A novel air injection scheme to achieve MILD combustion in a can-type gas turbine combustor

Journal Pre-proof A novel air injection scheme to achieve MILD combustion in a can-type gas turbine combustor Saurabh Sharma, Arindrajit Chowdhury, Sudarshan Kumar PII:

S0360-5442(19)32514-9

DOI:

https://doi.org/10.1016/j.energy.2019.116819

Reference:

EGY 116819

To appear in:

Energy

Received Date: 22 July 2019 Revised Date:

8 October 2019

Accepted Date: 21 December 2019

Please cite this article as: Sharma S, Chowdhury A, Kumar S, A novel air injection scheme to achieve MILD combustion in a can-type gas turbine combustor, Energy (2020), doi: https://doi.org/10.1016/ j.energy.2019.116819. This is a PDF file of an article that has undergone enhancements after acceptance, such as the addition of a cover page and metadata, and formatting for readability, but it is not yet the definitive version of record. This version will undergo additional copyediting, typesetting and review before it is published in its final form, but we are providing this version to give early visibility of the article. Please note that, during the production process, errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain. © 2019 Published by Elsevier Ltd.

A Novel Air Injection Scheme to Achieve MILD Combustion in a Can-type Gas Turbine Combustor Saurabh Sharmaa*, Arindrajit Chowdhurya, Sudarshan Kumarb a

Department of Mechanical Engineering, Indian Institute of Technology Bombay, Powai, Mumbai 400 076, India

b

Department of Aerospace Engineering, Indian Institute of Technology Bombay, Powai, Mumbai 400 076, India

*

Corresponding Author: Saurabh Sharma,

Department of Mechanical Engineering, Indian Institute of Technology Bombay, Powai, Mumbai, India, 400 076 Email: [email protected]

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Abstract This article presents the design and development of a can-type gas turbine combustor operating in flameless combustion mode with liquid fuels. The combustor operates with kerosene and thermal intensities varying from 5.1 – 7.5 MW/m3. A novel air-injection scheme is proposed, in which air is supplied from different injection holes namely, swirl air near fuel injection, primary, secondary and dilution air in the downstream. These air injection holes are arranged in a way to help create strong recirculation of hot combustion products leading to increased mixing and dilution of incoming fresh reactants. Direction, orientation and mass fraction of swirl, primary, secondary and dilution air are optimized through a series of reacting flow simulations aimed at maximizing the reactant dilution ratio, a key parameter to achieve flameless combustion. It is observed that if the momentum of the air is increased beyond a critical value, the combustor switches its operation into flameless mode due to increased mixing and dilution of fresh reactants with hot combustion products. Combustion occurs in a well-distributed reaction regime within the combustor volume. Measured NOx emissions are less than 5 ppm and acoustic emissions are significantly reduced during the combustor operation in flameless combustion mode. Keywords: Can combustor, MILD combustion, Swirl air injection, Emissions, Gas turbine combustion

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1. Introduction The norms for emission control are expected to become stringent in the near future. Flightpath’s 2050; a report by advisory council for aeronautics research in Europe (ACARE) proposes that the NOx emissions and noise from aviation engines should be reduced by 90 % and 65% respectively compared to a typical aircraft flying in 2000 [1]. New combustion technologies are therefore required to introduce a better compromise between high combustion efficiency and lower emissions. A few emission-reduction techniques, such as lean direct injection, trapped vortex combustor and flameless oxidation have been investigated for their application in gas turbines. The first two concepts are incapable of maintaining high overall efficiencies while keeping ultra-low levels of emissions. Flameless/MILD (moderate or intense low-oxygen dilution) combustion appears to be a suitable alternative due to its well-known characteristics such as well-distributed combustion, low NOx and acoustic emissions and enhanced combustion stability [2-4]. Low NOx emissions in flameless combustion mode are achieved by a) Reducing the peak temperature and its fluctuations in the combustor through homogeneous thermal field and, b) Reducing the oxidizer concentration by diluting the fresh reactants with hot combustion products. The requirement of wide operational limits at part load conditions in gas turbines requires enhanced recirculation of hot combustion products to achieve increased flame stability. Although flameless combustion has been successfully applied to industrial furnaces, its direct application to gas turbine combustors still lies in theoretical concepts [5]. It is due to, (a) Difficulties in achieving the desired recirculation and dilution at high heat release densities, (b) Requirement of wider operational range in terms of global equivalence ratio, φ (0.3-1.0), (c) Forward flow configuration of the combustor similar to a can-type combustor and (d) Minimum pressure losses in the combustor with high heat release densities and low residence times. These issues make it difficult to achieve high recirculation rates and lower oxidizer 3

concentration in the combustor volume; the requisite conditions for achieving flameless combustion mode. Flameless combustion was first described by Wunning et al. [4] as flameless oxidation (FLOX), in which a stable form of combustion with very low NOx emissions was reported for specific conditions of furnace temperature and recirculation ratios. Flat thermal field, low pollutant and acoustic emissions are observed due to the distributed nature of combustion process [4]. Flameless oxidation is also referred to as high temperature air combustion (HiTAC) [6], MILD combustion [7] and flameless combustion [7]. Flameless/MILD combustion was extensively studied and implemented for industrial furnaces and burners [2, 3, 6-15]. Plessing et al. [13] studied the characteristics of flameless oxidation in a reverse flow furnace fitted with a FLOX burner. They concluded that combustion process in flameless mode can be modeled as a well-stirred reactor (WSR), because the turbulent time scales were comparatively much smaller than the corresponding chemical time scales [13]. The same furnace was used by Dally et al. [9] to investigate the effect of fuel composition on the stability of MILD combustion. They concluded that addition of the inert gases to the fuel jet reduces the NOx emissions and extends the stability of MILD combustion over a wide range of operating conditions [9]. Dally et al. [8] studied the turbulent non-premixed flames using a jet in hot coflow (JHC) burner. They observed that the peak temperatures and CO emission levels were significantly reduced, when the O2 levels were dropped from 9% to 3% [8]. Medwell et al. [12] further studied the same burner through OH, temperature imaging and concluded that partial premixing takes place at higher Reynolds number, leading to local flame extinction at reduced O2 levels [12]. Kumar et al. [10, 11] studied the combustion characteristics in a forward flow MILD combustion burner at higher energy densities (~2 - 10 MW/m3) and

