- Email: [email protected]
Effect of the Dryout in Tube Bundles on the Heat Transfer Performance of Falling Film Evaporators
Available online at www.sciencedirect.com
ScienceDirect
Available online at www.sciencedirect.com Procedia Engineering 00 (2017) 000–000
ScienceDirect
www.elsevier.com/locate/procedia
Procedia Engineering 205 (2017) 2176–2183
10th International Symposium on Heating, Ventilation and Air Conditioning, ISHVAC2017, 1922 October 2017, Jinan, China
Effect of the Dryout in Tube Bundles on the Heat Transfer Performance of Falling Film Evaporators Li Yanga,b,c,d,*, Xiaoman Songa,b,c, Yong Xiee b
a School of Thermal Engineering, Shandong Jianzhu University, Jinan, 250101, China Key Laboratory of Renewable Energy Technologies for Buildings, Ministry of Education, Jinan, 250101, China c Shandong Key Laboratory of Renewable Energy Technologies for Buildings, Jinan, 250101, China d Bright air conditioning Co.,Ltd, Dezhou, 253002, China e Shandong Provincial Architectural Design Institute, Jinan, 250001, China
Abstract In this paper, the dryout occurrence was discussed on the tube bundles in falling film evaporators for refrigeration. A numerical simulation was carried out on the bundles of horizontal tubes based on the distributed parameter method. In the simulation, the liquid flow rate and maldistribution on the tube bundles were analyzed about their influences on the dry patch formation and heat transfer performance of evaporators. These results are helpful to guide the design of falling film evaporators. © 2017 The Authors. Published by Elsevier Ltd. © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Air Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Conditioning. Air Conditioning. Keywords: Falling film; Evaporation; Dryout; Heat transfer; Tube bundle
1. Introduction Falling film evaporators with horizontal tubes have been widely used in refrigeration, chemical, and desalinization industries. They possess higher heat transfer coefficients and operate with lower liquid charge comparing with flooded evaporators. However, the dryout may occur on the outside of several evaporation tubes by trickling distribution of liquid refrigerant due to liquid maldistribution on the tube bundle or too little local liquid
* Corresponding author. Tel.: +86 531 86361236. E-mail address: [email protected] 1877-7058 © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Air Conditioning.
1877-7058 © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Air Conditioning. 10.1016/j.proeng.2017.10.041
2
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al. / Procedia Engineering 00 (2017) 000–000
2177
Nomenclature din tube inner diameter (m) do tube outside diameter (m) dor tube outside root diameter (m) x distance along the tube (m) Δx tube length of unit volume m mass flow rate (kg·s-1) Cp specific heat (J·kg-1·K-1) hfg latent heat of evaporation (J·kg-1) h specific enthalpy (J·kg-1) Nx grid number in the direction of tube axis Ny number of tube rows in tube bundles Nk number of tube columns in tube bundles Subscripts e evaporating i unit volume inlet o unit volume outlet in evaporator inlet cw chilled water cwi chilled water inlet cwo chilled water outlet w tube wall wi internal tube wall wo external tube wall refi refrigerant inlet refo refrigerant outlet supply, which could result in a sharp falloff in the evaporator performance. On the other hand, liquid oversupply to avoid the phenomena may bring about a waste of energy to pump back the unevaporated liquid to the distributor. Hence, the minimum quantity of liquid supply should be necessary to eliminate liquid recycling and keep well heat transfer performance in the falling film evaporator. In falling film evaporators, the liquid film flow rate is less and less from the top tube bundles due to continuous evaporation. On the other hand, the temperature of the water in the tube is high at the entrance. As a result, dry patches occur when the flow rate on the tube falls below a certain limit. Obviously, they reduce the effective areas wetted by the liquid film on the tube and lead to the overall heat transfer performance worsening. However, some studies as to surface dryout are performed for subcooled liquid films at non-boiling conditions and on vertical surfaces in the open literature [1,2,3]. Few studies are carried out on horizontal tube bundles in spite of the importance of dryout. G. Ribatski et al. [4] investigated experimentally the onset of local dryout of saturated HFC-134a falling film with nucleate boiling on a vertical array of horizontal plain tubes. J.F. Roques [5,6] studied the onset of dryout on a vertical row of horizontal plain and enhanced tubes of Turbo-BII HP, Gewa-B and HighFlux with saturated HFC-134a respectively. They proposed a criterion for predicting the onset of dryout based on their test results of falling film heat transfer. However, the dryout occurrence factors and their effects on falling film heat transfer performance are insufficient, and there is still a great deal for further study. In our study, numerical study is carried out with saturated HFC-134a on the effect of surface dryout on the heat transfer performance of the horizontal falling film evaporator with liquid trickling distribution system. A numerical simulation is performed with distributed parameters on tube bundles in falling film evaporators. The enhanced tube of Turbo-EHP is employed. In this simulation, the analysis mainly focuses on the dryout occurrence factors, such as liquid flow rate and maldistribution on tube bundles, and their influence on the heat transfer performance of falling film evaporators. 2. Mathematical model In general, the liquid flow rate decreases due to continuous evaporation as it flows down the bundle of horizontal tubes. The temperature of chilled water varies along the tubes due to heat exchanging with the liquid refrigerant.
