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Chemical Kinetic and Thermodynamics of Flameless Combustion Methodology for Gas Turbine Combustors
Yeshayahou Levy†, G. Arvind Rao‡, Valery Sherbaum§
The Turbo and Jet Engine Laboratory, Faculty of Aerospace Engineering Technion, Israel Institute of Technology
Haifa 32000, Israel.
ABSTRACT The paper is concened with a novel combustion methodology based on the flameless combustion phenomenon. Flameless combustion is a newly discovered combustion regime and is one of the most promising technologies that can meet the stringent demands of reduced pollution and increased reliability in future gas turbine engines. In the present paper, chemical kinetics of a novel combustion methodology is carried out using the Chemical Reactor Modelling (CRM) approach. The proposed combustion scheme is designed with an objective of reducing oxygen concentration and temperature in the primary combustion zone, thus creating an environment where flameless combustion can be sustained. In addition, the fuel is injected in the recirculation zone where the oxygen concentration is low and the temperature is above the auto-ignition temperature. The thermodynamic and CRM analysis of the combustion methodology shows that transferring heat from primary combustion zone to the secondary (annulus) cooling air can reduce combustion temperature and oxygen concentration in the reactants, thus reducing NOx formation by a large margin. One more novelty of the proposed methodology is that the energy from fuel is released in two steps with intermediate cooling to reduce peak flame temperature within the combustor. The new proposed methodology with its internal conjugate heat transfer is compared vis-à-vis to the conventional methodology and the benefits are brought out explicitly. The NOx and CO emission of this new combustion methodology is found to be quite less as compared to the conventional gas turbine combustor.
For the proposed combustion methodology it is found that unlike in the conventional combustors, the NOx and CO emission behave similarly with various design parameters, thus providing the designer with new opportunities to reduce emissions from gas turbine combustors. Further, schematic of a practical engineering design working on the new proposed methodology is presented. This new approach of gas turbine combustor design is expected to change the way in which gas turbine combustors will be designed in the future.
NOMENCLATURE Cp specific heat at constant pressure [J/kg-K] D diameter of duct [m] K recirculation ratio [-] m mass flow rate [kg/s] Qex heat extraction transfer rate [W/kg] Qf calorific value of the fuel [MJ/kg] RR recirculation ratio [-] ℜ universal gas constant [= 8.3144 J⋅mol-1⋅K-1] T temperature [K] Greek scripts α primary to secondary mass flow ratio (primary air /secondary air) [-] β pre-combustion ratio [-] γ heat extraction ratio [-] ϕ equivalence ratio [-] Subscripts f fuel I ith species R refers to recirculated gases
† Associate Professor, Corresponding author; Tel.: +972-4-8293807; Fax: +972-4-8121604. E-mail: [email protected] (Y. Levy) ‡ Post Doctoral Fellow § Research Scientist
43rd AIAA/ASME/SAE/ASEE Joint Propulsion Conference & Exhibit 8 - 11 July 2007, Cincinnati, OH
AIAA 2007-5629
Copyright © 2007 by the American Institute of Aeronautics and Astronautics, Inc. All rights reserved.
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INTRODUCTION
With the global warming becoming one of the major problems faced by human kind and with the ever-increasing air traffic, emissions from aircraft can no longer be ignored. Since aircraft emit their pollutants in upper troposphere or lower layers of the stratosphere, the impact of these pollutants in deteriorating environment is more as compared to the land-based pollutants. Hence, there is an urgent need to reduce NOx, CO, UHC and other pollutants emitted from aircraft. Bahr [1] showed that NOx predominates both in the vicinity of the airport and during altitude cruise. For a long flight, the NOx emission would present a large fraction of the total emissions. It is anticipated that emission norms will become more stringent in the coming years, with the consequent need to reduce pollutants level drastically, especially NOx by more than 70% [2]. Figure 1 shows the main challenges faced by aviation today. The anticipated reduction at various fronts (noise, air pollution, fuel consumption and maintenance) required to meet the future challenges are also shown, the noise and air pollution being most prominent. It is expected that NOx, CO, UHC emissions should to be reduced by 70% of the present level by the year 2015, and by 80% of the present level by the year 2025 [2].
Figure 1. Main Challenges Faced by the Aviation Industry
The pollution created by land-based turbines is also a significant contributor towards air pollution because a significant number of thermal power stations use gas turbine engines to generate power. Some of the latest combustion Dry Low NOx (DLN) combustion technologies developed for land-based gas turbines include; Lean Premixed (LP) for gas fired units and Lean Premixed Prevaporized (LPP) combustors for liquid fuelled units. Additional methods are Staged Combustion, Rich-Burn/Quick-Quench/Lean-Burn (RQL), Catalytic Combustion and more. However, most of these technologies have met only limited success as far as their pollution control norms coupled with operational stability is concerned. In addition, none of the above-mentioned technologies has been ported to an aircraft jet engine. The inherent difficulties in applying these above-mentioned technologies to gas turbines have been discussed in detail by Arfi [3]. The mechanisms of pollutant formation and the methods employed in alleviating emissions from conventional gas turbine combustors has been given by Lefebvre [4]
There have been many combustion methodologies / schemes / techniques proposed to meet the stringent
emission norms. Flameless Combustion (FC) is a promising combustion regime that forms highly transparent flames (and hence the name Flameless Combustion) with low pressure oscillations, distributed combustion and uniform temperature profile within the combustor. Flameless Combustion is also referred to in the literature as FLameless OXidation (FLOX®) [5], High Temperature Air Combustion (HiTAC) [6, 7], Moderate and Intense Low oxygen Dilution (MILD) combustion [8, 9], and heat-recirculating combustion [10].
