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Study on the combustion characteristics of a premixed combustion system with exhaust gas recirculation
Energy 61 (2013) 345e353
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Study on the combustion characteristics of a premixed combustion system with exhaust gas recirculation Byeonghun Yu a, Sung-Min Kum b, Chang-Eon Lee a, Seungro Lee a, * a b
Department of Mechanical Engineering, Inha University, 100 Inha-ro, Nam-gu, Incheon 402-751, Republic of Korea School of Mechanical and Automotive Engineering, Halla University, Wonju 220-712, Republic of Korea
a r t i c l e i n f o
a b s t r a c t
Article history: Received 15 February 2013 Received in revised form 26 August 2013 Accepted 28 August 2013 Available online 24 September 2013
The boiler of a premixed combustion system with EGR (exhaust gas recirculation) is investigated to explore the potential for increasing thermal efficiency and lowering pollutant emissions. To achieve this purpose, a thermodynamic analysis is performed to predict the effect of EGR on the thermodynamic efficiency for various equivalence ratios. Experiments of a preheated air condensing boiler with EGR were conducted to measure the changes in the thermal efficiency and the characteristics of the pollutant emission. Finally, a 1-D premixed code was calculated to understand the effect of the EGR method on the NO reduction mechanism. The results of the thermodynamic analysis show that the thermodynamic efficiency is not changed because the temperature and the amount of the exhaust gas are unchanged, even though the EGR method is implemented in the system. However, when the EGR method is used with an equivalence ratio near 1.00, it is experimentally verified that the thermal efficiency increases and the NOx concentration decreases. Based on the results from numerical calculations, it is shown that the NO production rates of N þ O2 4 NO þ O and N þ OH 4 NO þ H are remarkably changed due to the decrease in the flame temperature and the NO mole fraction is decreased. Ó 2013 Elsevier Ltd. All rights reserved.
Keywords: EGR (Exhaust gas recirculation) Premixed combustion NOx emission Thermal efficiency Production rate
1. Introduction As industrial development has rapidly increased in recent years, the problems of excessive energy consumption and environmental pollution due to increasing generation of exhaust gas have become a social issue. Various types of combustion systems have already been studied and developed to solve these problems, and research in renewable energy fields such as hydrogen, solar and bio energy is being conducted. However, the results are not as satisfactory as was expected. Thus, the most realistic and economical method to solve the problems of excessive energy consumption and environmental pollution is to conduct research in combustion systems that use fossil fuels with the goal of achieving high thermal efficiencies and low pollutant emissions. Combustion with EGR (exhaust gas recirculation) is one among recent, representative low-pollutant combustion methods and has widely been applied. EGR has the advantage of being applicable to any type of combustion system because it reuses the exhaust gas produced from the combustion process. However, most research regarding EGR is limited to automotive internal combustion
* Corresponding author. Tel.: þ82 32 867 4522; fax: þ82 32 876 7838. E-mail addresses: [email protected], [email protected] (S. Lee). 0360-5442/$ e see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.energy.2013.08.057
engines, and no research has been conducted on small-scale household combustion systems with premixed flames. Rather, most research presently focuses on large-scale industrial combustion systems with non-premixed flames. For research on internal combustion engines with the EGR method, Park et al. [1] showed that EGR with optimized ignition timing for a spray-guided combustion system in a gasoline direct injection engine can reduce harmful NOx and THC emissions effectively. Yadav et al. [2] reported that dual fuel operation with hydrogen induction coupled with exhaust gas recirculation results in lowered emission levels and improved performance levels for a direct injection diesel engine, Lakshmanan and Nagarajan [3] showed reduced NOx emissions and improved part load performances for an acetylene induction diesel engine with cooled EGR. In research on non-premixed flames applied EGR method, Shinomori et al. reported that NOx emission for small boiler can be reduced approximately 80% when the amount of a recirculating gas was increased to 50% [4]. Kim et al. investigated the NO emission characteristics for the oxy-fuel combustors using FGR (flue gas recirculation) technology experimentally [5]. In this study, they showed that the reduction ratio of NO emission was more than approximately 85% when the combustor was operated at the 40% FGR ratio and they also verified that the FGR technology was quite effective for reducing the NO emission in the oxy-fuel combustor. Baltasar et al. presented an
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experimental and numerical study of the effect of FGR (flue gas recirculation) on flame characteristics and pollutant emissions for a small-scale laboratory furnace [6]. In this study, they showed that the flue gas decreased NOx emissions with FGR without significant effects on flame stability, overall combustion efficiency and CO emission. In addition, Jiang et al. reported the result of NOx reduction for an industrial reheating furnace with flue gas recirculation [7]. A typical, present-day combustion system that has been adapted to use premixed flames is described by an equivalence ratio, which is the ratio of the stoichiometric airefuel ratio to the actual air-fuel ratio. Most premixed combustion systems are operated near an equivalence ratio of 0.70 to reduce pollutant emissions such as nitrogen oxides (NOx) and to ensure complete combustion of the fuel. However, combustion systems at this operating condition have a limited maximum theoretical thermal efficiency [8]. Table 1 shows the characteristics of combustion systems by equivalence ratio. Specifically, the thermal efficiency of a combustion system is represented by the ratio of the useful output of a device to the input in terms of energy. Thus, as shown in Table 1, if the system were operating below an equivalence ratio of 1.00 with the same input heat, the thermal efficiency decreases with decreasing flame temperature and the amount of exhaust gas, i.e., the displacement volume, increases. However, if the system were operating near an equivalence ratio of 1.00, from an energy-saving point of view, the thermal efficiency of the system increases with increasing flame temperature and the amount of exhaust gas decreases. Of course, increasing the flame temperature can cause durability problem, and has a disadvantage in that it increases NOx emissions. Generally, in combustion systems, the trade-offs between the thermal efficiency and the pollutant emissions are described by the equivalence ratio. Thus, if the combustion system is operating near an equivalence ratio of 1.00, an increase in the thermal efficiency must be coupled with the introduction of low-pollutant emission combustion methods to decrease pollutant emissions. In previous works [9,10], Yu et al. suggested the possibility of using the EGR method for a premixed combustion system, especially a small-scale household boiler and the optimal burner material for the EGR boiler. Specifically, we have studied the effect of EGR on the thermal efficiency and pollutant emission characteristics for various EGR ratios and equivalence ratios, and it was shown that the EGR method is advantageous from the standpoint of reducing emission concentrations and ensuring the combustion stability of the burner. However, although we have confirmed the possibility of using the EGR method for a small-scale, premixed combustion system, these results do not accurately represent a real combustion system because the main burner and EGR burner were used separately to produce the exhaust gas. Thus, in this study, EGR method is applied to a real, household premixed combustion system to improve and supplement the problem in the previous work. In view of the above considerations and the extension of the previous work, the purpose of this study is the fundamental understanding of the effect of EGR on the premixed combustion system. To achieve this purpose, the thermodynamic analysis is first performed to predict the effect of EGR on the thermodynamic efficiency for the preheated air condensing boiler with various Table 1 Characteristics of a combustion system by equivalence ratio.
equivalence ratios. Second, the EGR method that is applied to a preheated air condensing boiler, which is a representative household premixed combustion system, is experimentally tested to measure the change in the thermal efficiency and the characteristics of the pollutant emission for the system with various equivalence ratios and EGR ratios. Here, it notes that since the thermodynamic efficiency and the thermal efficiency are the same thing, differing only because the first one is calculated on a thermodynamic basis while the second one estimated experimentally. Finally, a 1-D premix code with a detailed chemical reaction mechanism is constructed to numerically understand the fundamentals of the effect of EGR on the characteristics of pollutant emissions. 2. Research methods 2.1. Thermodynamic analysis As mentioned above, a thermodynamic analysis is performed to predict the effect of EGR on the thermodynamic efficiency for the preheated air condensing boiler with various equivalence ratios. For the thermodynamic analysis calculations, the fuel is assumed to be methane for the convenience of the analysis, and the pressure of the exhaust gas and the ambient temperature are assumed to be 1 atm and 273 K. The stoichiometric reaction of methane, the inlet and the outlet heat of the system and the thermodynamic efficiency used in the thermodynamic analysis are shown in Eqs. (1)e(4). Here, the calculation of the thermodynamic efficiency is considered by including the heat recuperated through condensation of water vapor based on the exhaust gas temperature [11,12].
NCH4 CH4 þ
2
F
ðO2 þ 3:76N2 Þ0NCO2 CO2 þ NH2 O H2 O þ NO2 O2
þ NN2 N2 ; (1) Qin ¼ HPROD Tref HREAC Tref ;
(2)
Qout ¼ HPROD Tref HPROD ðTout Þ;
(3)
hTD
HPROD Tref HPROD ðTout Þ Qout : ¼ 1 ¼ 1 Qin T H T H PROD
ref
REAC
(4)
ref
In Eqs. (1)e(4), N is the number of moles of the species (a subscript indicates the species name), Q is the inlet (with a subscript in) and outlet (with a subscript out) heat, H is the enthalpy of the product (with a subscript PROD) and reactant (with a subscript REAC), T is the product (with a subscript PROD) and ambient (with a subscript ref) temperature, and hTD is the thermodynamic efficiency by the thermodynamic analysis. For the calculation, the equivalence ratio is varied between 0.60, 0.80 and 1.00, and the temperature of exhaust gas is changed from 275 K to 420 K. 2.2. Experimental method
Item
F z 1:00
F << 1:00
Amount of exhaust gas Flame temperature Thermal efficiency Low pollutant Durability of system
Decrease Increase Advantage Disadvantage Disadvantage
Increase Decrease Disadvantage Advantage Advantage
Fig. 1 shows the schematic illustration of the condensing boiler used in this study. As Fig. 1 shows, the premixed porous-media burner is installed at the top of the boiler; the burner is a metal fiber burner with a porosity of approximately 89%. The total burner size is 187 mm 73 mm, and the size of combustion surface is
B. Yu et al. / Energy 61 (2013) 345e353
Fig. 1. Schematic illustration of the condensing boiler.
