A heating furnace burner structure optimization method based on a flow-heat coupling model
By optimizing the burner structure of the heating furnace using a fluid-thermal coupling model, the problems of uneven combustion and high NOx emissions in traditional designs were solved, enabling precise control of the combustion process and improving equipment safety.
Patent Information
- Authority / Receiving Office
- CN · China
- Patent Type
- Applications(China)
- Current Assignee / Owner
- Filing Date
- 2026-04-17
- Publication Date
- 2026-07-10
AI Technical Summary
Traditional burner designs struggle to accurately predict the complex turbulent combustion, radiative heat transfer, and coupled heat transfer processes between combustion and furnace tubes within the furnace, leading to problems such as poor flame shape, uneven temperature distribution, localized hot spots, high NOx emissions, furnace tube coking, and concentrated thermal stress.
A combustion furnace structure optimization method based on a fluid-thermal coupling model is adopted. By establishing a three-dimensional physical model, the Navier-Stokes equations are used to simulate the gas flow in the furnace and tubes. Combined with turbulence model, EDM combustion model, P-1 radiation model and NOx generation model, fluid-thermal coupling simulation is carried out to optimize the fuel gas inlet diameter and air preheating temperature to regulate the combustion process.
It achieves precise control of the combustion process, reduces NOx emissions, improves the uniformity of temperature distribution in the furnace, extends equipment service life, and reduces the risk of local high temperature and thermal stress concentration.
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Figure CN122365762A_ABST
Abstract
Description
Technical Field
[0001] This invention relates to the field of burner structure optimization technology, and in particular to a method for optimizing the structure of a heating furnace burner based on a fluid-thermal coupling model. Background Technology
[0002] Industrial heating furnaces are core energy-consuming equipment in high-energy-consuming process industries, accounting for a very high proportion of the total energy consumption of the entire plant. Among them, the burner is a key component that determines combustion efficiency, the temperature field distribution inside the furnace, and the generation of pollutants, especially NOx.
[0003] Traditional burner designs rely heavily on empirical formulas and steady-state empirical models, making it difficult to accurately predict the complex turbulent combustion, radiative heat transfer, and coupled heat transfer processes between combustion and furnace tubes within the furnace. This results in problems such as poor flame shape, uneven temperature distribution, localized hot spots, high NOx emissions, coking of furnace tubes, and thermal stress concentration during actual operation.
[0004] Therefore, there is a need for an analysis and optimization method that can accurately quantify the comprehensive influence of burner structural parameters on the internal flow field, temperature field, and composition field of the furnace, in order to achieve precise control of the combustion process. Summary of the Invention
[0005] The purpose of this invention is to address the shortcomings of existing technologies by proposing a method for optimizing the structure of a furnace burner based on a fluid-thermal coupling model.
[0006] To achieve the above objectives, the present invention adopts the following technical solution: A method for optimizing the structure of a furnace burner based on a fluid-thermal coupling model includes the following steps: S1: Establish a three-dimensional physical model of the heating furnace burner. Using the symmetrical modeling method, the furnace tubes and U-shaped bends are regularly arranged and connected to the radiant manifold to form a radiant coil model. The Navier-Stokes equations are used to simulate the gas flow in the furnace and tubes. The basic control equations consist of continuity, momentum, energy and composition equations. S2: The low NOx burners are symmetrically installed on the side wall of the radiation chamber of the three-dimensional physical model of the heating furnace burner. The structure of the low NOx burner is simplified to a sleeve structure. In the simplified structure, the inner ring is the fuel gas inlet and the outer ring is the air inlet. S3: The furnace wall, burner flue, and furnace tube wall in contact with the outside air in the three-dimensional physical model of the heating furnace burner are all simplified as adiabatic walls. The furnace tube is coupled with the furnace chamber for heat transfer, and the tube wall is the coupling wall. The thickness of the furnace tube wall is defined by Shell Conduction. The working fluid inside the furnace tube is gaseous hydrogenated naphtha to be heated. S4: A fluid-thermal coupling model of turbulent combustion and heat transfer in the furnace is constructed using the turbulence model, EDM combustion model, P-1 radiation model and NOx generation model. S5: The model was validated based on laboratory-scale industrial heating furnace experiments. The experimental furnace consisted of a rectangular combustion chamber and a coaxial jet burner. Methane was used as fuel. Neither the fuel gas nor the air was preheated. Samples were taken at a height of z=350mm on the side wall of the furnace. The temperature and flue gas component concentration inside the furnace were measured by thermocouples and gas sampling probes. S6: The effect of different excess air coefficients on the temperature of the radiation chamber was simulated and analyzed using a fluid-thermal coupling model; S7: The effect of different air preheating temperatures on the temperature of the radiation chamber was simulated and analyzed using a fluid-thermal coupling model; S8: The influence of different heat loads on the temperature of the radiant chamber is simulated and analyzed using a fluid-thermal coupling model; S9: Adjust the fuel gas inlet diameter of the heater burner based on the excess air coefficient, air preheating temperature, and heat load control method simulated by the fluid-thermal coupling model.
[0007] Preferably, in step S4, the flow inside the heating furnace is turbulent flow. Therefore, the turbulence model uses the Standard k-ε model with standard wall functions. This model is based on a semi-empirical formula of turbulent kinetic energy and dissipation rate and is applicable to incompressible flow.