The operating range of the flameless combustion is shown in Fig. 2 [11]. It can be seen that Flameless
Combustion takes place at low oxygen concentration and high inlet temperature (above the auto-ignition temperature of the fuel air mixture) marked by the region ‘C’ in the figure below. This unique combination of high inlet air temperature and low oxygen concentration results in a lower adiabatic flame temperature as compared to the conventional diffusion flame. The Damköhler number is less and therefore the combustion is of distributed type.
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Figure 2: Different combustion regimes [11]
The high temperature and low oxygen content oxidizer can be achieved by dilution of fuel with vitiated combustion gases prior to combustion. High temperature combustion enhances thermal field uniformity, flame stability, and produces smaller temperature gradients relative to conventional combustion [12]. Flameless Combustion is characterized by
• Distributed combustion zone with uniform temperature and less number of hot spots, a prerequisite for reduced NOx production
• Spontaneous auto-ignition when the fuel air mixture reaches its auto-ignition temperature • Lower adiabatic flame temperature • Highly transparent flame with low acoustic / pressure oscillation • Recirculation of combustion products, which in addition to increasing oxidizer temperature reduces the
oxygen concentration; thus increasing the chemical reaction time and hence lowering the Damköhler number, a required condition for practically obtaining flameless combustion
FLAMELESSS COMBUSTION IN GAS TURBINE COMBUSTOR Application of FC has been successfully demonstrated in industrial furnaces [5, 10]. However, operational parameters for gas turbine combustors are quite different from those of industrial furnaces and makes the application of FC to gas turbine difficult. Several teams work worldwide in design & development of FC gas turbine combustors. Levy et al [13, 14] worked on a reversed flow combustor for small and medium size gas turbine engines (FLOXCOM). The FLOXCOM was successful in achieving very low levels of NOx and was among the first to demonstrate application of FC to gas turbines. A schematic of the FLOXCOM combustor is shown in Fig.3. Luckerath et al [15, 16] have demonstrated flameless combustion in a straight through-flow cylindrical combustor at high pressure up to 30 bars. The combustor had 12 equally spaced circumferential nozzles, with air and fuel injected from each nozzle at high velocity. The basic mechanism used by Luckerath et al [15] is to shield the fuel stream with cold air until there is sufficient mixing between recirculated combustion gases and the cold shielding air. Guoqiang et al [17] and Overman et al [18] have also used a similar technique for shielding the fuel with cold air. However, they had to operate near the lean blow out limit to achieve the required flame properties.
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(A) (B)
Figure 3: The reverse flow FLOXCOM combustor. A) Schematic of the flow inside the combustor[13]; B) Schematic of the FLOXCOM combustor in a gas turbine engine [19]
Challenges The initial studies on FC were performed on non-adiabatic type combustion systems, like the industrial furnaces. The reduced temperature level due to heat extraction from within the furnace allows operation of these furnaces at near stoichiometric conditions with the consequent lower O2 concentration in the burnt gases. Also these furnaces operate at higher overall equivalence ratio. In the case of adiabatic combustors, like that in a gas turbine engine, the thermodynamic parameters of the recirculating and the exhaust gases are determined only by the aerodynamic and thermodynamic processes inside the combustor.
The emission from a combustor is a function of temperature, residence time, concentration profile of various chemical species and the combustion process. For combustion to be complete, the ignition delay or ignition induction time should be carefully tailored to match the residence time. In addition, stabilizing the flame in a gas turbine combustor is difficult due to its smaller volume. Elevated pressures increase the chemical reaction rates and shorten the reaction time. Hence some of the main difficulties in applying the FC principles to a gas turbine engine can be summarized as:
• Gas turbine combustors operate at a low overall equivalence ratio and therefore O2 concentration in the combustor is very high. Since FC is a low O2 concentration combustion phenomenon, creating a suitable operational zone is difficult.
• The operating temperature and pressure is high as compared to an industrial burner, hence the operating conditions are much harsher.
• The gas turbine cycle efficiency increases with Turbine Inlet Temperature (TIT), pushing the average operational temperature even higher. This high operational temperature results in more NOx.
• The operational range of FC combustors is narrow and hence cannot satisfy the dynamic requirements of a gas turbine engine, especially that of an aircraft
• Since a large amount of combustion gases are circulated within the combustor, the volume required by FC combustors is large as compared to the conventional combustors, an important constraint for aircraft applications. This also results in lower volumetric heat densities as compared to existing conventional gas turbine combustors.
