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expanded graphite composite
Accepted Manuscript Experimental and numerical investigation on the novel latent heat exchanger with paraffin/expanded graphite composite Wenzhu Lin, Qianhao Wang, Xiaoming Fang, Xuenong Gao, Zhengguo Zhang PII: DOI: Reference:
S1359-4311(18)33962-0 https://doi.org/10.1016/j.applthermaleng.2018.08.103 ATE 12606
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
26 June 2018 29 August 2018 29 August 2018
Please cite this article as: W. Lin, Q. Wang, X. Fang, X. Gao, Z. Zhang, Experimental and numerical investigation on the novel latent heat exchanger with paraffin/expanded graphite composite, Applied Thermal Engineering (2018), doi: https://doi.org/10.1016/j.applthermaleng.2018.08.103
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Experimental and numerical investigation on the novel latent heat exchanger with paraffin/expanded graphite composite Wenzhu Lin a, Qianhao Wang a, Xiaoming Fang a,b, Xuenong Gao a,b, Zhengguo Zhang a,b,* a
Key Laboratory of Enhanced Heat Transfer and Energy Conservation, The
Ministry of Education, School of Chemistry and Chemical Engineering, South China University of Technology, Guangzhou 510640, China b
Guangdong Engineering Technology Research Center of Efficient Heat
Storage and Application, South China University of Technology, Guangzhou 510640, China
The corresponding author: Tel: +86-20-87112845; Fax: 86-20-87113870; E-mail: [email protected]*
Abstract In this paper, a novel latent heat exchanger with two flow channels was investigated experimentally and numerically. The phase change material was filled in the annular tube, while the working fluid flow in the shell and tube side runner. The design of multi flow channels allow it apply into energy storage system with different working fluid. The paraffin/expanded graphite composite with phase changing temperature around 50℃ was filled in the annular tubes. Heating and discharging tests have been performed in different flow rate. The numerical model was built and validated with 1
experimental data. The experimental results indicates that overall energy shows decreasing tendency with higher flow rate while the average power increases with inlet velocity. The liquid fraction and temperature distribution of PCM was studied numerically by validated simulation model. In addition, it can conclude that the heat resistance of system lay in the PCM side, the improvement of system overall thermal conductivity still need further investigation.
Keywords: Latent heat storage; phase change material; heat exchanger; paraffin
1. Introduction In the process of energy conversion and utilization, it often exist mismatch between supply and demand in time and space distribution, such as the peak-valley difference of power load, solar energy, intermittence of wind energy, etc [1]. The energy storage technology can adjust the supply and demand contradiction, thus achieve the purpose of energy saving and environmental protection [2]. The phase change material (PCM) can absorb or release latent heat during the phase change period, it has the advantages of large energy storage density, small temperature change and easy control, which make it one of the most potential heat storage materials at present [3]. The PCM can be applied into many fields, lots of work were done to investigate the latent heat storage system with phase change material [4]. The most commonly used PCM in energy storage system was paraffin [5], which has the advantages of high latent heat, good stability and relatively low price [6]. Medrano et al. [7] studied the RT35 in five different heat 2
exchanger’s melting and solidification performance and found the compact heat exchanger has the highest average thermal power. Cano et al. [8] investigated four different commercial paraffin in home-made heat exchanger and the result shows that Rubitherm RT42 has the best thermal performance since it was able to accumulate the highest amount of energy. Palomba et al. [9] studied commercial paraffin blend in a compact fin and tube heat exchanger and confirmed that heat storage density increasesed about 50% compared to sensible water storages. The shell-and-tube heat exchangers (STHEs) are widely used in the energy storage investigation [10]. Adine et al. [11] numerically studied the thermal performance of P116 and n-octadecane in single tube-shell heat exchanger. Gasia et al. [12] investigated the dynamic melting in a cylindrical shell-and-tube heat exchanger using water as PCM. Tao et al. [13] investigated the performance of high temperature molten salt in pipes latent heat storage unit. All researches revealed that STHEs were suitable for energy storage due to its flexible adaptability. However, low thermal conductivity of PCM still limit its application. In order to solve this problem, tubes with fins have been extensively studied, including double-pipe unit with a helical fin [14], Y-shaped fins with one and two bifurcations [15], square finned heat pipes [16], et al. From these studies, it has been proved that fins can significantly enhance the heat transfer of PCM due to the extended surface. Other structure such as the corrugated enclosures [17], double screw heat exchanger [18], flat plate [19] were also studied in thermal storage. The design and optimization of latent heat system were proposed to improve heat transfer performance, the material selection [20], different boundary condition [21], the optimization of design for tube and shell structure [22, 23] in latent heat storage were investigated. Nevertheless, 3
the heat exchanger studied before is usually in small scale and has only one flow channel, which limit the usage in the heat storage system with different working fluid. Therefore, the multi flow channels latent heat exchanger still need further investigation. In this paper, the large commercial latent heat exchanger (LHE) with two flow channels was experimentally and numerically studied, the paraffin/expanded graphite (EG) composite was selected as PCM filled in the annular tubes, the working fluid can flow both in shell side and tube side. The design of two-flow channels increase its operational flexibility, which solve the shortcomings of aforementioned LHE. It can apply into energy storage system with different working fluid, such as gas-water in industrial waste heat recovery and refrigerant-water in heat pump system. In addition, the large commercial scale system studied in this experiment can be more instrumental in practical application. The heat pump (HP) system was used to investigate the performance of LHE. HP is a heat-generating device that can be used in domestic water, space heating or air conditioning [24]. It has higher energy utilization efficiency than conventional heating or cooling systems [25]. The heat exchanger is used as water storage tank, due to the high latent heat of PCM, it can improve the heat capacity and save the space. In addition, with the electricity peak load shifting policy [26], the heat pump can heat and store energy at off-peak time and release when need, which can bring economic benefits. The charging and discharging performance of system in different flow rate at shell and tube flow channel was experimentally investigated, and the numerical model was built and validated by the experiment data, the melting and solidification of material was studied by simulation method.
