, as shown, and/or additional inlet holes  69  may be provided on the manifold  66  to assure a minimum flow. The throttle may be controlled by control logic  70  integrated into or connected to an engine control/sensor system that determines and sets optimum throttle settings based on engine speed, sensed vibrations, and/or other operating conditions including ambient conditions. This optimizes acoustic damping in conjunction with cooling of the flow directing structure  12  for different stages of engine operation as later described. An array of such resonators  60 ,  62  may be provided around the flow directing structure  12 . This may be a circular array where the flow directing structure has a circular cross section. However, a variety of configurations of different sized resonators is possible. A thermal barrier coating  72  such as a ceramic layer may be provided on the hot side of the wall  74  lining the working gas path  18 . 
       FIG. 6  shows a resonator embodiment  76  with a larger chamber  48 C formed by a larger pocket  32  between ribs  30  covered by a resonator plate  64 C. The casting of the flow directing structure  12  may provide a plurality of pocket sizes for a range of damping frequencies. Wider rib spacing providing wider pockets  32  may be cast on wall portions of the flow directing structure  12  that need less reinforcement than other portions. Cylindrical or conic portions of the wall  74  may need less reinforcement that flat portions of the wall. Different resonator chambers may have different volumes within the same chamber height by using different spacings between the ribs  30  rather than different chamber depths. This is illustrated by chamber  48 B of  FIG. 5  compared with  48 C of  FIG. 6 , which have the same height H2 but different volumes. Such a height limitation prevents the larger resonators from extending outward into the air plenum  37  and impeding the flow therein. 
       FIG. 7  shows a resonator embodiment  78  with a relatively large chamber  48 D formed by a perforated resonator plate  64 C covering a large pocket  32  between ribs  30  of a given height H2, and further providing an intermediate rib  31  with less height across the pocket  32 , wherein the intermediate rib  31  does not reach the height of the resonator plate. This provides a damping frequency close to that of  FIG. 6 , based on the total volume of the chamber  48 D, while providing more structural reinforcement compared to the resonator  76  of  FIG. 6 . 
       FIG. 8  schematically illustrates a plurality of resonators of different sizes arranged on an outer surface of a wall of a flow directing structure  12 . A plurality of resonator sizes may be provided to damp a respective plurality of unwanted acoustic frequencies. 
     This example shows a row of each type of resonator previously illustrated herein— 60 ,  62 ,  76 , and  78 . The rows of resonators may encircle the flow directing structure or may be oriented along the working gas flow path or in any other direction. The shapes, sizes, number, arrangement, and orientation of resonators may be designed for each engine to optimize damping and rib reinforcement. The resonators do not need to be rectangular. They can be any shapes needed to fit around a flow directing structure that may be conic or irregular. Thus, the chamber shapes may include, but are not limited to, trapezoidal, triangular, hexagonal, and irregular. One or more airflow control manifolds may cover all or some of the resonators  60 ,  62 ,  76 ,  78 . In this non-limiting example, independent throttling T 1 , T 2 , and T 3  is provided by three manifolds  66  as shown in  FIGS. 5 ,  6 , and  7  respectively (not shown here for clarity). Each manifold/throttle T 1 , T 2 , T 3  variably meters compressed air inflow to a subset of the resonators, where each subset damps a different frequency or a different range of frequencies from the other subsets. In another embodiment, the manifolds may provide fixed metering, without an active throttle valve but providing different fixed metering for each subset of resonators. The different chamber volumes of the different sets of resonators may be sized to provide damping over a wide range, for example with peak frequency responses distributed over a range of 300-4000 Hz. 
       FIG. 9  shows frequency/airflow damping response curves  80 ,  82  for an exemplary resonator that is large enough to damp intermediate frequencies (50-1000 Hz). It is most effective at about 700 Hz with minimal purge air.  FIG. 10  shows estimated corresponding damping curves  80 ,  82  for a smaller resonator with peak effectiveness at about 1100 Hz with minimal purge air. The control logic  70  may control the throttle valves T 1 , T 2 , T 3  to vary each inflow based on such damping frequency/airflow response curves to optimize acoustic damping, cooling, and combustion temperature in the engine under varying operating conditions. 
     At low engine loads, more compressed air  50  is available than is needed for combustion. If this excess air enters the combustor, it cools the combustion zone  34 , increasing carbon monoxide (CO) emissions. Thus, at low engine loads, one or more of the resonator throttles  68 , T 1 , T 2 , T 3  may be opened, causing more compressed air  50  to bypass the combustor inlet, thus increasing combustion temperature to an optimum range, and reducing CO emissions. At higher loads including full rated power, maximum air is needed for combustion to avoid excessive temperatures in the combustion zone that increase nitrogen oxide emissions (NOx). Thus, under high loads, the resonator throttles may be closed, thus providing more compressed air to the combustion zone, which reduces its temperature to optimum range, and reducing emissions of NOx. This also maximizes damping effectiveness at high engine power when it is most needed, as exemplified by the sets of function curves in  FIGS. 9 and 10 . 
     Specific unwanted frequencies that occur at partial engine loads can be damped by minimizing the flow to certain subset(s) of resonators while not minimizing the flow to other resonators to avoid over-cooling the combustion zone. For example, for prolonged operation at ¾ load, the control logic may set throttles T 1  and T 2  half closed, and throttle T 3  fully closed to maximize damping of a specific intermediate frequency by larger resonators  78 . However, at ½ load, the control logic may fully open all throttles to minimize CO emissions, especially if ½ load is a short-term transitional stage. The apparatus herein provides a wide range of such options that can be selected by an operator and/or by predetermined control logic  70  to optimize the combination of noise reduction, emission reduction, and cooling over a wide range of operating conditions. 
     Some resonators may be specialized to damp specific frequencies that occur only during specific operating conditions. Under other conditions, the peak resonance (the peak of trace  80  in  FIGS. 9 and 10 ) in these specialized resonators may be partially or largely quenched by increasing the inflow (trace  82  in  FIGS. 9 and 10 ) to reduce resonance that could otherwise be audible and unneeded in such other operating conditions. 
     One benefit of this resonator design is elimination of welds on the hot wall  74 . These welds are a limiting factor in the prior design of  FIG. 4  (welds  56 ,  58  on wall  28 ). 
     The attachment points of the resonator cover plates  64 A,  64 B, and the manifold  66  are separated by a distance (exemplified by H1 and/or H2 herein) from the hot wall by the ribs  30 . The acceleration geometry  20  provides more pressure differential than is minimally needed to purge and cool the resonators  60 ,  62 . Reducing the airflow  50  entering in the manifold  66  under some operating conditions improves engine efficiency, cooling, and damping, and reduces emissions. Reducing the pressure of the compressed air in the manifolds under all conditions allows the coolant/purge inlet and exit holes  40 ,  44  to be larger, thus less susceptible to particulate clogging, and causes the film cooling  52  to flow more slowly and thus adhere better to the hot inner surface of the component wall  74  without overshoot. 
     Separate fabrication and attachment of side walls for the resonators is not needed when reinforcing ribs  30  are already provided on the casting of the flow directing structure to oppose the pressure differential previously described. In that case, these resonators take advantage of existing pockets between the ribs of the castings of the transition piece and exit piece. 
     While various embodiments of the present invention have been shown and described herein, it will be obvious that such embodiments are provided by way of example only. Numerous variations, changes and substitutions may be made without departing from the invention herein. Accordingly, it is intended that the invention be limited only by the spirit and scope of the appended claims.