175 mm 56.6 mm. The main heat exchanger is installed under the burner. The heat exchanger has a rectangular structure and is made from stainless steel (SUS 304). The main heat exchanger consists of upper and lower heat exchangers. The upper heat exchanger is of a fin-and-tube type for sensible heat exchange, and the lower heat exchanger is of a plate type for latent heat recovery. The preheated air heat exchanger is installed at the stack where the exhaust gas is emitted. The purpose of the preheated air heat exchanger is to exchange heat between the exhaust gas and the air supplied to the burner. In addition to the burner and the heat exchanger, a venturi and fan are installed to supply the fuel, air and exhaust gas for the recirculation. The capacity of the condensing boiler used in this study is 30,000 kcal/h. Fig. 2 shows the schematic diagram of the experimental apparatus. In this study, natural gas is used as the fuel. The desired amount of fuel is controlled by solenoid valve installed in the boiler and the fuel is supplied to the venturi via a wet gas meter
347
(W-NK-10, SINAGAWA) which measures the fuel flow. An instrumentation accuracy of wet gas meter is 0.1% of the full scale. The surrounding air at the outside the boiler is used as the supplied air. The surrounding air is heated by the heat exchanger installed at the stack of the boiler. The preheated air is supplied by the fan from the direction of the axis of the venturi and it passes through the venturi installed on the fan. At that time, the preheated air supplied by the fan is formed the pressure drop inside the venturi. The exhaust gas generated the burner of the boiler is existed via the sensible, the latent and the preheated air heat exchangers. The supply position of the EGR gas is between the latent heat exchanger and the preheated air heat exchanger. The EGR ratio is controlled by both the fan and the control valve installed at the middle of the EGR supply tube. If the amount of EGR gas is increased with increase of EGR ratio, RPM of the fan will be increased in order to increase the pressure drop formed inside the venturi. Supplying the EGR gas without any additional equipment seems to be an advantage in terms of future commercialization of the product. Meanwhile, the fuel and EGR gas are supplied by the pressure drop of the venturi which is formed by the preheat air, and then the fuel, the EGR gas and the preheated air are mixed to supply to the burner. In addition, the O2 concentration of the mixture is measured to monitor the EGR ratio. Furthermore, the water of the boiler heated by the sensible and the latent heat exchangers is supplied to the boiler. The return and the supply heating waters are controlled by heat exchanger with the feed water and a digital flow meter to keep temperatures at the 60 C and 80 C. The feed water is supplied by the pump via the flow meter in order to calculate the thermal efficiency, and the water is existed via the plate heat exchanger to heat exchanger with the heating waters. The NOx concentration is measured at the outlet of the boiler based on the KS (Korean Industrial Standards) [13]. Water is removed from the exhaust gas with a water trap, and the NO, NO2 and O2 concentrations are measured by an electrochemical gas analyzer (Testo 340, TESTO). This instrument is capable of measuring NO concentrations in a range of 0e3000 ppm, NO2 concentrations in a range of 0e500 ppm and O2 concentrations in a
Fig. 2. Schematic diagram of the experimental apparatus.
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B. Yu et al. / Energy 61 (2013) 345e353
range of 0e25%. The measurement accuracies of these three concentrations are 5%, 5% and 1%. The NOx concentration is corrected to a 0% O2 level in the product stream, and the NOx concentration is the sum of the amounts of NO and NO2. The water flow rate is measured using a digital flow meter (RT375MI-LPM1, BlueeWhite), and the water temperature is measured using a K-type thermocouple connected to a data logger (HP-34970A) at the inlet and the outlet. The range of the digital flow meter is 3.00e30.00 L/minute, and the accuracy is 2%. The data logger measures temperatures for the K-type thermocouple in the range of 200 to 1200 C with an accuracy of 1.0 C. For all reported results, the tests are run five times under identical conditions to examine the repeatability of the data. The final reported value reflects the average of the repeated test runs. In this study, the effects of various EGR ratios and equivalence ratios on the thermal efficiency and NOx emission are experimentally tested. Thus, the equivalence ratios are varied from 0.78, which is the current operating condition of the boiler, to 0.85, 0.90 and 0.95. The EGR ratios are changed from 0% to 6%, 9%, 12% and 15% at each equivalence ratio. The thermal efficiency of the boiler used in this study can be written as Eq. (5):
h¼
_ w cp;w Tw;o Tw;i m : Q_
(5)
in
_ w is the mass flow rate of the feed water, cp,w is the where, m constant-pressure specific heat of the water, Tw is the inlet (with a subscript i) and outlet (with a subscript o) temperature of the feed water, and Q_ in is the input heat capacity of the fuel. For the calculation of the thermal efficiency, the inlet and outlet temperatures of the feed water and the mass flow rate of the feed water are measured when the return and the supply heating waters of the temperatures are 60 C and 80 C. Table 2 shows the thermal efficiency uncertainty analysis for the 30,000 kcal/h boiler capacity with an EGR ratio of 15% and an equivalence ratio of 0.90. The uncertainty analysis of the effectiveness is conducted according to the method used by Kline and McKlintock [14]. As shown in Table 2, the uncertainty of the thermal efficiency is 4.08%. At this condition, the uncertainty of the _ w ) is the largest contributor to the overall mass flow rate of water (m uncertainty, and the smallest parameter of uncertainty is the volume flow rate of fuel (q_ f ).
K X dT _ dT 1 d lA dT þ A rcp;k Yk Vk M dx cp dx dx cp dx k¼1
K A X A þ hk u_ k Wk q_ rad ¼ 0; cp cp k¼1
r¼
pW : RT
(9)
In these equations, the subscript k refers to the kth species, x is _ is the mass flux rate, T is the temperature, the spatial coordinate, M Yk is the mass fraction, p is pressure, u is the velocity, r is the mass density, Wk is the molecular weight, R is the universal gas constant, l is the thermal conductivity, cp is the constant-pressure specific heat, u_ k is the species molar production rate, hk is the specific enthalpy, Vk is the diffusion velocity, and A is the cross-sectional area of the stream tube encompassing the flame. For the boundary conditions, T and Yk are defined at the cold (upstream) boundary and the subscript u refers to the unburned gas, whereas all gradients vanish at the downstream boundary. The conditions are written as follows [15]:
Tðx ¼ NÞ ¼ Tu ; Yk ðx ¼ NÞ ¼ Yk;u ;
(10)
dT dYk ðx ¼ þNÞ ¼ 0; ðx ¼ þNÞ ¼ 0: dx dx
(11)
The main contribution to the radiative heat loss is assumed to be caused by the CO2, H2O, CO and CH4 species, and radiative heat loss, based on the optically thin approximation, is calculated as follows [16]:
4 ; q_ rad ¼ 4sKp T 4 TN
(12)
Kp ¼ PCO2 KCO2 þ PH2 O KH2 O þ PCO KCO þ PCH4 KCH4 :
(13)
where s is the StefaneBoltzmann constant, T and TN are the local and ambient temperatures, and Kp is the Plank mean absorption coefficient. Pk and Kk are the partial pressure and the Plank mean absorption coefficient of the kth species. The approximate Plank mean absorption coefficient is obtained as follows:
Kk ¼
2.3. Numerical method
(8)
5 X
Akj T j
ðk ¼ CO2 ; CO; H2 O; CH4 Þ:
(14)
j¼0
Calculations are performed using a flame-fixed coordinate system. The equations governing steady, isobaric, quasi-1D flame propagation are written as follows:
_ ¼ ruA; M
(6)
_ dYk ¼ d ðrAY V Þ þ Au_ W M k k k k dx dx
ðk ¼ 1; 2; 3; :::; KÞ;
Table 2 Thermal efficiency uncertainty analysis. Xi Variables
Value
dXi
dXi vh h vXi
Unit
C Tw,i Tw,o C _w m kg/min L/min q_ f Total h uncertainty: dh=h ¼ 4.08
25.0 75.3 9.5 56.7 (%)
0.10 0.35 0.40 0.09
0.18 0.62 3.75 0.14
(7)
100 ð%Þ
where Akj is the polynomial coefficient of the kth species expressed as a function of temperature [17]. For this study, the premixed code originally developed by Kee et al. [15] for a 1D, freely propagating, laminar premixed flame is used to predict the effects of EGR. The model used a hybrid Newton/ time integration solution algorithm with an adaptive grid system and combined established thermodynamic and transport properties with the chemical kinetic reaction rates from Chemkin-II [18] and the Transport package [19]. All of the calculations are conducted using a GRI-v3.0 mechanism [20] involving 53 species and 325 elementary reactions. The fuel is assumed to be methane for the convenience of the calculation. The numerical calculation is performed to understand the detail NO formation mechanism under the effect of EGR. Thus, the calculation conditions are selected such that the equivalence ratios for the calculation are 0.78, which is the current operating condition for the boiler, and 0.90; the corresponding EGR ratios at these equivalence ratios are 0% and 15%. The calculation domain is
B. Yu et al. / Energy 61 (2013) 345e353
10.0 cm, corresponding to the inner space of the boiler, which is the distance between the burner and the main heat exchanger. In addition, the inlet temperature and pressure for the mixture of fuel and air are 298 K and 1 atm. 3. Results and discussion 3.1. Effect of the equivalence ratio and EGR by thermodynamic analysis The results of this section are coming from the thermodynamic analysis of the system. First, Fig. 3 is the results of the investigation of the effect of the EGR method with various equivalence ratios on the thermodynamic efficiency of the system for a premixed combustion system. In Fig. 3, the x- and y-axes represent the exhaust gas temperature and thermodynamic efficiency of the system, respectively. In addition, the line means an absence of EGR and the symbol means a presence of EGR. Considering the absence of EGR, as the exhaust gas temperature is increased, the thermodynamic efficiency decreases for a constant equivalence ratio. This is because the outlet heat (Qout) of the system increases with increasing exhaust gas temperature, as can be expected from Eqs. (3) and (4). Especially, the slope of the thermodynamic efficiency curves at 330 K is changed by the latent heat by the condensation of H2O in EGR gas [8]. The rate of condensation of H2O is determined by the partial pressure of H2O in EGR gas, and condensation of H2O will be occurred from condensing temperature corresponding to the partial pressure of H2O. When the exhaust gas temperature is constant, the calculated thermodynamic efficiency gradually increases as the equivalence ratio approaches 1.00. This can be explained using Eq. (15), derived from Eq. (4).