THE NEW COMBUSTION METHODOLOGY
The key to have a low NOx FC gas turbine combustor is to have a low O2 concentration within the combustion zone and a relatively even temperature distribution within the combustion chamber. The absence of hot spots and low temperature quenching zones is necessary for reducing NOx, CO and UHC. In order to meet the above mentioned stringent demands, a novel combustion methodology is proposed hereby. The schematic of the combustion methodology is shown in Fig. 4. The junctions have been marked in squares whereas the stations are marked in ellipses. Every junction has inlet and / or exit stations.
The inlet air is split into primary air and secondary air at junction “1”. The primary air is then injected into the combustor at junction “2”, where it is mixed with the hot recirculated combustion products. Part of the total combustion takes place in the path between junction “2” & “3”. At junction “3”, the combustion gases are split into two streams, one going to the recirculation zone, while the other towards the exhaust where it is mixed with the secondary air at junction “5”. In the recirculation zone, the combustion products drawn from junction “3” are cooled to some extent because of the heat transfer from recirculation zone to the secondary air stream, but temperature of
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combustion products is kept high enough for auto-ignition of the fuel-air mixture. Fuel is added in the recirculation zone, junction 4, the temperature here is above auto-ignition point and the O2 concentration is quite low. Only a certain fraction of the total fuel added at junction “4” is burnt in the pre-combustion zone (between junction “4” & “2”, precombustion zone), while rest is burnt in the main combustion zone between junction “2” & “3”. The secondary air is heated due to heat extraction from the recirculation zone. The heated secondary air is mixed with part of the high temperature combustion products at exit (junction “5”). The ratio of fuel burning in precombustion (between junction “4” & “2”) is dictated by: 1) the temperature and O2 concentration at junction “4”, which will determine the chemical reaction time for the process and, 2) the average velocity and the physical distance between junction “4” & “2”, which will determine the residence time available for the reaction. The remaining fuel is then burnt in the main combustion zone. Salient features of this concept as compared to the other proposed combustors for gas turbines are:
1. Fuel is injected into the O2 deficient and high CO2, H2O concentration recirculation zone, thus injecting the fuel in an optimum environment where FC can take place.
2. Certain amount of energy is transferred from primary combustion zone to the secondary cooling air, thus reducing combustion temperature and hence limiting the NOx formation.
3. Energy is added in two steps, partially in the pre-combustion region (between junctions “4” & “2”) of the recirculation zone and partially in the main combustion zone (between junctions “2” & “3”). Thus limiting the maximum temperature rise below the temperature above which NO formation would increase exponentially (≅ 2000K).
4. Intermediate cooling is used after heat addition to limit the temperature rise e.g. the recirculated combustion products are cooled to some extent between junction “3” & “4” (for example by using fins) before adding fuel at junction “4”, and the recirculated gases (after precombustion) is further cooled by the fresh air (at junction “2”) before the onset of main combustion.
Even though conventionally gas turbine combustors are adiabatic in terms of no heat transfer from the primary
combustion zone, it is found from the present exercise that heat transfer from main combustion zone to secondary cooling air has many benefits, as illustrated from the analysis in the subsequent sections. Heat can be extracted from the recirculated gases e.g. by the use of fins on both internal and external side of the recirculation zone or using jet impingement for enhancing the heat transfer coefficient of secondary cooling air.
Most of the other contemporary concepts are based on shielding the fuel by cold air till a sufficient amount of premixing is achieved; this technique can not be classified strictly under the Flameless regime described in Fig. 2 because the O2 concentration is more than that required for FC. The proposed concept (shown in Fig. 4) exploits all the facets of FC because it uses low O2 concentration in the reaction zone, high concentration of CO2 and H2O in the reaction zone, rapid mixing between various streams, and intermediate cooling to limit the temperature rise and NOx formation.
Figure 4: Schematic of the proposed new flameless combustion methodology
The NOx emissions is dominated by NO (called as Thermal NO) produced by the oxidation of atmospheric N2 in the high temperature zones of the flame. This endothermic process takes place only above 1850K, however it increases exponentially with the temperature. Thus it is necessary to limit the peak temperatures within the combustor to around 2000K. The thermal NO formation is given by the extended Zeldovich 4 step mechanisms given as O2 = 2O N2 + O = NO +N N + O2 = NO +O N + OH = NO +H The NO formation reaction is a relatively slow reaction and therefore the residence time should be less in order to reduce the NOx formation.
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THERMODYNAMIC ANALYSIS
The detail procedure for the thermodynamic analysis of the proposed methodology has been described in [20]. The Method of Successive Substitution [21] was used to solve the governing equations, which include heat balance equations and expressions for evaluating various terms involved in these equations at every junction of the cycle shown in Fig.4. Variables like temperatures and species concentrations are assigned initial values and then proceeding through the system of equations, all variables are recalculated and successively substituted until satisfactory convergence is achieved and all the constraints are met.