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al./ Procedia Engineering 00 (2017) 000–000
2178
3
Thus, the heat transfer performance may be affected by different tubes in tube bundles. The calculation for the falling film evaporator is based on the following assumptions, 1) The heat exchanger is insulated from the surroundings; 2) The thermal properties of the refrigerant are constant; 3) The liquid refrigerant is saturated at the inlet of the evaporator; 4) The falling film evaporation is in steady state, and the influence of liquid film wave is negligible. a
1
2
...
3
...
Nx
1 2 3
j
b1
2
3
... ...
Nk
2 3
. . . . . .
. . . . .
j
Ny
Ny
i
1
k
Fig. 1 Grid structure employed in the finite different analysis (a)lateral view and (b)front view Refrigerant T ref mrefi Water T cwi mcw
T cwo mcw
Falling film tubes
Falling film tubes
T ref mrefo
Fig. 2 Unit volume of fluids and tube wall
Flooded tubes
Fig. 3 A schematic diagram of bottom-to-top tube row layout
Fig. 1 shows the grid in the finite difference analysis with the lateral and front views. The grid points about j and k point to the row and column in the tube bundles respectively, and the grid point i denotes each discrete element along the tube length direction. In each calculated unit, the relation between the fluids inside and outside the tube and tube wall is shown in Fig. 2. The energy equation in the discrete unit can be expressed as follows, For the chilled water inside the tube, m cw C p , cw
dT cwi = πd in α cwi (Twi − Tcwi ) dx
For the refrigerant outside the tube,
π d o Δ x α refi (T wo − T refi ) = ( m refi − m refo ) hfg
For the tube wall, mcw C p ,cw (Tcwi − Tcwo ) =
2πλw Δx(Twi − Two ) ln(d o d in )
(1) (2) (3)
In terms of the mass balance over the unit volume in the bottom-to-top two-tube-pass arrangement, the boundary condition can be expressed as follows: (4) Tref=Te, mrefi=mref,in/(Nk×Nx) for k=1, 2, …, Nk, and i=1, 2, …, Nx at j=1 In the above equations, αref is the heat transfer coefficient of falling liquid, and Kff is the falling film factor defined as [6], α ref (5) K ff = α
pb
where, αpb is the nucleate pool boiling heat transfer coefficient calculated at the same heat flux with αref. It can be obtained by Webb & Pais correlation [7] for enhanced tubes. Therefore, the falling film factor can be used to compare the pool boiling performance with the falling film one, which varies versus the film Reynolds number in a certain heat flux when some dry patches are formed on tube surfaces. So, Kff should be predicted at non-dryout and partial dryout cases, respectively. During operation, the liquid film around a tube probably breaks down if the liquid flow rate is not sufficient to wet the tube surface. The onset of dryout threshold Reynolds number, with which the falling film heat transfer coefficient starts to decrease sharply, is considered to be a linear function of the heat flux [6]: Re ref, threshold = 2(c ⋅ qo + d ) (6)
4
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al. / Procedia Engineering 00 (2017) 000–000
2179
where, the empirical constants c and d are 2.92×10-3 and 29.8 for the Turbo-EHP tube, respectively. When the film Reynolds number reaches the threshold Reynolds number at a certain heat flux (heat transfer without dryout), the falling film factor Kff,plateau can be calculated as below. q q tp K ff, plateau = (1 + b1 )(b2 + b3 ( o ) + b4 ( o ) 2 ) (7) tpo qcrit qcrit where, the tube pitch tp is nondimensionalized with the minimum tube pitch tpo=22.25 mm. For the enhanced tube of Turbo-EHP, the empirical constants b1~b4 are -0.361, 2.891, -16.314 and 59.906, respectively. The critical heat flux qcrit can be obtained from the Kutateladze correlation [8]. Therefore, the heat transfer coefficient for the falling liquid film without dryout αref can be obtained by multiplying Kff,plateau by αpb from Eq. (5). On the other hand, if the refrigerant flow is below the threshold Reynolds number, the falling film heat transfer coefficient decreases rapidly with the liquid flow rate due to the dry patches formation. In the case, the heat transfer coefficient will be calculated as below. 2
qo Re ref = Re ref, threshold qplateau Re ref, threshold = 2 c ⋅ qplateau + d
(
(8)
)
(9)
