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Evaporator performance enhancement by pulsation width modulation (PWM)
Applied Thermal Engineering 99 (2016) 825–833
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Applied Thermal Engineering j o u r n a l h o m e p a g e : w w w. e l s e v i e r. c o m / l o c a t e / a p t h e r m e n g
Research Paper
Evaporator performance enhancement by pulsation width modulation (PWM) X. Wang a,*, K. Tang b, P.S. Hrnjak a,c a b c
Mechanical Science and Engineering, University of Illinois, Urbana, IL 61801, USA Institute of Refrigeration and Cryogenics, Zhejiang University, Hangzhou 310027, China CTS, Urbana, IL 61802, USA
H I G H L I G H T S
• • • • •
Effects of pulsating flow on heat transfer performance and pressure drop in an air-refrigerant evaporator have been explored. Pulsating flow has enhanced overall and refrigerant-side heat transfer coefficient by 27% and 123% respectively, compared to continuous flow. Pulsation period has a big effect on heat transfer performance; a larger enhancement at a shorter period was observed. The temporal pressure drop for pulsating flow changes periodically following the pulsation period. Pulsation effect on pressure drop becomes weaker at a larger mass flux and a shorter pulsation period.
A R T I C L E
I N F O
Article history: Received 8 November 2014 Accepted 17 December 2015 Available online Keywords: Pulsating flow Pulsation width modulation Evaporator control Heat transfer coefficient Two-phase flow regime
A B S T R A C T
In this work, the effects of pulsation width on heat transfer coefficient and pressure drop in an air-to-refrigerant evaporator is studied experimentally. The pulsation width can be controlled to a minimum period of 2 s. A general model is also developed to predict heat transfer performance and pressure drop. The overall heat transfer coefficient, refrigerant side heat transfer coefficient, and pressure drop are measured and compared to a baseline without pulsating flow. The results show that the overall and refrigerant-side heat transfer coefficients have improved about 27% and 123%, respectively, with pulsating flow. The refrigerant mass flux and pulsation period have an important effect on the heat transfer enhancement. The average pressure drop of the pulsating flow differs little from that of continuous flow, but the temporal pressure drop of pulsating flow is quite different and varies with mass flux and pulsation period. © 2016 Elsevier Ltd. All rights reserved.
1. Introduction Flow pulsation, sometimes referred to as mechanical excitation or fluid-borne oscillation, is well known to affect heat transfer behavior. For example, if a single-phase convective heat transfer condition is subjected to pulsating flow, the pulsations reduce the timeaveraged thickness of the boundary layer and hence the thermal resistance, which improves heat transfer performance. If flow boiling is subjected to pulsating flow, the pressure change during the pulsating flow might lead to a local cavitation effect, which will also affect heat transfer performance. There has been significant previous research on pulsating flow and heat transfer, and some research shows a significant enhancement of heat transfer performance by using pulsating flow, while some shows little or no effect. In a turbulent flow, Martinelli et al. [1] found no difference in heat transfer
* Corresponding author. Tel.: +1 217 693 2358; fax: +1 217 244 6534. E-mail address: [email protected] (X. Wang). http://dx.doi.org/10.1016/j.applthermaleng.2015.12.049 1359-4311/© 2016 Elsevier Ltd. All rights reserved.
between steady continuous flow and pulsating flow; however, West and Taylor found an increase of 60–70% [2]. In a laminar flow, Morris [3] and Webb [4] found no change in heat transfer. Gao and Zeng [5] found that with a self-oscillator nozzle in a system, heat transfer performance increased about 10–30%. In their later work [6], they found that the geometric structure of the self-oscillator nozzle can affect the pulsating flow and also the heat transfer coefficient. The pressure increase caused by the self-oscillator nozzle was found negligible when compared to the improvement of heat transfer coefficient. Zohir [7] conducted experiments to compare parallel flow and counter flow when pulsing flow was generated, and he found that pulsing flow can increase heat transfer coefficient by up to about 20% in parallel flow and up to about 90% in counter flow. In Zohir’s recent work [8], he found that heat transfer performance was improved with pulsing flow, and the enhancement reached a peak value when the flow regime was in transition between laminar and turbulent flow. For all pulsation frequencies, the highest enhancement was about 10 times for single-phase counter flow and about 8 times for parallel flow. Guo et al. [9] investigated the pulsating flow effect
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for a pipe partially filled with a porous medium using numerical methods, and they found an increase in effective thermal conductivity along the tube length direction and an increase in convective heat transfer coefficient. Their results show that the maximum heat transfer enhancement was achieved by optimizing the pulse frequency. Recently, Rahgoshay et al. [10] conducted research on nanofluids in a pulsating flow using numerical method, and the result shows that the heat transfer coefficient increased slightly with increasing frequency and amplitude. Bose [11] found that a pulsing flow can increase heat transfer or decrease heat transfer, depending on the pulsating frequency and amplitude. Apparently, there exists a critical frequency which is a function of wave-form and Reynolds number; however, there was only a decrease in heat transfer performance if the pulse amplitude was very low. Most prior research on pulsed-flow heat transfer used water, oil and water mixtures, or air as working fluids, and there are contradictions as to the effect of pulsating flow on heat transfer performance. To date there has been no research on two-phase refrigerant pulsating flows. Moreover, the effects of pulsation are more complex for pulsations within a practical evaporator, for which the use of a multi-tube construction and the presence of tube bends at the end of each tube certainly affect the resulting flow. However, it is anticipated that pulse width modulation might be effective for evaporator performance enhancement. In this work, refrigerant R134a is used as the working fluid, and a pulsating flow generated by solenoid-valve system in a real evaporator. A general model is also built to predict heat transfer performance while pulsating flow occurs in the evaporator.
2. Experimental approach and facilities When a flow pulsates with a prescribed period, in this case controlled by a solenoid valve, the mass flow rate and other parameters are expected to manifest the same periodicity. In these experiments the mass flow rate in the evaporator ideally varies periodically as shown in Fig. 1. The pulsating flow period is τ, and the on-time ratio is μ, which is the ratio of on-time to the duration of the entire period. During one period, the mass flow rate is mr during the ontime of μτ and 0 for the off-time of (1 − μ)τ. The time-averaged mass flow rate is equal to μmr, which will be considered as the continuousflow baseline for purposes of comparison. If μ = 0.5, the on-time is equal to the off-time in a period, and during the on-time the mass flow rate for pulsating flow is twice that of the continuous flow baseline. For experimental control and operation, two identical loops were created, with two identical evaporators. Downstream from both evaporators, refrigerant flowed into a common flow-conditioning system, consisting of a condenser, a pump, and a preheater. Just upstream of the evaporators, the flow split and the refrigerant flow was directed into one evaporator or another, as determined by two
Fig. 1. Pulsating flow mass flow rate vs time.
Fig. 2. Schematic drawing of the system.
solenoid valves. When one solenoid valve was open the other was close, and vice versa (see Fig. 2). In this way, the flow through the shared portion of the loops was steady, but the flow through the evaporators was pulsed, and when pulsation occurred, the pressure upstream of the solenoid valves was essentially constant. So during the experiment, the two loops were operated alternately; there was always one loop open and one loop closed. Referring again to Fig. 1, when HX1 experiences on-time, HX2 experiences off time, and when HX1 is off, HX2 is on. The two loops are never closed at the same time. This operating procedure ensured a steady state inlet pressure and a continuous mass flow rate, mr. The experiment apparatus is designed to measure heat transfer coefficient and pressure drop for pulsating and continuous flow, and in addition to the two refrigerant loops, there is a chilledwater loop operated at steady state to cool the condenser (see Fig. 2). The refrigerant is circulated by a gear pump, absorbing heat from the air through evaporators and releasing heat to the chilled water through a condenser. As described, the refrigerant loops have two parallel, identical subloops, each with one evaporator (HX1 or HX2). The air flow rate and inlet temperature to each evaporator are equal for both evaporators. 2.1. Refrigerant loop In Fig. 2, the red loop represents refrigerant, which is pumped from a subcooler by a gear pump with mass flow rate measured by a flow meter, to a controlled heater which can adjust the refrigerant flow to a desired vapor quality. After that, the refrigerant loop is divided into two identical subloops, subloop1 and subloop2. Each subloop includes a solenoid valve, a heat exchanger, a power controllable heater, and a sight glass. All elements of the two subloops are identical as shown in Fig. 2. After the refrigerant passes through a sight glass, these two subloops are combined and connected to a common loop, then the refrigerant flows through another flow meter to a condenser, a receiver and the subcooler, the pump, and finally back to the point where the lines split to complete the loop. The refrigerant mass flow rate can be adjusted by adjusting the pump speed controller and also by adjusting the valve in the bypass, the flow rate was measured by a Coriolis-effect flow meter before it was divided into two subloops. The vapor quality of the refrigerant
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flow was set to a certain value by heating the subcooled refrigerant using a heater, controlled by adjusting the input voltage. The input power, refrigerant temperature before the heater, and the refrigerant pressure after the heater are measured for vapor quality control. Pulsation frequency is controlled by controlling the solenoid valve on and off periods; the period can be set from 0.1 s to 60 s. The system works such that when one solenoid valve in one subloop is on, the other solenoid valve in the other loop is off automatically. Temperatures of refrigerant before and after each heat exchanger in each loop are measured and differential pressures across each heat exchanger are also measured. After the refrigerant leaves the heat exchanger, it is heated to a prescribed superheat with a heater controlled in the same way as the heater upstream of the heat exchanger. Flow was observed in the sight glass to ensure that no liquid exists in the flow before the two subloop refrigerant flows are combined and pass through the downstream flow meter. 2.2. Air loop Air loops (wind tunnels) are shown in Fig. 2, with an orange arrow line representing the air flow loop. There are two identical wind tunnels with identical elements in each of them. The heat exchanger is located on the top of the wind tunnel, with thermocouple grids above and below it. The heat exchanger, shown in Figs. 3 and 4, is a plain-aluminum-fin-and-copper-tube heat exchanger, with an area ratio of fin to tube equal to 17.15. The detailed parameters are given in Table 1. A nozzle is used to measure the air flow rate, and a blower is used to circulate air in the wind tunnel, controlled by a speed controller. The inlet air is measured at different locations with 9 thermocouples evenly located in a grid, which is 0.03 m from the top of each heat exchanger. The thermocouples are connected as one channel to the data logger, providing the average inlet temperature. Temperature of the outlet air after is also measured,
Fig. 4. Heat exchanger (side view).