[0008] Furthermore, in S6, as the excess air coefficient increases, a temperature boundary region appears between the burners, making combustion more concentrated. At the same time, the excess air reduces the combustion rate and lengthens the flame.
[0009] Further preferred: In step S6, simulation results show that the maximum pipe wall temperature within the excess air coefficient range of 1.05-1.25 meets the safety standard.
[0010] As a preferred embodiment of the present invention: in step S6, after the excess air coefficient is increased to 1.2, the rate of decrease in NO emissions slows down.
[0011] As a further preferred embodiment of the present invention: In S7, as the preheating temperature increases, the temperature difference between air and fuel increases, resulting in uneven temperature distribution in the furnace and a gradual decrease in the combustion zone. The preheating of the air at the inlet leads to a decrease in air density, further reducing the air mass flow rate in the furnace, thereby causing the flame combustion range to shrink.
[0012] As a further aspect of the present invention: In S7, as the preheating temperature increases, the temperature difference between air and fuel increases, resulting in uneven temperature distribution in the furnace and a gradual decrease in the combustion zone. The preheating of the air at the inlet leads to a decrease in air density, further reducing the air mass flow rate in the furnace, thereby causing the flame combustion range to shrink.
[0013] Based on the aforementioned scheme: In S7, as the preheating temperature increases, the surface temperature distribution of the radiant tube wall becomes more and more uniform, the high-temperature area on the surface of the furnace tube becomes smaller and smaller, as the air preheating temperature increases, the CO2 volume fraction at the outlet of the radiant chamber continuously decreases, and as the preheating temperature increases, the NO emission continuously increases.
[0014] Based on the aforementioned scheme, the preferred embodiment is as follows: In S8, the core flame temperature in the radiant chamber increases significantly with the increase of heat load. The increase of heat load is accompanied by the increase of fuel flow rate and combustion intensity, which accelerates the reaction rate and makes the flame shape longer and wider. Increasing the heat load will lead to uneven temperature distribution on the radiant tube wall, which will increase the fuel flow rate in the radiant chamber and more fuel will react with air to generate more CO2.
[0015] Further optimization based on the aforementioned scheme: In S9, when the diameter of the fuel gas inlet is increased, the jet velocity of the fuel gas can be reduced, so that the fuel gas and air flow rates are matched, allowing sufficient time for thorough mixing. This reduces the intensity of the reaction between the fuel gas and air, causing the maximum flame temperature in the furnace to gradually decrease. As the fuel gas and air are mixed more evenly, the production of nitrogen oxides in the flue gas is suppressed due to the reduction in the maximum flame temperature. Once the fuel gas inlet diameter reaches a suitable level, further increasing the inlet diameter leads to poorer mixing of fuel gas and air, resulting in a gradual increase in NO emissions.
[0016] The beneficial effects of this invention are as follows:
[0017] 1. This method uses turbulence models, combustion models, and NOx formation models to simulate the flow field, temperature field, and component field inside the furnace. Through simulation analysis of the excess air coefficient, air preheating temperature, and heat load, the influence of different parameters on the combustion process and NOx emissions can be obtained, making the control of the burner more precise. The simulation analysis shows that an excess air coefficient in the range of 1.05-1.25 can effectively reduce NOx emissions while ensuring the safety of the tube wall temperature. Increasing the fuel gas inlet diameter to 80 mm can promote the uniform mixing of fuel and air, reduce the maximum flame temperature, and significantly inhibit NOx formation.
[0018] 2. This method, through modeling and analysis of the coupled heat transfer between the furnace tube and the furnace chamber, predicts and optimizes the temperature distribution of the radiant tube wall using the model. Under excessively high heat loads, the simulation shows that the furnace tube wall temperature will exceed the safety limit; while under appropriate loads, the furnace tube temperature distribution is more uniform and the working fluid outlet temperature is stable. Combined with the control of air preheating temperature, it can effectively avoid local high temperatures, reduce the risk of furnace tube coking and thermal stress concentration, thereby extending the service life of the equipment and improving the overall operational safety. Attached Figure Description
[0019] Figure 1 This is a schematic diagram of the process for optimizing the structure of a heating furnace burner based on a fluid-thermal coupling model proposed in this invention. Figure 2 This is a diagram of a heating furnace structure for a heating furnace burner structure optimization method based on a fluid-thermal coupling model proposed in this invention. Figure 3 This invention provides a design parameter table for hydrotreated naphtha media in a furnace burner structure optimization method based on a fluid-thermal coupling model. Figure 4 This invention provides a table of fuel gas composition for a furnace burner structure optimization method based on a fluid-thermal coupling model. Figure 5 This is a table of kinetic parameters for the Westbrook and Dryer two-step combustion mechanism of a furnace burner structure optimization method based on a fluid-thermal coupling model proposed in this invention. Figure 6 This is a comparison of experimental data for a furnace burner structure optimization method based on a fluid-thermal coupling model proposed in this invention. Figure 7 This is a diagram showing the temperature distribution of the radiant chamber under different excess air coefficients in a furnace burner structure optimization method based on a flow-thermal coupling model proposed in this invention. Figure 8 The graph shows the maximum and average temperatures of the radiant chamber under different excess air coefficients for the furnace burner structure optimization method based on the fluid-thermal coupling model proposed in this invention. Figure 9 This is a radiant tube temperature distribution diagram under different excess air coefficients for a furnace burner structure optimization