Then, q q2 2 qplateau − 2c o qplateau − 2d o = 0 (10) Re ref Re ref The falling film factor Kff with partial dryout can be predicted with the liquid film Reynolds number Reref approximately, K ff,plateau K ff = Re threshold
Re ref
(11)
Similarly, falling film heat transfer coefficient αref can be solved by multiplying Kff, by αpb from Eq. (5) when there is partial dryout on tube surfaces. Therefore, the dry patches areas on evaporation tubes Adryout can be predicted by Eq. (12), Adryout = ndryoutπd o Δx (12)
In this equation, ndryout is the grid point number of falling film evaporation tubes on which the dryout occurred. Usually, the falling film evaporator is operated in semi-falling film and semi-flooded conditions in order to keep a well overall performance. The lower portion of the falling film evaporator shell is flooded with liquid refrigerant to wet the lower tubes while the tubes in the upper portion are wetted only by liquid refrigerant from the trickling distributor. Therefore, the percentage of dry patches areas in outside surface of all falling film tubes is, Pdryout=(Adryout/Afalling)×100 (13) where, Afalling is the outside surface areas of all falling film tubes in falling film evaporators. The usual internal heat transfer coefficient αcw can be expressed as below, α cw =
Ci λcw 1 3 μ cw 0.14 Re 0cw.8 Prcw ( ) d in μw
(14)
where Ci is a coefficient related to the internal tube surface structure [9]. The overall heat transfer coefficient based on the outside surface area of the tubes is, 1 (15) U= 1 do
α cw d in
+
do d d 1 ln or + + Rfi o 2λw d in α ref d in
where Rfi ( m2 ·℃·W-1) is the fouling resistance of surfaces inside the tube. Here, its order of magnitude is 10-5 m2· ℃·W-1 in terms of the ARI Standard 550/590-98[10]. In falling film evaporators, there is usually two-tube-pass on the side of chilled water. In this simulation, the bottom-to-top tube pass arrangement is adopted, as shown in Fig. 3. Turbo-EHP tube is employed with outer and inside diameters of 19.05 mm and 16.54 mm respectively, and its length is 3.97 m and the number of total evaporation tubes is 236. It is assumed that the refrigerant HFC-134a
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al./ Procedia Engineering 00 (2017) 000–000
2180
5
evaporates at 6 ℃, chilled water mass flow rate inside the tubes is 52.49 kg·s-1 and its inlet and outlet temperatures are 12 ℃ and 7 ℃, respectively. 3. Simulation and discussion 3.1. The effect of mass flow rate
1.60
1130
1.50
1110
1.40
1090
1.30
1070
1.20
1050
1.10
Total heat transfer rate Falling film factor
1030 1010
50.0
Percentage of dry patches areas(%)
1150
Falling film factor
Total heat transfer rate(kW)
Fig. 4 shows the effect of liquid refrigerant mass flow rate to the top row of the tube bundle on the overall heat transfer performance of evaporators. It can be observed that the total heat transfer rate at first increases and then almost keeps constant with the increase of the refrigerant flow rate. Thus, the falling film heat transfer performance can be enhanced by increasing the refrigerant flow rates when they are in low values. The falling film factor is smaller than 1.0 when the refrigerant flow rate is lower than 3.1 kg·s-1 approximately. Therefore, the falling film heat transfer performance is much greater than the flooded performance when the refrigerant flow rate is higher than a certain value, while the performance is worse than the flooded one when liquid refrigerant flow rate is lower than this value at the same heat flux. The variation of the percentage of dry patches areas in outside surface areas is displayed in Fig. 5. Dry patches areas on falling film tubes gradually decrease and then nearly maintain a value of zero with the increasing refrigerant flow rate in a certain tube bundle arrangement. The percentage of dry patches areas corresponds to the total heat transfer rate at the same refrigerant flow rates, and they present contrary trends with liquid refrigerant flow rate. It is indicated that the heat transfer performance of evaporators can be enhanced by increasing the refrigerant flow rate within a certain range due to the decreasing dry patches areas on falling film tubes. Once the dry patches decrease to the minimum value of zero, the evaporators just reach the best thermal performance.