4
2. Thermal storage system The latent heat exchanger in this research is for residential water, so the paraffin with phase change range between 45 ~ 50℃ was chosen to be the material. But paraffin was limited to its low thermal conductivity [27], which can be improved by absorbing liquid paraffin into expanded graphite to form stable and high thermal conductivity composite, the detailed preparation method can reference in former paper [6]. The prepared PCM composite is pressed into the annular tube of heat exchanger and the thermo-physical properties of paraffin/EG in this work is listed in Table 1.
Table 1. The thermo-physical properties of paraffin/EG Property
Value
Density (kg/m3)
300
Enthalpy (J/g)
155
Melting temperature (℃)
50.87
Solidification temperature (℃)
44.93
Thermal conductivity (W/m.K)
2.35
Specific heat (kJ/kg.K)
1.7
The heat exchanger is shell-and-tube structure with PCM in annular tube. The detailed structure can be seen in Fig. 1 and the cross section of inside tubes is shown in Fig. 2. The heat exchanger can be divided into three main parts: inside runner, annular tubes and shell side. The paraffin/EG composite is filled in annular tubes while the heat transfer fluid (HTF) —water flow in shell and tube side. The tubes are distributed in triangle uniformly while the segmental 5
baffles located in shell side to support tube bundle and guide the flow of working fluid. The streamline of fluid can be seen in Fig. 1(d), in tube runner, fluid flows vertically through the tube and reverse in chamber. The baffle in shell side make sure transverse flow fully exchange heat with PCM. Detailed geometry parameters of tested heat exchanger are indicated in Table 2.
Fig. 1. The structure of latent heat exchanger. (a) Inside tubes runner, (b) PCM in annular tubes, (c) Shell side runner, (d) The streamline of working fluid in shell and tube side.
6
Fig. 2. The cross section of inside tubes bundle.
Table 2. Geometry parameters of the latent heat exchanger Item
Dimensions and descriptions
Tube-side parameters di (mm)
12
Tube pinch t (mm)
39
Tube bundle pinch Sn (mm)
50
li (mm)
1000
Number of tubes
82
Inlet and outlet distance (mm)
213
Annular tubes parameters da (mm)
32
la (mm)
1000
Shell-side parameters do (mm)
480
lo (mm)
1000 7
Inlet and outlet distance (mm)
900
Baffle parameters Baffle pinch B (mm)
115
Baffle height (mm)
360
Thickness (mm)
3
Material
Stainless steel
Figure. 3 is the layout of experimental set-up and the schematic is shown in Fig. 4, the set-up mainly consists of a heat pump, water storage tanks and latent heat exchanger. The air source heat pump is used to heat or cool water in the buffer water tank, when the water reach the desired temperature, it was pumped into the heat exchanger and charging and discharging at constant mass flow rate is started. Test was stopped when the outlet HTF has two degree temperature difference with inlet water, which remain in 55℃ in heating process and 30℃ in discharging experiment. The HTF flow rates changed from 0.3 to 1.2 m3/h. The inlet, outlet temperature and volume flow rate of HTF were monitored by temperature sensor and flowmeter, which has the accuracy of ± 0.1℃ and ± 3%. Temperature of inlet HTF has little fluctuation, about ± 1.0 ℃. The wall of the heat exchanger and pipelines was carefully encapsulated with insulation material in order to reduce heat loss.
8
Fig. 3. Photograph of the experimental set-up
Fig. 4. Schematic diagram of the experimental setup
3. Data reduction The HTF flowing in the shell and tube runner to exchange heat with energy storage system, the following figures have been calculated to characterize and compare the heat transfer performance. 9
— Charge/discharge energy, to calculate the total amount of thermal energy stored in LHS unit, the following integral method is implemented:
E mc p Tin Tout d 0
(1)
— Average power, calculated as energy to whole duration time:
P
0
mc p Tin Tout d
(2)
— Shell side heat transfer coefficient, the segmental baffles’ shell side heat transfer coefficient can be calculated by following equation [28]: h0 0.36
de
Re00.55 Pr01/3 ( / w )0.14
(3)
Where λ is the thermal conductivity of the shell side fluid, d e is the equivalent diameter, Re0 is the Reynold number of shell side, Pr is the Prandtl number. η is the viscosity of the fluid. The equivalent diameter changes as the arrangement of tubes, when the piping arrangement was triangle, the equivalent diameter was calculated as [29]: 3 2 2 4 t di 2 4 de di
Re0
(4)
d eu0
(5)
u0 is the velocity of the HTF and it can be calculated as the volume flow rate to the flow area: u0
V0 S0
The flow area can be calculated as the following equation [30]:
10
(6)
d S0 BD 1 i t
(7)
—Tube side heat transfer coefficient: the tube side heat transfer coefficient can be calculated by Sieder-Tate equation [31]: Nui i di
hi
(8)
where the Nusselt number and Reynold number can be calculated as: di 13 0.14 Nui 1.86 (Rei Pri ) ( ) ( ) l w 1 3
Rei
di ui
(9) (10)
4. Mathematical modeling 4.1 Physical model The geometrical model for numerical simulation was built by software SolidWorks. The layout pattern and geometrical parameter of latent heat exchanger can be seen in Fig. 1 and Table. 2. In order to simplify the calculation while still keep the accuracy of results, necessary and reasonable assumptions were made: (1) the heat loss of system was neglected; (2) the natural convection of working fluid is neglected; (3) the thermal property of PCM and HTF is constant.
4.2 Grid generation and independence The tetrahedral grid was chosen to mesh shell side domain due to the complicated structure, and domain for PCM and inside tubes was generated by hexahedral grid. A careful check for grid independence was conducted to obtain accurate simulation result. The relative deviation of outlet 11
temperature and heat transfer rate of models with 1.12×107 and 1.50×107 cells were both less than 1.0%. Considering the calculation resource cost and solution accuracy, the final 1.12×107 grid cell was adopted for numerical investigation.
