Research Paper
Journal of Chemical Engineering of Japan, Vol. 49, No. 5, pp. 435–444, 2016
Enhancement on Thermal-Hydraulic Performance of Serpentine Air Preheater Ting Zhang1,2, Chun J. Liu1,2, Kai Guo1,2 and Zhe Q. Huang1,2 School of Chemical Engineering and Technology, Tianjin University, 92 Weijin Road, Nankai District, Tianjin, 30072, P. R. China 2 State Key Laboratory of Chemical Engineering, Tianjin University, 92 Weijin Road, Nankai District, Tianjin, 30072, P. R. China 1
Keywords: Serpentine Air Preheater, Thermal-Hydraulic Performance, Computational Fluid Dynamics (CFD), Structural Optimization, Geometric Correlation The present study numerically and experimentally investigates the turbulent flow and heat transfer enhancement mechanisms of a serpentine air preheater, and further, the influence of geometrical parameters on thermal enhancement and flow resistance as well. The numerical simulation results coincided well with the experimental data. The results indicated that the structure with smaller path height ratios (HR=H/W) had promising enhancement capabilities, but larger pressure drop at the same time. Here, the factor HR is the ratio of height to width of the effective heat transfer flow path. The larger effective path length (Le) had a positive effect on the thermal-hydraulic performance at lower Re number (Re2,200). With the nonlinear multiple regression methods, correlation for the averaged Nusselt number (Nu) was fitted by geometrical parameters Le, HR and Re number.
Introduction The concept of energy saving and environment protection attracts more and more attention with the increase in energy consumption and the deterioration of the living environment. Waste heat recovery is an important aspect for energy saving owing to the wide utilization of heat exchangers in various industries, such as the systems of air conditioning, petrochemical industry, boiler, metallurgy and chemical industry. Energy could be saved and the release of harmful gases could be reduced if gas is preheated by the waste heat of flue. Figure 1 shows the operation process of an air preheater in boiler combustion. Air spreads across one blower, enters the air preheater, and then goes into the boiler to combust with fuel after absorbing a quantity of exhaust heat. As the heat removal is done, flue gas with high temperature emitted by the boiler goes to the air preheater, and then releases into the atmosphere after exchanging a quantity of heat with low temperature air. By using the air preheater, boiler efficiency could be enhanced by about 1% for every 22°C increase in temperature by heating combustion air (Zeng et al., 2012). Since the thermal resistance of the gas-side is the main reason of poor thermal performance, various enhancement techniques have been developed. The techniques can be divided into active and passive methods (Webb and Bergles, 1983; Bergles, 1988; Balaras, 1990; Bergles, 2001). In general, the mechanism of the augmentaReceived on April 30, 2015; accepted on July 8, 2015 DOI: 10.1252/jcej.15we104 Correspondence concerning this article should be addressed to Z. Q. Huang (E-mail address:
[email protected]). Vol. 49 No.©5 2016 Copyright 2016The Society of Chemical Engineers, Japan
tion techniques is to disrupt the boundary layer growth, increase the turbulent intensity and generate secondary flow (Wang et al., 2002; Depaiwa et al., 2010). Compared with the active method, the passive enhancement strategy consumes power for air flow transportation without any additional forces and has rapidly developed, such as for surface treatment (Jones et al., 2009; Takata et al., 2009; Kananeh et al., 2010; Shahi, 2010; Al-Janabi et al., 2011), surface roughness (Bahrami and Yovanovich, 2005; Burgess and Ligrani, 2005; Dierich and Nikrityuk, 2013), surface extension (Tian and He, 2009; T’Joen et al., 2011), element spoiler (Liao and Xin, 2000; Krishna et al., 2009; Wongcharee and Eiamsa-ard, 2011) and vortex generator (Aris et al., 2011; Zhou and Feng, 2014; Khoshvaght-Aliabadi et al., 2015). The insertion of baffles increases the heat exchange area in a given volume by periodically altering the flow direction. As an effective element spoiler, it is used in a variety of applications to enhance the flow behavior and heat transfer. Chu et al. (2014) proposed four modified inlet manifolds to solve the large fluid flow maldistribution of multi-channel plate heat exchangers. Their results indicated that the uniform flow distribution among each channel could be improved due to the diversion of installed baffles. Wang et al.