The main focus was to study the variation of temperature and species concentrations (O2, CO2, N2 and H2O) at each and every junction as this will dictate the formation of pollutants in the combustor. Only one step global reaction was considered for the thermodynamic computations. The calculations have been carried out for an inlet air mass flow rate of 1kg/s. The mass flow ratio between the primary and the secondary cooling air is dictated by the maximum permissible equilibrium temperature, which for this study is forced to be limited at 2050K. As elaborated earlier, the main aim is to have a uniform temperature distribution throughout the combustor and to have a low O2 and higher CO2 concentration in the combustion zone. It has been proved by Arfi [3] that a higher concentration of CO2 in the reactants decreases NO formation to some extent, e.g. the PSR studies carried out by Arfi [3] show that the total fixed Nitrogen (TFN) formation is reduced by two orders of magnitude when the reactants are diluted with 10% CO2 for a stoichiometric mixture with initial temperature of 300 K and with a residence time of 100 ms. The temperatures and concentrations of O2 and CO2 at every location in the combustor with a heat extraction of 80 KJ are shown in Fig 5. The value of 80 KJ/Kg of total inlet air was selected as a typical realistic value for such a system. The operating conditions are typical of that encountered in a gas turbine engine operating at 15 bars pressure. It should be noted that the analysis has been carried out with Methane (CH4) as the fuel and recirculation ratio value of 2.0. The effect of various parameters like heat extraction, recirculation ratio, etc on the cycle are described in the subsequent sections.
It can be seen from Fig 5 that while the maximum equilibrium temperature is below 2100K, the O2 concentration
in the recirculation zone is below 6% and the CO2 concentration is more than 7%. Such combination of temperature and species concentration is suitable for flameless combustion and it is expected that the flame would be of distributed type. The schematic of temperature variation along the combustor for this case, as compared to a conventional gas turbine combustor for the same inlet and outlet conditions, is shown in Fig 6. It can be seen that the peak temperature and the temperature gradient for the proposed new methodology are quite less as compared to that of a conventional combustor (the numbering correspond to junctions in Fig 4). Figure 6 shows that there are two small peaks of temperature as compared to the single large temperature peak in a conventional combustor, an indication that the new proposed cycle would reduce the formation of NO substantially as compared to other contemporary combustion technologies.
Figure 5: Calculated Chemical Equilibrium Temperature and Specie Concentrations at various Junctions for the Proposed Flameless Combustion Methodology (K = 1, β = 0.25, γ = 0.5)
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T1 = 700K O2 = 21% CO2=0.0% T51 = 765K
O2 = 23% CO2=0.0%
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Figure 6: Schematic representation of the temperature variation along combustor for the new methodology as compared to a conventional combustor
CHEMICAL KINETIC ANALYSIS
After the thermodynamic analysis of the newly proposed combustion methodology, it is important to perform the chemical kinetic in order to study the various characteristics of the proposed FC gas turbine combustor like residence time required for combustion, the concentration of species at various junctions inside the combustor, effect of recirculation ratio on the chemical kinetics inside the combustor and subsequent pollutant formation, effect of heat extraction, etc. The Chemical Reactor Modeling (CRM) approach has been used to study the above mentioned combustor characteristics. Bhargawa et al [22], Rutar & Malte [23], and Russco et al [24] have established the validity of CRM approach in modeling chemical kinetics inside combustors / furnaces to predict combustors emissions. Russco et al [24] used the CRM to evaluate exhaust emissions of a micro gas turbine combustor. They used CFD analysis of the combustor with 2 step global mechanism as the basis for building the CRM network. However since the present research is aimed at preliminary design of a combustor based on the newly proposed methodology, the CRM analysis will be used to build the reactor and subsequently modeled by CFD at a later stage.
The CRM constructed to model the chemical kinetics is shown in Figure 7. The reactor network consists of 11 Perfectly Stirred Reactors (PSR), 1 Plug Flow Reactor (PFR) and 2 inlet streams. The reactor network is solved using CHEMKIN® commercial software. The junctions corresponding to the thermodynamic cycle (of Fig. 4) have been shown in the Fig. 7 below. The fuel is injected at junction “4” (PSR10), in the recirculation zone. The recirculation zone is simulated by the four PSRs (PSR7-PSR10) with a cumulative residence time of 1.8ms. The fuel gets partially burnt in the PSR 10 before getting mixed with the fresh air in Junction 2 (PSR1). The region between junctions “2” to “3” are simulated by PSR1-PSR6, with a cumulative residence time of 3ms. The combustion gases are split after PSR6 (junction “3”). The recirculation ratio is dictated by the split in the mass flow rate between PSR6 to PSR7 and PSR6 to PSR11. The PSR 11 represents junction “5” of Fig. 4 where the combustion gases and the secondary bypass air mix together. The Plug Flow Reaction (PFR) simulates the kinetics in the flow after mixing between combustion gases and secondary air till the turbine inlet. The heat loss is distributed evenly across the four PSRs in the recirculation zone. The secondary air temperature increases due to heat extraction from the PSRs in the recirculation zone. The GRI 3.0 chemical kinetic mechanism [25] is used to model the kinetics in all the PSRs and PFR.