mF cp T out Tref Qout h ¼ 1 ¼ 1 : Qin Qin
(15)
where, mF is the mass flow rate of the exhaust gas at equivalence ratio and cp is the constant-pressure specific heat of the exhaust gas. According to Eq. (15), when the equivalence ratio of the system is changed, the inlet heat (Qin) is not changed. Additionally, if the temperature of the exhaust gas is constant, the temperature difference (Tout Tref) will not change. Thus, the only factor affecting the thermodynamic efficiency of the system when the equivalence
Fig. 3. Thermodynamic efficiency versus exhaust gas temperature for various equivalence ratios depending on the presence or absence of the EGR method.
349
ratio of the system is changed is the mass flow rate of the exhaust gas (mF). As a result, as the equivalence ratio approaches 1.00 with a constant exhaust gas temperature, the thermodynamic efficiency increases due to the decreasing mass flow rate of the exhaust gas. Furthermore, when the EGR method is applied to the system, the thermodynamic efficiency of the system is the same if the equivalence ratio and the temperature of the exhaust gas were the same, as shown in Fig. 3. This can be easily explained with the concept of the premixed combustion system with the EGR method. Fig. 4 shows the concept of a premixed combustion system with the EGR method such as the system used in this study. As can be seen in Fig. 4, ① indicates the mixed gas supplied the system that is mixed with fuel and air, ② indicates the inner recirculation gas that is the summation of mixed gas and EGR gas, and ③ indicates the exhaust gas that is the difference between the inner recirculation gas and the EGR gas. As shown in Fig. 4, the amount of exhaust gas is always the same as the amount of the mixed gas for a premixed combustion system with EGR, regardless of the amount of EGR gas. Thus, as mentioned in Fig. 3, even if the EGR method is adopted for the system, the thermodynamic efficiency does not change because the temperature and the amount of the exhaust gas do not change regardless of EGR ratio. 3.2. Effect of equivalence and the EGR ratios on the experimental setup Figs. 5e7 show the experiment results for the EGR method for the preheated air condensing boiler at a boiler capacity of 30,000 kcal/h. Note that the dotted lines in Figs. 5e7 represent the measured NOx concentration and the thermal efficiency at the current operating condition of the boiler, which has an equivalence ratio (F ¼ 0.78). Fig. 5 shows the measured NOx concentration versus various equivalence ratios with varying EGR ratios. As can be seen in Fig. 5 as the equivalence ratio approaches 1.00 to increase the thermal efficiency of the system, the NOx concentration increases due to the increase in flame temperature. However, as the EGR ratio increases, the NOx concentration decreases due to the decrease in flame temperature. Specifically, if the EGR ratio is greater than 12% at F ¼ 0.85 and 0.90, a NOx concentration could be obtained that is similar to or lower than the current operating condition of the boiler, which is 29 ppm. Fig. 6 shows the thermal efficiency versus various equivalence ratios with varying EGR ratios. As shown in Fig. 6, as the equivalence ratio approaches 1.00, the thermal efficiency of the system gradually increases due to the increase in flame temperature and reaches 90.5%, which is approximately 1.8% higher than the 88.7% of the current operating condition of the boiler (F ¼ 0.78). In addition, the thermal efficiency of the system increases slowly with an increasing EGR ratio, e.g., 90.4% at F ¼ 0.85, 90.9% at F ¼ 0.90, and
Fig. 4. Concept of premixed combustion system with EGR method.
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Fig. 5. Measured NOx concentration versus equivalence ratio for different EGR ratios.
91.2% at F ¼ 0.95. This is because the temperature of the mixture and the heat transfer rate are increased by higher temperature and the amount of exhaust gas for the recirculation and by increasing flow rate of the mixture, respectively. Fig. 7 shows the relationship between thermal efficiency and measured NOx concentration for various equivalence ratios and EGR ratios. As shown in Fig. 7, if the EGR ratio of the system is greater than 12% at F ¼ 0.85 or 15% at F ¼ 0.90, the NOx concentration of the system will be lower and the thermal efficiency will be higher compared with the current operating conditions of the boiler (F ¼ 0.78). However, in the case of F ¼ 0.95, even if the adopted EGR ratio is up to 12%, which is a possible EGR ratio in the current experimental apparatus, it is undesirable in terms of low pollutant emission. In other words, although the thermal efficiency of the system will also be higher, NOx concentration will be higher than that of the current system. Thus, in order to get the high thermal efficiency of the system, if the boiler is operated at F ¼ 0.90 with 15% of EGR ratio, the obtained thermal efficiency will be 90.9% and NOx concentration of the system will be 25 ppm at that
Fig. 6. Thermal efficiency versus equivalence ratio for different EGR ratios.
Fig. 7. Relationship between the thermal efficiency and measured NOx concentration for various equivalence ratios and EGR ratios.
condition. With respect to low pollutant emission, if the EGR ratio is 15% at F ¼ 0.85, the obtained thermal efficiency and NOx concentration of the system will be 90.4% and 22 ppm, respectively. Thus, it confirms that even if the premixed combustion system such as the boiler is operated near F ¼ 1.00 to increase the thermal efficiency, the EGR method can be used to simultaneously achieve high thermal efficiency and low pollutant emission performances for the premixed combustion system, as it can reduce pollutant emissions such as NOx that increase due to the higher flame temperature. 3.3. Numerical analysis of the NO formation mechanism As mentioned in Sections 2, 3, for the calculation condition, the selected equivalence ratios are 0.78 and 0.90 when the EGR method is not applied. Additionally, an EGR ratio of 15% is calculated at F ¼ 0.90 because this condition yields a high thermal efficiency and low pollutant emission for the system based on the experimental results. Figs. 8 and 9 show the predicted NO mole fraction and flame temperature as a function of axial distance. As shown in Fig. 8, the NO mole fraction increases rapidly in the flame region and reaches the maximum value, after which it remains constant. This is because the flame temperature rapidly increases at the flame zone and then decreases by radiative heat loss, as can be seen in Fig. 9. In detail, the NO mole fraction is the highest at F ¼ 0.90 and an EGR ratio of 0%, where the flame temperature is also the highest. However, in the cases of F ¼ 0.78 with an EGR ratio of 0% and F ¼ 0.90 with an EGR ratio 15%, both NO mole fractions have a similar value. From this result, it can be seen that the NO formation is closely related to the maximum flame temperature in the flame region. In addition, the predicted NO mole fraction has quantitative differences with the experimental results, but has the same trends qualitatively. Fig. 10 shows the NO net production rate and key elementary reactions for the NO formation as a function of axial distance. To understand the NO formation mechanisms in detail, considering only elementary reactions that show relatively remarkable production (consumption) rates, the production rates of both the net and elemental reactions are given as a function of the axial distance. The key elementary reactions for the NO production are classified as R179, R180, R189 and R214 and are R186 and R249 for NO consumption based on the calculation result. The key elementary reactions for NO species are as follows:
B. Yu et al. / Energy 61 (2013) 345e353
351
Fig. 8. Comparison of predicted NO mole fraction at each condition.
3.3.1. Key elementary reaction for NO production
N þ O2 5NO þ O
ðR179Þ
N þ OH5NO þ H
ðR180Þ
N þ OH5NO þ H
ðR189Þ
HNO þ H5H2 þ NO
ðR214Þ
3.3.2. Key elementary reaction for NO consumption
HO2 þ NO5NO2 þ OH
ðR186Þ
CH2 þ NO5H þ HNCO
ðR249Þ
where, the number of the reaction indicates the elementary reaction number for the GRI-v3.0 mechanism.
Fig. 10. Effects of exhaust gas recirculation on the NO production rates: (a) F ¼ 0.78 with an EGR ratio of 0%, (b) F ¼ 0.90 with an EGR ratio of 0% and (c) F ¼ 0.90 with an EGR ratio of 15%.
Fig. 9. Comparison of flame temperature at each condition.