Fig. 1 Boiler system with air preheater 435
(2014) studied the application of baffles in a novel flat plate photobioreactor. The large surface area/volume ratio enhanced mixing performance and light utilization efficiency. Baffle blocks installed on stilling basins helped to compensate for a slight deficiency of tail water for low flow rate. As for a high flow rate, they helped to deflect the flow away from the river bed (Abdelhaleem, 2013). Wang et al. (2013) infixed the internal baffles into the foam column to promote protein adsorption by increasing the residence time of the bubbles. Nemati Taher et al. (2012) investigated the flow behavior of the shell side in the shell and tube heat exchangers and revealed that the fluid was forced to flow across the tubes. By using baffles, the convection was enhanced and the residence time was extended as well. The enhancement technology of serpentine channel plays a similar role with baffles. Compared with straight ones (Ary et al., 2012; Khairul et al., 2013; Aliabadi and Alizadeh, 2015), the heat transfer rate was improved by generating secondary convective transport, such as rotational flow and swirl flow. Choi and Anand (1995) and Choi et al. (1996) provided the periodically fully developed module shown in Figure 2(a). Due to the function of the serpentine module, the flow made a quite complex route in the channel including separation, impingement, deflection, reattachment and recirculation. Several studies (Amano, 1985; Amano et al., 1987) interpreted the mechanisms of heat transfer by numerical and experimental methods. Chen et al. (2014) discussed the effect of design variables based on the objection of total thermal resistance and pressure drop. However, the reported studies were mostly based on two-dimensional channel and periodical module. In the current work, a novel kind of double channel (Figure 2(b)) assembled baffles and serpentine channel is introduced. The flow behavior and thermal performance in the novel serpentine air preheater are experimentally and numerically investigated. Geometrical parameters of serpentine air preheater are optimized. With the nonlinear multiple regression methods, correlation for the Nusselt number (Nu) is set up in terms of HR, Le and Re number.
1. Experimental Setup and Procedure Figure 3 shows a schematic diagram of experimental setup. The experimental system mainly consists of compressor, electrical heater, air preheater and measuring devices. Vortex shedding flow meter (VSF) is used to measure the volume flow rates of working fluid. The flow rate is controlled by valves. If necessary, mass flow rates can be calculated by the density of working fluid at the average temperature of inlet and outlet entrance. The total pressure drop between inlet and outlet entrance at both air-side and gas-side channels are measured with probes. The values are acquired using digital differential manometers. The T-type thermocouples are used to measure the temperature of inlet and outlet entrance. In the process of experiment, exhaust gas from combustion is replaced by high temperature air. On one side, air 436
Fig. 2 Serpentine channel model, (a) single channel (b) double channel
Fig. 3 Structure diagram of experimental set-up
flows into the air-side channel from the compressor and goes out after absorbing the heat of the gas-side channel. On the other side, air provided by another compressor is preheated by an electrical heater and the heated air is regarded as the “exhaust gas”. The “exhaust gas” flows through the gas-side channel, and then it is released into the atmosphere. During the process, energy transfers from the “exhaust gas” to the air. In the experiment, the inlet temperature of air is about 23°C; the volume flow rate of air varies from 30 to 45 m3/h; the inlet temperature of flue changes from 160 to 320°C and the volume flow rate of flue changes from 20 to 60 m3/h. In order to ensure a steady state, the experiment lasted for approximately five hours and the experimental data were recorded once an hour. All of the outside surface of the air preheater was covered by over 100 mm thick thermal insulation to minimize the losses of heat. Thus, the heat losses to surroundings could be ignored and the experimental data could be obtained on the basis of overall heat balance. The uncertainty of T-type thermocouples is around ±0.20°C, while the accuracy of pressure drop, measured by digital differential manometers, is ±0.05% full scale (0–1.6 KPa). The maximum error of VSF is less than 1%. All of the experimental data was reduced using dimensionless parameters Nu and f. The uncertainties in the experimental data were determined according to the procedure proposed by Kline and McClintock (1953). The estimated uncertainties for Nu and f are 5.95% and 4.43%, respectively.
