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Numerical model and simulation of a vehicular heat storage system with phase-change material
Applied Thermal Engineering 113 (2017) 1496–1504
Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
Numerical model and simulation of a vehicular heat storage system with phase-change material Sangki Park a, Seungchul Woo a, Jungwook Shon b, Kihyung Lee a,⇑ a b
Department of Mechanical Engineering, Hanyang University, 1271 Sa1-dong, Sangrok-gu, Gyeonggi-do, Republic of Korea Powertrain HEX Team, Hanon Systems, 95, Sinilseo-ro, Daedeok-gu, Daejeon, Republic of Korea
h i g h l i g h t s Numerical analysis and evaluation of a heat storage system with a phase-change material. Numerical simulation of the New European Driving Cycle was carried out for a vehicle. The warm-up time was reduced by 40.5% and the fuel consumption was reduced by 2.71%.
a r t i c l e
i n f o
Article history: Received 3 August 2016 Revised 14 October 2016 Accepted 22 November 2016 Available online 23 November 2016 Keywords: Phase-change material Heat storage system One-dimensional numerical analysis Diesel engine
a b s t r a c t For heat storage applications designed to recover and recycle waste heat energy, it is usually advantageous to store heat in a phase-change material. One-dimensional numerical analysis and evaluation of a heat storage system that uses a phase-change material to store latent heat in addition to sensible heat was carried out, and it was found that up to 30% of the total heat energy generated by the fuel and subsequently lost to cooling can be recovered. A heat storage system was installed to reduce warm-up time by releasing heat directly into the engine coolant during cold start, and the corresponding reduction in fuel consumption was measured. With the addition of a heat storage device, the warm-up time to 95 °C was reduced by between 18.1% and 27.1%. A numerical simulation of the New European Driving Cycle was carried out for a vehicle equipped with a 1.6-L diesel engine and a heat storage system. Analytical results showed that the warm-up time was reduced by 40.5% and the fuel consumption was reduced by 2.71% compared to a vehicle without a heat storage system installed. Ó 2016 Elsevier Ltd. All rights reserved.
1. Introduction Modern vehicle engines are designed to consume unnecessarily large amounts of fuel in order to warm up cooling water during initial cold start, which increases emissions and lowers fuel economy. In order to resolve this issue, alternative warming methods have been proposed, such as heating and storing cooling water with an electric heater and then resupplying to the engine during ignition or installing a separate heat exchanger in the exhaust pipe to capture waste heat [1,2]. The major components of a heat storage system can be broadly categorized into a phase-change material, a heat exchanger, and an insulation vessel. The most important criteria to evaluate their performance are their heat storage and heat radiation capabilities. The heat exchanger, which transfers heat energy into and out of the ⇑ Corresponding author at: Department of Mechanical Engineering, Hanyang University, 1271 Sa1-dong, Sangrok-gu, Gyeonggi-do 426-791, Republic of Korea. E-mail address: [email protected] (K. Lee). http://dx.doi.org/10.1016/j.applthermaleng.2016.11.162 1359-4311/Ó 2016 Elsevier Ltd. All rights reserved.
phase-change material, determines the speed of heat storage and radiation. The insulation vessel preserves stored energy until the moment when it is needed and plays an important role [3]. Phase-change materials take advantage of latent heat in a phase transition to store energy for a given temperature difference. Storing energy only as sensible heat results in a dramatically lower heat capacity. For example, the specific heat of paraffin wax is 2.1 kJ/kg °C, and its latent heat is 140 kJ/kg. Compared with an equivalent mass of water, which has a specific heat of 4.2 kJ/kg °C, if the temperature increases by 15 °C, the additional heat stored in paraffin wax after melting and reaching equilibrium is 2.7 times greater [4]. When a solid phase-change material is heated to its melting point, its phase changes to liquid. During this process, latent heat (also known as enthalpy of fusion) is absorbed, and the temperature of the material remains constant, as shown in Fig. 1. Likewise, in the reverse situation when the phase changes from liquid to solid at the freezing point, latent heat is discharged with no change
S. Park et al. / Applied Thermal Engineering 113 (2017) 1496–1504
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Fig. 1. Energy charge characteristics in a phase-change material [5].
in temperature. The same phenomenon occurs in the liquid-gas phase transition. In this study, one-dimensional numerical analysis and evaluation were carried out in order to design a heat storage system that uses a phase-change material to recover around 30% of the total heat energy produced by fuel and subsequently lost to engine cooling.