4.3 Numerical method The commercial software FLUENT was used to simulate the flow and heat transfer performance of latent heat exchanger, the Energy and Solidification and Melting model were chosen to simulate the heat transfer and PCM phase change performance. With the calculation of equation (5) and (10), while the flow rate change from 0.3 to 1.2 m3/h, the tube side Reynold number range from 325.5 to 1301.8, which is the laminar flow. Due to the complicated structure of shell side runner, it is often considered to be turbulence region when Reynold number is higher than 100 [32]. From equation (5), the shell side Reynold number is ranging from 240.3 to 916.3. So the laminar model was chosen for tube runner simulation and the standard k-ε model [33] was selected for shell side flow calculation. The SIMPLE algorithm scheme was adopted while the second order upwind scheme was used for spatial discretization. The convergence criterion for energy and continuity was less than 10 -7 and 10-4, respectively. A parallel computation with 24 numbers of processes was performed on workstation with 32 Intel Xeon-core CPUs and 32 GB RAM. It took approximately 36h to complete the calculation of one case.
4.4. Boundary conditions The inlet boundary was set to be mass flow inlet for shell and tube side, which can be calculated by equation 11, the volume flow rate was based on 12
experimental data, ranging from 0.3 to 1.2 m3/h, the mass flow rate is about 0.083 to 0.333 kg/s. The charging temperature was set to be 55℃ while the discharging temperature was 30℃. The outlet boundary was set to be pressure outlet and the shell wall and baffles were considered to be adiabatic. m V
(11)
5. Results and discussion 5.1 Shell side charging/discharging performance Fig. 5 shows the temperature profile for shell side charging and discharging experiments, differing from flow rate of HTF. The charging curve in Fig. 5(a) has a rapid increase at the first period, corresponding to the sensible heating of PCM in solid state. Then the PCM start melting and absorbing energy as latent heat, the temperature increases relatively slowly. For discharging process plotted in Fig. 5(b), the PCM suffer from three period: liquid sensible cooling, solid-liquid phase change, solid sensible cooling. The slope of the curve reduces in the solid-liquid period due to the slower dynamic of latent heat releasing process. All the curve show the same tendency: the higher the flow rate, the more rapid the charging and discharging test accomplish, and consequently the change of temperature inside the storage system. This is because the shell side heat transfer coefficient increases from 390.3 to 780.7 W.m-2.K-1 while the flow rate changing from 0.3 to 1.2 m3/h as equation 3 indicates. Increasing heat transfer coefficient lead to higher heat transfer rate between HTF and system. 13
(a)
(b) Fig. 5. Temperature profiles of shell side test in different flow rate. (a) charging; (b) discharging.
The cumulative and recovered energy and average power for charging and discharging tests at different flow rates is shown in Fig. 6. The total 14
energy decrease with higher flow rate and the average power increase with higher flow rate. This is because with higher flow rate, the heat transfer rate in shell side increase, corresponding to higher heat exchange power. With increasing heat transfer, shorter time was needed when the outlet temperature reaches the target value. Due to the relatively low thermal conductivity of PCM composite, the heat absorbing and releasing of system was inadequate, so the overall energy shows the decreasing tendency with higher flow rate. The charging energy lies in the range about 7.5 -13.5 MJ while the recovered energy was 7.0 -10.1 MJ. The average power increase significantly with higher flow rate, ranging from 4 kW to 14 kW, indicating the system has high working power.
(a)
15
(b) Fig. 6. Total energy and average power in shell side at different flow rate. (a) charging; (b) discharging.
5.2 Tube side charging/discharging performance The temperature profiles for thermocouple in the outlet of tube side at different flow rate is shown in Fig. 7, which show the same tendency with shell runner performance. In Fig. 7(a), the slope of temperature evolution reduces at around 48-50℃, which indicate the PCM is suffering from the phase changing period, consistent with the melting temperature of paraffin in Table.1. The discharging temperature can be seen in Fig. 7(b), also have a turning point when PCM suffer solidification process. From equation 8, the tube side heat transfer coefficient changes from 241.7 to 383.7 W.m-2.K-1 while flow rate ranging from 0.3 -1.2 m3/h. The fluid heat transfer rate increases with higher flow rate and the charging and discharging process was faster with increasing volume inlet.
16
(a)
(b) Fig. 7. Temperature profiles of tube side test in different flow rate. (a) charging; (b) discharging.
The total energy latent system absorbing and releasing when water flows in tube side is reported in Fig. 8, the total cumulative energy is around 7.2 to 10.1 MJ and recovered energy ranging from 6.3 to 9.9 MJ. The 17
average power increase with higher flow rate, ranging from 2.5 to 10.0 kW for heating and discharging process. It can be noted that the average power increase little when flow rate change from 0.8 to 1.2 m3/h, which can be explained by low thermal conductivity of PCM coupled to the heat exchanger efficiency. The thermal resistance lays in PCM side in higher flow rate, which is consistent with previously mentioned conclusion in shell side. Thus, increasing the thermal conductivity of PCM still needs further investigation.
(a)
18
(b) Fig. 8. Total energy and average power in tube side at different flow rate. (a) charging; (b) discharging.
5.3 Comparison between shell and tube runner As can be seen in Fig. 9, the discharging average power is always higher than the charging power, which is caused by higher temperature difference between inlet water and material’s phase changing range. The heating water is about 55℃, has smaller difference with PCM melting temperature around 51℃, and the discharging water remain at 30℃, has relatively higher difference with solidification point about 45 ℃ , the instantaneous power for discharging process is higher as equation 2 indicates. In addition, it is clearly shown that the shell charging power is higher than tube, it can be explained by the volume for water storage in shell side is much larger. The volume in shell side is about 0.115m3 and the tube side is around 0.027m3. Due to the relatively low thermal conductivity of PCM, tube side working fluid charging the water in shell side is much slower than the shell side directly fills the hot water in cavity. Considering the practical application, it can conclude that the shell side charging and tube side discharging is a better operating condition to achieve higher power and better economic benefits.
19
Fig. 9.
Comparison between shell and tube side average power
5.4 Model validation The numerical results of latent heat exchanger is compared with experimental data to validate the simulation model. For simplicity, the 1.0 m3/h volume flow rate was considered. The numerical and experimental outlet temperature for tube side in 1.0 m3/h flow rate is shown in Fig. 10(a), the maximum devitation for charging process is 6.27% and the discharging process is 7.27%. Shell side outlet water in 1.0 m3/h is indicated in Fig. 10(b), the variation of outlet temperature versus time was in a good agreement with the measured values, the maximum deviations of temperature were 3.47% for charging process and 7.91% for discharging. The numerical results meet well with experimental data, apart from unavoidable measurement uncertainty for experiment, the simplification in boundary condition and the difference of initialization conditions can be responsible for the deviations. Therefore, it can be concluded that the 20
numerical model is sufficiently accurate to investigate the heat transfer performance of investigated latent heat exchanger.