method based on a flow-thermal coupling model proposed in this invention. Figure 10 The graph shows the relationship between (a) the volume fraction of O2 and CO2 in the flue gas and (b) the NO emission and excess air coefficient in a furnace burner structure optimization method based on a flow-heat coupling model proposed in this invention. Figure 11This is a diagram showing the temperature distribution of the radiant chamber under different preheating temperatures in a furnace burner structure optimization method based on a flow-thermal coupling model proposed in this invention. Figure 12 The diagram shows the highest and average temperatures of the radiant chamber under different preheating temperatures, based on a heat flow coupling model for optimizing the structure of a furnace burner. Figure 13 This is a diagram showing the temperature distribution of radiant tubes under different preheating temperatures in a furnace burner structure optimization method based on a flow-thermal coupling model proposed in this invention. Figure 14 The graphs show the relationship between (a) CO2 volume fraction and (b) NO emissions in flue gas with air preheating temperature for a furnace burner structure optimization method based on a flow-heat coupling model proposed in this invention. Figure 15 This is a temperature distribution diagram of the radiant chamber under different heat loads for a furnace burner structure optimization method based on a flow-thermal coupling model proposed in this invention. Figure 16 The diagram shows the highest and average temperatures of the radiant chamber under different heat loads for the furnace burner structure optimization method based on the flow-heat coupling model proposed in this invention. Figure 17 This is a temperature distribution diagram of a radiant tube under different heat loads, representing a furnace burner structure optimization method based on a flow-thermal coupling model proposed in this invention. Figure 18 The figure shows the volume fraction of CO2 and NO emissions in the flue gas versus the heat load for a furnace burner structure optimization method based on a flow-heat coupling model proposed in this invention. Detailed Implementation
[0020] The technical solution of the present invention will be further described in detail below with reference to specific embodiments.
[0021] Example 1: A method for optimizing the structure of a furnace burner based on a fluid-thermal coupling model, such as... Figure 1 As shown, it includes the following steps: S1: Establish a three-dimensional physical model of the heating furnace burner. Using a symmetrical modeling method, regularly arrange and connect the furnace tubes and U-shaped bends to the radiant manifold to form a radiant coil model, such as... Figure 2 As shown in (b), the Navier-Stokes equations are used to simulate gas flow in the furnace and tubes. The basic governing equations consist of continuity, momentum, energy, and composition equations;
[0022] The continuity equation is: In the formula, , , for , , Velocity components in three directions, unit: m / s; time: t, unit: s; density: ρ, unit: kg / m³. 3 .
[0023] The momentum conservation equation is:
[0024]
[0025]
[0026] In the formula, P represents the pressure on the infinitesimal element, with the unit being Pa. , , The viscous stress on the surface of the micro-element caused by molecular viscosity. The component, in Pa. , , The unit mass force on the infinitesimal element in three directions, with units of m / s². 2 If only gravity acts on it, and the z-axis points vertically upward, then = =0, = .
[0027] The energy conservation equation is: In the formula, For effective thermal conductivity, Let be the diffusion flux of the component. The three terms on the right side of the equation are the thermal conductivity term, the component diffusion term, and the viscous dissipation term, respectively. This is the source term, which includes the heat of chemical reaction and other volumetric heat sources. Wherein: .
[0028] The component conservation equation is: In the formula, This represents the turbulent mass diffusivity of gas species k. This represents the net yield of gaseous substance k produced due to a chemical reaction.
[0029] S2: Low-NOx burners are symmetrically installed on the sidewalls of the radiation chamber in the 3D physical model of the heating furnace burner. The structure of the low-NOx burner is as follows: Figure 2 As shown in (c), the low-NOx burner structure is simplified as follows: Figure 2 (d) The sleeve structure is simplified so that the inner ring is the fuel gas inlet and the outer ring is the air inlet; according to the actual low NOx burner size, the outer diameter of the simplified burner is 350mm and the inner diameter is 40mm. S3: The furnace wall material of the 3D physical model of the heating furnace burner consists of a metal outer wall, refractory bricks, and insulation cotton. Therefore, the furnace wall, burner flue, and furnace tube wall in contact with external air are simplified as insulating walls with a wall emissivity of 0.85. The fuel gas flow rate is 0.51425 m³ / s, the air flow rate is 5.961 m³ / s, the excess air coefficient is 1.15, and both fuel gas and air enter the burner at a temperature of 298 K. The furnace tubes are coupled with the furnace chamber for heat transfer. The thermal conductivity of the furnace tube material is 26 W / m·K, the density is 7850 kg / m³, the tube wall is the coupling wall, the wall emissivity is 0.85, and the thickness is 5.74 mm. Shell Conduction is used to define the furnace tube wall thickness. The working fluid inside the furnace tube is gaseous hydrotreated naphtha to be heated, with a working fluid flow rate of 33.72 kg / s. Figure 3 Detailed design physical properties of the medium at the inlet and outlet; S4: A fluid-thermal coupling model of turbulent combustion and heat transfer in the furnace is constructed using a turbulent model (preferably the Standard k-ε model), an EDM (vortex dissipation) combustion model, a P-1 radiation model, and a NOx generation (thermal and rapid) model. The flow within the heating furnace is turbulent; therefore, the Standard k-ε model with standard wall functions is used. This model is based on semi-empirical formulas for turbulent kinetic energy and dissipation rate, is applicable to incompressible flows, and is widely used in engineering. The turbulent kinetic energy and dissipation rate are calculated as follows:
[0030]
[0031] In the above equation, It is the turbulent kinetic energy caused by turbulent fluctuations. , , , and It is an empirical constant obtained from experiments with air and water, and its value is: =1.44, =1.92, =0.09, =1.0, =1.3.