As shown above key elementary reactions and in Fig. 10, it is clear that the key elementary reactions for the NO production rate do not change when the equivalence ratio is changed or when the EGR method is applied, but the magnitude of the NO production
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B. Yu et al. / Energy 61 (2013) 345e353
rate does change. In terms of changing of the NO net production rate, the NO species is initially destructed at the front of the flame and is produced due to the increase in flame temperature. The net production rate of NO at F ¼ 0.90 with an EGR ratio of 0% had the highest flame temperature and the largest magnitude, but the net production rate of NO at F ¼ 0.78 with an EGR ratio of 0% and F ¼ 0.90 with an EGR ratio of 15% had similar maximum flame temperatures and similar magnitudes. In the elementary reactions, when the equivalence ratio is increased from F ¼ 0.78 to F ¼ 0.90 with an EGR ratio of 0%, the production rate of R179 and R180 increases. When the EGR method is applied for F ¼ 0.90 with EGR ratios of 0% and 15%, the production rates of R179 and R180 are remarkably different. Specifically, in all cases, the key elementary reaction for NO consumption is R186 at the front of the flame. As can be seen in R186, NO reacts with HO2 and forms NO2 and OH. NO2 formed by R186 reacts with H and then generates NO, as shown in R189. Continuing, NO is formed by R179 and R180 mainly at the flame region. Overall, the key reactions that contribute to the NO production are R179 and R180 when the equivalence ratio is increased, which belong to the thermal NOx mechanism (i.e., Zeldovich mechanism) formed in the high-temperature flame region. NO is actively produced through R179 and R180 by reacting with OH formed in flame region and O2 from the oxidizer. Thus, R179 and R180 at F ¼ 0.90 with an EGR ratio of 0% had the highest flame temperature and had the largest NO contributions; both reactions at F ¼ 0.78 with an EGR ratio of 0% and F ¼ 0.90 with an EGR ratio of 15% had a similar flame temperature as well as similar NO contributions. Furthermore, NO formed at the flame region reacts with CH2 separated from the fuel, which consumes NO and transforms it to H and HNCO by R249. Additionally, NO is formed by R214, which HNO formed at the front of flame region reacted with H. In summary, when the EGR method is applied to the system, the NO production rates of elementary reactions such as R179 and R180 are decreased due to the decrease in flame temperature; the NO net production rate and NO mole fraction are consequently decreased.
temperature. In particular, it is verified that if the operating condition is set to an EGR ratio greater than 12% at F ¼ 0.85 or greater than 15% at F ¼ 0.90, the NOx concentration will be similar to or lower than that of the current operating condition of the boiler (F ¼ 0.78). (3) As the equivalence ratio approaches 1.00 with an EGR ratio of 0%, the thermal efficiency of the system gradually increases due to the increase in flame temperature and reaches 90.5%, which is approximately 1.8% higher than 88.7% of the current operating condition of the boiler (F ¼ 0.78). In addition, the thermal efficiency of the system increases slowly with an increasing EGR ratio, e.g., 90.4% at F ¼ 0.85, 90.9% at F ¼ 0.90, and 91.2% at F ¼ 0.95, because the temperature of the mixture and the heat transfer rate are increased by the high temperature and the amount of exhaust gas for the recirculation and by flow rate of the mixture, respectively. (4) Based on the results of the numerical calculation, it is important that the key elementary reactions for the NO production rate do not change even if the equivalence ratio is changed and the EGR method is applied, but the magnitude of the NO production rate does change. In the elementary reactions, when the equivalence ratio is increased from F ¼ 0.78 to F ¼ 0.90 with an EGR ratio of 0%, the production rate of R179 and R180 increase. When the EGR method is applied for F ¼ 0.90 with EGR ratios of 0% and 15%, the NO production rates of R179 and R180 are remarkably different due to the decrease in flame temperature; the NO net production rate and NO mole fraction are consequently decreased. Acknowledgment This work was supported by the National Research Foundation of Korea (NRF) grant funded by the Korea government (MEST) (No. 2012R1A2A2A01013884) and INHA UNIVERSITY Research Grant.
4. Conclusions References The boiler of a premixed combustion system with exhaust gas recirculation is investigated to explore the potential for increasing thermal efficiencies and lowering pollutant emissions. To achieve this purpose, a thermodynamic analysis is performed to predict the effect of EGR on the thermodynamic efficiency for the preheated air condensing boiler with various equivalence ratios. The EGR method is applied to a preheated air condensing boiler, which is a representative household premixed combustion system, and experimental tests are performed to measure the changing of the thermal efficiency and the characteristics of pollutant emission for the system with various equivalence ratios and EGR ratios. Finally, a 1-D premixed code with detailed chemical reaction mechanisms is calculated to numerically understand the effect of EGR on NO reduction mechanism. The results are as follows: (1) According to the results of the thermodynamic analysis, when the equivalence ratio approaches 1.00 with a constant exhaust gas temperature, the thermodynamic efficiency increases due to the decreasing mass flow rate of the exhaust gas. It also confirms that the thermodynamic efficiency is not changed because the temperature and the amount of the exhaust gas do not change, even though the EGR method is implemented in the system. (2) When the equivalence ratio is increased to nearly 1.00 to raise the thermal efficiency of the premixed combustion system, the NOx concentration with the EGR method decreases due to the decrease in the maximum flame
[1] Park C, Kim S, Kim H, Moriyoshi Y. Stratified lean combustion characteristics of a spray-guided combustion system in a gasoline direct injection engine. Energy 2012;41:401e7. [2] Yadav VS, Soni SL, Sharma D. Performance and emission studies of direct injection C.K. engine in duel fuel mode (hydrogen-diesel) with EGR. Int J Hydrog Energy 2012;37:3807e17. [3] Lakshmanan T, Nagarajan G. Study on using acetylene in dual fuel mode with exhaust gas recirculation. Energy 2011;36:3547e53. [4] Shinomori K, Katou K, Shimokuri D, Ishizuka S. NOx emission characteristics and aerodynamic structure of a self-recirculation type burner for small boilers. Proc Combust Inst 2011;33:2735e42. [5] Kim HK, Kim Y, Lee SM, Ahn KY. NO reduction in 0.03e0.2 MW oxy-fuel combustor using flue gas recirculation technology. Proc Combust Inst 2007;31:3377e84. [6] Baltasar J, Carvalho MG, Coelho P, Costa M. Flue gas recirculation in a gas-fired laboratory furnace: measurements and modelling. Fuel 1997;76(10):919e29. [7] Jiang Q, Zhang C, Jiang J. An industrial reheating furnace with flue gas recirculation modeled by linear transfer functions. Combust Sci Technol 2004;176: 1437e64. [8] Kuck J. Efficiency of vapour-pump-equipped condensing boilers. Appl Therm Eng 1996;16(3):233e44. [9] Yu B, Kum SM, Lee CE, Lee S. Effects of exhaust gas recirculation on the thermal efficiency and combustion characteristics for premixed combustion system. Energy 2013;49:375e83. [10] Yu B, Kum SM, Lee CE, Lee S. Combustion characteristics and thermal efficiency for premixed porous-media types of burners. Energy 2013;53:343e50. [11] Turns SR. An introduction to combustion: concepts and applications. 2nd ed. Singapore: McGraw-Hill; 2006. [12] Moran MJ, Shapiro HN. Fundamentals of engineering thermodynamics. 6th ed. New Jersey, USA: John Wiley&Sons, Inc; 2008. [13] Korean Agency for Technology and Standards. KS B 8127-Condensing gas burning hot water boiler; 2010. Seoul, Korea. [14] Kline SJ, McKlintock FA. Describing uncertainty in single sample experiments. Mech Eng 1953;75:3e8.
B. Yu et al. / Energy 61 (2013) 345e353 [15] Kee RJ, Grcar JF, Smooke MD, Miller JA. A Fortran program for modeling steady laminar one-dimensional premixed flame. SAND85e8240; 1994. [16] Tien L. Thermal radiation properties of gases. Adv Heat Transfer 1968;5: 253e324. [17] Ju Y, Guo H, Maruta K, Liu F. On the extinction limit and flammability limit of non-adiabatic stretched methane-air premixed flames. J Fluid Mech 1997;32: 315e34. [18] Kee RJ, Rupley FM, Miller JA. Chemkin-II: a Fortran chemical kinetics package for the analysis of gas phase chemical kinetics. SAND89e8009B; 1989. [19] Kee RJ, Dixon-Lewis G, Warnatz GJ, Coltrin ME, Miller JA. A Fortran computer code package for the evaluation of gas-phase multi-component transport. SAND86e8246; 1994. [20] GRI Mech. Ver. 3.0, [cited 2012 Aug 06]. Available from: http://www.me. berkeley.edu/gri_mech/.