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thermal power levels (3 - 150 kW). The flameless combustion mode was achieved through internal recirculation of the hot combustion products through geometric modifications in the combustor. Verissimo et al. [14] studied the effect of global equivalence ratio on the emissions characteristics of FLOX combustor in a closed environment. They reported that NOx emissions are reduced significantly due to increased mixing in the combustor [14]. Kruse et al. [16] reported that NOx emissions increases at elevated pressures due to increased residence time and different NOx formation pathways. Ye et al. [17] investigated the combustion of pre-vaporized liquid fuels at high pressures in the same test rig [16]. They reported that fuel type affects the stability of combustion for similar levels of emissions [17]. The studies of Sorrentino et al. [18] in a cyclonic combustor showed that cyclonic flow field helped achieve complete combustion. However, high residence times (~0.5 s) are not suitable for the typical gas turbine applications [18]. Preliminary efforts on application of flameless combustion to gas turbines have been reported in ref. [19-21]. Luckerath et al. [20] reported a narrow operational range with low CO and NOx emissions for flameless combustion regime in a hexagonal combustor at high pressures (~ 20 bar). Lammel [19] adopted a premixed fuelair configuration called HiPerMix and reported ~ 10 ppm NOx emissions along with a substantial pressure loss [20]. Melo et al. [21] experimentally verified the findings of Levy et al. [5] by developing a prototype annular combustor with very low NOx emissions. However, the levels of CO and unburnt HC emissions were relatively too high for gas turbine applications. Reddy et al. [22-25] studied the flameless combustion of liquid fuels in a swirl stabilized high heat intensity combustor (21 MW/m3 ~85 kW) with tangential air injection. Various investigations on fuel flexibility with flameless combustion mode were reported by Ye et al.

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[26], Reddy et al. [22], Azevedo et al. [27] and Sharma et al. [28]. Biodiesel showed different emission characteristics compared to kerosene and diesel due to significantly different thermophysical properties of the biodiesel [22]. Sharma et al. [29] studied the effect of the spray parameters on the combustion characteristics of a flameless combustor. They reported that fine spray improves the combustion characteristics due to enhanced mixing. Although significant work has been reported in the field of flameless combustion, its application to a gas turbine combustor with a can-type or can-annular type configuration is still in preliminary stage. It is due to the difficulties involved in achieving the desired recirculation of hot combustion products within the combustor at high heat release densities. The requirement of low oxygen concentration environment to achieve flameless combustion makes its application difficult in a typical gas turbine combustor due to lean operational range (φ = 0.3 - 0.4). It is also observed that there are few studies available in literature on direct use of liquid fuels in a flameless combustor [22-25, 28-31] in contrast to prevaporized liquid and gaseous fuels. The present work deals with the design and development of a novel can-type gas-turbine combustor operating in flameless combustion mode with liquid fuel (kerosene). The present combustor has novel geometric features which allows its operation in flameless mode. A unique air injection scheme is proposed to achieve the desired recirculation inside the combustor at high heat release densities and low global equivalence ratios (φ = 0.3 - 0.4). Initially, the evolution of geometry is discussed to maximize the recirculation of hot combustion products inside the combustor. Then, detailed computational studies are performed on the optimized combustor to understand the flow physics inside the combustor. Experimental studies are carried out to demonstrate the achievement of flameless combustion mode in the proposed combustor and measure the temperature profiles, pollutant and acoustic

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emissions. This is one of the first studies focused on highlighting the development of a cantype combustor operating in flameless mode with liquid fuel. The geometric layout of the combustor is such as it can be integrated to a conventional gas turbine engine directly. Ultralow gaseous and acoustic emissions with wider operational limits are expected to be advantageous for gas turbine applications.

2. Cold flow spray studies The spray characteristics of a solid cone-type, pressure swirl nozzle are measured for nonreacting conditions using a laser based shadowgraphy setup. A Quantel made Nd-YAG laser with model number EVERGREEN EVG00145 is used for spray illumination. This laser produces an infrared wavelength of 1064 nm, which is converted to visible 532 nm using a polarizer and second harmonic generator [32]. A single nozzle, N1 is used at two different injection pressures to achieve two different fuel flow rates or thermal inputs. Kerosene is considered as the fuel for the present study. Since higher injection pressure improves the combustion characteristics [29], the atomizer is calibrated and tested at 9 and 14 bar injection pressures to achieve the thermal inputs of 20.6 and 30 kW. Various details of atomizer such as the flow rates and fuel injection velocities are summarized in Table 1. Table 1: Summary of atomizer details, injection pressures, flow rates and flow velocities, 9N1: N1 atomizer at 9 bar injection pressure, dnozzle = 0.187 mm. Sr No. 9N1