[0032] The combustion of fuel gas in the furnace involves complex chemical reactions, the intensity of which is influenced by turbulence. Since the burner used in this embodiment is a non-premixed burner, the EDM model in Fluent is well-suited for studying non-premixed combustion in gas-phase combustion. Therefore, this embodiment uses the EDM model to simulate the actual combustion process in the furnace. The principle of this model is that the formation rate Ri,r of substance i in the reaction is determined by the smaller value in the equation.
[0033] In the formula, The mass fraction of the reactants, The product mass fraction Let r be the stoichiometric coefficient of reactant i in reaction r. Let be the stoichiometric coefficient of product j in reaction r, and A and B be empirical constants, usually taken as 4 and 0.5 respectively.
[0034] The fuel gas used in the heating furnace is methane gas, a mixture of various alkanes, alkenes, and hydrogen, with methane and hydrogen being the main components. The volume fraction of the mixed gas is as follows: Figure 4 As shown.
[0035] The gaseous fuel combustion reaction mechanism used in this embodiment is the Westbrook and Dryer two-step combustion mechanism, and the specific kinetic parameters are as follows: Figure 5 As shown.
[0036] Heat transfer within the furnace chamber is primarily radiative, and this embodiment employs the P-1 radiation model. A grey gas weighted average (WSGGM) model is introduced to calculate the absorption coefficient of the physical properties of the materials involved in radiative heat transfer.
[0037] Radiative heat flow The equation is: In the formula, Let G be the divergence, and G be the incident radiation. The scattering coefficient is... Let be the absorption coefficient, C be the linear anisotropic phase function coefficient, and the incident radiation G transport equation be: In the formula, Customize the radiant heat source for the user. Let be the Stefan-Boltzmann constant. Combining the two equations above, we can obtain:
[0038] Applying the above formula to the energy equation allows us to calculate the amount of heat generated by radiation.
[0039] NOx generated during combustion originates from three sources: thermal NOx, rapid NOx, and fuel NOx. Fuel NOx is generated by the reaction of nitrogen atoms in the fuel with hydrocarbons. Since the fuel composition in this study is a nitrogen-free hydrocarbon fuel, only thermal and rapid NOx are considered. For thermal and rapid NOx, only the transport equations concerning the NO mass fraction need to be solved: .
[0040] Source term in the formula The sum of thermal and rapid NO: In the formula This indicates the molecular weight of NO.
[0041] S5: The model was verified based on an experiment of a 20KW laboratory-scale industrial heating furnace. The experimental furnace consisted of a cuboid combustion chamber and a coaxial jet burner. Methane was used as fuel, with a fuel gas flow rate of 0.65655 kg / h, an air-fuel ratio of 0.9, and an air flow rate of 12.5926 kg / h. Neither the fuel gas nor the air was preheated. Samples were taken at a height of z=350 mm on the side wall of the furnace. The temperature and flue gas component concentration inside the furnace were measured using thermocouples and gas sampling probes. A comparison between the numerical simulation results and experimental data of temperature and O2, CO2, and NO concentration distributions during the combustion process in the experimental furnace is shown below. Figure 6 The simulated boundary conditions were consistent with the experimental conditions. The results show that the simulation results agree well with the experimental monitoring values, with a maximum error within 10%. The RANS model can accurately reveal the combustion heat transfer process in the heating furnace.
[0042] S6: The effect of different excess air coefficients on the temperature of the radiation chamber was simulated and analyzed using a fluid-thermal coupling model; Figure 7 The model simulation analysis shows the temperature distribution of the radiant chamber under different excess air coefficients. When the excess air coefficient is 1.05, the flame combustion area is wide and the high temperature area of the furnace is large. The flue gas temperature in the furnace is relatively high. It can be clearly observed that when the excess air coefficient is 1.05, flue gas backflow occurs at the burner nozzle. The mixed backflow gas is further burned on the side of the burner nozzle, which increases the temperature. However, long-term backflow combustion will lead to coking of the furnace tube.
[0043] As the excess air coefficient increases, temperature boundary zones can be observed between burners, indicating that combustion is more concentrated. However, at the same time, excess air reduces the combustion rate and lengthens the flame.
[0044] As the excess air coefficient increases, the high-temperature zone in the furnace gradually decreases, and the flue gas temperature in the furnace also decreases, resulting in a decrease in the average temperature of the furnace. This is because the excess unreacted air carries away the heat generated by combustion.