The combustor must provide appropriate residence time for optimum combustion. If the residence time is less, CO
formed in the primary combustion zone is quickly quenched by the secondary air, thereby increasing CO formation drastically. Also the combustion may not be complete if the residence time is too short. On the other hand, if the residence time is longer, the NOx formation increases because the thermal NO formation process is a relatively slow reaction. The CRM is simulated for various conditions to study the characteristics of the combustor. The results obtained from simulations for operational conditions typically encountered in gas turbine application are now discussed.
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m1 = 1kg/s m12 = 700 gm/s T1 = 700K m15 = 300 gm/s P1 = 15 atm RR = 2.0 mf = 21gm/s T6 = 1500K ϕ = 0.36 (global value)
Table1: Operational conditions for the simulation as typically encountered in gas turbine application
Figure 7: Schematic representation of Chemical Reactor Model for studying the chemical kinetics inside the proposed
novel combustion methodology
The variation of O2, H2O & CO2 concentrations at various junctions along the combustor (represented by non-dimensionalized combustor length) is shown in Fig. 8. The dotted line represents concentration in the recirculation zone. It can be seen that concentrations of CO2 and H2O are quite high within the recirculation zone before addition of the fuel at junction “4”. On the other hand, the molar fraction of O2 is quite low (around 10%) in the recirculation zone. Such a combination of species concentration is good for sustaining the FC within the combustor. Thus, contrary to the conventional combustor where fuel is added in the oxygen rich zone, adding fuel in the recirculation zone will form a distributed combustion zone and an even temperature distribution within the combustor.
Figure 8: Schematic representation of species concentration (O2, H2O & CO2 along the combustor. The junctions corresponding to Fig. 4 have been marked in the squares.
Apart from the temperature and species concentration, the concentration of radicals (O, H, OH, CH3) also plays an
important role in FC, as explained by Levy et al. [26]. They found that the radicals accelerate ignition and reduce the ignition delay time, hence reducing the residence time required for combustion to be complete. This is important because lower residence time means smaller volume required for the combustor. The variation of O, H and OH radicals is shown in Figure 9. The distribution is shown only in the main section of the combustor. The dotted lines
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represent concentration within the recirculation zone. It is seen from the figure that OH radical (which represents the temperature distribution) is evenly distributed between junction “2” and “3”. The O and H radicals have a short lifetime and therefore their concentration falls off rapidly as compared to OH.
Figure 9: Variation of OH, H, O radicals at various junctions along the combustor. The junctions corresponding to Figure 4 have been marked in the squares
The concentration of NOx (NO2, NO & N2O) is shown in Fig. 10. The NO, formed by the Zeldovich mechanism
described earlier, is the main contributor in NOx, and is an order of magnitude more than NO2 and N2O. Whereas the NO level is around 15 ppm, the N2O and NO2 are in the sub ppm levels. It can also be seen from Figure 10 that while N2O and NO2 decreases along the combustor length, the thermal NO formation continues. This proves that NO formation reaction is a slow reaction and NO increase with increase in the residence time of the combustor.
Figure 10: Schematic representation of species concentration (NO, NO2 & N2O) along the combustor. The junctions corresponding to Figure 4 have been marked in the squares
One of the other main pollutants from gas turbine engines apart from the oxides of nitrogen is the carbon monoxide
(CO). The variation of CO within the combustor is shown in Fig. 11. It can be seen that the CO is basically formed in the high temperature combustion zone. However it is interesting to see that most of the CO is oxidized in the downstream and within the recirculation zone due to the relatively high temperature in the recirculation zone, hence preventing the CO from getting quenched. The concentration of CO at the combustor exit is found to be in the sub ppm
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level, an indication that the combustion was complete within the given residence time. This is very important because in a gas turbine combustor, the combustor length is primarily decided by the residence time required for oxidation of CO. The even temperature distribution within the proposed FC combustor reduces residence time required for CO oxidation, hence allowing the combustor volume to be smaller, an important criteria for aircraft application where volume inside the gas turbine engine comes at a high premium.
Figure 11: Schematic representation of CO concentration at various junctions along the combustor. The junctions corresponding to Figure 4 have been marked in the squares
SENSITIVITY ANALYSIS
From the analysis in the earlier section, it can be said that the newly proposed FC combustor has the capability to meet stringent demands posed by the projected future environmental regulations. The reactor network described in Fig. 7 is now used to perform the sensitivity analysis of the proposed FC combustor to various design parameters in order to study the combustor characteristic and to optimize the combustion methodology. Effect of Recirculation Ratio (K) The main novelty of this combustion methodology is to inject the fuel within the recirculation zone, where the O2 concentration level is low and the concentration of CO2 and H2O is quite high. Thus injecting the fuel in a high temperature low O2 concentration region and creating an environment where Flameless Combustion can sustain itself. The recirculation ratio is defined as the mass of recirculated gases that enter the main combustion area at junction “2” divided by to the mass of fresh air and fuel. The recirculation ratio will dictate the mass flow inside the recirculation zone of the combustor, the species concentration, the temperature distribution, and the volume required for the combustor. Thus the recirculation ratio has a significant effect on the overall combustor design and performance.