2. Model Description 2.1 Physical model and geometry parameters The inlet velocity of the experiment is presumed to be evenly distributed, such that the physical model is simplified regardless of the entrance structure. Two variables including path height ratio HR (HR=H/W) and effective path length Le are introduced. The ratio of HR denotes the ratio of height H to width W and the detailed schematic diagram Journal of Chemical Engineering of Japan
Momentum equation:
∂ ( ρuiu j ) ∂xi =−
∂u j 2 ∂P ∂ ∂ui ∂uk − δ μ + + μ ∂x j ∂xi eff ∂x j ∂xi 3 ij eff ∂xk
(2)
The turbulent influences are taken into account by introducing the effective viscosity μeff, which is the sum of turbulent viscosity μt and molecular viscosity μ. μeff = μ + μt
Here, the turbulent viscosity is solved by the RNG k–ε method. The scale elimination in RNG theory results in a differential equation for turbulent viscosity μt.
Fig. 4 Schematic of the test section Table 1 Geometrical size of serpentine air preheater Item
Unit
Value
Total length of effective heat transfer/Lt Length of effective heat transfer flow path/Le Width of effective heat transfer flow path/W Height of effective heat transfer flow path/H Thickness of heat transfer plate/δ Material of heat transfer plate Area of effective heat transfer/A
mm mm mm mm mm
6,950 340 340 20 0.3 Stainless steel 2.2
m2
is shown in Figure 4. The 3D model is generated to simulate an accurate geometric representation. The main geometrical sizes of physical model are listed in Table 1. 2.2 Turbulence model In the reported studies of Choi and Anand (1995), Choi et al. (1996) and Wang and Falconer (1998), the single serpentine structure was simulated and analyzed using the standard k–ε model. The results indicated that the simulated turbulent flow field was consistent with the tested value. As an improvement of standard k–ε model, the RNG k–ε model provides an option to account for the effects of swirl or rotation by modifying the turbulent viscosity appropriately. This additional feature makes the RNG k–ε model more accurate and reliable for a wide class of complex flows, such as rotating uniform shear flow, free flow, boundary flow and separation (Choudhury, 1993). Hence, the RNG k–ε turbulence model with swirl dominated flow is selected in the simulation. 2.3 Governing equations The working fluid is assumed to be incompressible fluid, and the properties of the fluid are supposed to be only dependent on temperature. The viscous dissipation and radiation are ignored Chen et al. (2014). The flow in the channel is turbulent and the process is in a steady state. The governing equations including continuity, momentum and energy equations could be summarized as follows. Continuity equation: ∂ ( ρui ) = 0 ∂xi Vol. 49 No. 5 2016
(1)
ρ2 k d εμ
vˆ = 1.72 dvˆ 3 ˆ − 1 v C + v
(3)
k μeff 0 = vˆ ⋅ μ; μeff = μeff 0 f αs , Ω, ε Here, k and ε are the turbulence kinetic energy and its dissipation rate. They are obtained from the following transport equations. ∂ ∂ ∂k ( ρkui ) = + Gk − ρε αk μeff ∂xi ∂x j ∂x j
(4)
∂ ∂ ∂ε ε ε2 + C1ε ρGk − C2∗ε ρ ( ρεui ) = αε μeff (5) ∂xi ∂x j ∂x j k k Here, Gk represents the generation of turbulence kinetic energy due to the mean velocity gradients, and η Cμ ρη3 1 − 2 η 0 ε C2∗ε ≡ C2 ε + 3 k 1 + βη η ≡ Sk / ε, η0 = 4.38, β = 0.012.