Pinj (bar) 9


 (kg/hr) 1.70

  (kW) 20.6

 (MW/m3) 5.1

 (m/s) 22.04

14N1

14

2.50

30.0

7.5

32.40

Shadowgraphy studies are performed on a separate non-reacting flow set-up to measure the spray parameters such as, Sauter mean diameter (SMD), spray velocity and droplet number

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density (DND). The details of the set-up are discussed elsewhere [28] and the results obtained from the cold-flow study are summarized in Table 2. Table 2: Summary of spray characteristics with kerosene fuel Pinj (bar)

Cone angle (º)

SMD(µm)

9

45

35

DND w.r.t. SMD(104/cm3) 1.05

14

56

34

1.71

3. Evolution of the combustor geometry Fig. 1 shows the schematic diagram of a conventional can-type combustor used in gas turbines. It consists of three different zones namely, primary, secondary and dilution zones. Air is supplied through different air injection holes such as swirler, primary holes, secondary holes and dilution holes. The fuel is combusted in the primary and secondary zones followed by the cooling of the combustion products in the dilution zone. The desired air flow is ensured using an air casing around the can. Recirculation zones are established mainly in the primary zone with the help of an air swirler as shown in the Fig. 1. For the present study, a similar layout of the can-type combustor is considered for gas turbines. The fuel supply along with different air injections are kept almost similar to that of a conventional gas turbine combustor design as seen in Fig. 1. Efforts are made to enhance the internal recirculation of the hot combustion products in the combustor through various aerodynamic means.

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Fig.1. Schematic diagram of a conventional can combustor A typical can-type gas turbine combustor geometry operating in diffusion combustion mode is considered in the present work. Fuel and air are injected using swirl injection from an upstream location as shown in Fig. 2a. The geometry is modified aerodynamically with an aim to increase the mixing and dilution of the reactants with the hot combustion products. Different configurations have been investigated computationally to maximize the reactant dilution ratio, Rdil, a non-dimensional parameter. Rdil is a critical parameter to achieve flameless combustion mode [24]. It is defined as the ratio of the net mass flow interaction at a given plane between reactants and exhaust gases to the total mass flow of the reactants [24]. It is defined as,

 =

|  |(    )

(1)

(    )

.

Where, m axial

is the total mass flow of hot combustion products at a given plane in the .

combustor, normal to the axis of the combustor and defined as m axial = ∫∫ ρ vaxial dxdz , here

vaxial is the axial velocity.
 !"
 are the oxidizer and fuel mass flow rates respectively. Axial velocity and density at each cell center were obtained from the numerical calculations and Rdil was calculated using Eq. 1 at different axial locations inside the combustor. 9

(a)

(b)

(c)

(d) Fig. 2. Selection of the combustor geometry

The combustor geometry is optimized with an aim to increase the reactant dilution ratio based on the computational studies. The reactant dilution ratio is enhanced by altering the injection of the primary, secondary and dilution air in the combustor through aerodynamic means to minimize the pressure losses in the combustor. Initially, a preliminary combustor design shown in Fig. 2a is investigated computationally. Air is supplied from four different sections; 10

(i) Swirl, (ii) Primary, (iii) Secondary and (iv) Dilution air, similar to a conventional gas turbine combustor. Swirl air is injected along with the fuel using a swirler. Primary air is injected through tangential air inlets in the combustor. Secondary and dilution air are injected through tangential air inlets towards upstream direction to enhance the recirculation of hot combustion products. It was observed from the flow field analysis that the recirculation zones are established within the combustor; however, they were located very close to the combustor wall region. The combustor design was then modified by maintaining the same orientation of the air flow for primary, secondary and dilution air as shown in Fig. 2b. The reactant dilution ratio, Rdil improved marginally compared to the earlier design, however, the location of the recirculation zones remained in the near wall region. More efforts were invested for improving the recirculation of hot combustion products to help avoid wall cooling through the relocation of the recirculation zones near the wall region. The location of the dilution air injection holes was shifted further downstream to the curved portion of the combustor near the exit as shown in Fig. 2c. The combustor shape was modified further by adopting a slightly conical geometry. It was aimed at providing more space to aid increased recirculation of the hot combustion products in the combustor. To maintain the same combustor volume, the length of the combustor was slightly reduced. This helped reduce the effect of air injection in the near wall region. The boundary of the recirculation zones was extended further downstream due to the shifting of the dilution holes to a downstream location. A maximum value of Rdil of 2.5, 2.8 and 3.0 (as discussed in section 3.3) was obtained for the cases of Fig. 2a (D1), 2b (D2) and 2c (D3) respectively. It was observed that the effect of the air injection was localized in the near wall region, which ultimately led to the increased cooling effect in the near wall region. It was then proposed to inject the air in such a way as to avoid the wall cooling and increase the mixing and dilution of fresh reactants with the hot combustion products. The primary and secondary air were injected in a way to enable create a strong

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swirl with flow moving upstream. The dilution air was injected in a direction opposite to the flow of reactants. The swirl direction of air and fuel were maintained in the same orientation. The hot air from primary and secondary holes was injected in a manner to ensure that wall cooling is reduced and better recirculation zones are created inside the combustor. For the ease of the fabrication of the proposed combustor and avoidance of unnecessary wall cooling, minor steps are provided for the primary and secondary air injection holes as shown in the Fig. 2d. The flow field was observed to improve significantly along with the Rdil. A maximum value of Rdil = 4.4 was obtained for the optimized combustor, D4. Based on the flow field and the Rdil distribution, the combustor design D4 is chosen for detailed computational and experimental studies in the present work.