Figure 12 shows the variation of temperature with the recirculation ratio at station “23”, “35” and “42” in the
combustor. It is seen from the figure that whereas the temperature in the pre-combustion zone decreases with RR, the temperature at station “23” increases with RR. Thus the peak temperature in the combustor reduces with RR and the combustor average temperature increases with RR. Also as the recirculation ratio is increased, the temperature gradients (T35-T23), (T42-T23), (T42-T35) within the combustor are reduced and the temperature distribution becomes uniform.
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Figure 12: Variation of Temperatures at station “23”, “35”, & “42” in the combustor for various recirculation ratios
The effect of RR on O2, H2O and CO2 concentration at stations “23”, “35” and “42” is shown in Fig. 13. It can be
seen that if the recirculation ratio is more than about 1.5, the combined concentration of combustion dilutants (H2O and CO2), is more than the oxidizer (O2) concentration. At all other stations, the O2 concentration is much less than the dilutants concentration. This slows down the reaction by increasing the chemical reaction time when compared to the physical mixing time. Thus it lowers the Damköhler Number, leading to Flameless Combustion. Also as stated earlier, the production of NO and thermal gradients are suppressed due to increase in CO2. In addition, the thermal gradients within the combustor are suppressed due to the higher heat capacity of H2O and CO2, consequently reducing the peak temperatures encountered within the combustor.
Figure 13: Variation of Oxygen Concentration and dilutants concentration (CO2 + H2O) at station “23”, “35”, & “42” for various recirculation ratios
The effect of recirculation ratio on NOx formation is shown in Fig. 14. It can be seen that NOx formation reduces
exponentially with RR till a recirculation ratio of 2.5 and later achieves an asymptote. As explained earlier, thermal NO dominates the overall NOx. The NOx is reduced as a result of reduction in the peak temperature due to increase in the CO2 and H2O concentration, and reduction in the O2 concentration. At higher RR, the NO concentration is comparable to N2O and NO2 concentration. This proves that the proposed combustor is capable of achieving low NOx combustion without sacrificing much on the operational characteristics of the combustor, as required in an aircraft gas turbine engine.
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O2_23O2_35O2_42CO2+H2O_23CO2+H2O_35CO2+H2O_42
Tem
pera
ture
(K)
1400
1600
1800
2000
2200
2400
1 1.5 2 2.5 3 3.5 4
T_23T_35T_42
Recirculation Ratio
35
42
23
12
Figure 14: Variation of NOx concentration at stations “23”, “35”, & “42” in the combustor for various recirculation
ratios
The effect of RR on CO formation is shown in Fig. 15. As in the case of NOx, the CO decreases with RR up to about 2.5 and then achieves an asymptotic value. Even at low RR, the CO is in the sub ppm level. In the conventional combustion methodology, it is well know that any operational / design parameter that reduces NOx, results in an increased production of CO. Thus there is only a small operating regime in which both NOx and CO are not too high. However in the proposed combustion methodology, it is proved from comparison of Fig. 14 and 15 that both NOx and CO can be kept to very low values over a wide range of operational parameters.
Figure 15: Variation of CO concentration at stations “23”, “35”, & “42” in the combustor for various recirculation ratios
Effect of Primary to Secondary air Ratio (α) The split between primary air and secondary air mass flow rate is crucial in dictating the peak temperatures and various species concentrations encountered within the combustion zone. Thus α affects the pollutant formation substantially, especially thermal NO. The secondary air is also required for the cooling of the combustor liner and will dictate the heat absorbed from the recirculation zone of the combustor. Figure 16 shows the effect of primary to secondary air ratio (α) on the temperatures encountered at various junctions within the combustor for a recirculation ratio of 2.0. It can be seen that as the mass of air in the primary stream increases, the temperature encountered at all junctions reduce because of the increase in the total mass flow rate (both in the fresh air and recirculated combustion gases) and the decrease in the equivalence ratio.
0.0E+00
5.0E-05
1.0E-04
1.5E-04
2.0E-04
2.5E-04
3.0E-04
3.5E-04
1 1.5 2 2.5 3 3.5 4
NOx_23NOx_35NOx_42NOx_6
Mol
ar F
ract
ion
Recirculation Ratio
1.0E-07
1.0E-06
1.0E-05
1.0E-04
1.0E-03
1.0E-02
1.0E-01
1 1.5 2 2.5 3 3.5 4
CO_1CO_6CO_10CO_12 M
olar
Fra
ctio
n
Recirculation Ratio
42
23
35
6
13
Figure 16: Effect of Primary to Secondary Air Ratio on Temperatures at station“42”, “35” “23”& “6” in the combustor
The overall effect of α on NOx formation is shown in Fig. 17. The NOx formation reduces exponentially with
increase in α because of decrease in the temperature and increase in the mass flow rate. Similarly it is found that CO also reduces substantially with increase in α (as shown in Fig. 18). It can be seen that contrary to the conventional combustors, the CO formation reduces with reduction in the temperature. This is because of the higher average combustor temperature due to higher recirculation ratio. Thus it can be seen that unlike in the conventional combustor, the NOx and CO in the proposed combustor behave similarly with different design parameters. This is very important because in the conventional combustion methodologies, the behavior of CO and NOx with design parameters is opposite, therefore fundamentally hampering any attempt to reduce emissions. However in the proposed combustor, this fundamental difference is not present, thus giving new opportunities to the designer in reducing emission.