Nomenclature A: cross sectional area, cm2 cp: specific heat at constant pressure, kcal/kg- C or ergs/g-K H: enthalpy, kJ/kmol h: specific enthalpy, ergs/g K: Plank mean absorption coefficient _ mass flow rate, kg/min m: _ mass flux rate, g/s M: N: number of moles of the species (as CH4, CO2, H2O, O2, N2) P: partial pressure, dynes/cm2 p: pressure, dynes/cm2
Q: heat, kJ/kmol Q_ : heat capacity of fuel, kcal/min _ volume flow rate, L/min q: R: universal gas constant, ergs/mol-K T: temperature, K or C u: velocity, cm/s V: diffusion velocity, cm/s W: molecular weight, g/mol Y: mass fraction F: equivalence ratio h: efficiency, % l: thermal conductivity, ergs/cm-K-s r: mass density, g/cm3 s: StefaneBoltzmann constant u_ : species molar production rate, mol/cm3-s Subscripts f: fuel i or in: inlet k: kth species o or out: outlet PROD: product REAC: reactant ref: reference TD: thermodynamic w: water x: spatial coordinate N: ambient
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Energy journal homepage: www.elsevier.com/locate/energy
Study on the combustion characteristics of a premixed combustion system with exhaust gas recirculation Byeonghun Yu a, Sung-Min Kum b, Chang-Eon Lee a, Seungro Lee a, * a b
Department of Mechanical Engineering, Inha University, 100 Inha-ro, Nam-gu, Incheon 402-751, Republic of Korea School of Mechanical and Automotive Engineering, Halla University, Wonju 220-712, Republic of Korea
a r t i c l e i n f o
a b s t r a c t
Article history: Received 15 February 2013 Received in revised form 26 August 2013 Accepted 28 August 2013 Available online 24 September 2013
The boiler of a premixed combustion system with EGR (exhaust gas recirculation) is investigated to explore the potential for increasing thermal efficiency and lowering pollutant emissions. To achieve this purpose, a thermodynamic analysis is performed to predict the effect of EGR on the thermodynamic efficiency for various equivalence ratios. Experiments of a preheated air condensing boiler with EGR were conducted to measure the changes in the thermal efficiency and the characteristics of the pollutant emission. Finally, a 1-D premixed code was calculated to understand the effect of the EGR method on the NO reduction mechanism. The results of the thermodynamic analysis show that the thermodynamic efficiency is not changed because the temperature and the amount of the exhaust gas are unchanged, even though the EGR method is implemented in the system. However, when the EGR method is used with an equivalence ratio near 1.00, it is experimentally verified that the thermal efficiency increases and the NOx concentration decreases. Based on the results from numerical calculations, it is shown that the NO production rates of N þ O2 4 NO þ O and N þ OH 4 NO þ H are remarkably changed due to the decrease in the flame temperature and the NO mole fraction is decreased. Ó 2013 Elsevier Ltd. All rights reserved.
Keywords: EGR (Exhaust gas recirculation) Premixed combustion NOx emission Thermal efficiency Production rate
1. Introduction As industrial development has rapidly increased in recent years, the problems of excessive energy consumption and environmental pollution due to increasing generation of exhaust gas have become a social issue. Various types of combustion systems have already been studied and developed to solve these problems, and research in renewable energy fields such as hydrogen, solar and bio energy is being conducted. However, the results are not as satisfactory as was expected. Thus, the most realistic and economical method to solve the problems of excessive energy consumption and environmental pollution is to conduct research in combustion systems that use fossil fuels with the goal of achieving high thermal efficiencies and low pollutant emissions. Combustion with EGR (exhaust gas recirculation) is one among recent, representative low-pollutant combustion methods and has widely been applied. EGR has the advantage of being applicable to any type of combustion system because it reuses the exhaust gas produced from the combustion process. However, most research regarding EGR is limited to automotive internal combustion
* Corresponding author. Tel.: þ82 32 867 4522; fax: þ82 32 876 7838. E-mail addresses: [email protected], [email protected] (S. Lee). 0360-5442/$ e see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.energy.2013.08.057
engines, and no research has been conducted on small-scale household combustion systems with premixed flames. Rather, most research presently focuses on large-scale industrial combustion systems with non-premixed flames. For research on internal combustion engines with the EGR method, Park et al. [1] showed that EGR with optimized ignition timing for a spray-guided combustion system in a gasoline direct injection engine can reduce harmful NOx and THC emissions effectively. Yadav et al. [2] reported that dual fuel operation with hydrogen induction coupled with exhaust gas recirculation results in lowered emission levels and improved performance levels for a direct injection diesel engine, Lakshmanan and Nagarajan [3] showed reduced NOx emissions and improved part load performances for an acetylene induction diesel engine with cooled EGR. In research on non-premixed flames applied EGR method, Shinomori et al. reported that NOx emission for small boiler can be reduced approximately 80% when the amount of a recirculating gas was increased to 50% [4]. Kim et al. investigated the NO emission characteristics for the oxy-fuel combustors using FGR (flue gas recirculation) technology experimentally [5]. In this study, they showed that the reduction ratio of NO emission was more than approximately 85% when the combustor was operated at the 40% FGR ratio and they also verified that the FGR technology was quite effective for reducing the NO emission in the oxy-fuel combustor. Baltasar et al. presented an
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experimental and numerical study of the effect of FGR (flue gas recirculation) on flame characteristics and pollutant emissions for a small-scale laboratory furnace [6]. In this study, they showed that the flue gas decreased NOx emissions with FGR without significant effects on flame stability, overall combustion efficiency and CO emission. In addition, Jiang et al. reported the result of NOx reduction for an industrial reheating furnace with flue gas recirculation [7]. A typical, present-day combustion system that has been adapted to use premixed flames is described by an equivalence ratio, which is the ratio of the stoichiometric airefuel ratio to the actual air-fuel ratio. Most premixed combustion systems are operated near an equivalence ratio of 0.70 to reduce pollutant emissions such as nitrogen oxides (NOx) and to ensure complete combustion of the fuel. However, combustion systems at this operating condition have a limited maximum theoretical thermal efficiency [8]. Table 1 shows the characteristics of combustion systems by equivalence ratio. Specifically, the thermal efficiency of a combustion system is represented by the ratio of the useful output of a device to the input in terms of energy. Thus, as shown in Table 1, if the system were operating below an equivalence ratio of 1.00 with the same input heat, the thermal efficiency decreases with decreasing flame temperature and the amount of exhaust gas, i.e., the displacement volume, increases. However, if the system were operating near an equivalence ratio of 1.00, from an energy-saving point of view, the thermal efficiency of the system increases with increasing flame temperature and the amount of exhaust gas decreases. Of course, increasing the flame temperature can cause durability problem, and has a disadvantage in that it increases NOx emissions. Generally, in combustion systems, the trade-offs between the thermal efficiency and the pollutant emissions are described by the equivalence ratio. Thus, if the combustion system is operating near an equivalence ratio of 1.00, an increase in the thermal efficiency must be coupled with the introduction of low-pollutant emission combustion methods to decrease pollutant emissions. In previous works [9,10], Yu et al. suggested the possibility of using the EGR method for a premixed combustion system, especially a small-scale household boiler and the optimal burner material for the EGR boiler. Specifically, we have studied the effect of EGR on the thermal efficiency and pollutant emission characteristics for various EGR ratios and equivalence ratios, and it was shown that the EGR method is advantageous from the standpoint of reducing emission concentrations and ensuring the combustion stability of the burner. However, although we have confirmed the possibility of using the EGR method for a small-scale, premixed combustion system, these results do not accurately represent a real combustion system because the main burner and EGR burner were used separately to produce the exhaust gas. Thus, in this study, EGR method is applied to a real, household premixed combustion system to improve and supplement the problem in the previous work. In view of the above considerations and the extension of the previous work, the purpose of this study is the fundamental understanding of the effect of EGR on the premixed combustion system. To achieve this purpose, the thermodynamic analysis is first performed to predict the effect of EGR on the thermodynamic efficiency for the preheated air condensing boiler with various Table 1 Characteristics of a combustion system by equivalence ratio.
equivalence ratios. Second, the EGR method that is applied to a preheated air condensing boiler, which is a representative household premixed combustion system, is experimentally tested to measure the change in the thermal efficiency and the characteristics of the pollutant emission for the system with various equivalence ratios and EGR ratios. Here, it notes that since the thermodynamic efficiency and the thermal efficiency are the same thing, differing only because the first one is calculated on a thermodynamic basis while the second one estimated experimentally. Finally, a 1-D premix code with a detailed chemical reaction mechanism is constructed to numerically understand the fundamentals of the effect of EGR on the characteristics of pollutant emissions. 2. Research methods 2.1. Thermodynamic analysis As mentioned above, a thermodynamic analysis is performed to predict the effect of EGR on the thermodynamic efficiency for the preheated air condensing boiler with various equivalence ratios. For the thermodynamic analysis calculations, the fuel is assumed to be methane for the convenience of the analysis, and the pressure of the exhaust gas and the ambient temperature are assumed to be 1 atm and 273 K. The stoichiometric reaction of methane, the inlet and the outlet heat of the system and the thermodynamic efficiency used in the thermodynamic analysis are shown in Eqs. (1)e(4). Here, the calculation of the thermodynamic efficiency is considered by including the heat recuperated through condensation of water vapor based on the exhaust gas temperature [11,12].
NCH4 CH4 þ
2
F
ðO2 þ 3:76N2 Þ0NCO2 CO2 þ NH2 O H2 O þ NO2 O2
þ NN2 N2 ; (1) Qin ¼ HPROD Tref HREAC Tref ;
(2)
Qout ¼ HPROD Tref HPROD ðTout Þ;
(3)
hTD
HPROD Tref HPROD ðTout Þ Qout : ¼ 1 ¼ 1 Qin T H T H PROD
ref
REAC
(4)
ref
In Eqs. (1)e(4), N is the number of moles of the species (a subscript indicates the species name), Q is the inlet (with a subscript in) and outlet (with a subscript out) heat, H is the enthalpy of the product (with a subscript PROD) and reactant (with a subscript REAC), T is the product (with a subscript PROD) and ambient (with a subscript ref) temperature, and hTD is the thermodynamic efficiency by the thermodynamic analysis. For the calculation, the equivalence ratio is varied between 0.60, 0.80 and 1.00, and the temperature of exhaust gas is changed from 275 K to 420 K. 2.2. Experimental method
Item
F z 1:00
F << 1:00
Amount of exhaust gas Flame temperature Thermal efficiency Low pollutant Durability of system
Decrease Increase Advantage Disadvantage Disadvantage
Increase Decrease Disadvantage Advantage Advantage
Fig. 1 shows the schematic illustration of the condensing boiler used in this study. As Fig. 1 shows, the premixed porous-media burner is installed at the top of the boiler; the burner is a metal fiber burner with a porosity of approximately 89%. The total burner size is 187 mm 73 mm, and the size of combustion surface is
B. Yu et al. / Energy 61 (2013) 345e353
Fig. 1. Schematic illustration of the condensing boiler.