(a)
(b) Fig. 10. Comparison of outlet temperature between simulation and experiment. (a) tube side in 1.0 m3/h; (b) shell side in 1.0 m3/h.
21
5.5 Liquid fraction of PCM Due to the latent heat exchanger is a closed system, the temperature of inside PCM cannot be monitored, since the numerical model has been proved accurate enough, the simulation method is used to investigate the detailed performance of PCM composite. The liquid fraction of PCM in charging/discharging process of shell runner test is illustrated in Fig. 11. It is seen that with the increase in flow rate melt fraction (Fig. 11a) and solidification fraction (Fig. 11b) is increasing progressively. The slope changes very fast at the first stage of experiment, after a certain time, the slope change slowly. That means melt and solidification rate is relatively high at early stage, which can be explained by the heat transfer is controlled by conduction and convection mode. At the later stage, the PCM mainly stay in the thermally stratified condition, and the changing rate decrease [34]. The total melting time is about 2000s for 1.2 m3/h, longer than solidification duration about 1500s, which meets well with the higher power in discharging process indicated from experiment data. Fig. 12 shows the liquid fraction during tube side charging and discharging progress, showing same tendency with shell runner. The total melting time for 1.2 m3/h is about 2500s, longer than shell side 2000s, which also fit the result indicated in Fig. 9, that shell side charging power is higher than tube runner.
22
(a)
(b) Fig. 11. Liquid fraction of PCM. (a) shell side charging; (b) shell side discharging.
23
(a)
(b) Fig. 12. Liquid fraction of PCM. (a) tube side charging; (b) tube side discharging.
5.6 Temperature distribution of PCM Fig. 13 and Fig. 14 show the temperature contours of latent heat 24
exchanger when flow time is 500s. The initial temperature was 303.15 K, inlet flow rate was 1.2 m3/h and 328.15 K, corresponding to the heating process in experiment. Fig. 13 clearly shows that in shell side, with the guide of segmental baffles, the water slowly flows across the tube bundle and gradually heating the PCM in annular tubes. PCM temperature change along shell fluid streamline, consistent with baffle configuration. In Fig. 14, tube side water flow across the inlet section, reverse in the tube chamber and flow into the outlet section, the temperature change along vertical direction. Also, it can be noted that in 500s, water near shell wall mostly stay in 303.15K, fluid flowing in tube side need more effort to heat the shell side fluid. Due to the relativity lower thermal conductivity of PCM, average power in tube side charging decrease, confirming the result in Fig. 9 that shell side charging has higher power than tube side.
Fig. 13. Temperature contours in shell side charging (Time=500s, V=1.2m3/h). 25
Fig. 14. Temperature contours in tube side charging (Time=500s, V=1.2m3/h).
6. Conclusions In this paper, a novel structure LHE with multi flow channels was investigated, which fix the shortcoming of most studied one flow latent heat system. The novel multi runner LHE can apply into the heat storage application with different working fluid, such as gas-water in industrial waste heat recovery and refrigerant-water in heat pump system, etc. The experiment was performed to investigate the charging and discharging performance of shell and tube side runner. Water for charging and discharging was set to be 55℃ and 30℃ while flow rate changing from 0.3 to 1.2 m3/h. Numerical simulation was conducted to further investigate the detailed performance of PCM. It can be concluded from the following 26
results: 1. Under the constant inlet temperature, as the water flow rate increases, the shell and tube side’s charging/discharging rate also increase, shorter time was needed to complete the heat transfer between the working fluid and system. The shell side charging energy lies from 7.5 to 13.5 MJ while the recovered energy was about 7.0 to 10.1 MJ. Tube side total absorbing energy is around 7.2 to 10.1 MJ and releasing energy is around 6.3 to 9.9 MJ. The overall energy shows the decreasing tendency with higher flow rate while the average power increases with higher volume flow inlet, showing good performance in energy storage and release. 2. Due to the restriction of experiment, the detailed performance of PCM cannot be investigated, therefore, the numerical model was built and proved to be accurate enough to investigate liquid fraction of composite material. The numerical results show that with higher flow rate, the melting and solidification rate of PCM significantly increase and PCM temperature change along with baffle in shell side and vertical direction in tube runner. 3. The experimental and numerical result all appeal that the thermal resistance of system lay in the PCM side, in order to achieve higher power in application, the improvement of thermal conductivity in PCM side still need further investigation.
Nomenclature cp
heat capacity (kJ·kg-1·K-1)
B
baffle spacing (mm)
m
mass flow rate (kg·s-1) 27
E
energy (MJ)
P
average power (kW)
di, da, d0 t, Sn li, la, l0
diameter (mm) tube pinch (mm) length of tube (mm)
m
mass flow rate (kg·s-1)
Nu
Nusselt number
Pr
Prandtl number
Re
Reynold number
T
temperature (K)
u
velocity (m·s-1)
h
heat transfer coefficient (W·m-2·K-1)
V
volume flow rate (m3·s-1)
Greek symbol ρ
density (kg·m-3)
η
dynamic viscosity (Pa·s)
λ
thermal conductivity (W·m-1·K-1)
τ
time (s)
Subscripts i
inside tube
in
inlet
a
annular tube
out
outside
o
shell side
28
Acknowledgments This work was supported by the National Natural Science Foundation of China (No. U1507201), Guangdong Natural Science Foundation (2014A030312009) and the Applied Science and Technology Project of Guangdong Province (No. 2016B020243008)
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Numerical comparison of shell-side performance for shell and tube heat exchangers with trefoil-hole, helical and segmental baffles. Applied Thermal Engineering. 2016;109:175-85. 31. Wu Y-T, Liu S-W, Xiong Y-X, Ma C-F, Ding Y-L. Experimental study on the heat transfer characteristics of a low melting point salt in a parabolic trough solar collector system. Applied Thermal Engineering. 2015;89:748-54. 32. Qian S. Heat Exchanger Design Manual Chemical Industry Press. 2002;M. 33. Shih T-H, Liou WW, Shabbir A, Yang Z, Zhu J. A new k-ϵ eddy viscosity model for high reynolds number turbulent flows. Computers & Fluids. 1995;24(3):227-38. 34. Begum L, Hasan M, Vatistas GH. Energy storage in a complex heat storage unit using commercial grade phase change materials: Effect of convective heat transfer boundary conditions. Applied Thermal Engineering. 2018;131:621-41.