Figure 17: Effect of Primary to Secondary Air Ratio on NOx concentration (NO, N2O, NO2) at station“42”, “34”,“23” and “6” within the combustion cycle
1400
1600
1800
2000
2200
2400
1 1.2 1.4 1.6 1.8 2 2.2 2.4
Primary to secondary air ratio
Tem
pera
ture
(K)
6
23
35
42
T_42 T_35 T_23 T_6
0.0E+00
5.0E-05
1.0E-04
1.5E-04
2.0E-04
2.5E-04
3.0E-04
3.5E-04
4.0E-04
4.5E-04
1 1.2 1.4 1.6 1.8 2 2.2 2.4
NOx1NOx6NOx10NOx12
Primary to secondary air ratio
Mol
e Fr
actio
n
42
35
23
6
NOx_42 NOx_35 NOx_23 NOx_6
14
Figure 18: Effect of Primary to Secondary Air Ratio on CO concentration at station“42”, “34”, “23” and “6” within
the combustion cycle Effect of Heat Extraction Heat transfer from recirculation zone to the secondary air can have significant effect on the thermodynamics of entire combustor in terms providing the ability to increase equivalence ratio within the primary zone while keeping the same peak combustion temperature. This reduces oxidizer concentration and increases the concentration of CO2 & H2O. Heat extraction also suppresses formation of NO by lowering the peak temperature for constant equivalence ratio or by increasing concentration of CO2 and H2O due to higher equivalence ratio. Also when the combustion takes place at near stoichiometric ratio, the stability of combustion is enhanced, an essential requirement to be satisfied by any combustor being considered for gas turbine applications. As discussed by Arfi [3], increase in CO2 concentration reduces temperature rise because of its high heat capacity, also the kinetic effect of CO2 may favor recombination reaction of HCN species to N2, thus reducing the total NOx emission. The effect of heat extraction on the total NOx and CO emission of the combustor for 15% O2 is shown in Fig. 19. The NOx level decreases with increase in the heat transfer from the recalculating gases to the secondary cooling air. The CO level also decreases with increase in the heat transfer rate, however the effect is not as prominent as for NOx. This again proves that for the newly proposed combustor, the CO and NOx behave similarly for different design parameters.
Figure 19: Effect of heat extraction from recirculating gases on combustor emissions
1.0E-07
1.0E-06
1.0E-05
1.0E-04
1.0E-03
1.0E-02
1.0E-01
1 1.2 1.4 1.6 1.8 2 2.2 2.4
CO1CO6CO10CO12
Primary to secondary air ratio
Mol
e Fr
actio
n
CO_42 CO_35 CO_23 CO_6
42
35
23
6
1.E-08
1.E-07
1.E-06
1.E-05
1.E-04
0 20 40 60 80 100 120
CO
NOx
Heat Extraction (KJ)
ppm
@ 1
5% O
2
0.1
0.01
1
10
100
15
A Practical FC Combustor for Gas Turbine
A schematic of the conceptual gas turbine combustor operating on the proposed combustion methodology is illustrated in Fig. 20. The recirculation zone is also shown in the figure. The fuel is injected in the toroidal recirculation zones formed due to the high primary air inlet velocity. The various junctions and stations described in the thermodynamic model (Fig. 4) are also depicted in the Fig. 20 for better understanding. Figure 20 also depicts the main combustion zone and the mixing zone formed due to mixing of secondary cooling air with the combustion gases.
In the conventional gas turbine combustors secondary air is continuously mixed with combustion gases over the
entire combustor length for film cooling, thus prohibiting formation of an oxygen deficient zone within the combustor. However for establishing the flameless regime in a combustor, it is essential to use a heat transfer mechanism for combustor liner cooling without any mass exchange. For this particular design, fins can also be used for transferring heat from the recirculation zone to the secondary cooling air.
Preliminary analyses of the heat extraction process in the proposed combustor suggest that around 10% of the total
energy released by fuel can be transferred to the secondary cooling air. The heat transfer mechanism involves all the three modes of heat transfer, i.e. conduction, convection and radiation. In a conventional combustor, the flame is anchored in a relatively small volume of the primary combustion zone, and therefore the contribution of heat transfer to the combustor walls is mainly by convection. Since in the proposed combustor, the combustion is of distributed type and the combustion volume is larger, and also since a large volume of the combustor is occupied by high temperature recirculating combustion gases, the contribution of radiation heat transfer will be much more than that for a conventional combustor. Even though the fins for heat transfer enhancement cause disruption of flow in the primary and secondary streams, it is anticipated that the pressure drop across this novel combustor would be of the same order as it is for the conventional combustor because of the following reasons • The total pressure loss in a conventional combustor is due to diffusion, mixing, swirling and friction. In the
proposed combustor, it is not necessary to diffuse the compressed air to very low inlet velocities because high air inlet velocities can be used. In fact, higher air velocity will enhance the recirculation ratio due to its ejector action and will also reduce the Damköhler number.