Here, S is the strain rate. Energy equation:
∂ ∂ ∂T [u ( ρE +p)] = k ∂xi i ∂xi eff ∂xi
(6)
2.4 Boundary conditions To be consistent with the experiment, heat transferred by the coupled heat exchange plate and the other walls are considered to be adiabatic. The boundary conditions are described as follows. Gas-side inlet (velocity inlet):
uh = uhin , vh = wh = 0, Th = Thin ; Air-side inlet (velocity inlet):
uc = ucin , vc = wc = 0, Tc = Tcin ;
437
Gas-side outlet (outflow): ∂uh ∂v ∂w h ∂T = h = = h = 0; ∂x ∂x ∂x ∂x
Air-side outlet (outflow): ∂uc ∂v ∂w c ∂T = c = = c = 0; ∂x ∂x ∂x ∂x
2.5 Numerical method The CFD solver Fluent 6.3 ANSYS, Inc. is utilized in our calculation. The governing equations are discretized by the finite volume method. The SIMPLE algorithm is used to implement the coupling between pressure and velocity. The convergence criteria are that the residuals of related velocity equations are below 10−5, while the residual of related energy equation is below 10−8.
3. Grid Generation and Independency Validation A preprocessor Gambit 2.4.6 (ANSYS, Inc.) is used to mesh the computational domain for the solver. The computational domain is discretized by hexahedral grids. For the near-wall regions, fine grids are used to encrypt the calculation points, while the coarse grids are used in other regions. In order to obtain an accurate and economic grid system, grid independence is conducted. Mesh densities with 760,000, 980,000, 1,340,000, 1,713,000, 2,250,000 and 3,400,000 cells are chosen. Figure 5 shows the variations of heat transfer coefficient with the mesh density. It is obvious that the heat transfer coefficient slightly increases as the grid number surpasses 1,713,000. Therefore, the mesh density with 1,713,000 is adopted.
4. Results and Discussion 4.1 Parameter definitions 4.1.1 Pressure drop According to Yang and Tao (2006), the formula calculating the friction factor f can be represented as follows. f=
Δpde 2 2 ρumax Lt
(7)
Here, ∆p is the pressure drop between the inlet and outlet entrance; de is the hydraulic diameter of the entrance; ρ is the density of air; and umax is the maximum velocity of air. Lt is the total length of effective heat transfer along the streamwise direction and can be defined as Lt =nLe+(n−1)H, where n is the number of bending board. 4.1.2 Heat transfer The physical properties of the fluids are determined by the average temperature of inlet and outlet entrance. Tb, i =
(Ti ,in +Ti , out ) (i = a; g ) 2
(8)
Here, the subscripts a and g represent the air-side and the gas-side, respectively. The heat transfer rate can be expressed as follows.
φ i = qm, icipΔti (i = a; g )
(9)
Here, qm is the mass flow rate; cp is the specific heat at constant pressure; and ∆ti is the temperature difference between the inlet and outlet. Note that the heat balance between the two sides in experiments should satisfy the following criterion.
φa − φ g ≤ 5% φm Here, ϕm is the average heat transfer rate and is defined by Eq. (10).
φm =
1 ( φ − φ ) 2 a g
(10)
Then, the heat transfer coefficient h is calculated by the following equation.
h=
φm A(Tw − Tb )
(11)
Here, A is the surface area of heat transfer clapboard; Tw is the local temperature of the heat transfer clapboard; and Tb is the average temperature of the air in the channel. The Nu and Re are calculated based on the hydraulic diameter of the entrance as Eqs. (12) and (13). Nu =
hde λ
(12)
Re =
ρumaxde μ
(13)
Here, λ is thermal conductivity; and μ is dynamic viscosity.
Fig. 5 Grid independency test 438
4.2 Validation of numerical model In order elucidate the availability of the CFD model, a comparison between the simulation results and experimental data are carried out. Figure 6 presents the fitted curves of Nu and f against Re, respectively. As can be seen, the heat transfer coefficient increases with the increase of Re number, and f shows the adverse trend. The maximum deviations for Nu and f are found to be less than 8.12%. It indicates that the simulation agrees well with the experimental data, and thus, Journal of Chemical Engineering of Japan
the CFD model could be regarded as an available tool for predicting the heat exchanger performance. 4.3 Effect of enhancement by serpentine structure Under the same operating conditions and heat transfer area, the heat transfer performance of the serpentine air preheater is compared with the multi-channel plate air preheater. As shown in Figure 7, the Nu of serpentine type is almost twice that of multi-channel type for high Re numbers. Thus, the serpentine channel could be of practical interest for wide application in industry.