175 mm 56.6 mm. The main heat exchanger is installed under the burner. The heat exchanger has a rectangular structure and is made from stainless steel (SUS 304). The main heat exchanger consists of upper and lower heat exchangers. The upper heat exchanger is of a fin-and-tube type for sensible heat exchange, and the lower heat exchanger is of a plate type for latent heat recovery. The preheated air heat exchanger is installed at the stack where the exhaust gas is emitted. The purpose of the preheated air heat exchanger is to exchange heat between the exhaust gas and the air supplied to the burner. In addition to the burner and the heat exchanger, a venturi and fan are installed to supply the fuel, air and exhaust gas for the recirculation. The capacity of the condensing boiler used in this study is 30,000 kcal/h. Fig. 2 shows the schematic diagram of the experimental apparatus. In this study, natural gas is used as the fuel. The desired amount of fuel is controlled by solenoid valve installed in the boiler and the fuel is supplied to the venturi via a wet gas meter
347
(W-NK-10, SINAGAWA) which measures the fuel flow. An instrumentation accuracy of wet gas meter is 0.1% of the full scale. The surrounding air at the outside the boiler is used as the supplied air. The surrounding air is heated by the heat exchanger installed at the stack of the boiler. The preheated air is supplied by the fan from the direction of the axis of the venturi and it passes through the venturi installed on the fan. At that time, the preheated air supplied by the fan is formed the pressure drop inside the venturi. The exhaust gas generated the burner of the boiler is existed via the sensible, the latent and the preheated air heat exchangers. The supply position of the EGR gas is between the latent heat exchanger and the preheated air heat exchanger. The EGR ratio is controlled by both the fan and the control valve installed at the middle of the EGR supply tube. If the amount of EGR gas is increased with increase of EGR ratio, RPM of the fan will be increased in order to increase the pressure drop formed inside the venturi. Supplying the EGR gas without any additional equipment seems to be an advantage in terms of future commercialization of the product. Meanwhile, the fuel and EGR gas are supplied by the pressure drop of the venturi which is formed by the preheat air, and then the fuel, the EGR gas and the preheated air are mixed to supply to the burner. In addition, the O2 concentration of the mixture is measured to monitor the EGR ratio. Furthermore, the water of the boiler heated by the sensible and the latent heat exchangers is supplied to the boiler. The return and the supply heating waters are controlled by heat exchanger with the feed water and a digital flow meter to keep temperatures at the 60 C and 80 C. The feed water is supplied by the pump via the flow meter in order to calculate the thermal efficiency, and the water is existed via the plate heat exchanger to heat exchanger with the heating waters. The NOx concentration is measured at the outlet of the boiler based on the KS (Korean Industrial Standards) [13]. Water is removed from the exhaust gas with a water trap, and the NO, NO2 and O2 concentrations are measured by an electrochemical gas analyzer (Testo 340, TESTO). This instrument is capable of measuring NO concentrations in a range of 0e3000 ppm, NO2 concentrations in a range of 0e500 ppm and O2 concentrations in a
Fig. 2. Schematic diagram of the experimental apparatus.
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B. Yu et al. / Energy 61 (2013) 345e353
range of 0e25%. The measurement accuracies of these three concentrations are 5%, 5% and 1%. The NOx concentration is corrected to a 0% O2 level in the product stream, and the NOx concentration is the sum of the amounts of NO and NO2. The water flow rate is measured using a digital flow meter (RT375MI-LPM1, BlueeWhite), and the water temperature is measured using a K-type thermocouple connected to a data logger (HP-34970A) at the inlet and the outlet. The range of the digital flow meter is 3.00e30.00 L/minute, and the accuracy is 2%. The data logger measures temperatures for the K-type thermocouple in the range of 200 to 1200 C with an accuracy of 1.0 C. For all reported results, the tests are run five times under identical conditions to examine the repeatability of the data. The final reported value reflects the average of the repeated test runs. In this study, the effects of various EGR ratios and equivalence ratios on the thermal efficiency and NOx emission are experimentally tested. Thus, the equivalence ratios are varied from 0.78, which is the current operating condition of the boiler, to 0.85, 0.90 and 0.95. The EGR ratios are changed from 0% to 6%, 9%, 12% and 15% at each equivalence ratio. The thermal efficiency of the boiler used in this study can be written as Eq. (5):
h¼
_ w cp;w Tw;o Tw;i m : Q_
(5)
in
_ w is the mass flow rate of the feed water, cp,w is the where, m constant-pressure specific heat of the water, Tw is the inlet (with a subscript i) and outlet (with a subscript o) temperature of the feed water, and Q_ in is the input heat capacity of the fuel. For the calculation of the thermal efficiency, the inlet and outlet temperatures of the feed water and the mass flow rate of the feed water are measured when the return and the supply heating waters of the temperatures are 60 C and 80 C. Table 2 shows the thermal efficiency uncertainty analysis for the 30,000 kcal/h boiler capacity with an EGR ratio of 15% and an equivalence ratio of 0.90. The uncertainty analysis of the effectiveness is conducted according to the method used by Kline and McKlintock [14]. As shown in Table 2, the uncertainty of the thermal efficiency is 4.08%. At this condition, the uncertainty of the _ w ) is the largest contributor to the overall mass flow rate of water (m uncertainty, and the smallest parameter of uncertainty is the volume flow rate of fuel (q_ f ).
K X dT _ dT 1 d lA dT þ A rcp;k Yk Vk M dx cp dx dx cp dx k¼1
K A X A þ hk u_ k Wk q_ rad ¼ 0; cp cp k¼1
r¼
pW : RT
(9)
In these equations, the subscript k refers to the kth species, x is _ is the mass flux rate, T is the temperature, the spatial coordinate, M Yk is the mass fraction, p is pressure, u is the velocity, r is the mass density, Wk is the molecular weight, R is the universal gas constant, l is the thermal conductivity, cp is the constant-pressure specific heat, u_ k is the species molar production rate, hk is the specific enthalpy, Vk is the diffusion velocity, and A is the cross-sectional area of the stream tube encompassing the flame. For the boundary conditions, T and Yk are defined at the cold (upstream) boundary and the subscript u refers to the unburned gas, whereas all gradients vanish at the downstream boundary. The conditions are written as follows [15]:
Tðx ¼ NÞ ¼ Tu ; Yk ðx ¼ NÞ ¼ Yk;u ;
(10)
dT dYk ðx ¼ þNÞ ¼ 0; ðx ¼ þNÞ ¼ 0: dx dx
(11)
The main contribution to the radiative heat loss is assumed to be caused by the CO2, H2O, CO and CH4 species, and radiative heat loss, based on the optically thin approximation, is calculated as follows [16]:
4 ; q_ rad ¼ 4sKp T 4 TN
(12)
Kp ¼ PCO2 KCO2 þ PH2 O KH2 O þ PCO KCO þ PCH4 KCH4 :
(13)
where s is the StefaneBoltzmann constant, T and TN are the local and ambient temperatures, and Kp is the Plank mean absorption coefficient. Pk and Kk are the partial pressure and the Plank mean absorption coefficient of the kth species. The approximate Plank mean absorption coefficient is obtained as follows:
Kk ¼
2.3. Numerical method
(8)
5 X
Akj T j
ðk ¼ CO2 ; CO; H2 O; CH4 Þ:
(14)
j¼0
Calculations are performed using a flame-fixed coordinate system. The equations governing steady, isobaric, quasi-1D flame propagation are written as follows:
_ ¼ ruA; M
(6)
_ dYk ¼ d ðrAY V Þ þ Au_ W M k k k k dx dx
ðk ¼ 1; 2; 3; :::; KÞ;
Table 2 Thermal efficiency uncertainty analysis. Xi Variables
Value
dXi
dXi vh h vXi
Unit
C Tw,i Tw,o C _w m kg/min L/min q_ f Total h uncertainty: dh=h ¼ 4.08
25.0 75.3 9.5 56.7 (%)
0.10 0.35 0.40 0.09
0.18 0.62 3.75 0.14
(7)
100 ð%Þ
where Akj is the polynomial coefficient of the kth species expressed as a function of temperature [17]. For this study, the premixed code originally developed by Kee et al. [15] for a 1D, freely propagating, laminar premixed flame is used to predict the effects of EGR. The model used a hybrid Newton/ time integration solution algorithm with an adaptive grid system and combined established thermodynamic and transport properties with the chemical kinetic reaction rates from Chemkin-II [18] and the Transport package [19]. All of the calculations are conducted using a GRI-v3.0 mechanism [20] involving 53 species and 325 elementary reactions. The fuel is assumed to be methane for the convenience of the calculation. The numerical calculation is performed to understand the detail NO formation mechanism under the effect of EGR. Thus, the calculation conditions are selected such that the equivalence ratios for the calculation are 0.78, which is the current operating condition for the boiler, and 0.90; the corresponding EGR ratios at these equivalence ratios are 0% and 15%. The calculation domain is
B. Yu et al. / Energy 61 (2013) 345e353
10.0 cm, corresponding to the inner space of the boiler, which is the distance between the burner and the main heat exchanger. In addition, the inlet temperature and pressure for the mixture of fuel and air are 298 K and 1 atm. 3. Results and discussion 3.1. Effect of the equivalence ratio and EGR by thermodynamic analysis The results of this section are coming from the thermodynamic analysis of the system. First, Fig. 3 is the results of the investigation of the effect of the EGR method with various equivalence ratios on the thermodynamic efficiency of the system for a premixed combustion system. In Fig. 3, the x- and y-axes represent the exhaust gas temperature and thermodynamic efficiency of the system, respectively. In addition, the line means an absence of EGR and the symbol means a presence of EGR. Considering the absence of EGR, as the exhaust gas temperature is increased, the thermodynamic efficiency decreases for a constant equivalence ratio. This is because the outlet heat (Qout) of the system increases with increasing exhaust gas temperature, as can be expected from Eqs. (3) and (4). Especially, the slope of the thermodynamic efficiency curves at 330 K is changed by the latent heat by the condensation of H2O in EGR gas [8]. The rate of condensation of H2O is determined by the partial pressure of H2O in EGR gas, and condensation of H2O will be occurred from condensing temperature corresponding to the partial pressure of H2O. When the exhaust gas temperature is constant, the calculated thermodynamic efficiency gradually increases as the equivalence ratio approaches 1.00. This can be explained using Eq. (15), derived from Eq. (4).