1.00 0.90
2.0
3.0
4.0 5.0 6.0
7.0
30.0
20.0
10.0
0.0
8.0 9.0 10.0 11.0 12.0 13.0
2.0
Refrigerant mass flow rate(kg/s)
Fig. 4 Variations of total heat transfer rate and average falling film factor with refrigerant mass flow rate
40.0
3.0
4.0
5.0
6.0
7.0
8.0
9.0 10.0 11.0 12.0 13.0
Refrigerant mass flow rate(kg/s)
Fig. 5 Percentage of dry patches areas in the outside surface areas of all falling film tubes versus refrigerant mass flow rate
3.2. The effect of liquid maldistribution on the tube bundle The actual maldistribution of liquid refrigerant on the tube bundle usually leads to locally uneven among the tubes in the falling film evaporator. The distributor plays an important role in liquid refrigerant distribution, and its effect on heat transfer efficiency of the evaporator is analyzed in the following.
a
b
Fig. 6 Schematic drawing of liquid maldistribution on the undersurface of distributor (a) along the width direction, and (b) along the length direction
Fig. 6 shows the schematic drawing of liquid maldistribution along the width direction (i.e. k direction) and length direction (i.e. i direction) on the undersurface of distributor, where the filled grids denote the local zone of the decrease of liquid refrigerant flow rate compared with that in the uniform distribution conditions while the blank grids denote the local zone of the increase of liquid refrigerant flow rate due to the liquid maldistribution.
6
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al. / Procedia Engineering 00 (2017) 000–000
2181
Liquid maldistribution along the tube bundle width For analyzing the flow maldistribution, a parameter ΦW is defined to designate liquid refrigerant maldistribution on the tube bundle along its width. ΦW =
mmean − mmiddle- W 2mmean
(16)
where, mmiddle-W is the local mass flow rate of liquid refrigerant on the top of the middle tube array along the tube bundle width direction under non-uniform liquid flow conditions, and it is uniformly distributed onto the local middle tube array as shown in Fig. 6(a), and mmean is the mass flow rate of liquid refrigerant flowing onto the top of the same middle tube array as mmiddle-W in the tube bundle under uniform liquid flow conditions. Liquid flow maldistribution has an important effect on the overall characteristics of evaporators. The variations of the heat transfer capacity and average falling film factor of the evaporator are shown in Fig. 7. The heat transfer capacity and falling film factor decrease with the increasing ΦW. Furthermore, the falling film evaporation performance is worse than the flooded evaporation at the same heat flux when ΦW is larger than 0.35. This is mainly due to the increasing dry patches on falling film tubes as shown in Fig. 8. 1150
1070
1.3
1050 1030
1.2
1010
1.1
990
1.0
970 950
Total heat transfer rate
930
Falling film factor
Falling film factor K ff
1.4
1090
0.9
Percentage of dry patches areas (%)
1.5
1110 Total heat transfer rate (kW)
60
1.6
1130
50
40
30
20
10
0.8
910 890
0
0.7 0
10
20
Φ W (%)
30
40
0
50
10
20
30
40
50
Φ W (%)
Fig. 7 Total heat transfer rate and average falling film factor profiles versus ΦW Fig. 8 Percentage of dry patches areas in outside surface areas of all falling film tubes versus ΦW
tube(2,14,1) tube(2,14,4)
tube(1,6,1)
tube(1,6,4)
Fig. 9 Locations of the four tubes chosen in the tube bundle
In addition, liquid flow maldistribution greatly influences the heat transfer performance of tubes at different locations in tube bundles. Four tubes are chosen in the tube bundle as shown in Fig. 9. Fig. 10(a) displays the influence of flow maldistribution on the chilled water temperature in the tube (1, 6, 4). It can be observed that the chilled water temperature gradually increases with ΦW ranging from 0 to 0.45. Furthermore, the temperature of the chilled water almost keeps constant in this tube when ΦW is larger than 0.35. Therefore, the more the ΦW, the more dry patches appear on this tube, it makes the heat transfer performance drop. However, the chilled water temperature drops with the ΦW increasing in the tube (1, 6, 1) as shown in Fig. 10(b). It is indicated that the dry patches areas are slightly reduced, and the heat transfer capacity is raised between chilled water and liquid refrigerant with the increase of ΦW. It is due to the flow maldistribution that the refrigerant flow rate is slightly increased on this tube surface compared with uniform flow distribution. In addition, in uniform flow distribution the partial dry patches are formed on the tube (1, 6, 4) and tube (1, 6, 1) surfaces.