Table 2 Specification of geometrical parameters No. 1 2 3 4 5 6 7 8 9 10
HR
Le [mm]
W [mm]
H [mm]
0.0446 0.0800 0.126 0.184 0.253 0.0800 0.0800 0.0800 0.0800 0.0800
985 985 985 985 985 1,320 985 784 650 554
335 250 199 165 140 250 250 250 250 250
15 20 25 30 35 20 20 20 20 20
Fig. 6 Comparison between the simulation results and experimental data (a) Averaged Nu vs. Re, (b) f vs. Re
Fig. 7 Comparison of serpentine channel and multi-channel type Vol. 49 No. 5 2016
Fig. 8 Variation of heat transfer coefficient with Re under different HR (a) Air-side inlet flux=30 m3/h, (b) Air-side inlet flux=37.5 m3/h, (c) Air-side inlet flux=45 m3/h 439
Fig. 9 Variation of pressure drop with Re under different HR (a) Airside inlet flux=30 m3/h, (b) Air-side inlet flux=37.5 m3/h, (c) Air-side inlet flux=45 m3/h
4.4 Optimization of geometry parameters To investigate the influence of geometry parameters on heat transfer performance, the operating conditions should be constant, including the inlet flux, channel velocity and total heat exchange area as well. Table 2 presents the simulation models with different Le and HR. The Re number of gas-side ranges from 800 to 6,000, while the inlet flux of airside is set to 30.0 m3/h, 37.5 m3/h and 45.0 m3/h, respectively. In this work, the thermal-hydraulic performance is presented in terms of the heat transfer coefficient and pressure drop. Further, the surface performance factor is employed as the criterion of geometric optimization. The factor is defined by Eq. (14) (Kays and London, 1984).
Nu / RePr1/3 j = f f 440
(14)
Fig. 10 Variation of heat transfer coefficient with Re under different Le (a) Air-side inlet flux=30 m3/h, (b) Air-side inlet flux=37.5 m3/h, (c) Air-side inlet flux=45 m3/h
4.4.1 Effect of HR According to the models of Nos. 1–5, the HR of structure varies from 0.0446 to 0.253 with the same Le. Figure 8 displays the increasing tendency of heat transfer coefficient with the rise of Re number. The channel with HR of 0.0446 leads to the highest heat transfer, while those of 0.0800 and 0.126 reach the second and third highest heat transfer. Additionally, the effect of various HR on pressure drop is depicted in Figure 9. The dependence of pressure drop on HR is not remarkable with low velocities. However, a sharp change in heat transfer coefficient and pressure drop occurs when Re is larger than 4,000. As the HR decreases, the parameter of H gets smaller and W gets larger, leading to a decrease in Lt. The temperature difference between the fluid and heat exchange plate diminishes, which leads to the increase of effective heat transfer Journal of Chemical Engineering of Japan
Fig. 11 Variation of pressure drop with Re under different Le (a) Airside inlet flux=30 m3/h, (b) Air-side inlet flux=37.5 m3/h, (c) Air-side inlet flux=45 m3/h
Fig. 12 Variation of local heat transfer coefficient along the streamwise direction Vol. 49 No. 5 2016
Fig. 13 Variation of surface performance factor with different HR (a) Air-side flux=30 m3/h, (b) Air-side flux=37.5 m3/h, (c) Airside flux=45 m3/h
per unit area. Meanwhile, the fluid separates from the heat exchange plate, and then reattaches due to the variation of velocity direction in the bending regions. According to the above mentioned analysis, the thermal performance is enhanced and the influence of HR on heat transfer coefficient is more significant at a lower HR. Therefore, the serpentine air preheater with the lowest HR possesses the best heat capabilities. Nevertheless, the temperature of air in the channels with better heat capabilities gets higher and the viscosity of air is thus heightened. When the velocity is high (Re>4,000), the body resistance of air is improved by the increasing drag force caused by the viscosity and disturbance at the bending sections. 441
Fig. 15 Comparison between numerical data and Eq. (15)