3.1 Computational studies The development of present combustor is aimed at improving the internal recirculation of hot combustion products through a new and unique air injection scheme. The fresh air is injected in a way to create a strong swirl inside the combustor. Fig. 2d shows the dimensional details of the computational domain. The combustor geometry is optimized with the help of a series of numerical computations aimed at increasing the recirculation of the hot exhaust products, a pre-requisite to achieve the flameless combustion mode [24]. A brief summary of the details of air injection holes and the respective air distribution is given in Table 3. Table 3: Geometric details of the air injection in the computational domain: velocity (m/s) 9N1/14N1. Here 9N1 corresponds to 20.6 kW thermal input and 14N1 corresponds to 30 kW thermal input in the same combustor. Parameter

Air swirler

Primary air

Secondary air

Dilution air

Diameter (mm)

5×3

5

5

7

Air distribution (%) Velocity (m/s)

15

30

20

35

33.59/50.39

51.32/76.98

34.22/51.32

30.55/45.83

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The numerical calculations are performed for a global mixture equivalence ratio, φ = 0.3. This condition is chosen to simulate the typical conditions of a gas turbine combustor. Air distribution for primary, secondary, swirl and dilution air is chosen based on a number of numerical calculations aimed at increasing the reactant dilution ratio.

3. 2 Numerical techniques A general purpose computational fluid dynamics code, FLUENT 16.1 is used to solve the three-dimensional Navier-Stokes equation in a finite volume domain using double-precision pressure-based solver. Turbulence is modeled using seven equations Reynold stress model as it accurately predicts the high swirl flows [33]. Non-premixed combustion model is used to simulate the liquid fuel combustion. Chemical equilibrium assumption is used in the present study for modelling the combustion reaction. A probability density function (PDF) mixture containing 20 intermediate species is used for simulating kerosene-air combustion. Kerosene is simulated as a single surrogate compound C12H23. The interaction between liquid spray and a continuous gas phase is modeled using discrete phase model. The liquid spray is modeled as solid-cone injection of kerosene with spherico-symmetric droplets. Appropriate values of spray parameters, 1) Sauter mean diameter (SMD), 2) cone angle and 3) fuel flow rate are used for the solid-cone spray. These parameters are experimentally measured through detailed studies on injector using shadowgraphy [28]. Mass flow inlet boundary conditions are applied to the inlet of air injection holes. The wall boundary is defined as isothermal and pressure-outlet boundary condition is applied at the combustor exit. The combustor geometry is meshed using tetrahedral elements. The number of elements is varied from 4 million to 10 million for grid independence study and it is observed that 9 million elements are required to obtain grid-independent results. Grid convergence index (GCI) is used to perform the grid independence calculations [22, 34]. The 13

steady state solution is assumed to be converged when the residuals for all the parameters are less than 10-5.

3.3 Reactant dilution ratio The distribution of the reactant dilution ratio, Rdil for all the combustors is studied computationally as shown in Fig.3. The air swirl angle is maintained at 60° for all the combustor configurations. It is clear from Fig. 3 that Rdil for combustors D1 - D3 increases initially and reaches a maximum in the middle of the combustor and decreases further downstream. However, for design D4, Rdil reaches a maximum value of 4.4 at an axial location of 100 mm and becomes nearly constant thereafter. Avoidance of the formation of localized recirculation zones in the near wall region due to modified swirl air injection leads to improved recirculation for D4 combustor. Therefore, combustor D4 is considered for further numerical and experimental studies. Variation of Rdil for the optimized combustor with varying thermal inputs is shown in Fig. 4. The effect of air swirl angle is considered to further optimize the swirler geometry at the combustor inlet for improved recirculation.

Fig.3. Variation of Rdil along the axial direction for different designs; D1-D4: Fig.2a-2d. 14

Fig. 4. Variation of Rdil along the axial direction with different swirl angles; 9N1_45- 9N1 with 45° bottom swirl angle.

It is clear from Fig. 4 that Rdil increases along the axial direction up to a certain axial distance for all the conditions studied here. Reactant dilution ratio decreases slightly in the downstream and remains nearly constant for higher axial distances. It is interesting to note that Rdil maintains a value greater than 3 even at an axial location of 240 mm from the fuel injection point. A recirculation zone with high reactant dilution ratio is maintained along the maximum length of the combustor through the proposed altered air injection scheme. The reactant dilution ratio, Rdil for 45° swirl angle is slightly lower than for the 60° case. Therefore, a swirl angle of 60° is considered for the inlet air swirler. A maximum Rdil value of 4.40 is achieved at a location of 100 mm downstream for 9N1 with 60° swirl angle. Previous studies [24] concluded that higher Rdil values help achieve suitable conditions for flameless combustion. The present values of Rdil clearly suggest that it will help achieve the flameless combustion mode and reduce the peak temperature, fluctuations in peak temperature and pollutant emissions. An increase in the fuel injection pressure to 14 bar

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increases the thermal input as well as the total air injected into the combustor. The increase in the reactant flow rates reduces the reactant dilution ratio marginally as clear from Fig. 4.

Fig. 5. Variation of Rdil along the axial direction with the reverse orientation of the bottom air swirl. acws: anti-clockwise swirl

Fig. 5 shows the variation of Rdil for two different orientations of the bottom swirl air namely clockwise and anti-clockwise directions with respect to other air injections for a case of 14N1 and 60° swirl angle. Similar values of Rdil are obtained for axial distances up to 140 mm. However, there is a sharp reduction observed in Rdil values in the combustor downstream for anti-clockwise swirl case. The absence of the formation of recirculation zones in the downstream area is responsible for the reduced recirculation levels as discussed in later section. It is apparent from Rdil distribution that same orientation of bottom swirl air and primary, secondary and dilution air (either clockwise or anti-clockwise) helps achieve higher levels of recirculation in the combustor. Any change in the direction of air injection would lead to a significant change in the reactant dilution ratio. This aspect will become clearer in the forthcoming discussion on axial-velocity distribution in the combustor volume.