31
Highlights 1.
A novel latent heat exchanger with two flow channels was developed and tested.
2.
Experimental tests about charging and discharging performance of novel heat exchanger was presented.
3.
The detailed performance of phase change material was investigated numerically.
32
S1359-4311(18)33962-0 https://doi.org/10.1016/j.applthermaleng.2018.08.103 ATE 12606
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
26 June 2018 29 August 2018 29 August 2018
Please cite this article as: W. Lin, Q. Wang, X. Fang, X. Gao, Z. Zhang, Experimental and numerical investigation on the novel latent heat exchanger with paraffin/expanded graphite composite, Applied Thermal Engineering (2018), doi: https://doi.org/10.1016/j.applthermaleng.2018.08.103
This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Experimental and numerical investigation on the novel latent heat exchanger with paraffin/expanded graphite composite Wenzhu Lin a, Qianhao Wang a, Xiaoming Fang a,b, Xuenong Gao a,b, Zhengguo Zhang a,b,* a
Key Laboratory of Enhanced Heat Transfer and Energy Conservation, The
Ministry of Education, School of Chemistry and Chemical Engineering, South China University of Technology, Guangzhou 510640, China b
Guangdong Engineering Technology Research Center of Efficient Heat
Storage and Application, South China University of Technology, Guangzhou 510640, China
The corresponding author: Tel: +86-20-87112845; Fax: 86-20-87113870; E-mail: [email protected]*
Abstract In this paper, a novel latent heat exchanger with two flow channels was investigated experimentally and numerically. The phase change material was filled in the annular tube, while the working fluid flow in the shell and tube side runner. The design of multi flow channels allow it apply into energy storage system with different working fluid. The paraffin/expanded graphite composite with phase changing temperature around 50℃ was filled in the annular tubes. Heating and discharging tests have been performed in different flow rate. The numerical model was built and validated with 1
experimental data. The experimental results indicates that overall energy shows decreasing tendency with higher flow rate while the average power increases with inlet velocity. The liquid fraction and temperature distribution of PCM was studied numerically by validated simulation model. In addition, it can conclude that the heat resistance of system lay in the PCM side, the improvement of system overall thermal conductivity still need further investigation.
Keywords: Latent heat storage; phase change material; heat exchanger; paraffin
1. Introduction In the process of energy conversion and utilization, it often exist mismatch between supply and demand in time and space distribution, such as the peak-valley difference of power load, solar energy, intermittence of wind energy, etc [1]. The energy storage technology can adjust the supply and demand contradiction, thus achieve the purpose of energy saving and environmental protection [2]. The phase change material (PCM) can absorb or release latent heat during the phase change period, it has the advantages of large energy storage density, small temperature change and easy control, which make it one of the most potential heat storage materials at present [3]. The PCM can be applied into many fields, lots of work were done to investigate the latent heat storage system with phase change material [4]. The most commonly used PCM in energy storage system was paraffin [5], which has the advantages of high latent heat, good stability and relatively low price [6]. Medrano et al. [7] studied the RT35 in five different heat 2
exchanger’s melting and solidification performance and found the compact heat exchanger has the highest average thermal power. Cano et al. [8] investigated four different commercial paraffin in home-made heat exchanger and the result shows that Rubitherm RT42 has the best thermal performance since it was able to accumulate the highest amount of energy. Palomba et al. [9] studied commercial paraffin blend in a compact fin and tube heat exchanger and confirmed that heat storage density increasesed about 50% compared to sensible water storages. The shell-and-tube heat exchangers (STHEs) are widely used in the energy storage investigation [10]. Adine et al. [11] numerically studied the thermal performance of P116 and n-octadecane in single tube-shell heat exchanger. Gasia et al. [12] investigated the dynamic melting in a cylindrical shell-and-tube heat exchanger using water as PCM. Tao et al. [13] investigated the performance of high temperature molten salt in pipes latent heat storage unit. All researches revealed that STHEs were suitable for energy storage due to its flexible adaptability. However, low thermal conductivity of PCM still limit its application. In order to solve this problem, tubes with fins have been extensively studied, including double-pipe unit with a helical fin [14], Y-shaped fins with one and two bifurcations [15], square finned heat pipes [16], et al. From these studies, it has been proved that fins can significantly enhance the heat transfer of PCM due to the extended surface. Other structure such as the corrugated enclosures [17], double screw heat exchanger [18], flat plate [19] were also studied in thermal storage. The design and optimization of latent heat system were proposed to improve heat transfer performance, the material selection [20], different boundary condition [21], the optimization of design for tube and shell structure [22, 23] in latent heat storage were investigated. Nevertheless, 3
the heat exchanger studied before is usually in small scale and has only one flow channel, which limit the usage in the heat storage system with different working fluid. Therefore, the multi flow channels latent heat exchanger still need further investigation. In this paper, the large commercial latent heat exchanger (LHE) with two flow channels was experimentally and numerically studied, the paraffin/expanded graphite (EG) composite was selected as PCM filled in the annular tubes, the working fluid can flow both in shell side and tube side. The design of two-flow channels increase its operational flexibility, which solve the shortcomings of aforementioned LHE. It can apply into energy storage system with different working fluid, such as gas-water in industrial waste heat recovery and refrigerant-water in heat pump system. In addition, the large commercial scale system studied in this experiment can be more instrumental in practical application. The heat pump (HP) system was used to investigate the performance of LHE. HP is a heat-generating device that can be used in domestic water, space heating or air conditioning [24]. It has higher energy utilization efficiency than conventional heating or cooling systems [25]. The heat exchanger is used as water storage tank, due to the high latent heat of PCM, it can improve the heat capacity and save the space. In addition, with the electricity peak load shifting policy [26], the heat pump can heat and store energy at off-peak time and release when need, which can bring economic benefits. The charging and discharging performance of system in different flow rate at shell and tube flow channel was experimentally investigated, and the numerical model was built and validated by the experiment data, the melting and solidification of material was studied by simulation method.