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al./ Procedia Engineering 00 (2017) 000–000
2182
b
a 12.5 12.0
The temperature of chilled water (℃)
The temperature of chilled water (℃)
Φw=10% Φw=25% Φw=40%
Φw=15% Φw=30% Φw=45%
11.0
11.0
10.5
10.5
10.0
10.0 9.5 9.0 8.5
Φw=0 Φw=20% Φw=35%
8.0 7.5 0
1
Φw=10% Φw=25% Φw=40%
Φw=15% Φw=30% Φw=45%
2 3 4 5 6 Grid point in the direction of tube length
9.5 9.0 8.5 8.0 7.5 0
7
d
9.5
1
2 3 4 5 6 Grid point in the direction of tube length
7
9.5 9.0
The temperature of chilled water (℃)
9.0 The temperature of chilled water (℃)
Φw=0 Φw=20% Φw=35%
12.0 11.5
11.5
c
12.5
7
8.5 8.0 7.5 7.0 Φw=0 Φw=20% Φw=35%
6.5
Φw=10% Φw=25% Φw=40%
Φw=15% Φw=30% Φw=45%
8.5 8.0 7.5 7.0 Φw=0 Φw=20% Φw=35%
6.5
Φw=10% Φw=25% Φw=40%
Φw=15% Φw=30% Φw=45%
6.0
6.0 0
1
2 3 4 5 6 Grid point in the direction of tube length
7
0
1
2 3 4 5 6 Grid point in the direction of tube length
7
Fig. 10 Chilled water temperature distribution in different tubes (a) in the tube (1, 6, 4), (b) in the tube (1, 6, 1), (c) in the tube (2, 14, 4), and (d) in the tube (2, 14, 1)
Figs. 10(c) and (d) show the effect of flow maldistribution on the temperature of chilled water in two tubes in the second tube pass. The flow maldistribution makes against the thermal performance, and moreover, the deterioration of the thermal performance depends critically on the tube location in the whole tube row. The effect on the tube (2, 14, 4) is greater than that on the tube (2, 14, 1). By comparison of the calculated results as shown in Figs. 10 (a) and (c), the dry patches areas on the tube (1, 6, 4) are larger than that on the tube (2, 14, 4) in flow maldistribution conditions, and the deterioration of heat transfer on the tube (1, 6, 4) is greater than that on the tube (2, 14, 4) as well. It is indicated that the dry patches areas on the lower tubes are larger than the upper ones in tube bundles when the refrigerant flows unequally over tube bundle in real operation. Therefore, the flow maldistribution influence on the lower tubes is greater, and the heat transfer deterioration on them is greater than the upper ones as well. Liquid maldistribution along the tube bundle length The flow uniformity greatly influences the performance of evaporation on the tube bundle along its length as well. Comparing to the flow maldistribution along the tube bundle width, a similar factor ΦL is introduced to analyze flow maldistribution along the evaporator axis direction, m − mmiddle-L (17) ΦL = mean 2mmean where, mmiddle-L is the local mass flow rate of liquid refrigerant flowing at the top and middle along the tube length in the whole tube bundles, and it is uniformly distributed onto the local tube bundle location, that is, mmiddle-L is the mass flow rate of the liquid refrigerant flowing through the filled grids as shown in Fig. 6(b) for non-uniform liquid flow, and mmean is the mass flow rate of liquid refrigerant flowing at the top of the same tube length as mmiddle-L in the tube bundle for uniform refrigerant flow conditions. In the calculated results, comparing to the liquid maldistribution along the tube bundle width, a similar thermal performance can be obtained to the effects of flow maldistribution ΦL, as shown in Fig. 11. The total heat transfer rate decreases with ΦL increasing due to the increasing dryout areas in flow maldistribution conditions along the evaporator axis direction.