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3.4 Axial velocity distribution Axial velocity distribution in x-y plane of the combustor is obtained from the reacting flow simulations for the cases of 9N1 and 14N1. Fig. 6 shows the axial velocity distribution for the case of 9N1 with two different swirl angles to understand the effect of angle of the swirler on flow structure during the reacting flow.

(a)

(b)

(c)

Fig. 6. Axial velocity (m/s) distribution for 9N1 with (a) 60° air swirler (b) 45° air swirler (c) Schematic of mixing mechanism and recirculation zone formation

It can be concluded from Fig. 6a and 6b that extended recirculation zones are formed in the combustor. This happens due to the unique air injection scheme proposed in this work. Heated air is supplied at 800 K through primary, secondary and dilution air injection holes. The clockwise swirl created by primary and secondary air injection holes helps increase mixing and recirculation of the hot combustion products due to the high momentum of the incoming hot air. This reverse flow is further assisted by the downward injection of dilution air. It is to be noted that the orientation of the primary-secondary and swirl air must be in the same direction, either clockwise or anticlockwise. It is concluded that the recirculation zone 17

created by the bottom swirl air is extended up to combustor exit due to its interaction with different air injections as shown in a schematic diagram in Fig. 6c. The central vortex created by the bottom swirl air interacts with the circumferential vortex formed due to the primary and secondary air injections in the same direction (i.e. clockwise while looking from the top). The recirculation zone (green curve) is established as a result of the mixing between the central and circumferential vortex at the periphery of the central vortex. It is clear from Fig. 6a and 6b that increased swirl angle results in improved recirculation as higher values of negative velocities are observed for 60° swirl angle compared to 45° swirl angle. For a higher swirl angle, the interaction between the swirling air from the bottom and different air injections becomes more prominent due to increased spread of fuel-air cone. Fig. 7 shows the axial velocity distribution for 14N1 with 60° air swirler angle. Effect of the orientation of bottom swirl air with respect to primary and secondary air injection is investigated. A comparison of Fig. 7a and 7b shows that the reversal of the flow orientation of swirl air significantly affects the flow physics. It is apparent from the velocity distribution that the recirculation zone becomes significantly smaller and limited to smaller axial distances compared to the previously studied cases (Figs. 6 and 7a). The reduced intensity of swirl due to different directions of the injected air from bottom swirl injector and primary secondary air is responsible for this flow structure. In this case, bottom swirl air interacts and mixes in a cross-flow pattern with the primary and secondary air as shown in Fig. 7c. This results in a reduced strength of both central vortex formed due to the injection of bottom swirl air and the circumferential vortex formed due to primary and secondary air injections. The recirculation is dominated by the central vortex created by bottom swirl air. The recirculation zone is limited to a certain axial distance at the center of the combustor as shown in Fig. 7b.

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(a)

(b)

(c)

Fig. 7 Axial velocity (m/s) distribution for 14N1 with (a) clockwise air injection (b) anticlockwise air swirl (c) Mixing mechanism and recirculation zone formation To further understand the flow physics inside the combustor and the effect of bottom swirl air orientation, axial velocity is plotted along the radial direction at different axial locations. Fig. 8 shows the variation of axial velocity for the operating condition of 14N1 in both the orientations of the bottom swirl air, when compared with primary and secondary air injections.

(a)

(b)

Fig. 8. Axial velocity distribution along the radial direction for (a) 14N1 and (b) 14N1_acws

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Zero velocity at the combustor wall is due to the no-slip boundary condition. It can be observed from the axial velocity distribution shown in Fig. 8a that recirculation zones are present even at a downstream location of 220 mm. Positive velocity is observed at the center for all axial locations. Fig. 8b shows that for axial locations greater than 160 mm, recirculation zones are nearly absent inside the combustor. The interaction of the bottom swirl air and the swirl air from primary and secondary injection holes in opposite direction restricts the formation of recirculation zones beyond a certain axial distance. It is interesting to note from Fig. 8b that the region between r/R = 0.8 - 1.0 has more than double the negative velocities compared to Fig. 8a for higher axial distances. It can also be observed in Fig. 7b that the localized effect of dilution air injection (shown by the arrow) is higher compared to Fig. 7a due to limited size of the recirculation zones for anti-clockwise case. Therefore, the same orientation of the bottom swirl air and primary/secondary air is chosen in the present experimental study. Quantification of the increased dilution level using the novel air injection scheme proposed in the present work is conducted by numerically determining the mixing efficiency. Fig. 9 shows the variation of the mixing efficiency at different radial planes along the axial direction. Calculations are performed for both operating conditions of 9N1 and 14N1 with 60° swirl angle and a comparison with the previous work for flameless combustion is presented here. The mixing efficiency for fuel lean conditions is calculated as [28], # ($) = % &'() "*⁄(% &'(+ "*),- .

(2)

Y Y Where, ρ =density, u =axial velocity, F =actual fuel mass fraction, F ,s = stoichiometric C12H23 mass fraction and () = (+ if YF < YF,s and () = (+,0 (1 − (+ )/(1 − (+,0 ), if YF > YF,s.