4
2. Thermal storage system The latent heat exchanger in this research is for residential water, so the paraffin with phase change range between 45 ~ 50℃ was chosen to be the material. But paraffin was limited to its low thermal conductivity [27], which can be improved by absorbing liquid paraffin into expanded graphite to form stable and high thermal conductivity composite, the detailed preparation method can reference in former paper [6]. The prepared PCM composite is pressed into the annular tube of heat exchanger and the thermo-physical properties of paraffin/EG in this work is listed in Table 1.
Table 1. The thermo-physical properties of paraffin/EG Property
Value
Density (kg/m3)
300
Enthalpy (J/g)
155
Melting temperature (℃)
50.87
Solidification temperature (℃)
44.93
Thermal conductivity (W/m.K)
2.35
Specific heat (kJ/kg.K)
1.7
The heat exchanger is shell-and-tube structure with PCM in annular tube. The detailed structure can be seen in Fig. 1 and the cross section of inside tubes is shown in Fig. 2. The heat exchanger can be divided into three main parts: inside runner, annular tubes and shell side. The paraffin/EG composite is filled in annular tubes while the heat transfer fluid (HTF) —water flow in shell and tube side. The tubes are distributed in triangle uniformly while the segmental 5
baffles located in shell side to support tube bundle and guide the flow of working fluid. The streamline of fluid can be seen in Fig. 1(d), in tube runner, fluid flows vertically through the tube and reverse in chamber. The baffle in shell side make sure transverse flow fully exchange heat with PCM. Detailed geometry parameters of tested heat exchanger are indicated in Table 2.
Fig. 1. The structure of latent heat exchanger. (a) Inside tubes runner, (b) PCM in annular tubes, (c) Shell side runner, (d) The streamline of working fluid in shell and tube side.
6
Fig. 2. The cross section of inside tubes bundle.
Table 2. Geometry parameters of the latent heat exchanger Item
Dimensions and descriptions
Tube-side parameters di (mm)
12
Tube pinch t (mm)
39
Tube bundle pinch Sn (mm)
50
li (mm)
1000
Number of tubes
82
Inlet and outlet distance (mm)
213
Annular tubes parameters da (mm)
32
la (mm)
1000
Shell-side parameters do (mm)
480
lo (mm)
1000 7
Inlet and outlet distance (mm)
900
Baffle parameters Baffle pinch B (mm)
115
Baffle height (mm)
360
Thickness (mm)
3
Material
Stainless steel
Figure. 3 is the layout of experimental set-up and the schematic is shown in Fig. 4, the set-up mainly consists of a heat pump, water storage tanks and latent heat exchanger. The air source heat pump is used to heat or cool water in the buffer water tank, when the water reach the desired temperature, it was pumped into the heat exchanger and charging and discharging at constant mass flow rate is started. Test was stopped when the outlet HTF has two degree temperature difference with inlet water, which remain in 55℃ in heating process and 30℃ in discharging experiment. The HTF flow rates changed from 0.3 to 1.2 m3/h. The inlet, outlet temperature and volume flow rate of HTF were monitored by temperature sensor and flowmeter, which has the accuracy of ± 0.1℃ and ± 3%. Temperature of inlet HTF has little fluctuation, about ± 1.0 ℃. The wall of the heat exchanger and pipelines was carefully encapsulated with insulation material in order to reduce heat loss.
8
Fig. 3. Photograph of the experimental set-up
Fig. 4. Schematic diagram of the experimental setup
3. Data reduction The HTF flowing in the shell and tube runner to exchange heat with energy storage system, the following figures have been calculated to characterize and compare the heat transfer performance. 9
— Charge/discharge energy, to calculate the total amount of thermal energy stored in LHS unit, the following integral method is implemented:
E mc p Tin Tout d 0
(1)
— Average power, calculated as energy to whole duration time:
P
0
mc p Tin Tout d
(2)
— Shell side heat transfer coefficient, the segmental baffles’ shell side heat transfer coefficient can be calculated by following equation [28]: h0 0.36
de
Re00.55 Pr01/3 ( / w )0.14
(3)
Where λ is the thermal conductivity of the shell side fluid, d e is the equivalent diameter, Re0 is the Reynold number of shell side, Pr is the Prandtl number. η is the viscosity of the fluid. The equivalent diameter changes as the arrangement of tubes, when the piping arrangement was triangle, the equivalent diameter was calculated as [29]: 3 2 2 4 t di 2 4 de di
Re0
(4)
d eu0
(5)
u0 is the velocity of the HTF and it can be calculated as the volume flow rate to the flow area: u0
V0 S0
The flow area can be calculated as the following equation [30]:
10
(6)
d S0 BD 1 i t
(7)
—Tube side heat transfer coefficient: the tube side heat transfer coefficient can be calculated by Sieder-Tate equation [31]: Nui i di
hi
(8)
where the Nusselt number and Reynold number can be calculated as: di 13 0.14 Nui 1.86 (Rei Pri ) ( ) ( ) l w 1 3
Rei
di ui
(9) (10)
4. Mathematical modeling 4.1 Physical model The geometrical model for numerical simulation was built by software SolidWorks. The layout pattern and geometrical parameter of latent heat exchanger can be seen in Fig. 1 and Table. 2. In order to simplify the calculation while still keep the accuracy of results, necessary and reasonable assumptions were made: (1) the heat loss of system was neglected; (2) the natural convection of working fluid is neglected; (3) the thermal property of PCM and HTF is constant.
4.2 Grid generation and independence The tetrahedral grid was chosen to mesh shell side domain due to the complicated structure, and domain for PCM and inside tubes was generated by hexahedral grid. A careful check for grid independence was conducted to obtain accurate simulation result. The relative deviation of outlet 11
temperature and heat transfer rate of models with 1.12×107 and 1.50×107 cells were both less than 1.0%. Considering the calculation resource cost and solution accuracy, the final 1.12×107 grid cell was adopted for numerical investigation.