2. Modeling and simulation 2.1. Analysis of a commercial heat exchanger Heat transfer is a phenomenon where heat energy flows from high temperature to low temperature. The movement of heat can be categorized as conduction, convection, radiation, or any combination of the three. Furthermore, if substances are displaced or experience phase changes, the analysis can get quite complicated. However, analytical results similar to those obtained experimentally can often be obtained even for complicated heat transfer processes if appropriate models and empirical equations are used. Heat transfer processes are implemented in various commercially-available software programs such as Amesim by LMS, which is based on Modelica. A commercial heat exchanger of the inner tubes and outer fins was applied in this study. The side that makes contact with air has louver-fins, which are designed to increase the rate of heat exchange by creating turbulence. However, in our case, the phase-change material makes direct contact with the fins and tubes without air flow, so the heat exchanger can be modeled more simply with an equivalent flat fin design. The total heat transfer coefficient describing the transfer of heat from the internal fluid to the outside through the heat exchanger can be expressed as a sum of heat resistances. If the fins and tubes are brazed, the total thermal resistance is the sum of the convective thermal resistance inside the tubes and the conductive thermal resistance of the fins and tubes. However, the rate of heat
transfer between the heat exchanger and the phase-change material is different during melting and freezing [6]. Phase-change material was added to a commercial heat exchanger to obtain preliminary data in order to fabricate a heat exchanger for a heat storage device, and the melting time was analyzed. The optimum fin pitch, temperature, and flow rate were determined by systematically varying each parameter. A cooling water solution containing 50% ethylene glycol was used for analysis. The physical properties of the cooling water are presented in Table 1. In the commercial heat exchanger, the distance between fins was 6 mm, the length of fins was 41.5 mm, the fin thickness was 0.1 mm, and fins were arranged in 17 rows of 139 fins. The heat exchanger also contained a total of 64 tubes with a hydraulic diameter of 2.418 mm. It was constructed as a U-band type, which is the same as a heat exchanger having 16 tubes and length 390 mm. The total mass of phase change material was 0.714 kg, and the total heat transfer area was 1.385 m2. A heat storage device is expected to be most efficient when installed in the heater core line water, which is directly supplied to the engine. The flow rate of heater core water for a 2000 cc capacity engine is shown in Table 2. Since the number of revolutions per minute (rpm) of vehicles during cold starting usually falls within the range of 1000–2000, three speeds were studied: 1000, 1500, and 2000 rpm. Around 3 L of cooling water is stored in commercially available sensible heat storage devices; therefore, three commercial heat exchangers must be connected in parallel to store 3 kg of phasechange material. However, the same analysis can be performed by reducing the amount of phase change material and the flow rate of cooling water by 1/3 rather than connecting heat exchangers in parallel. In order to determine the convective heat transfer coefficient inside the heat exchanger, the hydraulic diameter Dh and flow speed V flow need to be calculated with Eqs. (1) and (2), respectively:
Dh ¼ 4
Afree ðNon-circular cross sectionÞ Pcontact
ð1Þ
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Table 1 Properties of the 50% ethylene glycol aqueous solution. Temperature (°C)
Density (kg/m3)
Specific heat (kJ/kg °C)
Heat conductivity (W/m °C)
Dynamic viscosity (kg/m s)
75 85 95
1042.79 1035.50 1027.79
3.495 3.535 3.574
0.408 0.411 0.413
0.001 0.001 0.001
Table 2 Heater core coolant flow rate for each engine speed. Speed (rpm)
Flow rate (L/min)
1000 1500 2000
8.6 12 15.5
V flow ¼
Q flow Afree
ð2Þ
where Afree is the area of free flow, Pcontact is the length of tubes in contact with fluid, and Q flow is the flow rate. Since cooling water moves through tubes in the form of a liquid, it is a single-phase flow. The Prandtl number N Pr and Reynolds number N Re are calculated with Eqs. (3) and (4), respectively:
NPr ¼
Cp l K
ð3Þ
NRe ¼
q V flow Dh l
ð4Þ
where C p is the specific heat, l is the dynamic viscosity coefficient, K is the thermal conductivity, q is the density, and l is the dynamic viscosity coefficient. The flow of cooling water inside tubes is turbulent, so the convective heat transfer coefficient is calculated with Eq. (5), which is based on the Dittus-Boelter equation: n NNu ¼ 0:023 N0:8 Re N Pr ¼
h Dh K
ð5Þ
n ¼ 0:4 : Internal temperature is higher than external temperature
is no contact heat resistance between them, so the conductive heat resistance is calculated with Eq. (7). The specific length of conduction through the tubes is the tube wall thickness, and the specific length of conduction through fins is their length. Since conductive heat resistance through a heat exchanger is determined by its physical shape, identical heat exchangers have the same conductive heat resistance.
RConv ection ¼
1 h Atotal
ð6Þ
where h is the convection heat transfer coefficient, and Atotal is the total heat transfer area.
RConduction ¼
Lc K Atotal
ð7Þ
where Lc is the characteristic length, K is the thermal conductivity coefficient, and Atotal is the total heat transfer area. Fins are in contact with tubes at both ends, and phase-change material does not flow, so it is assumed that the heat transfer efficiency is 100%, and heat is transferred by conduction only. Since tubes and fins transfer heat simultaneously, their resistances are added in parallel. The sum of the conductive heat resistances in the heat exchanger is therefore calculated as in Eq. (8).
Rtotal; conduction ¼
1 Rtube
þ
1
1
Rtube
ð8Þ
The total heat resistance is obtained by adding the convective heat resistance and conductive heat resistance of tubes and fins using Eq. (9). Once the total heat transfer is obtained by the above method, the only factor affecting the heat transfer to the solid material is the flow rate of cooling water, provided that the same heat exchanger is used [7].
n ¼ 0:3 : External temperature is higher than internal temperature
Rtotal ¼ Rconv ection ðCoolantÞ þ Rconduction ðHeat exchangerÞ
where N Nu is the Nusselt number, NRe is the Reynolds number, N Pr is the Prandtl number, h is the convection heat transfer coefficient, Dh is the hydraulic diameter, and K is the thermal conductivity. In order to add the convective and the conductive heat transfer coefficients in the form of heat resistances, the convective heat resistance should be calculated with Eq. (6). Fins and tubes of the heat exchanger are all made of AL1100 aluminum alloy, and there
After melting has started, heat transfer by natural convection in the liquid layer is added to the above resistance. Because the liquid layer develops as time elapses, an equation describing the variation in heat resistance with time is required [8]. Two of the six faces of the container encapsulating the phasechange material do not receive heat directly by contact to the outer case, so the space can be imagined as one-dimensional with a rect-
Fig. 2. Schematic diagram showing the increase in thickness of the melted layer of phase-change material with time.
ð9Þ
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Fig. 3. Visual measurement of phase-change material melting time.
angular cross section in contact with the fins and tubes. If the rate of phase change per hour is assumed to be constant, the amount of liquid can be found by calculating DS with a x2-x1 increase in the liquid layer thickness with time, as shown in Fig. 2. The increase in the convection heat resistance RC v L with time due to the increasing thickness of the liquid layer is given by Eq. (10). If it is assumed that heat energy in the cooling water is transferred to the phase-change material without loss as the phasechange material melts, the quantity of heat in the coolant DQ c can be expressed by Eq. (11), where the surface area of melted phase-change material DSarea is calculated using Eq. (12). Once the surface area is obtained, a x2-x1 increase in the thickness of
the melted layer can be calculated, from which the total amount of melted phase-change material can be determined.