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al. / Procedia Engineering 00 (2017) 000–000 1160
55
1140
50 45
Total heat transfer rate(kW)
1120
40
1100
35
Total heat transfer rate
1080
30
Percentage of dry patches areas
25
1060
20
1040
15
1020
10
1000
2183
Percentage of dry patches areas (%)
8
5
980
0 0
10
20
Φ L (%)
30
40
50
Fig. 11 Total heat transfer rate and percentage of dry patches areas in the outside surface areas of all falling film tubes versus ΦL
4. Conclusions • According to the simulated results, the capacity of heat transfer has decreased by 12% while the dry patches have reached half heat transfer surface of falling film tubes in the tube bundle. The occurrence of dry patches is mainly because of the less liquid refrigerant flow rate and the nonuniform liquid distribution on the tube bundles in practical operation. • At low refrigerant flow rates, the overall heat transfer performance is enhanced by increasing the flow rate, and the dry patch percentage in the total falling film tube surfaces is decreased at the same time. While the refrigerant flow rate rises above a certain value, the heat transfer performance hardly varies with the flow rate due to dry patches being too small. • The refrigerant flow maldistribution can lead to local dryout, and affect the evaporation heat transfer of the falling film evaporator. The capacity of heat transfer deteriorates with the extending degree of the refrigerant nonuniform flow. Moreover, the liquid maldistribution has different effects on heat transfer performances of the tubes at different locations in the tube bundle, and tubes at lower rows present worse heat transfer performance than the upper ones. Acknowledgements The authors gratefully acknowledge the support from the Doctoral Fund Project of Shandong Jianzhu University (XNBS1224), and Wang Liqiu Overseas Taishan Scholar Research Group on Heat and Moisture Transfer.
References [1] B.X. Wang, J.T. Zhang, X.F. Peng, Experimental study on the dryout heat flux of falling liquid film. International Journal of Heat and Mass Transfer, 43 (2000) 1897-1903. [2] M.S. Bohn, S.H. Davis, Thermocapillary breakdown of falling liquid films at high Reynolds numbers. International Journal of Heat and Mass Transfer, 36 (1993) 1875-1881. [3] Tang Zhiwei, Yang Xiaoke, Jiang Zhangyan, Experimental study on surface wave and film breakdown of falling liquid film flow. Heat Mass Transfer, 45 (2009) 673-677. [4] Gherhardt Ribatski, John R. Thome, Experimental study on the onset of local dryout in an evaporating falling film on horizontal plain tubes. Experimental Thermal & Fluid Science, 31 (2007) 483-493. [5] J.F. Roques, Falling film evaporation on a single tube and on a tube bundle, Ph.D. thesis, École Polytechnique Fédérale de Lausanne, Switzerland, 2004. [6] J.F. Roques and J. R. Thome, Falling films on arrays of horizontal tubes with R-134a, Part II: flow visualization, onset of dryout, and heat transfer predictions. Heat Transfer Engineering 28(5) (2007), 415-434. [7] Webb R L, Pals C., Nucleate boiling data for five refrigerants on plain, integral-fin and enhanced tube geometries. Int. J. Heat and Mass Transfer, 35(8) (1992) 1893-1904. [8] Kutateladze, S. S. On the transition to film boiling under nature convection. Kotloturbostoenie, 3(1948) 10-15. [9] Bell K J, Muller A C. Wolverine engineering data book II [M]. Wolverine Tube Co, 2001. [10] ARI Standard 550/590-98. Standard for water chilling packages using the vapor compression cycle, 1998, Air-Conditioning and Refrigeration Institute, 4301 North Fairfax Drive, Suite 425, Arlington, Va. 22203, U.S.A.