4.3 Numerical method The commercial software FLUENT was used to simulate the flow and heat transfer performance of latent heat exchanger, the Energy and Solidification and Melting model were chosen to simulate the heat transfer and PCM phase change performance. With the calculation of equation (5) and (10), while the flow rate change from 0.3 to 1.2 m3/h, the tube side Reynold number range from 325.5 to 1301.8, which is the laminar flow. Due to the complicated structure of shell side runner, it is often considered to be turbulence region when Reynold number is higher than 100 [32]. From equation (5), the shell side Reynold number is ranging from 240.3 to 916.3. So the laminar model was chosen for tube runner simulation and the standard k-ε model [33] was selected for shell side flow calculation. The SIMPLE algorithm scheme was adopted while the second order upwind scheme was used for spatial discretization. The convergence criterion for energy and continuity was less than 10 -7 and 10-4, respectively. A parallel computation with 24 numbers of processes was performed on workstation with 32 Intel Xeon-core CPUs and 32 GB RAM. It took approximately 36h to complete the calculation of one case.
4.4. Boundary conditions The inlet boundary was set to be mass flow inlet for shell and tube side, which can be calculated by equation 11, the volume flow rate was based on 12
experimental data, ranging from 0.3 to 1.2 m3/h, the mass flow rate is about 0.083 to 0.333 kg/s. The charging temperature was set to be 55℃ while the discharging temperature was 30℃. The outlet boundary was set to be pressure outlet and the shell wall and baffles were considered to be adiabatic. m V
(11)
5. Results and discussion 5.1 Shell side charging/discharging performance Fig. 5 shows the temperature profile for shell side charging and discharging experiments, differing from flow rate of HTF. The charging curve in Fig. 5(a) has a rapid increase at the first period, corresponding to the sensible heating of PCM in solid state. Then the PCM start melting and absorbing energy as latent heat, the temperature increases relatively slowly. For discharging process plotted in Fig. 5(b), the PCM suffer from three period: liquid sensible cooling, solid-liquid phase change, solid sensible cooling. The slope of the curve reduces in the solid-liquid period due to the slower dynamic of latent heat releasing process. All the curve show the same tendency: the higher the flow rate, the more rapid the charging and discharging test accomplish, and consequently the change of temperature inside the storage system. This is because the shell side heat transfer coefficient increases from 390.3 to 780.7 W.m-2.K-1 while the flow rate changing from 0.3 to 1.2 m3/h as equation 3 indicates. Increasing heat transfer coefficient lead to higher heat transfer rate between HTF and system. 13
(a)
(b) Fig. 5. Temperature profiles of shell side test in different flow rate. (a) charging; (b) discharging.
The cumulative and recovered energy and average power for charging and discharging tests at different flow rates is shown in Fig. 6. The total 14
energy decrease with higher flow rate and the average power increase with higher flow rate. This is because with higher flow rate, the heat transfer rate in shell side increase, corresponding to higher heat exchange power. With increasing heat transfer, shorter time was needed when the outlet temperature reaches the target value. Due to the relatively low thermal conductivity of PCM composite, the heat absorbing and releasing of system was inadequate, so the overall energy shows the decreasing tendency with higher flow rate. The charging energy lies in the range about 7.5 -13.5 MJ while the recovered energy was 7.0 -10.1 MJ. The average power increase significantly with higher flow rate, ranging from 4 kW to 14 kW, indicating the system has high working power.
(a)
15
(b) Fig. 6. Total energy and average power in shell side at different flow rate. (a) charging; (b) discharging.
5.2 Tube side charging/discharging performance The temperature profiles for thermocouple in the outlet of tube side at different flow rate is shown in Fig. 7, which show the same tendency with shell runner performance. In Fig. 7(a), the slope of temperature evolution reduces at around 48-50℃, which indicate the PCM is suffering from the phase changing period, consistent with the melting temperature of paraffin in Table.1. The discharging temperature can be seen in Fig. 7(b), also have a turning point when PCM suffer solidification process. From equation 8, the tube side heat transfer coefficient changes from 241.7 to 383.7 W.m-2.K-1 while flow rate ranging from 0.3 -1.2 m3/h. The fluid heat transfer rate increases with higher flow rate and the charging and discharging process was faster with increasing volume inlet.
16
(a)
(b) Fig. 7. Temperature profiles of tube side test in different flow rate. (a) charging; (b) discharging.
The total energy latent system absorbing and releasing when water flows in tube side is reported in Fig. 8, the total cumulative energy is around 7.2 to 10.1 MJ and recovered energy ranging from 6.3 to 9.9 MJ. The 17
average power increase with higher flow rate, ranging from 2.5 to 10.0 kW for heating and discharging process. It can be noted that the average power increase little when flow rate change from 0.8 to 1.2 m3/h, which can be explained by low thermal conductivity of PCM coupled to the heat exchanger efficiency. The thermal resistance lays in PCM side in higher flow rate, which is consistent with previously mentioned conclusion in shell side. Thus, increasing the thermal conductivity of PCM still needs further investigation.
(a)
18
(b) Fig. 8. Total energy and average power in tube side at different flow rate. (a) charging; (b) discharging.
5.3 Comparison between shell and tube runner As can be seen in Fig. 9, the discharging average power is always higher than the charging power, which is caused by higher temperature difference between inlet water and material’s phase changing range. The heating water is about 55℃, has smaller difference with PCM melting temperature around 51℃, and the discharging water remain at 30℃, has relatively higher difference with solidification point about 45 ℃ , the instantaneous power for discharging process is higher as equation 2 indicates. In addition, it is clearly shown that the shell charging power is higher than tube, it can be explained by the volume for water storage in shell side is much larger. The volume in shell side is about 0.115m3 and the tube side is around 0.027m3. Due to the relatively low thermal conductivity of PCM, tube side working fluid charging the water in shell side is much slower than the shell side directly fills the hot water in cavity. Considering the practical application, it can conclude that the shell side charging and tube side discharging is a better operating condition to achieve higher power and better economic benefits.
19
Fig. 9.