[0045] Figure 8 The maximum and average temperatures of the radiant chamber of the heating furnace under different excess air coefficients are given, from Figure 8 It can be clearly seen that as the excess air coefficient increases, the highest temperature inside the heating furnace radiant chamber continuously decreases, because the excess air carries away more heat.
[0046] However, when the excess air coefficient increased from 1.2 to 1.25, the highest temperature inside the furnace rose again. This was because the difference between the air velocity and the fuel gas velocity was large, and the flame center was always in a state of oxygen deficiency with excess fuel gas. The mixing of fuel gas and air was poor. When the excess air coefficient reached 1.25, the mixing of fuel gas and air was better, which promoted the combustion of fuel gas and air, thus causing the flame temperature to rise again. Meanwhile, the average furnace temperature continued to decrease due to the continuous increase in the excess air coefficient.
[0047] Figure 9 The temperature distribution of the furnace tube under different excess air coefficients is shown. It can be seen intuitively that as the excess air coefficient increases, the high-temperature area on the surface of the furnace tube decreases. This is because more cold air enters the furnace and absorbs the heat released by combustion, resulting in a decrease in flame temperature and less heating of the furnace tube.
[0048] Combustion is more gentle and complete, the temperature gradient in the furnace is reduced, and the temperature distribution on the surface of the furnace tube is more uniform. As the excess air coefficient increases, the maximum temperature of the radiant furnace tube wall gradually decreases. In engineering, the maximum safe temperature of the radiant furnace tube made of A335P9 material is 923K. Simulation results show that the maximum temperature of the tube wall in the range of 1.05-1.25 with an excess air coefficient meets the safety standard.
[0049] Figure 10 (a) represents the relationship between O2, CO2 and excess air coefficient in the flue gas emitted from the heating furnace. The CO emission from combustion is almost zero under different excess air coefficients, so the simulated values are considered to be based on the complete combustion of fuel under different excess air coefficients.
[0050] from Figure 10 As can be seen in (a), the volume fraction of CO2 decreases as the excess air coefficient increases. This is because the introduction of excess air increases the total amount of flue gas in the radiation chamber, thereby diluting the CO2 concentration in the flue gas.
[0051] The oxygen concentration at the outlet increases with the increase of the excess air coefficient. The increase of the excess air coefficient introduces excess oxygen, which is discharged through the outlet of the radiation chamber.
[0052] Figure 10 (b) gives the relationship between NO concentration in flue gas and excess air coefficient. Usually, the NO2 content in flue gas is less than one-tenth of the NO content, so the NO emission is directly regarded as the NOx emission.
[0053] pass Figure 10 (b) It can be directly observed that NO emissions decrease as the excess air coefficient increases. This is because the introduction of excess air lowers the furnace temperature and inhibits the generation of thermal NOx. After the excess air coefficient increases to 1.2, further increases in the excess air coefficient slow down the rate of decrease in NO emissions.
[0054] S7: The effect of different air preheating temperatures on the temperature of the radiation chamber was simulated and analyzed using a fluid-thermal coupling model; Figure 11 The diagram shows the temperature distribution in the radiant chamber under different air preheating temperatures. It can be seen that as the preheating temperature increases, the central flame becomes more concentrated, the flame core temperature is higher, but the flame length gradually shortens, and the high-temperature zone shrinks towards the burner. This is because the increased preheating temperature promotes the reaction rate between fuel and air, leading to an earlier ignition process for gaseous fuels. However, the increased preheating temperature also leads to a larger temperature difference between air and fuel, resulting in uneven temperature distribution within the furnace and a gradual reduction in the combustion zone. Air preheating at the inlet reduces air density, further decreasing the air mass flow rate within the furnace, thus reducing the flame combustion area.
[0055] Figure 12 This represents the maximum and average furnace temperature under different air preheating temperatures. As the preheating temperature increases, the maximum furnace temperature rises continuously. This increase in preheating temperature intensifies the reaction rate between the fuel gas and air, promoting the activation energy of the reaction. However, due to the increased air preheating temperature, the air density decreases, resulting in a reduced air mass flow rate into the furnace. This leads to incomplete combustion of the fuel gas within the furnace, causing a decrease in the overall average temperature.
[0056] Figure 13 This represents the surface temperature distribution of the radiant tube wall at different air preheating temperatures. From... Figure 13As can be seen, with the increase of preheating temperature, the surface temperature distribution of the radiant tube becomes more uniform, and the high-temperature area on the furnace tube surface becomes smaller. This is because the increase in preheating temperature makes the flame more concentrated, and the heating on both sides of the U-shaped tube is more uniform. At the same time, because preheating changes the air density, the total heat in the furnace decreases, resulting in a decrease in the surface temperature of the furnace tube, a decrease in heat transfer through the radiant tube wall, and a decrease in the outlet temperature of the working fluid inside the tube. With the increase of preheating temperature, the outlet working fluid temperature of the furnace tube decreases from 817K to 782K.
[0057] Figure 14 (a) The CO2 outlet volume fraction at different air preheating temperatures is given. Analysis Figure 14 (a) It can be seen that as the air preheating temperature increases, the CO2 volume fraction at the outlet of the radiation chamber continuously decreases. This is because the increase in air preheating temperature leads to a decrease in air density, resulting in insufficient oxygen concentration in the furnace and incomplete combustion of fuel gas, thus reducing the amount of CO2 generated.