Fig. 5. Outlet air thermocouple grid.
using 18 thermocouples that are attached to the grid at different locations as shown in Fig. 5. The grid is 0.015 m from the bottom of the heat exchanger. 2.3. Chilled water loop Fig. 3. Heat exchangers.
Table 1 Parameters of heat exchangers used in this work. Parameters
Heat exchanger
Tube
Fin
Structure
24 tubes staggered arranged, one inlet, one outlet
Plain aluminum fin
Dimensions
385 × 300 × 30 (mm) (length × width × thickness)
Round copper tube Di = 6.2 (mm) Do = 8.0 (mm) Length = 385 (mm)
Fin density: 596 fpm Fin thickness: 0.3 mm
Chilled water is used to cool the refrigerant vapor in the condenser and subcooler. The loop is shown in Fig. 2 in blue lines. The cooling capacity can be adjusted by adjusting the flow rate of chilled water by the adjusting valve. The temperature of chilled water is measured before and after it passes the condenser. 3. Experimental matrix, data reduction and uncertainty analysis The experimental matrix is listed in Table 2. The working fluid is R134a, and it has heat exchange with room air under continuous and pulsating flow modes.
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Table 2 Matrix for the experimental parameters.
Flow rate Temperature
Vapor quality Operating mode
Refrigerant side
Air side
Mass flux mr 75–400 kg/m2·s Evaportation temp 15 °C; superheat temp (heated by heater) ~ 40 °C Inlet of HX ~0.1; Outlet of HX ~0.9 Continous flow Pulsating flow: Period τ 2–20 s, ratio μ 0.5
Speed va 0.8–1.7 m/s Inlet Temp: room temperature (~26 °C) Non Continous flow
Fig. 6. Simplified sketch for data reduction.
A simplified sketch relevant for data reduction is shown in Fig. 6. The refrigerant properties (mass flux mr, inlet temperature Tri, pressure p ri ) are measured upstream of the pre-heater, and the temperature Trsup and pressure prsup also can be measured downstream of the super-heater. The input of pre-heater qhin and superheater qhsup are also measured. Therefore, the heat transfer rate, qr, according to an energy balance, is determined from Eq. (1). The airside heat transfer rate, qa, is also available from the air mass flow rate, ma, and the air temperature before and after the heat exchanger, Tai, and Tao, per Eq. (2), where cpa is air specific heat based on the average temperature of Tai and Tao. The overall heat transfer coefficient, U, is calculated based on the average heat transfer rate of qr and qa using the ε-NTU method, and the area for the calculation of U is based on heat transfer area of refrigerant side. The air side heat transfer coefficient ha is calculated and referred to KimYoun-Webb’s model [12] for the similar structure of heat exchanger based on the knowledge of air velocity (measured) and fin
Table 3 Range and accuracy of measured parameters.
Air side Refrigerant side
Measured parameters/unit
Range
Accuracy
Temperature/°C Pressure drop/kPa Mass flow rate/g/s Pressure (inlet of HX)/kPa Pressure (outlet of HX)/kPa Temperature/°C Pressure drop/kPa Heater power/W
−10 to 80 0 to 0.623 0 to 18 0 to 1034 0 to 689 −10 to 80 0 to 37.36 0 to 1000
±0.1/°C ±1% FS ±0.1% RS ±0.25% RS ±0.25% RS ±0.1/°C ±0.25% RS ±0.02 W
dimensions. The heat transfer coefficient on the refrigerant side, hr, can be derived by subtracting the heat transfer coefficient of air, ha, from the overall heat transfer coefficient, U, as shown in Eq. (3), where ri, ro is the tube inside and outside diameter, Ai and Ao is the refrigerant side heat transfer area and air side overall heat transfer area including fin and the base area, L is the tube length and k is the tube wall thermal conductivity, and η is the overall surface efficiency.
q a = m ac pa (Tai − Tao )
(1)
q r = m r (hr sup − hri ) − q hin − q h sup
(2)
ln (ro ri ) 1 1 1 = + + ηha Ao UAi hr Ai 2π kL
(3)
For pulsating flow, every derived parameter is based on the timeaveraged measured parameters. For example, in Eq. (2), temperature Tao is the average of Tao during on-time and off-time of a period in the pulsating flow. The accuracy and range of every measured parameter is listed in Table 3. This accuracy leads to an error range of ±26 W/m2·°C to ±36 W/m2·°C for overall heat transfer coefficient, and ±54 W/m2·°C to ±206 W/m2·°C for refrigerant-side heat transfer coefficient, and the maximum error for both coefficients is obtained under the condition of pulsating flow with mass flux of 200 kg/m2·s, with a pulsating period of 4 s. The error of all tested conditions is limited to within ±5% of the measured values. Refrigerant-side heat transfer rate is also compared to the air-side, and it was found that for all cases refrigerant-side heat transfer rate has a maximum difference of 13% from that of air-side. 4. Model for estimating heat transfer performance Because the understanding of pulsating flow is incomplete, the models in this work are from the literature, which mainly are models for steady continuous flow, as shown in Table 4. The model for pulsating flow on-time period is the same as for continuous flow, which neglects the interaction between the new flash in fluid and the existing left over fluid from last period at the moment of valve on, and the model for pulsating flow off-time period neglects the fluid’s inertial flow at locations further from the valve along the evaporator at the moment of valve off. All those assumptions for pulsating flow might underestimate heat transfer performance in the model. However, such a quasi-steady model might provide a basic understanding on the trends of how pulsating flow affects performance. 5. Results and analysis 5.1. Effect of mass flux on heat transfer performance The results shown in Fig. 7 are for a fixed evaporation temperature of 15 °C, inlet vapor quality of 0.1, air inlet temperature of about
Table 4 Model for pulsating and continuous flow mode. Case Continuous
Single phase Two-phase
Pulsation/off
Single phase Two-phase
Pulsation/on Air side
Model for heat transfer
Model for pressure drop
Model for void fraction
Gnielinski’s correlation Kattan et al. [14] Wojtan et al. [15] Laminar flow model Flow rate for liquid is 0, and nuclear boiling for liquid, laminar flow model for vapor, convective heat transfer is neglected Same as continuous flow Kim et al. [12]
Churchill and Bernstein [13]. Muller-Steinhagen and Heck [16]
Non Steiner modified Rouhani-Axelsson [17]
Pressure drop is neglected
Keep the same as the end of on-time Same as continuous flow Non
Same as continuous flow Non (measured directly)
X. Wang et al./Applied Thermal Engineering 99 (2016) 825–833
5.5
1.4
5.0 1.2
Heat transfer coefficient (HTC) [kW/(m2.oC)]
Overall heat transfer coefficient (U) [kW/(m2.oC)]
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1.0 0.8 0.6 Exp., Continuous Exp., Pulsating Model, Continuous Model, Pulsating
0.4 0.2 0.0 80
100
120 140 160 180 200 Mass flux (G) [kg/(m2.s)]
220
(a) Overall heat transfer coefficient
4.5 4.0 3.5 3.0 2.5 2.0 Exp., Continuous Exp., Pulsating Model., Continuous Model, Pulsating
1.5 1.0 0.5 0.0 80
100
120 140 160 180 200 Mass flux (G) [kg/(m2.s)]
220
(b) Heat transfer coefficient at refrigerant side
Fig. 7. Heat transfer performance comparison of pulsating, continuous in experiment and model.