RC v L ¼
1 Dtime hLP DSarea
ð10Þ
DQ c ¼
DT coolant RC v C þ RC d H þ RC v L
ð11Þ
DSarea ¼
hLP
Fig. 4. New European Driving Cycle.
DTime DT
DQ coolant
ðRC v C þ RC d H Þ
ð12Þ
1500
S. Park et al. / Applied Thermal Engineering 113 (2017) 1496–1504
Fig. 5. Vehicle model and heat storage control model used for analysis.
where RCv C is the heat resistance of coolant convection, RC d H is the heat resistance of heat exchanger conduction, hLP is the convection heat transfer coefficient, and DT is the temperature difference. The above equation was programmed into one-dimensional commercial Amesim software [9]. First, the melting time was calculated as a function of the flow rate, with the cooling water temperature held constant at 95 °C. Then, the melting time was calculated for a constant flow rate with the cooling water temperature set to 75 °C, 85 °C, or 95 °C. 2.2. Validation test of the analytical result using a real heat exchanger When the heat exchanger and phase-change material are located inside a sealed container, the amount of melted phasechange material cannot be measured directly. Therefore, to determine the actual development time of the liquid layer, a device with a fin pitch of 1.3 mm was used to heat the phase-change material evenly as in a commercial heat exchanger, but a window was installed to visualize the melting layer as shown in Fig. 3. Photographs of the liquid layer between the fins and tubes were taken at one frame per second using a digital camera. The thickness of the liquid layer in each frame was then measured using an image tool, and the total amount of melted material in the heat exchanger was calculated. For this experiment, cooling water was supplied at 95 °C with a flow rate of 5.2 L/min. 2.3. Analysis of fuel consumption in drive mode using simulation In order to investigate the improvement in fuel consumption when a heat storage device is installed, a chassis dynamometer should be used. However, since a dynamometer was not installed in the actual vehicle simulation bench, the improvement in fuel consumption in driving mode cannot be known experimentally and was instead modeled analytically. The New European Driving Cycle (NEDC) was adopted for the drive mode analysis, which is the standard test used to evaluate fuel consumption and exhaust material in Europe. In NEDC, the
first 780 s simulate driving in a city center, while the next 400 s simulate driving on national highways and express ways, as shown in Fig. 4. The experiment should be conducted when the temperature of the vehicle is stabilized at 25 °C, and the chassis dynamometer should account for the weight of the vehicle and the effect of air resistance. The cooling system and detailed specifications of a vehicle with a commercial 1.6 L capacity diesel engine, including appropriate gear ratios and transmission specifications, were used to construct an Amesim model for analysis, as shown in Fig. 5(a). The vehicle specifications are given in Table 3. For the tests, 8.1 L of cooling water was contained in the system including the heat storage device. Fuel consumption and cooling water temperature with and without a heat storage device installed during cold starting were compared in order to measure the effect of the heat storage device on the overall fuel consumption. A control algorithm for the heat storage system was implemented as shown in Fig. 5(b) to avoid overcooling the cooling water, and the time required for re-storing heat was measured.
Table 3 Specifications of the test vehicle for analysis. Type
Diesel engine (EURO5)
Number of cylinders Displacement (cc) Bore Stroke (mm) Compression ratio Max power (ps/rpm) Max torque (kg m/rpm) Vehicle weight (kg) Transmission gear ratio 1st 2nd 3rd 4th 5th 6th Reverse
4 1582 77.2 84.5 17.3:1 128/4000 26.5/1900–2750 1360 6 speed 4.212 2.637 1.800 1.386 1.000 0.772 3.385
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Fig. 6. Photographs showing increasing stearic acid melted layer thickness with time.
Fig. 7. Comparison of analytical and experimental stearic acid melting results.
3. Results and discussion 3.1. Experimental test of analytical result for phase-change material melting time Images of the melting phase-change material in a commercial heat exchanger with a 1.3-mm fin pitch are shown in Fig. 6. The time for the stearic acid to completely melt was 192 s. A comparison of the total amount of melted material in the analytical and laboratory tests is plotted in Fig. 7. The difference in the total melting time between the analytical and experimental results is 22.4%. In the experiment, a melted
layer was not observed for the first 12 s. In the numerical analysis, the heat loss to the outside was not considered; in the experiment, insulation was not included to enable visualization. For the experiment, the observation window was installed on the face of the heat exchanger parallel with the ground in order to prevent leakage of stearic acid. Furthermore, cooling water heat was lost to the outside through the aluminum case, which was welded directly to the heat exchanger case without insulation. All of these factors may have contributed to the slower melting in the experiment relative to the analysis. In the numerical analysis, the curve of the melted amount increases smoothly with time, while experimental data increase
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Fig. 8. Analytical cooling water warm-up results for a 1.6 L diesel passenger car in an NEDC driving cycle.
whereas most energy transferred to the stearic acid. In a practical heat exchanger, energy loss resulting from heat transfer to the outer case rather than to fins and tubes would be more significant. 3.2. Analysis of fuel consumption for an NEDC driving cycle
Fig. 9. Comparison of fuel consumption NEDC driving cycle between cold start and warm start.
more slowly in the initial stage and more rapidly later. It is apparent that, after some time, the temperature in the aluminum case increased past the stearic acid melting point, but only an insignificant amount of energy transferred to the aluminum,
Cooling water warm-up results for a vehicle capacity of 1.6 L are shown in Fig. 8. Since the cooling water temperature increases with the acceleration of the vehicle, the warm-up rate was not constant. The rise in temperature was also disturbed due to mixing inside the radiator from 81 °C when the thermostat started opening to 86 °C when it had completely opened. It took 566 s, or 48% of the full 1180-s cycle, for the cooling water temperature to climb from 25 °C to 70 °C, the temperature at which excess fuel spray during cold start shuts off. Since cooling water is also supplied to the engine oil cooler, the oil temperature also increases with the cooling water temperature. However, due to the relatively high specific heat of the engine oil, fast variations in the temperature curve with sudden changes in engine load were not observed. The optimum engine oil temperature to reduce friction in each component of the engine was between 95 and 100 °C, and the cool-
Fig. 10. Coolant warm-up time comparison for NEDC driving cycles.