[0058] Figure 14 (b) shows the NO emissions at different air preheating temperatures. Figure 14 (b) It can be clearly observed that NO emissions increase continuously with increasing preheating temperature. This is mainly because the increase in air preheating temperature raises the temperature inside the furnace, promoting the generation of thermal NOx. Simultaneously, it can be found that when the preheating temperature exceeds 433 K, the increase in NO emissions accelerates significantly. This is because excessively high preheating temperatures cause premature fuel combustion, leading to localized high temperatures within the radiant chamber and accelerating the formation of NO in the flue gas.
[0059] S8: The influence of different heat loads on the temperature of the radiant chamber is simulated and analyzed using a fluid-thermal coupling model; Figure 15 This indicates the temperature distribution within the radiant room under different heat loads. From... Figure 15 As can be seen, the core flame temperature within the radiant chamber increases significantly with increasing heat load. This increase in heat load, accompanied by increased fuel flow and combustion intensity, accelerates the reaction rate, resulting in a longer and wider flame. This is due to the expanded flame diffusion range caused by the increased fuel injection velocity. The highest and average temperatures within the radiant chamber increase synchronously, leading to a rise in the overall temperature within the chamber. This is clearly visible in Figure 26, where the high-temperature zone expands with increasing heat load. However, it is important to note that increased heat load places a greater burden on the furnace's operation. Prolonged high-load operation can easily cause equipment wear and tear, increasing maintenance costs.
[0060] Figure 16The maximum and average temperatures inside the radiant furnace chamber under different heat loads are shown. It can be seen that the increase in heat load has a more significant effect on the average temperature of the furnace, while the change in the maximum temperature of the furnace is smaller. Increasing the heat load increases both the fuel gas flow rate and the air flow rate entering the furnace, resulting in more intense combustion in the furnace and an overall increase in the average temperature of the furnace. The simultaneous increase in fuel gas and air flow rates has no significant effect on the maximum temperature of the flame combustion in the furnace.
[0061] Figure 17 This illustrates the surface temperature distribution of the radiant tube wall inside the furnace under different heat loads. Increasing the heat load leads to uneven temperature distribution on the radiant tube wall. This is because increasing the heat load causes more intense flue gas flow within the radiant chamber, intensifying the combustion reaction in the lower part of the chamber. High temperatures concentrate in the lower part of the chamber, resulting in the lower part of the radiant tube being significantly heated than other areas, thus exacerbating the unevenness of the surface temperature distribution on the radiant tube wall.
[0062] Especially at a heat load of 16.37 MW, the simulated highest temperature on the furnace tube wall reached 938 K, exceeding the maximum allowable temperature of 923 K required by the engineering specifications. This would severely exacerbate slagging and even damage the furnace tubes. The heating furnace operating at a heat load of 10.91 kW exhibited the most uniform temperature distribution on the radiant tube walls, with a working fluid outlet temperature reaching 789 K, which is 20 K lower than the working fluid outlet temperature of the furnace tubes operating at a rated capacity of 13.64 MW.
[0063] Figure 18 (a) shows the CO2 volume fraction under different heat loads. The CO2 volume fraction increases significantly with increasing heat load, leading to a greater fuel flow rate within the radiant chamber, resulting in more fuel reacting with air to generate more CO2. Although increasing heat load introduces more air, the absolute rate of CO2 formation exceeds the rate of increase in total flue gas volume. Figure 18 (b) shows the NO emissions under different heat loads. Increased heat load leads to a rise in temperature in the flame core region, promoting the generation of thermal NOx, thus resulting in a continuous increase in NO emissions. Excessive heat load has adverse effects on both the overall load on the furnace and flue gas emissions.
[0064] S9: Adjust the fuel gas inlet diameter of the heater burner based on the excess air coefficient, air preheating temperature, and heat load control method simulated by the fluid-thermal coupling model; By studying the impact of excess air coefficient on combustion in a reforming furnace, this embodiment uses a fluid-thermal coupling model to simulate and find that increasing the excess air coefficient effectively suppresses the generation of thermal NOx, thereby reducing nitrogen oxide emissions. After the excess air coefficient reaches 1.2, further increases in the excess air coefficient result in a significant decrease in the reduction of nitrogen oxide emissions. As the excess air coefficient increases, the excess unburned air carries away the high-temperature flue gas in the furnace, leading to a decrease in the average furnace temperature. Therefore, in this embodiment, the furnace burner needs to control the excess air coefficient below 1.2.
[0065] Simulations using a fluid-thermal coupling model revealed that as the excess air coefficient increased from 1.2 to 1.25, the maximum flame temperature inside the furnace increased slightly. Considering the velocity difference between the air inlet and the fuel gas inlet caused by the burner structure, the disturbance effect of the air inlet was enhanced as the excess air coefficient increased. The influence of the excess air coefficient and the adequacy of mixing between the fuel gas and the fuel gas are in competition.