mF cp T out Tref Qout h ¼ 1 ¼ 1 : Qin Qin
(15)
where, mF is the mass flow rate of the exhaust gas at equivalence ratio and cp is the constant-pressure specific heat of the exhaust gas. According to Eq. (15), when the equivalence ratio of the system is changed, the inlet heat (Qin) is not changed. Additionally, if the temperature of the exhaust gas is constant, the temperature difference (Tout Tref) will not change. Thus, the only factor affecting the thermodynamic efficiency of the system when the equivalence
Fig. 3. Thermodynamic efficiency versus exhaust gas temperature for various equivalence ratios depending on the presence or absence of the EGR method.
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ratio of the system is changed is the mass flow rate of the exhaust gas (mF). As a result, as the equivalence ratio approaches 1.00 with a constant exhaust gas temperature, the thermodynamic efficiency increases due to the decreasing mass flow rate of the exhaust gas. Furthermore, when the EGR method is applied to the system, the thermodynamic efficiency of the system is the same if the equivalence ratio and the temperature of the exhaust gas were the same, as shown in Fig. 3. This can be easily explained with the concept of the premixed combustion system with the EGR method. Fig. 4 shows the concept of a premixed combustion system with the EGR method such as the system used in this study. As can be seen in Fig. 4, ① indicates the mixed gas supplied the system that is mixed with fuel and air, ② indicates the inner recirculation gas that is the summation of mixed gas and EGR gas, and ③ indicates the exhaust gas that is the difference between the inner recirculation gas and the EGR gas. As shown in Fig. 4, the amount of exhaust gas is always the same as the amount of the mixed gas for a premixed combustion system with EGR, regardless of the amount of EGR gas. Thus, as mentioned in Fig. 3, even if the EGR method is adopted for the system, the thermodynamic efficiency does not change because the temperature and the amount of the exhaust gas do not change regardless of EGR ratio. 3.2. Effect of equivalence and the EGR ratios on the experimental setup Figs. 5e7 show the experiment results for the EGR method for the preheated air condensing boiler at a boiler capacity of 30,000 kcal/h. Note that the dotted lines in Figs. 5e7 represent the measured NOx concentration and the thermal efficiency at the current operating condition of the boiler, which has an equivalence ratio (F ¼ 0.78). Fig. 5 shows the measured NOx concentration versus various equivalence ratios with varying EGR ratios. As can be seen in Fig. 5 as the equivalence ratio approaches 1.00 to increase the thermal efficiency of the system, the NOx concentration increases due to the increase in flame temperature. However, as the EGR ratio increases, the NOx concentration decreases due to the decrease in flame temperature. Specifically, if the EGR ratio is greater than 12% at F ¼ 0.85 and 0.90, a NOx concentration could be obtained that is similar to or lower than the current operating condition of the boiler, which is 29 ppm. Fig. 6 shows the thermal efficiency versus various equivalence ratios with varying EGR ratios. As shown in Fig. 6, as the equivalence ratio approaches 1.00, the thermal efficiency of the system gradually increases due to the increase in flame temperature and reaches 90.5%, which is approximately 1.8% higher than the 88.7% of the current operating condition of the boiler (F ¼ 0.78). In addition, the thermal efficiency of the system increases slowly with an increasing EGR ratio, e.g., 90.4% at F ¼ 0.85, 90.9% at F ¼ 0.90, and
Fig. 4. Concept of premixed combustion system with EGR method.
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Fig. 5. Measured NOx concentration versus equivalence ratio for different EGR ratios.
91.2% at F ¼ 0.95. This is because the temperature of the mixture and the heat transfer rate are increased by higher temperature and the amount of exhaust gas for the recirculation and by increasing flow rate of the mixture, respectively. Fig. 7 shows the relationship between thermal efficiency and measured NOx concentration for various equivalence ratios and EGR ratios. As shown in Fig. 7, if the EGR ratio of the system is greater than 12% at F ¼ 0.85 or 15% at F ¼ 0.90, the NOx concentration of the system will be lower and the thermal efficiency will be higher compared with the current operating conditions of the boiler (F ¼ 0.78). However, in the case of F ¼ 0.95, even if the adopted EGR ratio is up to 12%, which is a possible EGR ratio in the current experimental apparatus, it is undesirable in terms of low pollutant emission. In other words, although the thermal efficiency of the system will also be higher, NOx concentration will be higher than that of the current system. Thus, in order to get the high thermal efficiency of the system, if the boiler is operated at F ¼ 0.90 with 15% of EGR ratio, the obtained thermal efficiency will be 90.9% and NOx concentration of the system will be 25 ppm at that
Fig. 6. Thermal efficiency versus equivalence ratio for different EGR ratios.
Fig. 7. Relationship between the thermal efficiency and measured NOx concentration for various equivalence ratios and EGR ratios.
condition. With respect to low pollutant emission, if the EGR ratio is 15% at F ¼ 0.85, the obtained thermal efficiency and NOx concentration of the system will be 90.4% and 22 ppm, respectively. Thus, it confirms that even if the premixed combustion system such as the boiler is operated near F ¼ 1.00 to increase the thermal efficiency, the EGR method can be used to simultaneously achieve high thermal efficiency and low pollutant emission performances for the premixed combustion system, as it can reduce pollutant emissions such as NOx that increase due to the higher flame temperature. 3.3. Numerical analysis of the NO formation mechanism As mentioned in Sections 2, 3, for the calculation condition, the selected equivalence ratios are 0.78 and 0.90 when the EGR method is not applied. Additionally, an EGR ratio of 15% is calculated at F ¼ 0.90 because this condition yields a high thermal efficiency and low pollutant emission for the system based on the experimental results. Figs. 8 and 9 show the predicted NO mole fraction and flame temperature as a function of axial distance. As shown in Fig. 8, the NO mole fraction increases rapidly in the flame region and reaches the maximum value, after which it remains constant. This is because the flame temperature rapidly increases at the flame zone and then decreases by radiative heat loss, as can be seen in Fig. 9. In detail, the NO mole fraction is the highest at F ¼ 0.90 and an EGR ratio of 0%, where the flame temperature is also the highest. However, in the cases of F ¼ 0.78 with an EGR ratio of 0% and F ¼ 0.90 with an EGR ratio 15%, both NO mole fractions have a similar value. From this result, it can be seen that the NO formation is closely related to the maximum flame temperature in the flame region. In addition, the predicted NO mole fraction has quantitative differences with the experimental results, but has the same trends qualitatively. Fig. 10 shows the NO net production rate and key elementary reactions for the NO formation as a function of axial distance. To understand the NO formation mechanisms in detail, considering only elementary reactions that show relatively remarkable production (consumption) rates, the production rates of both the net and elemental reactions are given as a function of the axial distance. The key elementary reactions for the NO production are classified as R179, R180, R189 and R214 and are R186 and R249 for NO consumption based on the calculation result. The key elementary reactions for NO species are as follows:
B. Yu et al. / Energy 61 (2013) 345e353
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Fig. 8. Comparison of predicted NO mole fraction at each condition.
3.3.1. Key elementary reaction for NO production
N þ O2 5NO þ O
ðR179Þ
N þ OH5NO þ H
ðR180Þ
N þ OH5NO þ H
ðR189Þ
HNO þ H5H2 þ NO
ðR214Þ
3.3.2. Key elementary reaction for NO consumption
HO2 þ NO5NO2 þ OH
ðR186Þ
CH2 þ NO5H þ HNCO
ðR249Þ
where, the number of the reaction indicates the elementary reaction number for the GRI-v3.0 mechanism.
Fig. 10. Effects of exhaust gas recirculation on the NO production rates: (a) F ¼ 0.78 with an EGR ratio of 0%, (b) F ¼ 0.90 with an EGR ratio of 0% and (c) F ¼ 0.90 with an EGR ratio of 15%.
Fig. 9. Comparison of flame temperature at each condition.