S. Park et al. / Applied Thermal Engineering 113 (2017) 1496–1504
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Fig. 11. Engine oil warm-up time comparison for NEDC driving cycles.
Fig. 12. Fuel consumption of NEDC driving cycle for different heat storage temperatures.
ing water temperature that resulted in the best fuel consumption was also within the same temperature range [10]. However, in the NEDC test, only a very small part of the drive time satisfied the above temperature range. During warm starting, the cooling water temperature is maintained at 93–100 °C, and the engine oil temperature is maintained at 100–110 °C. Fig. 9 shows a 7.28% reduction in the rate of fuel consumption to 41.94 g/km, compared to 45 g/km during cold starting. If the warm-up time to 70 °C is reduced with a heat storage device so that the cooling water and engine oil reach the optimum temperature more quickly, fuel consumption should decrease. As shown in Fig. 10, the time to reach a cooling water temperature of 70 °C decreased by 40.5% from 566 s to 337 s when 95 °C heat storage was applied and by 35.2–367 s when 70 °C heat storage was applied. The time to reach the optimum cooling water temperature of 95 °C decreased by around 3.1% from 1097 s to 1036 s when 95 °C heat storage was applied and by 2.6–1068 s when 70 °C heat storage was applied. A similar pattern was observed in the actual vehicle test, i.e., the time delay decreased as the temperature of heat storage increased. In Fig. 8, the rate of increase of the cooling water temperature sud-
denly dropped after reaching 81 °C as the thermostat started operating. The full amount of cooling water stored from the end of the thermostat to the radiator was around 2.5 L. When a heat storage device was not installed, the deviation from the original temperature rise was very small, because heating increased after finishing the drive in the city center. However, when a heat storage device was installed, less time was required to reach 81 °C, and the thermostat opened while still in the city center drive zone, causing a significant delay in the cooling water temperature increase. As shown in Fig. 11, the time to reach 95 °C decreased by 6.6%, from 1084 to 1012 s when 95 °C heat storage was applied and by 4.8% from 1084 to 1032 s when 70 °C heat storage was applied. As expected, the time to reach the optimum temperature for engine oil and cooling water was reduced with heat storage, improving fuel consumption. The fuel consumption during driving changed in a similar manner, as shown in Fig. 12. In the zone where the cooling water temperature increased rapidly, the fuel consumption decreased; in the zone where the cooling water temperature changed more slowly, the fuel consumption was nearly constant. The fuel consumption decreased by 2.71% from 45 g/km to 43.9 g/km when 95 °C heat storage was applied and by 2.45% to 43.78 g/km when 70 °C heat storage was applied. Fuel economy was 19.58 km/L without a heat storage device installed, 20.12 km/L with heat storage at 95 °C, and 20.07 km/L with heat storage at 70 °C. In the above two cases, the fuel economy calculated during a 100-km drive in the city center and expressway was around 20.56 km/L with a heat storage device installed and 20.29 km/L when a heat storage device was not installed, a 1.3% improvement. 4. Conclusion The goal of this study was to apply a numerical model of a heat storage device using a phase-change material to a conventional diesel engine and determine the improvement in fuel consumption during driving. An NEDC drive test analysis for an engine capacity of 1.6 L shows that the time to reach a cooling water temperature of 70 °C, where the mode changes from cold to warm, was reduced by 40.5% with a 95 °C heat storage device and by 35.2% with a 70 °C heat storage device. Furthermore, the time to reach 95 °C, which is the engine oil temperature for optimum fuel consump-
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S. Park et al. / Applied Thermal Engineering 113 (2017) 1496–1504
tion, was reduced by 6.6% with a 95 °C heat storage device and by 4.8% with a 70 °C heat storage device. In the NEDC drive test, fuel consumption was reduced in the zone where the overall cooling water temperature was high, while the difference in fuel consumption was small in the zone where variation in the cooling water temperature was small. With a 95 °C heat storage device, fuel consumption was reduced by 2.71%; with a 70 °C heat storage device, it was reduced by 2.45%. The fuel economy without a heat storage device installed was 19.58 km/L, while the fuel economy was 20.12 km/L with a heat storage device at 95 °C, and 20.07 km/L at 70 °C. However, most of the improvement occurred during initial warm up; therefore, if the driving distance or time is long, the percentage increase in fuel economy might be less.
Acknowledgment This work was supported by Mid-career Researcher Program through a National Research Foundation of Korea (NRF) grant funded by the Korea government (MSIP). No. NRF2014R1A2A2A01005055.