25 °C, and an increasing refrigerant mass flux from 100 kg/m2·s to 200 kg/m2·s, with a pulsating period τ = 4 s and time ratio μ = 0.5. Both the data for experiments and the model results are reported. Fig. 7(a) shows an overall heat transfer coefficient, and Fig. 7(b) shows heat transfer coefficient on the refrigerant side. Note that for every mass flux of refrigerant, the air velocity is fixed for pulsating flow and continuous flow in both the experiment and the model. However, the air velocity was controlled to increase from 0.8 m/s to 1.7 m/s when increasing the refrigerant mass flux. Both the experimental data and the model results suggest that the heat transfer performance is improved with increasing mass flux, for both pulsating flow and continuous flow in Fig. 7. This trend confirms common sense, because of an enlarged wetted area and flow regime changes associated with a larger mass flux. However, this improvement becomes weak at high mass flux ranges, which is especially notable from the model data in Fig. 7. There are two reasons for this behavior: (1) the air side heat transfer performance is limiting compared to the refrigerant side, and when the air velocity is increased, there is not so much increase in air side heat transfer coefficient, as shown in Fig. 8. Therefore, the overall heat transfer coefficient improvement is smaller for an increasing refrigerant mass flux; (2) the change of air temperature gets to a limited bar which is set less than the difference of the air inlet temperature and the refrigerant evaporation temperature. In Fig. 9, the air side temperature change factor
1.0
Air side temperature change factor (FTa)
Heat transfer coefficient (HTCa) o [W/m2. C]
100 90 80 70 60 50 40 30 20 10 0 0.6
is defined as FTa = ma × (Tai − Tao)/[ma-ave × (Tai − Tevap)]. When FTa is close to 1, it means that the difference between air inlet and outlet temperature is closed to the maximum change (Tai − Tevap), and heat transfer coefficient has a very limited space to increase. We can find in Fig. 9 that as the refrigerant mass flux increases, the factor FTa approaches 1. Therefore, the heat transfer coefficient increases less with increased mass flux. The heat transfer performance is enhanced by pulsating flow compared to continuous flow, within our experimental range, as shown in Fig. 7. This enhancement is reflected by the experimental data, both in the overall heat transfer coefficient and the refrigerant side heat transfer coefficient – both are higher than that of continuous flow. The enhancement ratios, γ, are defined as γU = Up/Uc and γHTC = HTCp/HTCc, in which γU, γHTC are the enhancement ratios for overall heat transfer coefficient and for refrigerant side heat transfer coefficient, respectively, and the subscripts p and c mean pulsating flow and continuous flow, respectively, HTC is for heat transfer coefficient in refrigerant-side. The enhancement ratios are shown in Fig. 10, where (a) and (b) are for the overall heat transfer and the refrigerant-side heat transfer enhancement, respectively. Both enhancement ratios from experimental data are bigger than 1. The maximum enhancement ratio for overall heat transfer coefficient and refrigerant-side heat transfer coefficient are 1.23 and 1.94, respectively, which occur when G = 100 kg/m2·s for both cases. The
0.8
1.0
1.2
1.4
1.6
1.8
Air velocity (Va) [m/s] Fig. 8. Air side heat transfer coefficient VS velocity.
2.0
0.8
0.6
0.4 Exp.Continuous Exp.Pulsating Model.Continuous Model. Pulsating
0.2
0.0 80
100
120 140 160 180 Mass flux (G) [kg/m2.s]
200
220
Fig. 9. Dimensionless temperature change VS refrigerant mass flux.
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2.2
1.3
2.0
Enhancement ratio (
HTC
)
Enhancement ratio ( U)
1.2
1.1
1.0
0.9
Exp. Model
0.8 80
100
1.8 1.6 1.4 1.2
Exp. Model
1.0
120 140 160 180 Mass flux (G) [kg/(m 2.s)]
200
80
220
(a) Overall heat transfer enhancement
100
120 140 160 180 Mass flux (G) [kg/m2.s]
200
220
(b) Refrigerant side heat transfer enhancement
Fig. 10. Enhancement ratio of heat transfer coefficient.
average enhancement ratio for γU and γHTC are about 1.13 and 1.56, respectively, within our experimental range. The enhancement ratio decreases with increasing mass flux, and it drops to 1.06 and 1.24 for overall and refrigerant-side heat transfer coefficient, respectively, when the mass flux is 200 kg/m2·s. According to flow regime expectations for continuous flow, there is an active range of mass flux within which flow regimes are sensitive to mass flux [14,15]. When the mass flux exceeds that range, the flow regime is unchanged for a large range of mass flux. Associated with the insensitivity of the flow regimes, the heat transfer coefficient increases in a smaller slope with an increasing mass flux. The flow regimes for pulsating flow might be very different from those of continuous flow; however, we might expect a similar insensitive range with increasing mass flux, which might explain a weak enhancement of heat transfer performance for pulsating flow at a larger mass flux. 5.2. Effect of pulsation period on heat transfer performance For a fixed pulsation ratio μ at 0.5, and as the pulsation period τ increases from 2 s to 20 s, the effect of pulsating flow on heat transfer coefficient is shown in Fig. 11, where the mass flux
G = 100 kg/m2·s, and (a) is for overall heat transfer coefficient, and (b) is for refrigerant side heat transfer coefficient. From Fig. 11, it is found that the heat transfer coefficient decreases with increasing pulsation period for both overall and refrigerant-side heat transfer coefficient. The heat transfer coefficient for pulsating flow is bigger than that for continuous flow within our experimental range. For the model, when the period is close to 20 s, the overall heat transfer coefficient becomes slightly lower than that for the continuous flow. The enhancement ratio is shown in Fig. 12, in which (a) is for overall heat transfer coefficient enhancement, and (b) is for refrigerant-side heat transfer coefficient enhancement. According to the experiment data, the enhancement ratio shows that there is a maximum increase of 27% and 123% for overall and refrigerant side heat transfer coefficients, respectively. The maximum enhancement occurs when the period is 2 s, and as the period increases, the enhancement becomes weak, and it drops to around 5% for overall heat transfer and 20% for refrigerantside heat transfer at a period τ = 20 s. From the model, it also can be found that the enhancement ratio decreases with increasing period; however, compared to the experimental data, it is quite different. The enhancement ratio decreases
Exp., Pulsating Exp., Continuous Model, Pulsating Model, Continuous
4.5 1.0 0.8 0.6 0.4
Exp., Continuous Exp., Pulsating Model, Continuous Model., Pulsating
0.2 0.0
0
5
10 15 Pulsation period ( ) [s]
(a)
20
Heat transfer coefficient (HTC) 2 o [kW/(m . C)]
Overall heat transfer coefficient (U) [kW/(m2.oC)]
1.2 4.0 3.5 3.0 2.5 2.0 1.5 1.0 0.5 0.0
0
5
10 15 Pulsation period ( ) [s]
(b)
Fig. 11. Heat transfer coefficient with different periods at G = 100 kg/m2·s.
20
X. Wang et al./Applied Thermal Engineering 99 (2016) 825–833
1.4
2.4
Exp. Model
2.1 HTC
Enhancement ratio (
Enhancement ratio ( U)
1.2
1.1
1.0
0.9
Exp. Model
)
1.3
831
0
5 10 15 Pulsation period ( ) [s]
20
1.8 1.5 1.2 0.9 0.6 0.3 0.0
0
5 10 15 Pulsation period ( ) [s]
(a)
20
(b)
Fig. 12. Enhancement ratio with different periods at G = 100 kg/m2·s.
5.3. Effect of pulsation on pressure drop
continuously for the experiment, but it remains constant and then decreases for model. This behavior is because, according to our simplified model, only when the period increases to some level and dryout occurs, the periods start to affect heat transfer performance. Consider Fig. 13, showing vapor quality distribution along the tube of heat exchanger at the last moment of a period, L = 0 associated with inlet of refrigerant, L = 1 associated with exit of refrigerant, and the calculated moment is t = τ−, the last moment in one period. The details of length and calculated moment are shown in Fig. 14. According to the model, when the period reaches roughly 16 s, there is dryout close to the tube exit, and as the period increases, the dryout length in the heat exchanger also increases. When τ = 30 s, more than half of the heat exchanger is under dry conditions. This clearly explains why an increasing period results in a heat transfer enhancement decrease. However, the model does not consider the effect of flow between on-time and off-time; thus it cannot catch the enhancement ratio drop for a short period as the experiment has shown. That might explain why for almost all cases, the model underestimates the experimental data.