20

Fig. 9. Variation of the mixing efficiency along the axial direction and their comparison

The mixing efficiency reaches nearly 100% at an axial distance greater than 140 mm for all thermal loads due to increased recirculation starting from the primary zone to the end of the dilution zone. The unique injection of high momentum hot air enhances the mixing and dilution of hot combustion products. This leads to complete mixing of hot combustion products with fresh reactants in the downstream region of the combustor. It can be noted from Fig. 9 that the mixing efficiency for the anti-clockwise case shows a slightly different trend. It happens due to a different flow feature for the respective case as seen in Fig. 7b, in which the recirculation zone is limited to a certain axial location in the downstream. However, the mixing efficiency is nearly same in the primary and secondary zones. The results are compared with the earlier work on a flameless combustor and a significant improvement is observed for the present work as clear from Fig. 9. Fig. 10 shows the variation of the combustion efficiency along the axial direction of the combustor, defined as, #4 ($) = 1 − (% &'(+ "*)/(% &'(+ "*),- .

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Fig. 10. Variation of combustion efficiency along the axial direction and their comparison

As shown in Fig. 10, the combustion efficiency is higher for 14N1 case as compared to 14N1_acws case in the combustor downstream due to better mixing and dilution for the former case as discussed in section 2.2. It is to be noted here that for all the cases, approximately 90% fuel is consumed within an axial distance of Y = 100 mm. The change is not significant further downstream and the fuel is consumed completely at Y = 140 mm. It is clear from the numerical calculations that the swirl flow in the optimized combustor geometry leads to the formation of large recirculation zones. The recirculation of hot combustion products is quantified using the parameter, reactant dilution ratio, Rdil. It is observed that Rdil should be greater than 2 for maximum portion of the combustor to achieve flameless combustion mode [4, 10]. Based on the Rdil distribution, the combustor is fabricated and experimental investigations are carried out, as discussed in the next section.

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4. Experimental studies The schematic diagram and a photograph of the experimental test rig is shown in Fig. 11. The combustor is placed vertically on a laboratory test stand. Fuel is stored in a pressurized stainless steel tank and injected from the bottom of the combustor through a pressure swirl atomizer. Air is supplied from the bottom air swirler and three rows of primary, secondary and dilution air injection. The air flow rate is controlled using three Aalborg-make electric mass flow controllers with an accuracy of ± 1.5 % of full scale. Air is heated to the temperatures of 800 K and 950 K for all the operating conditions using a 36 kW electric heater to match the typical gas turbine combustor conditions. The temperature inside the combustor is measured using OMEGA make R-type thermocouples of 0.25 mm wire diameter along with a data logging system.

(a)

23

(b) Fig. 11. (a) Schematic diagram and (b) the photograph of the experimental setup

To initially start the combustor operation, preheated air is supplied from the bottom air swirler and primary holes. Liquid kerosene is supplied at 5 bar injection pressure to ensure a rich mixture in the primary zone of the combustor. After an initial start-up period of 3 - 5 minutes, once the combustor walls are sufficiently heated, the fuel injection pressure is increased to the required value of 9 and 14 bar. Air supply is then turned on from all the injection holes. In this condition, the combustor operates at nearly stoichiometric conditions. The proper distribution of air is ensured with the help of attached jackets around the combustor. Combustor is operated in the conventional mode until the air flow rate is increased to achieve a global equivalence ratio, φ < 0.8. The combustion mode is shifted to flameless mode due to the increased momentum of the air flowing through primary, secondary and dilution zones. The concentration of pollutants in the exhaust gas is measured using a TESTO 350 flue gas analyzer. It has O2 sensor: 0 – 25 % range with 0.1 % accuracy, 24

CO sensor: 0 - 10,000 ppm range with an accuracy of ± 5 %, NO sensor: 0 - 5000 ppm and ± 5 ppm accurate and CxHy sensor: 0 - 50,000 ppm. Radiative wall heat flux is measured using a MEDTHERM Schmidt-Boelter gauge. Acoustic emissions are measured using a Lautron SL-4001 sound level meter (35 - 135 dB range and 0.1 dB resolution with 200 ms response time).

5. Experimental results 5.1 Temperature distribution Detailed temperature measurements within the combustor are conducted to understand the temperature distribution and pattern factor inside the combustor. Measurements are carried out for different thermal loads, mixture equivalence ratios and axial locations. All the temperature measurements are corrected for radiative heat losses from the thermocouple junction. The details of the temperature correction method are reported in our previous work [29]. The difference between the measured and corrected temperatures is less than 10 % of the measured value. Fig. 12 shows the measured (right side) and predicted temperature (left side) distribution at 200 mm downstream location of the combustor for 9N1 and 14N1 cases at a global equivalence ratio, φ = 0.3. It has been reported in the previous section that the combustor switches its operation to flameless combustion mode between φ = 0.8 - 0.7. Therefore, the temperature is measured at an extreme condition of φ = 0.8. It is clear from the comparison of numerical and experimental values that the numerical values slightly underpredict the measured values at all radial locations except the near wall region. The thermal field is observed to be nearly flat for both the thermal loads and temperatures are slightly higher for 14N1 case compared to 9N1 due to higher thermal input.

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Fig. 12. Measured and predicted temperatures at 200 mm downstream in the combustor (14N1_0.3 operating condition means 14 bar fuel injection pressure with a global equivalence ratio of 0.3.)