References [1] M. Hatami, D. Ganji, M. Gorji-Bandpy, A review of different heat exchangers designs for increasing the diesel exhaust waste heat recovery, Renew. Sustain. Energy Rev. 37 (2014) 168–181. [2] A. Patil, L. Navale, V. Patil, Simulative analysis of single cylinder four stroke C.I. engine exhaust system, Int. J. Sci. Spirituality Bus Technol. (IJSSBT) 2 (2013) 2277–7261. [3] S. Tiari, S. Qiu, M. Mahdavi, Numerical study of finned heat pipe-assisted thermal energy storage system with high temperature phase change material, Energy Con. Man. 89 (2015) 833–842. [4] F. Agyenim, N. Hewitt, P. Eames, M. Smyth, A review of materials, heat transfer and phase change problem formulation for latent heat thermal energy storage systems (LHTESS), Renew. Sustain. Energy Rev. 14 (2010) 615–628. [5] A. Sharma, V. Tyagi, C. Chen, D. Buddhi, Review on thermal energy storage with phase change materials and applications, Renew. Sustain. Energy Rev. 13 (2009) 318–345. [6] A. Yunus, Heat Transfer: A Practical Approach, McGraw-Hill International Editions, 2012. [7] F. Incropera, D. Dewitt, Fundamentals of Heat and Mass Transfer, fourth ed., John Wiley & Sons Inc., 1996. [8] P. Verma, Varun, S. Singal, Review of mathematical modeling on latent heat thermal energy storage systems using phase-change material, Renew. Sustain. Energy Rev. 12 (2008) 999–1031. [9] LMS Imagine, Lab Amesim 14.1 User’s Guide. Siemens Inc., Belgium, 2014. [10] K. Hyungik, S. Jungwook, L. Kihyung, A study of fuel economy and exhaust emission according to engine coolant and oil temperature, J. Therm. Sci. Technol. 8 (2013) 255–268.
Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
Numerical model and simulation of a vehicular heat storage system with phase-change material Sangki Park a, Seungchul Woo a, Jungwook Shon b, Kihyung Lee a,⇑ a b
Department of Mechanical Engineering, Hanyang University, 1271 Sa1-dong, Sangrok-gu, Gyeonggi-do, Republic of Korea Powertrain HEX Team, Hanon Systems, 95, Sinilseo-ro, Daedeok-gu, Daejeon, Republic of Korea
h i g h l i g h t s Numerical analysis and evaluation of a heat storage system with a phase-change material. Numerical simulation of the New European Driving Cycle was carried out for a vehicle. The warm-up time was reduced by 40.5% and the fuel consumption was reduced by 2.71%.
a r t i c l e
i n f o
Article history: Received 3 August 2016 Revised 14 October 2016 Accepted 22 November 2016 Available online 23 November 2016 Keywords: Phase-change material Heat storage system One-dimensional numerical analysis Diesel engine
a b s t r a c t For heat storage applications designed to recover and recycle waste heat energy, it is usually advantageous to store heat in a phase-change material. One-dimensional numerical analysis and evaluation of a heat storage system that uses a phase-change material to store latent heat in addition to sensible heat was carried out, and it was found that up to 30% of the total heat energy generated by the fuel and subsequently lost to cooling can be recovered. A heat storage system was installed to reduce warm-up time by releasing heat directly into the engine coolant during cold start, and the corresponding reduction in fuel consumption was measured. With the addition of a heat storage device, the warm-up time to 95 °C was reduced by between 18.1% and 27.1%. A numerical simulation of the New European Driving Cycle was carried out for a vehicle equipped with a 1.6-L diesel engine and a heat storage system. Analytical results showed that the warm-up time was reduced by 40.5% and the fuel consumption was reduced by 2.71% compared to a vehicle without a heat storage system installed. Ó 2016 Elsevier Ltd. All rights reserved.
1. Introduction Modern vehicle engines are designed to consume unnecessarily large amounts of fuel in order to warm up cooling water during initial cold start, which increases emissions and lowers fuel economy. In order to resolve this issue, alternative warming methods have been proposed, such as heating and storing cooling water with an electric heater and then resupplying to the engine during ignition or installing a separate heat exchanger in the exhaust pipe to capture waste heat [1,2]. The major components of a heat storage system can be broadly categorized into a phase-change material, a heat exchanger, and an insulation vessel. The most important criteria to evaluate their performance are their heat storage and heat radiation capabilities. The heat exchanger, which transfers heat energy into and out of the ⇑ Corresponding author at: Department of Mechanical Engineering, Hanyang University, 1271 Sa1-dong, Sangrok-gu, Gyeonggi-do 426-791, Republic of Korea. E-mail address: [email protected] (K. Lee). http://dx.doi.org/10.1016/j.applthermaleng.2016.11.162 1359-4311/Ó 2016 Elsevier Ltd. All rights reserved.
phase-change material, determines the speed of heat storage and radiation. The insulation vessel preserves stored energy until the moment when it is needed and plays an important role [3]. Phase-change materials take advantage of latent heat in a phase transition to store energy for a given temperature difference. Storing energy only as sensible heat results in a dramatically lower heat capacity. For example, the specific heat of paraffin wax is 2.1 kJ/kg °C, and its latent heat is 140 kJ/kg. Compared with an equivalent mass of water, which has a specific heat of 4.2 kJ/kg °C, if the temperature increases by 15 °C, the additional heat stored in paraffin wax after melting and reaching equilibrium is 2.7 times greater [4]. When a solid phase-change material is heated to its melting point, its phase changes to liquid. During this process, latent heat (also known as enthalpy of fusion) is absorbed, and the temperature of the material remains constant, as shown in Fig. 1. Likewise, in the reverse situation when the phase changes from liquid to solid at the freezing point, latent heat is discharged with no change
S. Park et al. / Applied Thermal Engineering 113 (2017) 1496–1504
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Fig. 1. Energy charge characteristics in a phase-change material [5].
in temperature. The same phenomenon occurs in the liquid-gas phase transition. In this study, one-dimensional numerical analysis and evaluation were carried out in order to design a heat storage system that uses a phase-change material to recover around 30% of the total heat energy produced by fuel and subsequently lost to engine cooling.