[0066] Air preheating temperature can effectively promote the reaction rate between fuel gas and air. In this embodiment, the air preheating temperature rises from 353K to 513K, the highest temperature inside the furnace increases from 1901K to 1946K, and the total heat transfer increases from 6521kW to 7252kW. This results in a 20°C increase in the flue gas outlet temperature of the radiation chamber, a maximum radiant tube wall temperature of 920K, an average radiant tube wall temperature increase of approximately 5°C, and a decrease in the uniformity of the tube wall temperature distribution. The medium outlet temperature rises from 817K to 826K. The safety standard temperature of the A335P9 material used for the radiant tube wall is 923K. When the air preheating temperature reaches 513K, the temperature of the radiant tube wall is already close to the allowable temperature limit of the tube wall. Therefore, the air preheating temperature should be below 513K.
[0067] Meanwhile, through calculations of flue gas emissions, it was found that NO emissions were lowest at 393K, at only 143ppm. Further increasing the air preheating temperature would promote the generation of thermal NOx and increase nitrogen oxide emissions. Therefore, based on the simulated data in this embodiment, it can be concluded that 393K is the optimal preheating temperature.
[0068] In this embodiment, the rated heat load of the heating furnace is 13.64MW, and 16.37MW represents the maximum heat load of the heating furnace. An intermediate transition heat load of 15.01MW is set to improve the parameter continuity of the heat load study. At the same time, heat loads of 12.28MW and 10.91MW are taken as low-load operating conditions. Through combustion simulation calculations of the heating furnace under five different heat loads, it was found that increasing the heat load significantly improved the heat transfer in the furnace, thereby increasing the furnace temperature. When the heat load increased from 10.91MW to 16.37MW, the average temperature of the heating furnace increased by about 60°C. When the heating furnace was running under the high heat load of 16.37MW, the highest temperature of the radiant tube wall reached 938K, exceeding the maximum allowable temperature of the radiant tube wall material.
[0069] The radiant tube wall temperature distribution is most uniform in the furnace operating at a heat load of 10.91 kW. The working fluid outlet temperature is 20 K lower than that in the furnace operating at a rated temperature of 13.64 MW, with an outlet temperature of 789 K. Increasing the furnace's heat load leads to uneven heat flow and temperature distribution on the tube walls, and also increases flue gas emissions. Therefore, in this embodiment, the furnace heat load needs to be strictly controlled below the maximum heat load of 16.37 MW, and the heat load should be reduced as much as possible based on the required outlet medium temperature.
[0070] Since there is a difference in size between the air inlet and the fuel gas inlet of the furnace burner, in this embodiment, the main gas inlet diameter of the furnace burner is 40mm. Increasing the fuel gas inlet diameter can reduce the fuel gas jet velocity, match the fuel gas and air flow rates, and allow sufficient time for thorough mixing. This reduces the intensity of the reaction between the fuel gas and air, resulting in a gradual decrease in the maximum flame temperature inside the furnace. As the fuel gas and air are mixed more evenly, the reduction in the maximum flame temperature also inhibits the generation of NOx in the flue gas. Once the fuel gas inlet diameter reaches a suitable level, further increasing the inlet diameter leads to poorer mixing of fuel gas and air, resulting in a gradual increase in NO emissions.
[0071] In this embodiment, the effects of fuel gas inlet diameters of 60mm, 80mm, 100mm and 120mm were simulated using a fluid-thermal coupling model. Increasing the fuel gas inlet diameter reduced the jet velocity and the intensity of the reaction between fuel gas and air, resulting in a gradual decrease in the highest flame temperature inside the furnace. Further increasing the fuel gas inlet diameter worsened the mixing of fuel gas and air, and NO emissions began to gradually increase.
[0072] Since the total heat transfer remains unchanged, the average temperature inside the furnace is almost unaffected, stabilizing at 967K. The flue gas outlet temperature and the medium outlet temperature in the radiant tubes show no significant changes. However, the formation of nitrogen oxides (NOx) in the flue gas is suppressed due to the more uniform mixing of fuel gas and air, and the reduction in the maximum flame temperature. When the fuel gas inlet diameter is increased to 80mm, NO emissions are only 35ppm. Further increasing the fuel gas inlet diameter worsens the mixing of fuel gas and air, and NO emissions begin to gradually increase. Therefore, in this embodiment, increasing the fuel gas inlet diameter to 80mm is the optimal choice for burner structural optimization.
[0073] The above description is only a preferred embodiment of the present invention, but the scope of protection of the present invention is not limited thereto. Any equivalent substitutions or modifications made by those skilled in the art within the scope of the technology disclosed in the present invention, based on the technical solution and inventive concept of the present invention, should be covered within the scope of protection of the present invention.