local heat transfer, but also promotes the flow redevelopment in the region of sequentially cyclical straight channel. As the heat transfer area is constant, the number of bending zones increases if Le decreases. The local heat transfer enhancement with lower Le improves the total thermal performance. 4.4.3 Optimization on Le and HR As can be seen from Figures 8–11, the enhancement of heat transfer accompanies flow resistance in most cases. Therefore, to evaluate the thermal-hydraulic performance of different geometric parameters, the comparison of HR and Le is shown in Figures 13 and 14. Taking the surface performance factor as the target parameter, the surface performance factor decreases as Re increases. The enhancement capability of HR=0.0446 is much better in lower Re region (Re2,200). Meanwhile, the most appropriate effective path length is 1,320 mm. To satisfy the industrial application, an empirical correlation for heat transfer characteristics in the novel serpentine air preheater is obtained. By fitting data of Nos. 1–10, the correlation takes the following form. Fig. 14 Variation of surface performance factor with different Le (a) Air-side flux=30 m3/h, (b) Air-side flux=37.5 m3/h, (c) Airside flux=45 m3/h
4.4.2 Effect of Le Compared with the models of Nos. 6–10, the effective heat length varies from 1320 to 554 mm with a constant value of HR. Figures 10 and 11 present the heat transfer coefficient and pressure drop as a function of Re. In the figures, the shorter Le provides considerable heat transfer enhancement, corresponding to an increasing pressure drop. The reason is that the convective heat transfer gradually changes to thermal conduction by thickening of the thermal boundary layer as path length increases, which indicates that a shorter Le would have better enhancement capability. As shown in Figure 12, the peak values of local heat transfer coefficient appear at the bending zone and the section downstream the bending zone. Additionally, the heat transfer coefficient increases after each bending zone. It reveals that the rapid flow mixing not only improves the 442
HR Nu = 37.702 Re 0.254 L−e 0.166 1.692exp − + 0.955 (15) 0.0218
As shown in Figure 15, the average and maximum deviation of Nu correlation (Eq. (15)) is 4.72% and 21.15% respectively.
Conclusions The simulation and experimental investigation on the thermal-hydraulic performance of a novel serpentine air preheater have been carried out. The simulation results include the hot and cold outlet temperature, pressure drop, heat flux of heat transfer plate, and the velocity distribution. Serpentine air preheaters with different geometry parameters are simulated to study the effects of the Le and HR on the thermal-hydraulic performance. The main conclusions are as follows. 1. The numerical results are consistent with the experiJournal of Chemical Engineering of Japan
2.
3.
mental data and the maximum deviation for the Nu and f are less than 8.12%. The serpentine air preheater with smaller HR shows better heat transfer enhancement but larger pressure drop at the same time. The larger Le is favorable in the improvement of the thermal-hydraulic performance in the small Re region (Re2,200). An empirical correlation for Nu is obtained in a range of Re (Re=800–4,000). Compared with the numerical data, the average and maximum deviation of the correlation is 4.72% and 21.15%, respectively.
Acknowledgement The authors wish to acknowledge financial support by the National Natural Science Foundation of China (Project No. 21406157) and National Basic Research Program of China (Project No. 2012CB720500).
Nomenclature A cp de f H j K k Le Lt
= = = = = = = = = =
n Nu Pr qm Re T umax λ W
= = = = = = = = =
heat transfer surface area specific heat hydraulic diameter friction factor height of effective heat transfer flow path Colburn factor overall heat transfer coefficient mean heat transfer coefficient length of effective heat transfer flow path total length of effective heat transfer along the streamwise direction number of bending board Nusselt number Prandtl number mass flow rate Reynolds number temperature maximum velocity thermal conductivity of plate width of effective heat transfer flow path
ρ ∆t ∆tm ∆p
= = = =
fluid density temperature difference between inlet and outlet log-mean temperature difference pressure drop
[kg/m3] [°C] [°C] [Pa]
ϕ a g v
= = = =
heat transfer rate overall surface efficiency of air-side overall surface efficiency of gas-side kinematic viscosity
[kw] [—] [—] [N/m2]
‹Subscripts› air = air side gas = gas side in = inlet out = outlet
Vol. 49 No. 5 2016
[m−2] [J/(kg·K)] [mm] [—] [m] [—] [W/(m2 ·K)] [W/(m2 ·K)] [m] [m] [—] [—] [—] [kg/s] [—] [°C] [m/s] [W/(m·K)] [m]
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