Heat transfer performance is enhanced dramatically by pulsating flow; however, the price might be a higher pressure drop. The average pressure drop within a heat exchanger for pulsating flow and continuous flow is shown in Fig. 15, for mass flux from 100–200 kg/m2·s, and a pulsation period of 4 s. It can be seen that the pressure drop increases when the mass flux increases for both continuous flow and pulsating flow. It also can be found that the pulsating pressure drop is almost the same as that for continuous flow. However, the temporal pressure drop for pulsating flow is very different from that for continuous flow, as shown in Fig. 16, for a mass flux of 100 kg/m2·s and pulsation period of 20 s. The pressure drop for continuous flow has a very small change with time, and it changes with a very similar period when pulsating flow occurs. The pressure drop during on-time is much higher than that for continuous flow and it is much lower during off-time. The pressure drop increases with mass flux for both on-time and off-time. Pressure drop of on-time is higher than that of continuous
Fig. 13. Vapor quality with periods, G = 100 kg/m2·s.
Fig. 14. Calculated moment t = τ−.
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1.6
25
1.4
20
Pressure drop ratio Dpp/Dpc
Pressure drop (Dp) [kPa]
Exp., Continuous Exp., Pulsating
15 10 5
1.2 1.0 0.8 0.6 0.4 Dpp,on/Dpc
0 80
100
120 140 160 180 Mass flux (G) [kg/m2.s]
200
0.0 80
220
Fig. 15. Average pressure drop with mass flux, τ = 4 s.
Pressure drop (Dp) [kPa]
15
Continuous τ = 20 s
10
5
0
40
80
120 160 Time (t) [s]
200
240
Fig. 16. Temporal pressure drop, G = 100 kg/m2·s.
20
10
120 140 160 180 Mass flux G (kg/m2.s)
200
Fig. 17. Pressure drop of on-time and off-time, τ = 4 s.
200
220
2s 14s
12 Pressure drop (Dp) [kPa]
Pressure drop (Dp) [kPa]
Continuous Pulsating,off-time Pulsating,on-time
100
120 140 160 180 Mass flux G (kg/m2.s)
be seen that the on-time pressure drop ratio decreases and the offtime pressure drop ratio increases with increasing mass flux, and the maximum pressure drop ratio is around 1.47 at a mass flux of 100 kg/m2·s. This indicates a 47% pressure drop increase for pulsating flow during on-time, and the pressure drop decreases to around 30% while increasing the mass flux to about 200 kg/m2·s. The pressure drop for on-time and off-time becomes closer to the pressure drop of continuous flow with increasing mass flux, which means the effect of pulsation on pressure drop gets weaker with increasing mass flux. The pulsation period has a big effect on pressure drop. A group of raw data of temporal pressure drop for period τ = 2 s and τ = 14 s is shown in Fig. 19, in which the black line and red line are for average of pulsating pressure drop at τ = 2 s and τ = 14 s, respectively. It can be seen that many data are located around the average value for τ = 2 s and very few data are close to the average value for τ = 14 s. In order to quantify this effect, a parameter called pressure drop distribution is defined as: number of pressure drop within 0.75~1.25 of average pressure drop/number of data pool. Pressure drop distribution can provide information as to how closely dates are grouped to the average value. The pressure drop distribution with period is shown in Fig. 20, which is based on 240 continuous measurements from experimental data having a period from 2 to 20 s. It can be found again that for a shorter period, data are more close to the average value, and for a longer period, data are further away. For example, about 27% of the data are located within the set range for a period of 2 s, but less than 3% of the data are in the set range
30
0 80
100
Fig. 18. Pressure drop ratio, τ = 4 s.
flow and the pressure drop of off-time is lower than continuous flow for increasing mass flux (see Fig. 17) for a mass flux of 100 kg/m2·s and a pulsation period of 4 s. The pulsating pressure drop divided by the continuous pressure drop, a pressure drop ratio, is obtained as shown in Fig. 18, where Dpp,off is pressure drop during the off-time of pulsation, Dpp,on is pressure drop during the on-time of pulsation, and Dpc is pressure drop during continuous flow. It can
0
Dpp,off/Dpc
0.2
220
9
6
3
0
0
20
40
60 80 Time (t) [s]
100
120
Fig. 19. Temporal pressure drop with τ, G = 100 kg/m2·s.
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trends, but it seems to underestimate the absolute value when compared with experimental results.
0.30 0.25 Pressure drop distribution
833
Acknowledgements
0.20
We gratefully acknowledge financial support by the Air Conditioning and Refrigerate Center at the University of Illinois at Urbana Champaign.
0.15 0.10
References
0.05 0.00 0
3
6
9 12 Period [s]
15
18
21
Fig. 20. Pressure drop distribution with τ, G = 100 kg/m2·s.
for period of 20 s. This might mean that pressure drop is more sensitive to pulsation with a long period. 6. Summary and conclusions The heat transfer coefficient and pressure drop for pulsating flow have been experimentally determined in an air-refrigerant (R134a) system. The effect of mass flux and pulsation period on heat transfer performance have been explored and compared to a simplified model developed in this work. The results show that pulsating flow has a significant heat transfer enhancement compared to continuous flow: an increase of up to 27% of overall heat transfer coefficient was observed, indicating an increase of 123% in the refrigerantside heat transfer coefficient. The heat transfer coefficient increases and the enhancement ratio decreases with increasing mass flux. The pulsation period also has a big effect on heat transfer performance; a large enhancement at a shorter period was observed. The average pressure drop for pulsating flow is almost the same as that for the continuous flow; however, the temporal pressure drop is different and it changes periodically following the pulsation period. The pulsation effect on pressure drop becomes weaker at a larger mass flux and a shorter pulsation period. The model shows similar
[1] R.C. Martinelli, L.M.K. Boelter, E.B. Weinbeg, S. Takahi, Heat transfer to a fluid flowing periodically at low frequency, Trans. ASME 65 (1943) 789–798. [2] F.B. West, A.T. Taylor, The effect of pulsation on heat transfer, Chem. Eng. Prog. 48 (1) (1952) 34–43. [3] R.R. Morris, Effect of pulsations on heat transfer in streamline flow (Ph.D. thesis), Chemical Engineering, University of Washington, 1950. [4] R.P. Webb, The effect of pulsations on heat transfer in streamline flow (Ph.D. thesis), Chemical Engineering, University of Washington, 1949. [5] H. Gao, D.L. Zeng, Experimental investigation of intensified heat exchange based on self-excited oscillation pulse jet, J. Eng. Therm. Energy Power 18 (2003) 349–351. [6] J. Zheng, D.L. Zeng, P. Wang, An experimental study of heat transfer enhancement with a pulsating flow, Heat Transf. Asian Res. 33 (5) (2004) 279–286. [7] A.E. Zohir, The influence of pulsation on heat transfer in a heat exchanger for parallel and counter water flows, N. Y. Sci. J. 4 (6) (2011) 61–71. [8] A.E. Zohir, Heat transfer characteristics in a heat exchanger for turbulent pulsating water flow with different amplitudes, J. Am. Sci. 8 (2) (2012) 241–250. [9] Z.X. Guo, S.Y. Kim, H.J. Sung, Pulsating flow and heat transfer in a pipe partially filled with a porous medium, Int. J. Heat Mass Transf. 40 (17) (1997) 4209–4218. [10] M. Rahgoshay, A.A. Rahjbar, A. Ramiar, Laminar pulsating flow of nanofluids in a circular tube with isothermal wall, Int. Commun. Heat Mass Transf. 39 (2012) 463–469. [11] A.K. Bose, Heat transfer in pulsating flow, Nature 174 (1954) 41. [12] N.H. Kim, B. Youn, R.L. Webb, Air-side heat transfer and friction correlations for plain fin-and-tube heat exchangers with staggered tube arrangements, J. Heat Transfer 121 (3) (1999) 662–667. [13] S.W. Churchill, M. Bernstein, A correlating equation for forced convection from gases and liquids to a circular cylinder in crossflow, J. Heat Transfer 99 (1977) 300–306. [14] N. Kattan, J.R. Thome, D. Favrat, Flow boiling in horizontal tubes: part 3 – development of a new heat transfer model based on flow pattern, J. Heat Transfer 120 (1998) 156–165. [15] L. Wojtan, T. Ursenbacher, J.H. Thome, Investigation of flow boiling in horizontal tubes: part II – development of a new heat transfer model for stratified-wavy, dryout and mist flow regimes, Int. J. Heat Mass Transf. 48 (2005) 2970–2985. [16] H. Muller-Steinhagen, K. Heck, A simple friction pressure drop correlation for two-phase flow in pipes, Chem. Eng. Process. 20 (1986) 297–308. [17] D. Steiner, Heat transfer to boiling saturated liquids, in: VDI-Wärmeatlas (VDI Heat Atlas), Verein Deutscher Ingenieure, VDI-Gesselschaft Verfahrenstechnic und Chemieingenieurswesen (GCV), Dusseldorf, 1993, pp. 15–24.