2. Modeling and simulation 2.1. Analysis of a commercial heat exchanger Heat transfer is a phenomenon where heat energy flows from high temperature to low temperature. The movement of heat can be categorized as conduction, convection, radiation, or any combination of the three. Furthermore, if substances are displaced or experience phase changes, the analysis can get quite complicated. However, analytical results similar to those obtained experimentally can often be obtained even for complicated heat transfer processes if appropriate models and empirical equations are used. Heat transfer processes are implemented in various commercially-available software programs such as Amesim by LMS, which is based on Modelica. A commercial heat exchanger of the inner tubes and outer fins was applied in this study. The side that makes contact with air has louver-fins, which are designed to increase the rate of heat exchange by creating turbulence. However, in our case, the phase-change material makes direct contact with the fins and tubes without air flow, so the heat exchanger can be modeled more simply with an equivalent flat fin design. The total heat transfer coefficient describing the transfer of heat from the internal fluid to the outside through the heat exchanger can be expressed as a sum of heat resistances. If the fins and tubes are brazed, the total thermal resistance is the sum of the convective thermal resistance inside the tubes and the conductive thermal resistance of the fins and tubes. However, the rate of heat
transfer between the heat exchanger and the phase-change material is different during melting and freezing [6]. Phase-change material was added to a commercial heat exchanger to obtain preliminary data in order to fabricate a heat exchanger for a heat storage device, and the melting time was analyzed. The optimum fin pitch, temperature, and flow rate were determined by systematically varying each parameter. A cooling water solution containing 50% ethylene glycol was used for analysis. The physical properties of the cooling water are presented in Table 1. In the commercial heat exchanger, the distance between fins was 6 mm, the length of fins was 41.5 mm, the fin thickness was 0.1 mm, and fins were arranged in 17 rows of 139 fins. The heat exchanger also contained a total of 64 tubes with a hydraulic diameter of 2.418 mm. It was constructed as a U-band type, which is the same as a heat exchanger having 16 tubes and length 390 mm. The total mass of phase change material was 0.714 kg, and the total heat transfer area was 1.385 m2. A heat storage device is expected to be most efficient when installed in the heater core line water, which is directly supplied to the engine. The flow rate of heater core water for a 2000 cc capacity engine is shown in Table 2. Since the number of revolutions per minute (rpm) of vehicles during cold starting usually falls within the range of 1000–2000, three speeds were studied: 1000, 1500, and 2000 rpm. Around 3 L of cooling water is stored in commercially available sensible heat storage devices; therefore, three commercial heat exchangers must be connected in parallel to store 3 kg of phasechange material. However, the same analysis can be performed by reducing the amount of phase change material and the flow rate of cooling water by 1/3 rather than connecting heat exchangers in parallel. In order to determine the convective heat transfer coefficient inside the heat exchanger, the hydraulic diameter Dh and flow speed V flow need to be calculated with Eqs. (1) and (2), respectively:
Dh ¼ 4
Afree ðNon-circular cross sectionÞ Pcontact
ð1Þ
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Table 1 Properties of the 50% ethylene glycol aqueous solution. Temperature (°C)
Density (kg/m3)
Specific heat (kJ/kg °C)
Heat conductivity (W/m °C)
Dynamic viscosity (kg/m s)
75 85 95
1042.79 1035.50 1027.79
3.495 3.535 3.574
0.408 0.411 0.413
0.001 0.001 0.001
Table 2 Heater core coolant flow rate for each engine speed. Speed (rpm)
Flow rate (L/min)
1000 1500 2000
8.6 12 15.5
V flow ¼
Q flow Afree
ð2Þ
where Afree is the area of free flow, Pcontact is the length of tubes in contact with fluid, and Q flow is the flow rate. Since cooling water moves through tubes in the form of a liquid, it is a single-phase flow. The Prandtl number N Pr and Reynolds number N Re are calculated with Eqs. (3) and (4), respectively:
NPr ¼
Cp l K
ð3Þ
NRe ¼
q V flow Dh l
ð4Þ
where C p is the specific heat, l is the dynamic viscosity coefficient, K is the thermal conductivity, q is the density, and l is the dynamic viscosity coefficient. The flow of cooling water inside tubes is turbulent, so the convective heat transfer coefficient is calculated with Eq. (5), which is based on the Dittus-Boelter equation: n NNu ¼ 0:023 N0:8 Re N Pr ¼
h Dh K
ð5Þ
n ¼ 0:4 : Internal temperature is higher than external temperature
is no contact heat resistance between them, so the conductive heat resistance is calculated with Eq. (7). The specific length of conduction through the tubes is the tube wall thickness, and the specific length of conduction through fins is their length. Since conductive heat resistance through a heat exchanger is determined by its physical shape, identical heat exchangers have the same conductive heat resistance.
RConv ection ¼
1 h Atotal
ð6Þ
where h is the convection heat transfer coefficient, and Atotal is the total heat transfer area.
RConduction ¼
Lc K Atotal
ð7Þ
where Lc is the characteristic length, K is the thermal conductivity coefficient, and Atotal is the total heat transfer area. Fins are in contact with tubes at both ends, and phase-change material does not flow, so it is assumed that the heat transfer efficiency is 100%, and heat is transferred by conduction only. Since tubes and fins transfer heat simultaneously, their resistances are added in parallel. The sum of the conductive heat resistances in the heat exchanger is therefore calculated as in Eq. (8).
Rtotal; conduction ¼
1 Rtube
þ
1
1
Rtube
ð8Þ
The total heat resistance is obtained by adding the convective heat resistance and conductive heat resistance of tubes and fins using Eq. (9). Once the total heat transfer is obtained by the above method, the only factor affecting the heat transfer to the solid material is the flow rate of cooling water, provided that the same heat exchanger is used [7].
n ¼ 0:3 : External temperature is higher than internal temperature
Rtotal ¼ Rconv ection ðCoolantÞ þ Rconduction ðHeat exchangerÞ
where N Nu is the Nusselt number, NRe is the Reynolds number, N Pr is the Prandtl number, h is the convection heat transfer coefficient, Dh is the hydraulic diameter, and K is the thermal conductivity. In order to add the convective and the conductive heat transfer coefficients in the form of heat resistances, the convective heat resistance should be calculated with Eq. (6). Fins and tubes of the heat exchanger are all made of AL1100 aluminum alloy, and there
After melting has started, heat transfer by natural convection in the liquid layer is added to the above resistance. Because the liquid layer develops as time elapses, an equation describing the variation in heat resistance with time is required [8]. Two of the six faces of the container encapsulating the phasechange material do not receive heat directly by contact to the outer case, so the space can be imagined as one-dimensional with a rect-
Fig. 2. Schematic diagram showing the increase in thickness of the melted layer of phase-change material with time.