Fig. 13. Measured mean temperatures at 200 mm downstream for different global equivalence ratios Temperature measurements at φ = 0.8 show a peak at the centreline of the combustor as the combustion switches its operation from the transition mode to flameless mode at this condition. A maximum temperature of 1795 K is reported for 14N1 case. The difference 26

between the centreline temperature and wall temperature for 14N1 and 9N1 cases at φ = 0.8 is, ∆Tmax = 327 K and ∆Tmax = 366 K respectively. This difference is observed to be ∆Tmax = 145 K and ∆Tmax = 175 K respectively at φ = 0.3, i.e. the thermal gradient reduces by nearly ~56% due to the transition of the combustor operation to flameless mode. This clearly suggests a change in the mode of the combustion at respective conditions. At lower values of φ, the increased momentum of the hot air helps increase the mixing with the combustion products, thereby diluting the fresh incoming reactants. This results in a slow reaction inside the combustor, resulting in a flat thermal field within the combustor volume. As the combustor switches its operation from transition mode to flameless mode between φ = 0.8 - 0.7, the measured temperature distribution for 14N1 case at different φ is shown in Fig. 13. Increased centreline temperatures are observed for higher values of φ and the measured temperature drops in the radial direction towards the wall. To gain more insight about the thermal field and the pattern factor, measurements are carried out at three axial locations of 120 mm (secondary zone), 200 mm (dilution zone) and at the combustor exit plane. The measured temperature is maximum in the secondary zone due to increased combustion intensity in this region and decreases along the axial direction as shown in Fig.14 [28]. The temperature distribution is nearly flat at the exit of the combustor for both the thermal inputs. The maximum temperatures at the center are 1493 K and 1479 K for 14N1 (left) and 9N1(right) respectively. The difference between the centreline and the wall temperature (∆Tmax) is 38 K and 69 K for 14N1 and 9N1 respectively. This thermal gradient at the exit is reduced by ~ 88 % and ~ 81 % in contrast to conventional mode for the 14N1 and 9N1 respectively. It can be argued from the temperature distribution at the exit plane that the present combustor can be an optimal choice for its use in the gas turbines as far as the life span of the turbine blades is considered. 27

Fig. 14. Measured temperature distribution at axial locations of 120 mm, 200 mm and exit of the combustor for 14N1 (left) and 9N1 (right).

5.2 Emission characteristics CO and NOx emissions are measured for the operating conditions of 9N1 and 14N1 at different overall equivalence ratios. The values of the measured emissions are presented and these values are corrected to 15% O2, a standard practice in the field of gas turbine engines. Fig. 15 shows the measured CO emissions for the operating conditions of 9N1 and 14N1 conditions with varying global equivalence ratios, φ. Slightly higher CO emissions are observed for leaner fuel-air mixtures due to lower residence times for the conversion of CO to CO2. However, due to increased uniformity in temperature and enhanced recirculation of the hot combustion products at lower equivalence ratios, the CO emissions are still significantly lower as compared to conventional combustion mode [35].

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Fig. 15. CO emission variation for 9N1 and 14N1 at different φ

Fig. 16. NOx emission variation for 9N1 and 14N1 at different φ

Fig. 16 shows that ultra-low NOx emissions are measured in the range of the operating parameters considered for the present work. It is interesting to note here that the NOx emissions drop suddenly (~68%) between φ = 0.8 - 0.7, as the combustion mode makes a 29

transition to the flameless mode. The increased momentum of the hot air from different air injection holes facilitates the dilution and recirculation within the combustor. This helps avoid the formation of any hot spots in the combustor. It leads to a uniformly distributed temperature resulting in low NOx emissions with flameless combustion mode. In this mode, the emissions are well below than the values reported in the literature. NOx emissions of 1 ppm are measured for very lean fuel-air mixtures. Effect of high level of air preheat is also investigated to counter the effect of slightly increased levels of CO emissions as seen in Fig. 15. The air preheat temperature is increased to 950 K for all the operating conditions of thermal inputs and global equivalence ratios, φ. 94 94

Fig. 17. Variation of CO and NOx emissions for 9N1 and 14N1 for Tair = 800 K and 950 K

Fig. 17 shows the variation of CO and NOx emissions respectively for air preheat temperatures of 800 K and 950 K. It is interesting to note that with an increase in fuel injection pressure from 9 to 14 bar, both CO and NO emissions are observed to increase. The

30

mean temperature also increases with the increase in injection pressure (thermal input). The increase in the thermal input or injection pressure reduces the residence time within the combustor along with increased mean temperature. Both these factors affect the CO and NOx formation in the combustor. Due to this, the reduced residence time results in slightly increased CO formation although an increase in the mean temperature is observed. The increase in air preheat temperature results in a marginal increase in the NOx emissions. The overall combustor temperature increases with increase in air preheat temperatures. The maximum temperature increases by ~ 45 K at the center due to increase in air preheat temperature. NOx emissions are below 15 ppm level for all the operating conditions in flameless mode. A maximum level of the measured CO emissions is 52 ppm at 30 kW thermal input and φ = 0.2 and as low as 15 - 20 ppm at φ = 0.8. To further understand the applicability of the present combustor in gas turbines, emissions from typical combustors in literature are tabulated and compared with the present work. The maximum heat release intensity for the present work is 7.5 MW/m3. This is in the lower range of gas turbine engine operation and expected to increase with an increase in the operating pressure of the combustor. Very few researchers have reported the flameless combustion of liquid fuels for combustors and most of the work is focused on gaseous fuels. The emissions for the present work are well within the range at high heat intensities with liquid fuels. Table 4: Summary of reported combustors and their comparison with the present work Work

Fuel

φ

Kumar et al. [10] Luckerath et al. [20] Melo et al. [21] Sharma et al. [28] Reddy et al. [22]

LPG Natural gas

Thermal intensity (MW/m3) 5.6 14

CO (ppm) at 15% O2

1.00 0.40

Air preheat temperature (K) 300 735

2900 <10

NOx (ppm) at 15% O2 7.5 1

Methane Kerosene

0.50 0.80

300 800

25 13.5

500 16

6 28

Kerosene

0.81

700

10

32

13

31

Ye et al. [17] Present Present

Prevaporized ethanol Kerosene Kerosene

0.58

873±50

7

0.78

4.16

0.50 0.30

800 800

7.5 7.5

38 59

5 3

It is clear from the detailed emission measurements that the developed combustor emits ultralow CO and NOx emissions for the complete operational range, φ = 0.8 − 0.3. The optimized combustor operates in MILD combustion regime due to increased internal recirculation of hot combustion products. The layout of the combustor is similar to a conventional can-type gas turbine combustor, which makes it a potential choice for further applications in gas turbines.