Contents lists available at ScienceDirect
Applied Thermal Engineering j o u r n a l h o m e p a g e : w w w. e l s e v i e r. c o m / l o c a t e / a p t h e r m e n g
Research Paper
Evaporator performance enhancement by pulsation width modulation (PWM) X. Wang a,*, K. Tang b, P.S. Hrnjak a,c a b c
Mechanical Science and Engineering, University of Illinois, Urbana, IL 61801, USA Institute of Refrigeration and Cryogenics, Zhejiang University, Hangzhou 310027, China CTS, Urbana, IL 61802, USA
H I G H L I G H T S
• • • • •
Effects of pulsating flow on heat transfer performance and pressure drop in an air-refrigerant evaporator have been explored. Pulsating flow has enhanced overall and refrigerant-side heat transfer coefficient by 27% and 123% respectively, compared to continuous flow. Pulsation period has a big effect on heat transfer performance; a larger enhancement at a shorter period was observed. The temporal pressure drop for pulsating flow changes periodically following the pulsation period. Pulsation effect on pressure drop becomes weaker at a larger mass flux and a shorter pulsation period.
A R T I C L E
I N F O
Article history: Received 8 November 2014 Accepted 17 December 2015 Available online Keywords: Pulsating flow Pulsation width modulation Evaporator control Heat transfer coefficient Two-phase flow regime
A B S T R A C T
In this work, the effects of pulsation width on heat transfer coefficient and pressure drop in an air-to-refrigerant evaporator is studied experimentally. The pulsation width can be controlled to a minimum period of 2 s. A general model is also developed to predict heat transfer performance and pressure drop. The overall heat transfer coefficient, refrigerant side heat transfer coefficient, and pressure drop are measured and compared to a baseline without pulsating flow. The results show that the overall and refrigerant-side heat transfer coefficients have improved about 27% and 123%, respectively, with pulsating flow. The refrigerant mass flux and pulsation period have an important effect on the heat transfer enhancement. The average pressure drop of the pulsating flow differs little from that of continuous flow, but the temporal pressure drop of pulsating flow is quite different and varies with mass flux and pulsation period. © 2016 Elsevier Ltd. All rights reserved.
1. Introduction Flow pulsation, sometimes referred to as mechanical excitation or fluid-borne oscillation, is well known to affect heat transfer behavior. For example, if a single-phase convective heat transfer condition is subjected to pulsating flow, the pulsations reduce the timeaveraged thickness of the boundary layer and hence the thermal resistance, which improves heat transfer performance. If flow boiling is subjected to pulsating flow, the pressure change during the pulsating flow might lead to a local cavitation effect, which will also affect heat transfer performance. There has been significant previous research on pulsating flow and heat transfer, and some research shows a significant enhancement of heat transfer performance by using pulsating flow, while some shows little or no effect. In a turbulent flow, Martinelli et al. [1] found no difference in heat transfer
* Corresponding author. Tel.: +1 217 693 2358; fax: +1 217 244 6534. E-mail address: [email protected] (X. Wang). http://dx.doi.org/10.1016/j.applthermaleng.2015.12.049 1359-4311/© 2016 Elsevier Ltd. All rights reserved.
between steady continuous flow and pulsating flow; however, West and Taylor found an increase of 60–70% [2]. In a laminar flow, Morris [3] and Webb [4] found no change in heat transfer. Gao and Zeng [5] found that with a self-oscillator nozzle in a system, heat transfer performance increased about 10–30%. In their later work [6], they found that the geometric structure of the self-oscillator nozzle can affect the pulsating flow and also the heat transfer coefficient. The pressure increase caused by the self-oscillator nozzle was found negligible when compared to the improvement of heat transfer coefficient. Zohir [7] conducted experiments to compare parallel flow and counter flow when pulsing flow was generated, and he found that pulsing flow can increase heat transfer coefficient by up to about 20% in parallel flow and up to about 90% in counter flow. In Zohir’s recent work [8], he found that heat transfer performance was improved with pulsing flow, and the enhancement reached a peak value when the flow regime was in transition between laminar and turbulent flow. For all pulsation frequencies, the highest enhancement was about 10 times for single-phase counter flow and about 8 times for parallel flow. Guo et al. [9] investigated the pulsating flow effect
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for a pipe partially filled with a porous medium using numerical methods, and they found an increase in effective thermal conductivity along the tube length direction and an increase in convective heat transfer coefficient. Their results show that the maximum heat transfer enhancement was achieved by optimizing the pulse frequency. Recently, Rahgoshay et al. [10] conducted research on nanofluids in a pulsating flow using numerical method, and the result shows that the heat transfer coefficient increased slightly with increasing frequency and amplitude. Bose [11] found that a pulsing flow can increase heat transfer or decrease heat transfer, depending on the pulsating frequency and amplitude. Apparently, there exists a critical frequency which is a function of wave-form and Reynolds number; however, there was only a decrease in heat transfer performance if the pulse amplitude was very low. Most prior research on pulsed-flow heat transfer used water, oil and water mixtures, or air as working fluids, and there are contradictions as to the effect of pulsating flow on heat transfer performance. To date there has been no research on two-phase refrigerant pulsating flows. Moreover, the effects of pulsation are more complex for pulsations within a practical evaporator, for which the use of a multi-tube construction and the presence of tube bends at the end of each tube certainly affect the resulting flow. However, it is anticipated that pulse width modulation might be effective for evaporator performance enhancement. In this work, refrigerant R134a is used as the working fluid, and a pulsating flow generated by solenoid-valve system in a real evaporator. A general model is also built to predict heat transfer performance while pulsating flow occurs in the evaporator.
2. Experimental approach and facilities When a flow pulsates with a prescribed period, in this case controlled by a solenoid valve, the mass flow rate and other parameters are expected to manifest the same periodicity. In these experiments the mass flow rate in the evaporator ideally varies periodically as shown in Fig. 1. The pulsating flow period is τ, and the on-time ratio is μ, which is the ratio of on-time to the duration of the entire period. During one period, the mass flow rate is mr during the ontime of μτ and 0 for the off-time of (1 − μ)τ. The time-averaged mass flow rate is equal to μmr, which will be considered as the continuousflow baseline for purposes of comparison. If μ = 0.5, the on-time is equal to the off-time in a period, and during the on-time the mass flow rate for pulsating flow is twice that of the continuous flow baseline. For experimental control and operation, two identical loops were created, with two identical evaporators. Downstream from both evaporators, refrigerant flowed into a common flow-conditioning system, consisting of a condenser, a pump, and a preheater. Just upstream of the evaporators, the flow split and the refrigerant flow was directed into one evaporator or another, as determined by two
Fig. 1. Pulsating flow mass flow rate vs time.
Fig. 2. Schematic drawing of the system.
solenoid valves. When one solenoid valve was open the other was close, and vice versa (see Fig. 2). In this way, the flow through the shared portion of the loops was steady, but the flow through the evaporators was pulsed, and when pulsation occurred, the pressure upstream of the solenoid valves was essentially constant. So during the experiment, the two loops were operated alternately; there was always one loop open and one loop closed. Referring again to Fig. 1, when HX1 experiences on-time, HX2 experiences off time, and when HX1 is off, HX2 is on. The two loops are never closed at the same time. This operating procedure ensured a steady state inlet pressure and a continuous mass flow rate, mr. The experiment apparatus is designed to measure heat transfer coefficient and pressure drop for pulsating and continuous flow, and in addition to the two refrigerant loops, there is a chilledwater loop operated at steady state to cool the condenser (see Fig. 2). The refrigerant is circulated by a gear pump, absorbing heat from the air through evaporators and releasing heat to the chilled water through a condenser. As described, the refrigerant loops have two parallel, identical subloops, each with one evaporator (HX1 or HX2). The air flow rate and inlet temperature to each evaporator are equal for both evaporators. 2.1. Refrigerant loop In Fig. 2, the red loop represents refrigerant, which is pumped from a subcooler by a gear pump with mass flow rate measured by a flow meter, to a controlled heater which can adjust the refrigerant flow to a desired vapor quality. After that, the refrigerant loop is divided into two identical subloops, subloop1 and subloop2. Each subloop includes a solenoid valve, a heat exchanger, a power controllable heater, and a sight glass. All elements of the two subloops are identical as shown in Fig. 2. After the refrigerant passes through a sight glass, these two subloops are combined and connected to a common loop, then the refrigerant flows through another flow meter to a condenser, a receiver and the subcooler, the pump, and finally back to the point where the lines split to complete the loop. The refrigerant mass flow rate can be adjusted by adjusting the pump speed controller and also by adjusting the valve in the bypass, the flow rate was measured by a Coriolis-effect flow meter before it was divided into two subloops. The vapor quality of the refrigerant
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flow was set to a certain value by heating the subcooled refrigerant using a heater, controlled by adjusting the input voltage. The input power, refrigerant temperature before the heater, and the refrigerant pressure after the heater are measured for vapor quality control. Pulsation frequency is controlled by controlling the solenoid valve on and off periods; the period can be set from 0.1 s to 60 s. The system works such that when one solenoid valve in one subloop is on, the other solenoid valve in the other loop is off automatically. Temperatures of refrigerant before and after each heat exchanger in each loop are measured and differential pressures across each heat exchanger are also measured. After the refrigerant leaves the heat exchanger, it is heated to a prescribed superheat with a heater controlled in the same way as the heater upstream of the heat exchanger. Flow was observed in the sight glass to ensure that no liquid exists in the flow before the two subloop refrigerant flows are combined and pass through the downstream flow meter. 2.2. Air loop Air loops (wind tunnels) are shown in Fig. 2, with an orange arrow line representing the air flow loop. There are two identical wind tunnels with identical elements in each of them. The heat exchanger is located on the top of the wind tunnel, with thermocouple grids above and below it. The heat exchanger, shown in Figs. 3 and 4, is a plain-aluminum-fin-and-copper-tube heat exchanger, with an area ratio of fin to tube equal to 17.15. The detailed parameters are given in Table 1. A nozzle is used to measure the air flow rate, and a blower is used to circulate air in the wind tunnel, controlled by a speed controller. The inlet air is measured at different locations with 9 thermocouples evenly located in a grid, which is 0.03 m from the top of each heat exchanger. The thermocouples are connected as one channel to the data logger, providing the average inlet temperature. Temperature of the outlet air after is also measured,
Fig. 4. Heat exchanger (side view).