ScienceDirect
Available online at www.sciencedirect.com Procedia Engineering 00 (2017) 000–000
ScienceDirect
www.elsevier.com/locate/procedia
Procedia Engineering 205 (2017) 2176–2183
10th International Symposium on Heating, Ventilation and Air Conditioning, ISHVAC2017, 1922 October 2017, Jinan, China
Effect of the Dryout in Tube Bundles on the Heat Transfer Performance of Falling Film Evaporators Li Yanga,b,c,d,*, Xiaoman Songa,b,c, Yong Xiee b
a School of Thermal Engineering, Shandong Jianzhu University, Jinan, 250101, China Key Laboratory of Renewable Energy Technologies for Buildings, Ministry of Education, Jinan, 250101, China c Shandong Key Laboratory of Renewable Energy Technologies for Buildings, Jinan, 250101, China d Bright air conditioning Co.,Ltd, Dezhou, 253002, China e Shandong Provincial Architectural Design Institute, Jinan, 250001, China
Abstract In this paper, the dryout occurrence was discussed on the tube bundles in falling film evaporators for refrigeration. A numerical simulation was carried out on the bundles of horizontal tubes based on the distributed parameter method. In the simulation, the liquid flow rate and maldistribution on the tube bundles were analyzed about their influences on the dry patch formation and heat transfer performance of evaporators. These results are helpful to guide the design of falling film evaporators. © 2017 The Authors. Published by Elsevier Ltd. © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Air Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Conditioning. Air Conditioning. Keywords: Falling film; Evaporation; Dryout; Heat transfer; Tube bundle
1. Introduction Falling film evaporators with horizontal tubes have been widely used in refrigeration, chemical, and desalinization industries. They possess higher heat transfer coefficients and operate with lower liquid charge comparing with flooded evaporators. However, the dryout may occur on the outside of several evaporation tubes by trickling distribution of liquid refrigerant due to liquid maldistribution on the tube bundle or too little local liquid
* Corresponding author. Tel.: +86 531 86361236. E-mail address: [email protected] 1877-7058 © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Air Conditioning.
1877-7058 © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Air Conditioning. 10.1016/j.proeng.2017.10.041
2
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al. / Procedia Engineering 00 (2017) 000–000
2177
Nomenclature din tube inner diameter (m) do tube outside diameter (m) dor tube outside root diameter (m) x distance along the tube (m) Δx tube length of unit volume m mass flow rate (kg·s-1) Cp specific heat (J·kg-1·K-1) hfg latent heat of evaporation (J·kg-1) h specific enthalpy (J·kg-1) Nx grid number in the direction of tube axis Ny number of tube rows in tube bundles Nk number of tube columns in tube bundles Subscripts e evaporating i unit volume inlet o unit volume outlet in evaporator inlet cw chilled water cwi chilled water inlet cwo chilled water outlet w tube wall wi internal tube wall wo external tube wall refi refrigerant inlet refo refrigerant outlet supply, which could result in a sharp falloff in the evaporator performance. On the other hand, liquid oversupply to avoid the phenomena may bring about a waste of energy to pump back the unevaporated liquid to the distributor. Hence, the minimum quantity of liquid supply should be necessary to eliminate liquid recycling and keep well heat transfer performance in the falling film evaporator. In falling film evaporators, the liquid film flow rate is less and less from the top tube bundles due to continuous evaporation. On the other hand, the temperature of the water in the tube is high at the entrance. As a result, dry patches occur when the flow rate on the tube falls below a certain limit. Obviously, they reduce the effective areas wetted by the liquid film on the tube and lead to the overall heat transfer performance worsening. However, some studies as to surface dryout are performed for subcooled liquid films at non-boiling conditions and on vertical surfaces in the open literature [1,2,3]. Few studies are carried out on horizontal tube bundles in spite of the importance of dryout. G. Ribatski et al. [4] investigated experimentally the onset of local dryout of saturated HFC-134a falling film with nucleate boiling on a vertical array of horizontal plain tubes. J.F. Roques [5,6] studied the onset of dryout on a vertical row of horizontal plain and enhanced tubes of Turbo-BII HP, Gewa-B and HighFlux with saturated HFC-134a respectively. They proposed a criterion for predicting the onset of dryout based on their test results of falling film heat transfer. However, the dryout occurrence factors and their effects on falling film heat transfer performance are insufficient, and there is still a great deal for further study. In our study, numerical study is carried out with saturated HFC-134a on the effect of surface dryout on the heat transfer performance of the horizontal falling film evaporator with liquid trickling distribution system. A numerical simulation is performed with distributed parameters on tube bundles in falling film evaporators. The enhanced tube of Turbo-EHP is employed. In this simulation, the analysis mainly focuses on the dryout occurrence factors, such as liquid flow rate and maldistribution on tube bundles, and their influence on the heat transfer performance of falling film evaporators. 2. Mathematical model In general, the liquid flow rate decreases due to continuous evaporation as it flows down the bundle of horizontal tubes. The temperature of chilled water varies along the tubes due to heat exchanging with the liquid refrigerant.