Comparison between shell and tube side average power
5.4 Model validation The numerical results of latent heat exchanger is compared with experimental data to validate the simulation model. For simplicity, the 1.0 m3/h volume flow rate was considered. The numerical and experimental outlet temperature for tube side in 1.0 m3/h flow rate is shown in Fig. 10(a), the maximum devitation for charging process is 6.27% and the discharging process is 7.27%. Shell side outlet water in 1.0 m3/h is indicated in Fig. 10(b), the variation of outlet temperature versus time was in a good agreement with the measured values, the maximum deviations of temperature were 3.47% for charging process and 7.91% for discharging. The numerical results meet well with experimental data, apart from unavoidable measurement uncertainty for experiment, the simplification in boundary condition and the difference of initialization conditions can be responsible for the deviations. Therefore, it can be concluded that the 20
numerical model is sufficiently accurate to investigate the heat transfer performance of investigated latent heat exchanger.
(a)
(b) Fig. 10. Comparison of outlet temperature between simulation and experiment. (a) tube side in 1.0 m3/h; (b) shell side in 1.0 m3/h.
21
5.5 Liquid fraction of PCM Due to the latent heat exchanger is a closed system, the temperature of inside PCM cannot be monitored, since the numerical model has been proved accurate enough, the simulation method is used to investigate the detailed performance of PCM composite. The liquid fraction of PCM in charging/discharging process of shell runner test is illustrated in Fig. 11. It is seen that with the increase in flow rate melt fraction (Fig. 11a) and solidification fraction (Fig. 11b) is increasing progressively. The slope changes very fast at the first stage of experiment, after a certain time, the slope change slowly. That means melt and solidification rate is relatively high at early stage, which can be explained by the heat transfer is controlled by conduction and convection mode. At the later stage, the PCM mainly stay in the thermally stratified condition, and the changing rate decrease [34]. The total melting time is about 2000s for 1.2 m3/h, longer than solidification duration about 1500s, which meets well with the higher power in discharging process indicated from experiment data. Fig. 12 shows the liquid fraction during tube side charging and discharging progress, showing same tendency with shell runner. The total melting time for 1.2 m3/h is about 2500s, longer than shell side 2000s, which also fit the result indicated in Fig. 9, that shell side charging power is higher than tube runner.
22
(a)
(b) Fig. 11. Liquid fraction of PCM. (a) shell side charging; (b) shell side discharging.
23
(a)
(b) Fig. 12. Liquid fraction of PCM. (a) tube side charging; (b) tube side discharging.
5.6 Temperature distribution of PCM Fig. 13 and Fig. 14 show the temperature contours of latent heat 24
exchanger when flow time is 500s. The initial temperature was 303.15 K, inlet flow rate was 1.2 m3/h and 328.15 K, corresponding to the heating process in experiment. Fig. 13 clearly shows that in shell side, with the guide of segmental baffles, the water slowly flows across the tube bundle and gradually heating the PCM in annular tubes. PCM temperature change along shell fluid streamline, consistent with baffle configuration. In Fig. 14, tube side water flow across the inlet section, reverse in the tube chamber and flow into the outlet section, the temperature change along vertical direction. Also, it can be noted that in 500s, water near shell wall mostly stay in 303.15K, fluid flowing in tube side need more effort to heat the shell side fluid. Due to the relativity lower thermal conductivity of PCM, average power in tube side charging decrease, confirming the result in Fig. 9 that shell side charging has higher power than tube side.
Fig. 13. Temperature contours in shell side charging (Time=500s, V=1.2m3/h). 25
Fig. 14. Temperature contours in tube side charging (Time=500s, V=1.2m3/h).
6. Conclusions In this paper, a novel structure LHE with multi flow channels was investigated, which fix the shortcoming of most studied one flow latent heat system. The novel multi runner LHE can apply into the heat storage application with different working fluid, such as gas-water in industrial waste heat recovery and refrigerant-water in heat pump system, etc. The experiment was performed to investigate the charging and discharging performance of shell and tube side runner. Water for charging and discharging was set to be 55℃ and 30℃ while flow rate changing from 0.3 to 1.2 m3/h. Numerical simulation was conducted to further investigate the detailed performance of PCM. It can be concluded from the following 26
results: 1. Under the constant inlet temperature, as the water flow rate increases, the shell and tube side’s charging/discharging rate also increase, shorter time was needed to complete the heat transfer between the working fluid and system. The shell side charging energy lies from 7.5 to 13.5 MJ while the recovered energy was about 7.0 to 10.1 MJ. Tube side total absorbing energy is around 7.2 to 10.1 MJ and releasing energy is around 6.3 to 9.9 MJ. The overall energy shows the decreasing tendency with higher flow rate while the average power increases with higher volume flow inlet, showing good performance in energy storage and release. 2. Due to the restriction of experiment, the detailed performance of PCM cannot be investigated, therefore, the numerical model was built and proved to be accurate enough to investigate liquid fraction of composite material. The numerical results show that with higher flow rate, the melting and solidification rate of PCM significantly increase and PCM temperature change along with baffle in shell side and vertical direction in tube runner. 3. The experimental and numerical result all appeal that the thermal resistance of system lay in the PCM side, in order to achieve higher power in application, the improvement of thermal conductivity in PCM side still need further investigation.
Nomenclature cp
heat capacity (kJ·kg-1·K-1)
B
baffle spacing (mm)
m
mass flow rate (kg·s-1) 27
E
energy (MJ)
P
average power (kW)
di, da, d0 t, Sn li, la, l0
diameter (mm) tube pinch (mm) length of tube (mm)
m
mass flow rate (kg·s-1)
Nu
Nusselt number
Pr
Prandtl number
Re
Reynold number
T
temperature (K)
u
velocity (m·s-1)
h
heat transfer coefficient (W·m-2·K-1)
V
volume flow rate (m3·s-1)
Greek symbol ρ
density (kg·m-3)
η
dynamic viscosity (Pa·s)
λ
thermal conductivity (W·m-1·K-1)
τ
time (s)
Subscripts i
inside tube
in
inlet
a
annular tube
out
outside
o
shell side
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Acknowledgments This work was supported by the National Natural Science Foundation of China (No. U1507201), Guangdong Natural Science Foundation (2014A030312009) and the Applied Science and Technology Project of Guangdong Province (No. 2016B020243008)
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Highlights 1.
A novel latent heat exchanger with two flow channels was developed and tested.
2.
Experimental tests about charging and discharging performance of novel heat exchanger was presented.
3.
The detailed performance of phase change material was investigated numerically.
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