Fig. 5. Outlet air thermocouple grid.
using 18 thermocouples that are attached to the grid at different locations as shown in Fig. 5. The grid is 0.015 m from the bottom of the heat exchanger. 2.3. Chilled water loop Fig. 3. Heat exchangers.
Table 1 Parameters of heat exchangers used in this work. Parameters
Heat exchanger
Tube
Fin
Structure
24 tubes staggered arranged, one inlet, one outlet
Plain aluminum fin
Dimensions
385 × 300 × 30 (mm) (length × width × thickness)
Round copper tube Di = 6.2 (mm) Do = 8.0 (mm) Length = 385 (mm)
Fin density: 596 fpm Fin thickness: 0.3 mm
Chilled water is used to cool the refrigerant vapor in the condenser and subcooler. The loop is shown in Fig. 2 in blue lines. The cooling capacity can be adjusted by adjusting the flow rate of chilled water by the adjusting valve. The temperature of chilled water is measured before and after it passes the condenser. 3. Experimental matrix, data reduction and uncertainty analysis The experimental matrix is listed in Table 2. The working fluid is R134a, and it has heat exchange with room air under continuous and pulsating flow modes.
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Table 2 Matrix for the experimental parameters.
Flow rate Temperature
Vapor quality Operating mode
Refrigerant side
Air side
Mass flux mr 75–400 kg/m2·s Evaportation temp 15 °C; superheat temp (heated by heater) ~ 40 °C Inlet of HX ~0.1; Outlet of HX ~0.9 Continous flow Pulsating flow: Period τ 2–20 s, ratio μ 0.5
Speed va 0.8–1.7 m/s Inlet Temp: room temperature (~26 °C) Non Continous flow
Fig. 6. Simplified sketch for data reduction.
A simplified sketch relevant for data reduction is shown in Fig. 6. The refrigerant properties (mass flux mr, inlet temperature Tri, pressure p ri ) are measured upstream of the pre-heater, and the temperature Trsup and pressure prsup also can be measured downstream of the super-heater. The input of pre-heater qhin and superheater qhsup are also measured. Therefore, the heat transfer rate, qr, according to an energy balance, is determined from Eq. (1). The airside heat transfer rate, qa, is also available from the air mass flow rate, ma, and the air temperature before and after the heat exchanger, Tai, and Tao, per Eq. (2), where cpa is air specific heat based on the average temperature of Tai and Tao. The overall heat transfer coefficient, U, is calculated based on the average heat transfer rate of qr and qa using the ε-NTU method, and the area for the calculation of U is based on heat transfer area of refrigerant side. The air side heat transfer coefficient ha is calculated and referred to KimYoun-Webb’s model [12] for the similar structure of heat exchanger based on the knowledge of air velocity (measured) and fin
Table 3 Range and accuracy of measured parameters.
Air side Refrigerant side
Measured parameters/unit
Range
Accuracy
Temperature/°C Pressure drop/kPa Mass flow rate/g/s Pressure (inlet of HX)/kPa Pressure (outlet of HX)/kPa Temperature/°C Pressure drop/kPa Heater power/W
−10 to 80 0 to 0.623 0 to 18 0 to 1034 0 to 689 −10 to 80 0 to 37.36 0 to 1000
±0.1/°C ±1% FS ±0.1% RS ±0.25% RS ±0.25% RS ±0.1/°C ±0.25% RS ±0.02 W
dimensions. The heat transfer coefficient on the refrigerant side, hr, can be derived by subtracting the heat transfer coefficient of air, ha, from the overall heat transfer coefficient, U, as shown in Eq. (3), where ri, ro is the tube inside and outside diameter, Ai and Ao is the refrigerant side heat transfer area and air side overall heat transfer area including fin and the base area, L is the tube length and k is the tube wall thermal conductivity, and η is the overall surface efficiency.
q a = m ac pa (Tai − Tao )
(1)
q r = m r (hr sup − hri ) − q hin − q h sup
(2)
ln (ro ri ) 1 1 1 = + + ηha Ao UAi hr Ai 2π kL
(3)
For pulsating flow, every derived parameter is based on the timeaveraged measured parameters. For example, in Eq. (2), temperature Tao is the average of Tao during on-time and off-time of a period in the pulsating flow. The accuracy and range of every measured parameter is listed in Table 3. This accuracy leads to an error range of ±26 W/m2·°C to ±36 W/m2·°C for overall heat transfer coefficient, and ±54 W/m2·°C to ±206 W/m2·°C for refrigerant-side heat transfer coefficient, and the maximum error for both coefficients is obtained under the condition of pulsating flow with mass flux of 200 kg/m2·s, with a pulsating period of 4 s. The error of all tested conditions is limited to within ±5% of the measured values. Refrigerant-side heat transfer rate is also compared to the air-side, and it was found that for all cases refrigerant-side heat transfer rate has a maximum difference of 13% from that of air-side. 4. Model for estimating heat transfer performance Because the understanding of pulsating flow is incomplete, the models in this work are from the literature, which mainly are models for steady continuous flow, as shown in Table 4. The model for pulsating flow on-time period is the same as for continuous flow, which neglects the interaction between the new flash in fluid and the existing left over fluid from last period at the moment of valve on, and the model for pulsating flow off-time period neglects the fluid’s inertial flow at locations further from the valve along the evaporator at the moment of valve off. All those assumptions for pulsating flow might underestimate heat transfer performance in the model. However, such a quasi-steady model might provide a basic understanding on the trends of how pulsating flow affects performance. 5. Results and analysis 5.1. Effect of mass flux on heat transfer performance The results shown in Fig. 7 are for a fixed evaporation temperature of 15 °C, inlet vapor quality of 0.1, air inlet temperature of about
Table 4 Model for pulsating and continuous flow mode. Case Continuous
Single phase Two-phase
Pulsation/off
Single phase Two-phase
Pulsation/on Air side
Model for heat transfer
Model for pressure drop
Model for void fraction
Gnielinski’s correlation Kattan et al. [14] Wojtan et al. [15] Laminar flow model Flow rate for liquid is 0, and nuclear boiling for liquid, laminar flow model for vapor, convective heat transfer is neglected Same as continuous flow Kim et al. [12]
Churchill and Bernstein [13]. Muller-Steinhagen and Heck [16]
Non Steiner modified Rouhani-Axelsson [17]
Pressure drop is neglected
Keep the same as the end of on-time Same as continuous flow Non
Same as continuous flow Non (measured directly)
X. Wang et al./Applied Thermal Engineering 99 (2016) 825–833
5.5
1.4
5.0 1.2
Heat transfer coefficient (HTC) [kW/(m2.oC)]
Overall heat transfer coefficient (U) [kW/(m2.oC)]
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1.0 0.8 0.6 Exp., Continuous Exp., Pulsating Model, Continuous Model, Pulsating
0.4 0.2 0.0 80
100
120 140 160 180 200 Mass flux (G) [kg/(m2.s)]
220
(a) Overall heat transfer coefficient
4.5 4.0 3.5 3.0 2.5 2.0 Exp., Continuous Exp., Pulsating Model., Continuous Model, Pulsating
1.5 1.0 0.5 0.0 80
100
120 140 160 180 200 Mass flux (G) [kg/(m2.s)]
220
(b) Heat transfer coefficient at refrigerant side
Fig. 7. Heat transfer performance comparison of pulsating, continuous in experiment and model.