ð9Þ
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Fig. 3. Visual measurement of phase-change material melting time.
angular cross section in contact with the fins and tubes. If the rate of phase change per hour is assumed to be constant, the amount of liquid can be found by calculating DS with a x2-x1 increase in the liquid layer thickness with time, as shown in Fig. 2. The increase in the convection heat resistance RC v L with time due to the increasing thickness of the liquid layer is given by Eq. (10). If it is assumed that heat energy in the cooling water is transferred to the phase-change material without loss as the phasechange material melts, the quantity of heat in the coolant DQ c can be expressed by Eq. (11), where the surface area of melted phase-change material DSarea is calculated using Eq. (12). Once the surface area is obtained, a x2-x1 increase in the thickness of
the melted layer can be calculated, from which the total amount of melted phase-change material can be determined.
RC v L ¼
1 Dtime hLP DSarea
ð10Þ
DQ c ¼
DT coolant RC v C þ RC d H þ RC v L
ð11Þ
DSarea ¼
hLP
Fig. 4. New European Driving Cycle.
DTime DT
DQ coolant
ðRC v C þ RC d H Þ
ð12Þ
1500
S. Park et al. / Applied Thermal Engineering 113 (2017) 1496–1504
Fig. 5. Vehicle model and heat storage control model used for analysis.
where RCv C is the heat resistance of coolant convection, RC d H is the heat resistance of heat exchanger conduction, hLP is the convection heat transfer coefficient, and DT is the temperature difference. The above equation was programmed into one-dimensional commercial Amesim software [9]. First, the melting time was calculated as a function of the flow rate, with the cooling water temperature held constant at 95 °C. Then, the melting time was calculated for a constant flow rate with the cooling water temperature set to 75 °C, 85 °C, or 95 °C. 2.2. Validation test of the analytical result using a real heat exchanger When the heat exchanger and phase-change material are located inside a sealed container, the amount of melted phasechange material cannot be measured directly. Therefore, to determine the actual development time of the liquid layer, a device with a fin pitch of 1.3 mm was used to heat the phase-change material evenly as in a commercial heat exchanger, but a window was installed to visualize the melting layer as shown in Fig. 3. Photographs of the liquid layer between the fins and tubes were taken at one frame per second using a digital camera. The thickness of the liquid layer in each frame was then measured using an image tool, and the total amount of melted material in the heat exchanger was calculated. For this experiment, cooling water was supplied at 95 °C with a flow rate of 5.2 L/min. 2.3. Analysis of fuel consumption in drive mode using simulation In order to investigate the improvement in fuel consumption when a heat storage device is installed, a chassis dynamometer should be used. However, since a dynamometer was not installed in the actual vehicle simulation bench, the improvement in fuel consumption in driving mode cannot be known experimentally and was instead modeled analytically. The New European Driving Cycle (NEDC) was adopted for the drive mode analysis, which is the standard test used to evaluate fuel consumption and exhaust material in Europe. In NEDC, the
first 780 s simulate driving in a city center, while the next 400 s simulate driving on national highways and express ways, as shown in Fig. 4. The experiment should be conducted when the temperature of the vehicle is stabilized at 25 °C, and the chassis dynamometer should account for the weight of the vehicle and the effect of air resistance. The cooling system and detailed specifications of a vehicle with a commercial 1.6 L capacity diesel engine, including appropriate gear ratios and transmission specifications, were used to construct an Amesim model for analysis, as shown in Fig. 5(a). The vehicle specifications are given in Table 3. For the tests, 8.1 L of cooling water was contained in the system including the heat storage device. Fuel consumption and cooling water temperature with and without a heat storage device installed during cold starting were compared in order to measure the effect of the heat storage device on the overall fuel consumption. A control algorithm for the heat storage system was implemented as shown in Fig. 5(b) to avoid overcooling the cooling water, and the time required for re-storing heat was measured.
Table 3 Specifications of the test vehicle for analysis. Type
Diesel engine (EURO5)
Number of cylinders Displacement (cc) Bore Stroke (mm) Compression ratio Max power (ps/rpm) Max torque (kg m/rpm) Vehicle weight (kg) Transmission gear ratio 1st 2nd 3rd 4th 5th 6th Reverse
4 1582 77.2 84.5 17.3:1 128/4000 26.5/1900–2750 1360 6 speed 4.212 2.637 1.800 1.386 1.000 0.772 3.385
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Fig. 6. Photographs showing increasing stearic acid melted layer thickness with time.
Fig. 7. Comparison of analytical and experimental stearic acid melting results.
3. Results and discussion 3.1. Experimental test of analytical result for phase-change material melting time Images of the melting phase-change material in a commercial heat exchanger with a 1.3-mm fin pitch are shown in Fig. 6. The time for the stearic acid to completely melt was 192 s. A comparison of the total amount of melted material in the analytical and laboratory tests is plotted in Fig. 7. The difference in the total melting time between the analytical and experimental results is 22.4%. In the experiment, a melted
layer was not observed for the first 12 s. In the numerical analysis, the heat loss to the outside was not considered; in the experiment, insulation was not included to enable visualization. For the experiment, the observation window was installed on the face of the heat exchanger parallel with the ground in order to prevent leakage of stearic acid. Furthermore, cooling water heat was lost to the outside through the aluminum case, which was welded directly to the heat exchanger case without insulation. All of these factors may have contributed to the slower melting in the experiment relative to the analysis. In the numerical analysis, the curve of the melted amount increases smoothly with time, while experimental data increase
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Fig. 8. Analytical cooling water warm-up results for a 1.6 L diesel passenger car in an NEDC driving cycle.