Claims
1. A method for optimizing the structure of a furnace burner based on a fluid-thermal coupling model, characterized in that, Includes the following steps: S1: Establish a three-dimensional physical model of the heating furnace burner. Using the symmetrical modeling method, the furnace tubes and U-shaped bends are regularly arranged and connected to the radiant manifold to form a radiant coil model. The Navier-Stokes equations are used to simulate the gas flow in the furnace and tubes. The basic control equations consist of continuity, momentum, energy and composition equations. S2: The low NOx burners are symmetrically installed on the side wall of the radiation chamber of the three-dimensional physical model of the heating furnace burner. The structure of the low NOx burner is simplified to a sleeve structure. In the simplified structure, the inner ring is the fuel gas inlet and the outer ring is the air inlet. S3: The furnace wall, burner flue, and furnace tube wall in contact with the outside air in the three-dimensional physical model of the heating furnace burner are all simplified as adiabatic walls. The furnace tube is coupled with the furnace chamber for heat transfer, and the tube wall is the coupling wall. The thickness of the furnace tube wall is defined by Shell Conduction. The working fluid inside the furnace tube is gaseous hydrogenated naphtha to be heated. S4: A fluid-thermal coupling model of turbulent combustion and heat transfer in the furnace is constructed using the turbulence model, EDM combustion model, P-1 radiation model and NOx generation model. S5: The model was validated based on laboratory-scale industrial heating furnace experiments. The experimental furnace consisted of a rectangular combustion chamber and a coaxial jet burner. Methane was used as fuel. Neither the fuel gas nor the air was preheated. Samples were taken at a height of z=350mm on the side wall of the furnace. The temperature and flue gas component concentration inside the furnace were measured by thermocouples and gas sampling probes. S6: The effect of different excess air coefficients on the temperature of the radiation chamber was simulated and analyzed using a fluid-thermal coupling model; S7: The effect of different air preheating temperatures on the temperature of the radiation chamber was simulated and analyzed using a fluid-thermal coupling model; S8: The influence of different heat loads on the temperature of the radiant chamber is simulated and analyzed using a fluid-thermal coupling model; S9: Adjust the fuel gas inlet diameter of the heater burner based on the excess air coefficient, air preheating temperature, and heat load control method simulated by the fluid-thermal coupling model.
2. The method for optimizing the structure of a heating furnace burner based on a fluid-thermal coupling model according to claim 1, characterized in that, In S4, the flow inside the heating furnace is turbulent flow. Therefore, the turbulence model uses the Standard k-ε model with standard wall functions. This model is based on a semi-empirical formula for turbulent kinetic energy and dissipation rate and is applicable to incompressible flow.
3. The method for optimizing the structure of a heating furnace burner based on a fluid-thermal coupling model according to claim 1, characterized in that, In S6, as the excess air coefficient increases, a temperature boundary region appears between the burners, making combustion more concentrated. At the same time, the excess air reduces the combustion rate and lengthens the flame.
4. The method for optimizing the structure of a heating furnace burner based on a fluid-thermal coupling model according to claim 3, characterized in that, In S6, simulation results show that the maximum pipe wall temperature in the range of 1.05-1.25 with an excess air coefficient meets the safety standard.
5. The method for optimizing the structure of a heating furnace burner based on a fluid-thermal coupling model according to claim 4, characterized in that, In S6, after the excess air coefficient increases to 1.2, further increasing the excess air coefficient slows down the reduction in NO emissions.
6. The method for optimizing the structure of a heating furnace burner based on a fluid-thermal coupling model according to claim 1, characterized in that, In S7, as the preheating temperature increases, the temperature difference between air and fuel increases, resulting in uneven temperature distribution in the furnace. The combustion zone begins to gradually decrease. The air preheating at the inlet causes the air density to decrease, further reducing the air mass flow rate in the furnace, thereby reducing the flame combustion range.
7. The method for optimizing the structure of a heating furnace burner based on a fluid-thermal coupling model according to claim 6, characterized in that, In S7, due to the increase in air preheating temperature, the air density decreases, and the mass flow rate of air entering the heating furnace decreases, which makes it impossible for the fuel gas in the furnace to burn completely, resulting in a decrease in the overall average temperature.
8. The method for optimizing the structure of a heating furnace burner based on a fluid-thermal coupling model according to claim 7, characterized in that, In S7, as the preheating temperature increases, the surface temperature distribution of the radiant tube becomes more uniform, the high-temperature area on the surface of the furnace tube becomes smaller, the CO2 volume fraction at the outlet of the radiant chamber continuously decreases as the air preheating temperature increases, and the NO emission continuously increases as the preheating temperature increases.
9. The method for optimizing the structure of a heating furnace burner based on a fluid-thermal coupling model according to claim 1, characterized in that, In S8, the core flame temperature in the radiant chamber increases significantly with the increase of heat load. The increase of heat load is accompanied by the increase of fuel flow and combustion intensity, which accelerates the reaction rate and makes the flame shape longer and wider. The increase of heat load will lead to uneven temperature distribution on the radiant tube wall, which will increase the fuel flow in the radiant chamber and more fuel will react with air to generate more CO2.
10. The method for optimizing the structure of a heating furnace burner based on a fluid-thermal coupling model according to claim 1, characterized in that, In S9, when the diameter of the fuel gas inlet increases, the jet velocity of the fuel gas can be reduced, so that the fuel gas and air flow rates are matched, allowing sufficient time for thorough mixing. This reduces the intensity of the reaction between the fuel gas and air, causing the maximum flame temperature in the furnace to gradually decrease. As the fuel gas and air are mixed more evenly, and the maximum flame temperature is reduced, the generation of NOx in the flue gas is suppressed. Once the fuel gas inlet diameter reaches a suitable level, further increasing the inlet diameter leads to poorer mixing of fuel gas and air, resulting in a gradual increase in NO emissions.