As shown above key elementary reactions and in Fig. 10, it is clear that the key elementary reactions for the NO production rate do not change when the equivalence ratio is changed or when the EGR method is applied, but the magnitude of the NO production
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rate does change. In terms of changing of the NO net production rate, the NO species is initially destructed at the front of the flame and is produced due to the increase in flame temperature. The net production rate of NO at F ¼ 0.90 with an EGR ratio of 0% had the highest flame temperature and the largest magnitude, but the net production rate of NO at F ¼ 0.78 with an EGR ratio of 0% and F ¼ 0.90 with an EGR ratio of 15% had similar maximum flame temperatures and similar magnitudes. In the elementary reactions, when the equivalence ratio is increased from F ¼ 0.78 to F ¼ 0.90 with an EGR ratio of 0%, the production rate of R179 and R180 increases. When the EGR method is applied for F ¼ 0.90 with EGR ratios of 0% and 15%, the production rates of R179 and R180 are remarkably different. Specifically, in all cases, the key elementary reaction for NO consumption is R186 at the front of the flame. As can be seen in R186, NO reacts with HO2 and forms NO2 and OH. NO2 formed by R186 reacts with H and then generates NO, as shown in R189. Continuing, NO is formed by R179 and R180 mainly at the flame region. Overall, the key reactions that contribute to the NO production are R179 and R180 when the equivalence ratio is increased, which belong to the thermal NOx mechanism (i.e., Zeldovich mechanism) formed in the high-temperature flame region. NO is actively produced through R179 and R180 by reacting with OH formed in flame region and O2 from the oxidizer. Thus, R179 and R180 at F ¼ 0.90 with an EGR ratio of 0% had the highest flame temperature and had the largest NO contributions; both reactions at F ¼ 0.78 with an EGR ratio of 0% and F ¼ 0.90 with an EGR ratio of 15% had a similar flame temperature as well as similar NO contributions. Furthermore, NO formed at the flame region reacts with CH2 separated from the fuel, which consumes NO and transforms it to H and HNCO by R249. Additionally, NO is formed by R214, which HNO formed at the front of flame region reacted with H. In summary, when the EGR method is applied to the system, the NO production rates of elementary reactions such as R179 and R180 are decreased due to the decrease in flame temperature; the NO net production rate and NO mole fraction are consequently decreased.
temperature. In particular, it is verified that if the operating condition is set to an EGR ratio greater than 12% at F ¼ 0.85 or greater than 15% at F ¼ 0.90, the NOx concentration will be similar to or lower than that of the current operating condition of the boiler (F ¼ 0.78). (3) As the equivalence ratio approaches 1.00 with an EGR ratio of 0%, the thermal efficiency of the system gradually increases due to the increase in flame temperature and reaches 90.5%, which is approximately 1.8% higher than 88.7% of the current operating condition of the boiler (F ¼ 0.78). In addition, the thermal efficiency of the system increases slowly with an increasing EGR ratio, e.g., 90.4% at F ¼ 0.85, 90.9% at F ¼ 0.90, and 91.2% at F ¼ 0.95, because the temperature of the mixture and the heat transfer rate are increased by the high temperature and the amount of exhaust gas for the recirculation and by flow rate of the mixture, respectively. (4) Based on the results of the numerical calculation, it is important that the key elementary reactions for the NO production rate do not change even if the equivalence ratio is changed and the EGR method is applied, but the magnitude of the NO production rate does change. In the elementary reactions, when the equivalence ratio is increased from F ¼ 0.78 to F ¼ 0.90 with an EGR ratio of 0%, the production rate of R179 and R180 increase. When the EGR method is applied for F ¼ 0.90 with EGR ratios of 0% and 15%, the NO production rates of R179 and R180 are remarkably different due to the decrease in flame temperature; the NO net production rate and NO mole fraction are consequently decreased. Acknowledgment This work was supported by the National Research Foundation of Korea (NRF) grant funded by the Korea government (MEST) (No. 2012R1A2A2A01013884) and INHA UNIVERSITY Research Grant.
4. Conclusions References The boiler of a premixed combustion system with exhaust gas recirculation is investigated to explore the potential for increasing thermal efficiencies and lowering pollutant emissions. To achieve this purpose, a thermodynamic analysis is performed to predict the effect of EGR on the thermodynamic efficiency for the preheated air condensing boiler with various equivalence ratios. The EGR method is applied to a preheated air condensing boiler, which is a representative household premixed combustion system, and experimental tests are performed to measure the changing of the thermal efficiency and the characteristics of pollutant emission for the system with various equivalence ratios and EGR ratios. Finally, a 1-D premixed code with detailed chemical reaction mechanisms is calculated to numerically understand the effect of EGR on NO reduction mechanism. The results are as follows: (1) According to the results of the thermodynamic analysis, when the equivalence ratio approaches 1.00 with a constant exhaust gas temperature, the thermodynamic efficiency increases due to the decreasing mass flow rate of the exhaust gas. It also confirms that the thermodynamic efficiency is not changed because the temperature and the amount of the exhaust gas do not change, even though the EGR method is implemented in the system. (2) When the equivalence ratio is increased to nearly 1.00 to raise the thermal efficiency of the premixed combustion system, the NOx concentration with the EGR method decreases due to the decrease in the maximum flame
[1] Park C, Kim S, Kim H, Moriyoshi Y. Stratified lean combustion characteristics of a spray-guided combustion system in a gasoline direct injection engine. Energy 2012;41:401e7. [2] Yadav VS, Soni SL, Sharma D. Performance and emission studies of direct injection C.K. engine in duel fuel mode (hydrogen-diesel) with EGR. Int J Hydrog Energy 2012;37:3807e17. [3] Lakshmanan T, Nagarajan G. Study on using acetylene in dual fuel mode with exhaust gas recirculation. Energy 2011;36:3547e53. [4] Shinomori K, Katou K, Shimokuri D, Ishizuka S. NOx emission characteristics and aerodynamic structure of a self-recirculation type burner for small boilers. Proc Combust Inst 2011;33:2735e42. [5] Kim HK, Kim Y, Lee SM, Ahn KY. NO reduction in 0.03e0.2 MW oxy-fuel combustor using flue gas recirculation technology. Proc Combust Inst 2007;31:3377e84. [6] Baltasar J, Carvalho MG, Coelho P, Costa M. Flue gas recirculation in a gas-fired laboratory furnace: measurements and modelling. Fuel 1997;76(10):919e29. [7] Jiang Q, Zhang C, Jiang J. An industrial reheating furnace with flue gas recirculation modeled by linear transfer functions. Combust Sci Technol 2004;176: 1437e64. [8] Kuck J. Efficiency of vapour-pump-equipped condensing boilers. Appl Therm Eng 1996;16(3):233e44. [9] Yu B, Kum SM, Lee CE, Lee S. Effects of exhaust gas recirculation on the thermal efficiency and combustion characteristics for premixed combustion system. Energy 2013;49:375e83. [10] Yu B, Kum SM, Lee CE, Lee S. Combustion characteristics and thermal efficiency for premixed porous-media types of burners. Energy 2013;53:343e50. [11] Turns SR. An introduction to combustion: concepts and applications. 2nd ed. Singapore: McGraw-Hill; 2006. [12] Moran MJ, Shapiro HN. Fundamentals of engineering thermodynamics. 6th ed. New Jersey, USA: John Wiley&Sons, Inc; 2008. [13] Korean Agency for Technology and Standards. KS B 8127-Condensing gas burning hot water boiler; 2010. Seoul, Korea. [14] Kline SJ, McKlintock FA. Describing uncertainty in single sample experiments. Mech Eng 1953;75:3e8.
B. Yu et al. / Energy 61 (2013) 345e353 [15] Kee RJ, Grcar JF, Smooke MD, Miller JA. A Fortran program for modeling steady laminar one-dimensional premixed flame. SAND85e8240; 1994. [16] Tien L. Thermal radiation properties of gases. Adv Heat Transfer 1968;5: 253e324. [17] Ju Y, Guo H, Maruta K, Liu F. On the extinction limit and flammability limit of non-adiabatic stretched methane-air premixed flames. J Fluid Mech 1997;32: 315e34. [18] Kee RJ, Rupley FM, Miller JA. Chemkin-II: a Fortran chemical kinetics package for the analysis of gas phase chemical kinetics. SAND89e8009B; 1989. [19] Kee RJ, Dixon-Lewis G, Warnatz GJ, Coltrin ME, Miller JA. A Fortran computer code package for the evaluation of gas-phase multi-component transport. SAND86e8246; 1994. [20] GRI Mech. Ver. 3.0, [cited 2012 Aug 06]. Available from: http://www.me. berkeley.edu/gri_mech/.
Nomenclature A: cross sectional area, cm2 cp: specific heat at constant pressure, kcal/kg- C or ergs/g-K H: enthalpy, kJ/kmol h: specific enthalpy, ergs/g K: Plank mean absorption coefficient _ mass flow rate, kg/min m: _ mass flux rate, g/s M: N: number of moles of the species (as CH4, CO2, H2O, O2, N2) P: partial pressure, dynes/cm2 p: pressure, dynes/cm2
Q: heat, kJ/kmol Q_ : heat capacity of fuel, kcal/min _ volume flow rate, L/min q: R: universal gas constant, ergs/mol-K T: temperature, K or C u: velocity, cm/s V: diffusion velocity, cm/s W: molecular weight, g/mol Y: mass fraction F: equivalence ratio h: efficiency, % l: thermal conductivity, ergs/cm-K-s r: mass density, g/cm3 s: StefaneBoltzmann constant u_ : species molar production rate, mol/cm3-s Subscripts f: fuel i or in: inlet k: kth species o or out: outlet PROD: product REAC: reactant ref: reference TD: thermodynamic w: water x: spatial coordinate N: ambient
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