whereas most energy transferred to the stearic acid. In a practical heat exchanger, energy loss resulting from heat transfer to the outer case rather than to fins and tubes would be more significant. 3.2. Analysis of fuel consumption for an NEDC driving cycle
Fig. 9. Comparison of fuel consumption NEDC driving cycle between cold start and warm start.
more slowly in the initial stage and more rapidly later. It is apparent that, after some time, the temperature in the aluminum case increased past the stearic acid melting point, but only an insignificant amount of energy transferred to the aluminum,
Cooling water warm-up results for a vehicle capacity of 1.6 L are shown in Fig. 8. Since the cooling water temperature increases with the acceleration of the vehicle, the warm-up rate was not constant. The rise in temperature was also disturbed due to mixing inside the radiator from 81 °C when the thermostat started opening to 86 °C when it had completely opened. It took 566 s, or 48% of the full 1180-s cycle, for the cooling water temperature to climb from 25 °C to 70 °C, the temperature at which excess fuel spray during cold start shuts off. Since cooling water is also supplied to the engine oil cooler, the oil temperature also increases with the cooling water temperature. However, due to the relatively high specific heat of the engine oil, fast variations in the temperature curve with sudden changes in engine load were not observed. The optimum engine oil temperature to reduce friction in each component of the engine was between 95 and 100 °C, and the cool-
Fig. 10. Coolant warm-up time comparison for NEDC driving cycles.
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Fig. 11. Engine oil warm-up time comparison for NEDC driving cycles.
Fig. 12. Fuel consumption of NEDC driving cycle for different heat storage temperatures.
ing water temperature that resulted in the best fuel consumption was also within the same temperature range [10]. However, in the NEDC test, only a very small part of the drive time satisfied the above temperature range. During warm starting, the cooling water temperature is maintained at 93–100 °C, and the engine oil temperature is maintained at 100–110 °C. Fig. 9 shows a 7.28% reduction in the rate of fuel consumption to 41.94 g/km, compared to 45 g/km during cold starting. If the warm-up time to 70 °C is reduced with a heat storage device so that the cooling water and engine oil reach the optimum temperature more quickly, fuel consumption should decrease. As shown in Fig. 10, the time to reach a cooling water temperature of 70 °C decreased by 40.5% from 566 s to 337 s when 95 °C heat storage was applied and by 35.2–367 s when 70 °C heat storage was applied. The time to reach the optimum cooling water temperature of 95 °C decreased by around 3.1% from 1097 s to 1036 s when 95 °C heat storage was applied and by 2.6–1068 s when 70 °C heat storage was applied. A similar pattern was observed in the actual vehicle test, i.e., the time delay decreased as the temperature of heat storage increased. In Fig. 8, the rate of increase of the cooling water temperature sud-
denly dropped after reaching 81 °C as the thermostat started operating. The full amount of cooling water stored from the end of the thermostat to the radiator was around 2.5 L. When a heat storage device was not installed, the deviation from the original temperature rise was very small, because heating increased after finishing the drive in the city center. However, when a heat storage device was installed, less time was required to reach 81 °C, and the thermostat opened while still in the city center drive zone, causing a significant delay in the cooling water temperature increase. As shown in Fig. 11, the time to reach 95 °C decreased by 6.6%, from 1084 to 1012 s when 95 °C heat storage was applied and by 4.8% from 1084 to 1032 s when 70 °C heat storage was applied. As expected, the time to reach the optimum temperature for engine oil and cooling water was reduced with heat storage, improving fuel consumption. The fuel consumption during driving changed in a similar manner, as shown in Fig. 12. In the zone where the cooling water temperature increased rapidly, the fuel consumption decreased; in the zone where the cooling water temperature changed more slowly, the fuel consumption was nearly constant. The fuel consumption decreased by 2.71% from 45 g/km to 43.9 g/km when 95 °C heat storage was applied and by 2.45% to 43.78 g/km when 70 °C heat storage was applied. Fuel economy was 19.58 km/L without a heat storage device installed, 20.12 km/L with heat storage at 95 °C, and 20.07 km/L with heat storage at 70 °C. In the above two cases, the fuel economy calculated during a 100-km drive in the city center and expressway was around 20.56 km/L with a heat storage device installed and 20.29 km/L when a heat storage device was not installed, a 1.3% improvement. 4. Conclusion The goal of this study was to apply a numerical model of a heat storage device using a phase-change material to a conventional diesel engine and determine the improvement in fuel consumption during driving. An NEDC drive test analysis for an engine capacity of 1.6 L shows that the time to reach a cooling water temperature of 70 °C, where the mode changes from cold to warm, was reduced by 40.5% with a 95 °C heat storage device and by 35.2% with a 70 °C heat storage device. Furthermore, the time to reach 95 °C, which is the engine oil temperature for optimum fuel consump-
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tion, was reduced by 6.6% with a 95 °C heat storage device and by 4.8% with a 70 °C heat storage device. In the NEDC drive test, fuel consumption was reduced in the zone where the overall cooling water temperature was high, while the difference in fuel consumption was small in the zone where variation in the cooling water temperature was small. With a 95 °C heat storage device, fuel consumption was reduced by 2.71%; with a 70 °C heat storage device, it was reduced by 2.45%. The fuel economy without a heat storage device installed was 19.58 km/L, while the fuel economy was 20.12 km/L with a heat storage device at 95 °C, and 20.07 km/L at 70 °C. However, most of the improvement occurred during initial warm up; therefore, if the driving distance or time is long, the percentage increase in fuel economy might be less.
Acknowledgment This work was supported by Mid-career Researcher Program through a National Research Foundation of Korea (NRF) grant funded by the Korea government (MSIP). No. NRF2014R1A2A2A01005055.
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