• Since there is less need for anchoring the flame (the flame is distributed), there are no swirling loses. • The mixing loses are also expected to be less as compared to a conventional combustor because the temperature
difference between the primary and the secondary air is lower. Also the number of holes required for mixing the primary and secondary streams would be much less as compared to that of a conventional combustor using film cooling.
• Thermal radiation would form a substantial part of the total heat transfer between the recirculation gases and the secondary cooling air due to the distributed flame, and uniform temperature. Thus the heat convected out by fins can be optimized for minimizing the overall incurred pressure losses.
• Alternative heat transfer techniques, like jet impingement, can also be used to enhance heat transfer between secondary air and the combustor wall.
Hence it can be said that the proposed combustor exploits all facets of flameless combustion, i.e. low O2
concentration, high concentration of CO2 and H2O in the reactants, distributed combustion and even temperature distribution. One of the other main constraints in a gas turbine combustor is CO quenching at the cold walls due to excessive cooling or improper residence time. However in the proposed conceptualized combustor working in the FC regime, the CO quenching phenomena is less likely to occur because the average temperature within the combustor is higher than the fuel air mixture auto-ignition temperature (an essential element for the establishment of a distributed combustion)
In addition, having a dedicated specified physical recirculation zone in the combustor will help to increase the
combustion stability. The unique combination of these favorable characteristics is expected to give a new direction to the philosophy of gas turbine combustion. A schematic representation of a prototype combustor with fins for heat transfer is shown in Fig. 21. Composite metallic fins have been considered for enhancing heat transfer from the recirculation zone to the secondary cooling air. It should be noted that the combustor shown in Fig. 21 is generic in nature and can be ported to can / can –annular or an annular type of combustor. Alternative cooling techniques like jet impingement can also be used to enhance the convective heat transfer coefficient.
16
Fuel Inlet
Fuel Inlet
Inlet Exit
Composite Metallic Fins
Secondary Air
Primary Air
Mixing holes
Fuel Inlet
Fuel Inlet
Inlet Exit
Composite Metallic Fins
Secondary Air
Primary Air
Mixing holes
Figure 20: Schematic of a Gas Turbine Combustor Operating on the Newly Proposed Cycle
Figure 21: Schematic of a Gas Turbine Combustor Operating on the Newly Proposed Cycle
CONCLUSIONS
A new methodology for gas turbine combustors operating on flameless combustion mode is presented. The new
scheme employs heat transfer from recirculating gases to the secondary cooling air and hence, is “non-adiabatic” as far as the primary combustion zone is concerned; however it keeps the “adiabatic” regime for the whole combustor. It is observed that heat transfer enables the designer to reduce O2 concentration in the reactants substantially, and helps to create an environment where flameless combustion can sustain itself. For the proposed combustor, the O2 concentration at various junctions within the combustor has been observed to be much lower than that for a conventional combustor, or any other contemporary concepts. Also CO2 and H2O concentration in the reactants are observed to be higher.
The chemical kinetics is simulated using the CRM approach. The GRI 3.0 mechanism is used in the simulations. The influence of important design parameters on the combustor kinetics has been studied. The emissions are found to be much lower, NOx being less than 20 ppm and CO being less than 1ppm for most of the cases. Hence it is expected that the proposed combustor will be able to meet the stringent demands of lower emissions and increased reliability posed to the future gas turbine combustors. It is found that the unlike in the conventional combustor, the NOx and CO behave similarly with different design parameters. This is very important because in the conventional combustion methodologies the behavior of CO and NOx with design parameters is opposite, therefore fundamentally hampering
Fuel
Fuel
4
4
1 3 52
Heat Transfer Surface
Primary air
Secondary air
Secondary air
Recirculation Zone
Recirculation Zone
Fuel
Fuel
44
44
11 33 5522
Heat Transfer Surface
Primary air
Secondary air
Secondary air
Recirculation Zone Recirculation Zone
Recirculation Zone Recirculation Zone
Combustion
Zone
Mixing
Zone
17
any attempt to reduce emissions. However in the proposed combustor, this fundamental difference is not present, thus giving new opportunities to the designer in reducing emission.
A practical design of a combustor operating on the newly proposed methodology is also presented. The proposed
methodology presents an attempt to build a combustor operating on the flameless combustion mode for gas turbines while exploiting all the facets of flameless combustion, which include, low O2 concentration in the reactants, high concentration of CO2 and H2O in the reactants, lower peak temperatures and thermal gradients, and a more even temperature distribution within the combustor. In addition due to the internal heat extraction, the combustion can take place at near stoichiometric conditions and hence will exhibit superior stability characteristics. The stability is further enhanced by the combustor’s dedicated recirculation zone. This unique combination of favorable characteristics is expected to give a new direction to the philosophy of gas turbine combustion, which is poised for a major technological change.
ACKNOWLEDGEMENT
The authors are thankful to the Israel Council for Higher Education, Govt. of Israel for supporting the research. The authors also thank Vladimir Erenburg for the help with computer systems.
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