5.3 Radiative heat flux measurements To ensure the flatness of thermal field inside the combustor, the radiative heat flux is measured using Schmidt-Boelter gauge on the inner part of the combustor walls. The gauge is placed at 200 mm downstream of the combustor such that the quartz surface of the gauge is in perfect alignment with the inner portion of the wall. The gauge is insulated from the combustor wall to avoid any direct heat conduction to the gauge. Fig. 18 shows the instantaneous measurements of the radiative flux for 14N1 operating condition at a preheat temperature of 800 K in both conventional and flameless mode. Initially, the measurements are carried out at stoichiometric conditions for the conventional mode. Lower radiative flux (~27 kW/m2) is measured due to the flame stabilization at the center. As the momentum of the hot air is increased (global equivalence ratio is reduced), the combustor switches its operation mode to the flameless mode leading to increased radiative heat flux from the combustion zone as clear from the Fig. 18.

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Fig. 18. Radiative heat flux measurement in different modes of combustion Radiative heat flux increases to a high value of ~ 55 kW/m2 (at φ = 0.3) as the combustor starts operating in flameless mode, thereby ensuring the flatness of the thermal field [31]. Results are compared with the existing literature on flameless combustion by Sharma et al. [28] (~ 60 kW/m2 at φ = 0.92) and Reddy et al. [34] (~ 43 kW/m2 at φ = 0.92). Weber et al. [31] also reported similar profiles for the radiative heat flux for flameless combustion mode with liquid fuels in an industrial furnace.

5.4 Acoustic emissions Fig. 19 shows the variation of the measured acoustic emissions from the exhaust of the combustor for all the operating conditions. The sound level meter is placed at a radial distance of 30 mm above the combustor exit and 100 mm away from the axis of the combustor.

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Fig. 19. Acoustic emission variation for 9N1 and 14N1 in different combustion modes For the non-reacting flow, the recorded sound pressure level is in the range of 75 - 76 dB for both the thermal loads and increases to 92 – 93 dB, when the fuel-air mixture is ignited. It remains in this range, while the combustor operates in the conventional mode. An increase in the sound pressure level is observed, when the air flow rate is increased (or the overall equivalence ratio is reduced from φ = 1 to 0.8). A transition zone appears with a further increase in the air flow rate and the acoustic emissions show a peak in this region. After a time period of around 60 seconds, the acoustic emissions show a steep reduction and acoustic emissions drops to lower values of 85 - 86 dB in the flameless mode. Reduced fluctuations in the heat release rate lead to lower acoustic emissions and a total drop of around 12 - 15 dB (~15%) is observed as the combustor starts operating in flameless mode. The results are compared with the previous work by Sharma et al. [28], where the flameless combustion with liquid fuels was achieved by modifying the combustor exit. For the present work, the acoustic emissions are lower than the previous studies due to lower air injection velocities (< 50 m/s) for all the combustion modes studied, as compared to previous work (~ 100 m/s) [28].

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6.0 Conclusions Computational and experimental investigations on a can-type combustor with a novel air injection scheme are presented in this paper. The air injection from different inlet holes is aimed at increasing the recirculation of the hot combustion products in the combustor. Following conclusions are derived based on the numerical and experimental investigations: i.

The novel air injection aimed at creating the strong swirl in the combustor helps achieve the desired mixing and dilution levels of the hot combustion products with the fresh reactants.

ii.

Extended recirculation zones are established along the axial direction compared to a conventional can-type combustor.

iii.

The transition to flameless combustion occurs when the momentum of the hot air increases beyond a certain point as mixing and dilution mechanism commences.

iv.

Improved pattern factor (flat thermal field) is observed at the combustor exit in the flameless combustion mode.

v.

Slightly higher CO emissions are measured compared to previous studies [28], however, the effect can be reduced by increasing the air preheat temperature to a higher level.

vi.

NOx emissions are reduced significantly in the flameless mode and the measured emissions are as low as 1 ppm for very lean fuel-air mixture conditions.

vii.

A significant reduction of around 15 dB in the acoustic emissions is observed in the flameless mode due to reduced fluctuations in the heat release rate.

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Acknowledgements: The Authors would like to acknowledge the support received for this work from Propulsion panel, Aeronautics Research and Development Board (ARDB), India through Grant-in-Aid scheme with grant no. ARDB/01/1041702/M/1.

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Highlights • • • • •

Investigated MILD combustion in a can type gas turbine combustor A novel air injection scheme is adopted to achieve MILD combustion Combustor switches to MILD combustion mode due to increased momentum of hot air Novel air injection scheme results in the establishment of large recirculation zones Combustor exhibits flat thermal field, low gaseous emissions and low noise level

Declaration of interests ☒ The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper. ☐The authors declare the following financial interests/personal relationships which may be considered as potential competing interests:

Declarations of interest: none