25 °C, and an increasing refrigerant mass flux from 100 kg/m2·s to 200 kg/m2·s, with a pulsating period τ = 4 s and time ratio μ = 0.5. Both the data for experiments and the model results are reported. Fig. 7(a) shows an overall heat transfer coefficient, and Fig. 7(b) shows heat transfer coefficient on the refrigerant side. Note that for every mass flux of refrigerant, the air velocity is fixed for pulsating flow and continuous flow in both the experiment and the model. However, the air velocity was controlled to increase from 0.8 m/s to 1.7 m/s when increasing the refrigerant mass flux. Both the experimental data and the model results suggest that the heat transfer performance is improved with increasing mass flux, for both pulsating flow and continuous flow in Fig. 7. This trend confirms common sense, because of an enlarged wetted area and flow regime changes associated with a larger mass flux. However, this improvement becomes weak at high mass flux ranges, which is especially notable from the model data in Fig. 7. There are two reasons for this behavior: (1) the air side heat transfer performance is limiting compared to the refrigerant side, and when the air velocity is increased, there is not so much increase in air side heat transfer coefficient, as shown in Fig. 8. Therefore, the overall heat transfer coefficient improvement is smaller for an increasing refrigerant mass flux; (2) the change of air temperature gets to a limited bar which is set less than the difference of the air inlet temperature and the refrigerant evaporation temperature. In Fig. 9, the air side temperature change factor
1.0
Air side temperature change factor (FTa)
Heat transfer coefficient (HTCa) o [W/m2. C]
100 90 80 70 60 50 40 30 20 10 0 0.6
is defined as FTa = ma × (Tai − Tao)/[ma-ave × (Tai − Tevap)]. When FTa is close to 1, it means that the difference between air inlet and outlet temperature is closed to the maximum change (Tai − Tevap), and heat transfer coefficient has a very limited space to increase. We can find in Fig. 9 that as the refrigerant mass flux increases, the factor FTa approaches 1. Therefore, the heat transfer coefficient increases less with increased mass flux. The heat transfer performance is enhanced by pulsating flow compared to continuous flow, within our experimental range, as shown in Fig. 7. This enhancement is reflected by the experimental data, both in the overall heat transfer coefficient and the refrigerant side heat transfer coefficient – both are higher than that of continuous flow. The enhancement ratios, γ, are defined as γU = Up/Uc and γHTC = HTCp/HTCc, in which γU, γHTC are the enhancement ratios for overall heat transfer coefficient and for refrigerant side heat transfer coefficient, respectively, and the subscripts p and c mean pulsating flow and continuous flow, respectively, HTC is for heat transfer coefficient in refrigerant-side. The enhancement ratios are shown in Fig. 10, where (a) and (b) are for the overall heat transfer and the refrigerant-side heat transfer enhancement, respectively. Both enhancement ratios from experimental data are bigger than 1. The maximum enhancement ratio for overall heat transfer coefficient and refrigerant-side heat transfer coefficient are 1.23 and 1.94, respectively, which occur when G = 100 kg/m2·s for both cases. The
0.8
1.0
1.2
1.4
1.6
1.8
Air velocity (Va) [m/s] Fig. 8. Air side heat transfer coefficient VS velocity.
2.0
0.8
0.6
0.4 Exp.Continuous Exp.Pulsating Model.Continuous Model. Pulsating
0.2
0.0 80
100
120 140 160 180 Mass flux (G) [kg/m2.s]
200
220
Fig. 9. Dimensionless temperature change VS refrigerant mass flux.
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2.2
1.3
2.0
Enhancement ratio (
HTC
)
Enhancement ratio ( U)
1.2
1.1
1.0
0.9
Exp. Model
0.8 80
100
1.8 1.6 1.4 1.2
Exp. Model
1.0
120 140 160 180 Mass flux (G) [kg/(m 2.s)]
200
80
220
(a) Overall heat transfer enhancement
100
120 140 160 180 Mass flux (G) [kg/m2.s]
200
220
(b) Refrigerant side heat transfer enhancement
Fig. 10. Enhancement ratio of heat transfer coefficient.
average enhancement ratio for γU and γHTC are about 1.13 and 1.56, respectively, within our experimental range. The enhancement ratio decreases with increasing mass flux, and it drops to 1.06 and 1.24 for overall and refrigerant-side heat transfer coefficient, respectively, when the mass flux is 200 kg/m2·s. According to flow regime expectations for continuous flow, there is an active range of mass flux within which flow regimes are sensitive to mass flux [14,15]. When the mass flux exceeds that range, the flow regime is unchanged for a large range of mass flux. Associated with the insensitivity of the flow regimes, the heat transfer coefficient increases in a smaller slope with an increasing mass flux. The flow regimes for pulsating flow might be very different from those of continuous flow; however, we might expect a similar insensitive range with increasing mass flux, which might explain a weak enhancement of heat transfer performance for pulsating flow at a larger mass flux. 5.2. Effect of pulsation period on heat transfer performance For a fixed pulsation ratio μ at 0.5, and as the pulsation period τ increases from 2 s to 20 s, the effect of pulsating flow on heat transfer coefficient is shown in Fig. 11, where the mass flux
G = 100 kg/m2·s, and (a) is for overall heat transfer coefficient, and (b) is for refrigerant side heat transfer coefficient. From Fig. 11, it is found that the heat transfer coefficient decreases with increasing pulsation period for both overall and refrigerant-side heat transfer coefficient. The heat transfer coefficient for pulsating flow is bigger than that for continuous flow within our experimental range. For the model, when the period is close to 20 s, the overall heat transfer coefficient becomes slightly lower than that for the continuous flow. The enhancement ratio is shown in Fig. 12, in which (a) is for overall heat transfer coefficient enhancement, and (b) is for refrigerant-side heat transfer coefficient enhancement. According to the experiment data, the enhancement ratio shows that there is a maximum increase of 27% and 123% for overall and refrigerant side heat transfer coefficients, respectively. The maximum enhancement occurs when the period is 2 s, and as the period increases, the enhancement becomes weak, and it drops to around 5% for overall heat transfer and 20% for refrigerantside heat transfer at a period τ = 20 s. From the model, it also can be found that the enhancement ratio decreases with increasing period; however, compared to the experimental data, it is quite different. The enhancement ratio decreases
Exp., Pulsating Exp., Continuous Model, Pulsating Model, Continuous
4.5 1.0 0.8 0.6 0.4
Exp., Continuous Exp., Pulsating Model, Continuous Model., Pulsating
0.2 0.0
0
5
10 15 Pulsation period ( ) [s]
(a)
20
Heat transfer coefficient (HTC) 2 o [kW/(m . C)]
Overall heat transfer coefficient (U) [kW/(m2.oC)]
1.2 4.0 3.5 3.0 2.5 2.0 1.5 1.0 0.5 0.0
0
5
10 15 Pulsation period ( ) [s]
(b)
Fig. 11. Heat transfer coefficient with different periods at G = 100 kg/m2·s.
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1.4
2.4
Exp. Model
2.1 HTC
Enhancement ratio (
Enhancement ratio ( U)
1.2
1.1
1.0
0.9
Exp. Model
)
1.3
831
0
5 10 15 Pulsation period ( ) [s]
20
1.8 1.5 1.2 0.9 0.6 0.3 0.0
0
5 10 15 Pulsation period ( ) [s]
(a)
20
(b)
Fig. 12. Enhancement ratio with different periods at G = 100 kg/m2·s.
5.3. Effect of pulsation on pressure drop
continuously for the experiment, but it remains constant and then decreases for model. This behavior is because, according to our simplified model, only when the period increases to some level and dryout occurs, the periods start to affect heat transfer performance. Consider Fig. 13, showing vapor quality distribution along the tube of heat exchanger at the last moment of a period, L = 0 associated with inlet of refrigerant, L = 1 associated with exit of refrigerant, and the calculated moment is t = τ−, the last moment in one period. The details of length and calculated moment are shown in Fig. 14. According to the model, when the period reaches roughly 16 s, there is dryout close to the tube exit, and as the period increases, the dryout length in the heat exchanger also increases. When τ = 30 s, more than half of the heat exchanger is under dry conditions. This clearly explains why an increasing period results in a heat transfer enhancement decrease. However, the model does not consider the effect of flow between on-time and off-time; thus it cannot catch the enhancement ratio drop for a short period as the experiment has shown. That might explain why for almost all cases, the model underestimates the experimental data.
Heat transfer performance is enhanced dramatically by pulsating flow; however, the price might be a higher pressure drop. The average pressure drop within a heat exchanger for pulsating flow and continuous flow is shown in Fig. 15, for mass flux from 100–200 kg/m2·s, and a pulsation period of 4 s. It can be seen that the pressure drop increases when the mass flux increases for both continuous flow and pulsating flow. It also can be found that the pulsating pressure drop is almost the same as that for continuous flow. However, the temporal pressure drop for pulsating flow is very different from that for continuous flow, as shown in Fig. 16, for a mass flux of 100 kg/m2·s and pulsation period of 20 s. The pressure drop for continuous flow has a very small change with time, and it changes with a very similar period when pulsating flow occurs. The pressure drop during on-time is much higher than that for continuous flow and it is much lower during off-time. The pressure drop increases with mass flux for both on-time and off-time. Pressure drop of on-time is higher than that of continuous
Fig. 13. Vapor quality with periods, G = 100 kg/m2·s.
Fig. 14. Calculated moment t = τ−.
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X. Wang et al./Applied Thermal Engineering 99 (2016) 825–833
1.6
25
1.4
20
Pressure drop ratio Dpp/Dpc
Pressure drop (Dp) [kPa]
Exp., Continuous Exp., Pulsating
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100
120 140 160 180 Mass flux (G) [kg/m2.s]
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Fig. 15. Average pressure drop with mass flux, τ = 4 s.
Pressure drop (Dp) [kPa]
15
Continuous τ = 20 s
10
5
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40
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Fig. 16. Temporal pressure drop, G = 100 kg/m2·s.
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Fig. 17. Pressure drop of on-time and off-time, τ = 4 s.
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2s 14s
12 Pressure drop (Dp) [kPa]
Pressure drop (Dp) [kPa]
Continuous Pulsating,off-time Pulsating,on-time
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120 140 160 180 Mass flux G (kg/m2.s)
be seen that the on-time pressure drop ratio decreases and the offtime pressure drop ratio increases with increasing mass flux, and the maximum pressure drop ratio is around 1.47 at a mass flux of 100 kg/m2·s. This indicates a 47% pressure drop increase for pulsating flow during on-time, and the pressure drop decreases to around 30% while increasing the mass flux to about 200 kg/m2·s. The pressure drop for on-time and off-time becomes closer to the pressure drop of continuous flow with increasing mass flux, which means the effect of pulsation on pressure drop gets weaker with increasing mass flux. The pulsation period has a big effect on pressure drop. A group of raw data of temporal pressure drop for period τ = 2 s and τ = 14 s is shown in Fig. 19, in which the black line and red line are for average of pulsating pressure drop at τ = 2 s and τ = 14 s, respectively. It can be seen that many data are located around the average value for τ = 2 s and very few data are close to the average value for τ = 14 s. In order to quantify this effect, a parameter called pressure drop distribution is defined as: number of pressure drop within 0.75~1.25 of average pressure drop/number of data pool. Pressure drop distribution can provide information as to how closely dates are grouped to the average value. The pressure drop distribution with period is shown in Fig. 20, which is based on 240 continuous measurements from experimental data having a period from 2 to 20 s. It can be found again that for a shorter period, data are more close to the average value, and for a longer period, data are further away. For example, about 27% of the data are located within the set range for a period of 2 s, but less than 3% of the data are in the set range
30
0 80
100
Fig. 18. Pressure drop ratio, τ = 4 s.
flow and the pressure drop of off-time is lower than continuous flow for increasing mass flux (see Fig. 17) for a mass flux of 100 kg/m2·s and a pulsation period of 4 s. The pulsating pressure drop divided by the continuous pressure drop, a pressure drop ratio, is obtained as shown in Fig. 18, where Dpp,off is pressure drop during the off-time of pulsation, Dpp,on is pressure drop during the on-time of pulsation, and Dpc is pressure drop during continuous flow. It can
0
Dpp,off/Dpc
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9
6
3
0
0
20
40
60 80 Time (t) [s]
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Fig. 19. Temporal pressure drop with τ, G = 100 kg/m2·s.
X. Wang et al./Applied Thermal Engineering 99 (2016) 825–833
trends, but it seems to underestimate the absolute value when compared with experimental results.
0.30 0.25 Pressure drop distribution
833
Acknowledgements
0.20
We gratefully acknowledge financial support by the Air Conditioning and Refrigerate Center at the University of Illinois at Urbana Champaign.
0.15 0.10
References
0.05 0.00 0
3
6
9 12 Period [s]
15
18
21
Fig. 20. Pressure drop distribution with τ, G = 100 kg/m2·s.
for period of 20 s. This might mean that pressure drop is more sensitive to pulsation with a long period. 6. Summary and conclusions The heat transfer coefficient and pressure drop for pulsating flow have been experimentally determined in an air-refrigerant (R134a) system. The effect of mass flux and pulsation period on heat transfer performance have been explored and compared to a simplified model developed in this work. The results show that pulsating flow has a significant heat transfer enhancement compared to continuous flow: an increase of up to 27% of overall heat transfer coefficient was observed, indicating an increase of 123% in the refrigerantside heat transfer coefficient. The heat transfer coefficient increases and the enhancement ratio decreases with increasing mass flux. The pulsation period also has a big effect on heat transfer performance; a large enhancement at a shorter period was observed. The average pressure drop for pulsating flow is almost the same as that for the continuous flow; however, the temporal pressure drop is different and it changes periodically following the pulsation period. The pulsation effect on pressure drop becomes weaker at a larger mass flux and a shorter pulsation period. The model shows similar
[1] R.C. Martinelli, L.M.K. Boelter, E.B. Weinbeg, S. Takahi, Heat transfer to a fluid flowing periodically at low frequency, Trans. ASME 65 (1943) 789–798. [2] F.B. West, A.T. Taylor, The effect of pulsation on heat transfer, Chem. Eng. Prog. 48 (1) (1952) 34–43. [3] R.R. Morris, Effect of pulsations on heat transfer in streamline flow (Ph.D. thesis), Chemical Engineering, University of Washington, 1950. [4] R.P. Webb, The effect of pulsations on heat transfer in streamline flow (Ph.D. thesis), Chemical Engineering, University of Washington, 1949. [5] H. Gao, D.L. Zeng, Experimental investigation of intensified heat exchange based on self-excited oscillation pulse jet, J. Eng. Therm. Energy Power 18 (2003) 349–351. [6] J. Zheng, D.L. Zeng, P. Wang, An experimental study of heat transfer enhancement with a pulsating flow, Heat Transf. Asian Res. 33 (5) (2004) 279–286. [7] A.E. Zohir, The influence of pulsation on heat transfer in a heat exchanger for parallel and counter water flows, N. Y. Sci. J. 4 (6) (2011) 61–71. [8] A.E. Zohir, Heat transfer characteristics in a heat exchanger for turbulent pulsating water flow with different amplitudes, J. Am. Sci. 8 (2) (2012) 241–250. [9] Z.X. Guo, S.Y. Kim, H.J. Sung, Pulsating flow and heat transfer in a pipe partially filled with a porous medium, Int. J. Heat Mass Transf. 40 (17) (1997) 4209–4218. [10] M. Rahgoshay, A.A. Rahjbar, A. Ramiar, Laminar pulsating flow of nanofluids in a circular tube with isothermal wall, Int. Commun. Heat Mass Transf. 39 (2012) 463–469. [11] A.K. Bose, Heat transfer in pulsating flow, Nature 174 (1954) 41. [12] N.H. Kim, B. Youn, R.L. Webb, Air-side heat transfer and friction correlations for plain fin-and-tube heat exchangers with staggered tube arrangements, J. Heat Transfer 121 (3) (1999) 662–667. [13] S.W. Churchill, M. Bernstein, A correlating equation for forced convection from gases and liquids to a circular cylinder in crossflow, J. Heat Transfer 99 (1977) 300–306. [14] N. Kattan, J.R. Thome, D. Favrat, Flow boiling in horizontal tubes: part 3 – development of a new heat transfer model based on flow pattern, J. Heat Transfer 120 (1998) 156–165. [15] L. Wojtan, T. Ursenbacher, J.H. Thome, Investigation of flow boiling in horizontal tubes: part II – development of a new heat transfer model for stratified-wavy, dryout and mist flow regimes, Int. J. Heat Mass Transf. 48 (2005) 2970–2985. [16] H. Muller-Steinhagen, K. Heck, A simple friction pressure drop correlation for two-phase flow in pipes, Chem. Eng. Process. 20 (1986) 297–308. [17] D. Steiner, Heat transfer to boiling saturated liquids, in: VDI-Wärmeatlas (VDI Heat Atlas), Verein Deutscher Ingenieure, VDI-Gesselschaft Verfahrenstechnic und Chemieingenieurswesen (GCV), Dusseldorf, 1993, pp. 15–24.