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al./ Procedia Engineering 00 (2017) 000–000
2178
3
Thus, the heat transfer performance may be affected by different tubes in tube bundles. The calculation for the falling film evaporator is based on the following assumptions, 1) The heat exchanger is insulated from the surroundings; 2) The thermal properties of the refrigerant are constant; 3) The liquid refrigerant is saturated at the inlet of the evaporator; 4) The falling film evaporation is in steady state, and the influence of liquid film wave is negligible. a
1
2
...
3
...
Nx
1 2 3
j
b1
2
3
... ...
Nk
2 3
. . . . . .
. . . . .
j
Ny
Ny
i
1
k
Fig. 1 Grid structure employed in the finite different analysis (a)lateral view and (b)front view Refrigerant T ref mrefi Water T cwi mcw
T cwo mcw
Falling film tubes
Falling film tubes
T ref mrefo
Fig. 2 Unit volume of fluids and tube wall
Flooded tubes
Fig. 3 A schematic diagram of bottom-to-top tube row layout
Fig. 1 shows the grid in the finite difference analysis with the lateral and front views. The grid points about j and k point to the row and column in the tube bundles respectively, and the grid point i denotes each discrete element along the tube length direction. In each calculated unit, the relation between the fluids inside and outside the tube and tube wall is shown in Fig. 2. The energy equation in the discrete unit can be expressed as follows, For the chilled water inside the tube, m cw C p , cw
dT cwi = πd in α cwi (Twi − Tcwi ) dx
For the refrigerant outside the tube,
π d o Δ x α refi (T wo − T refi ) = ( m refi − m refo ) hfg
For the tube wall, mcw C p ,cw (Tcwi − Tcwo ) =
2πλw Δx(Twi − Two ) ln(d o d in )
(1) (2) (3)
In terms of the mass balance over the unit volume in the bottom-to-top two-tube-pass arrangement, the boundary condition can be expressed as follows: (4) Tref=Te, mrefi=mref,in/(Nk×Nx) for k=1, 2, …, Nk, and i=1, 2, …, Nx at j=1 In the above equations, αref is the heat transfer coefficient of falling liquid, and Kff is the falling film factor defined as [6], α ref (5) K ff = α
pb
where, αpb is the nucleate pool boiling heat transfer coefficient calculated at the same heat flux with αref. It can be obtained by Webb & Pais correlation [7] for enhanced tubes. Therefore, the falling film factor can be used to compare the pool boiling performance with the falling film one, which varies versus the film Reynolds number in a certain heat flux when some dry patches are formed on tube surfaces. So, Kff should be predicted at non-dryout and partial dryout cases, respectively. During operation, the liquid film around a tube probably breaks down if the liquid flow rate is not sufficient to wet the tube surface. The onset of dryout threshold Reynolds number, with which the falling film heat transfer coefficient starts to decrease sharply, is considered to be a linear function of the heat flux [6]: Re ref, threshold = 2(c ⋅ qo + d ) (6)
4
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al. / Procedia Engineering 00 (2017) 000–000
2179
where, the empirical constants c and d are 2.92×10-3 and 29.8 for the Turbo-EHP tube, respectively. When the film Reynolds number reaches the threshold Reynolds number at a certain heat flux (heat transfer without dryout), the falling film factor Kff,plateau can be calculated as below. q q tp K ff, plateau = (1 + b1 )(b2 + b3 ( o ) + b4 ( o ) 2 ) (7) tpo qcrit qcrit where, the tube pitch tp is nondimensionalized with the minimum tube pitch tpo=22.25 mm. For the enhanced tube of Turbo-EHP, the empirical constants b1~b4 are -0.361, 2.891, -16.314 and 59.906, respectively. The critical heat flux qcrit can be obtained from the Kutateladze correlation [8]. Therefore, the heat transfer coefficient for the falling liquid film without dryout αref can be obtained by multiplying Kff,plateau by αpb from Eq. (5). On the other hand, if the refrigerant flow is below the threshold Reynolds number, the falling film heat transfer coefficient decreases rapidly with the liquid flow rate due to the dry patches formation. In the case, the heat transfer coefficient will be calculated as below. 2
qo Re ref = Re ref, threshold qplateau Re ref, threshold = 2 c ⋅ qplateau + d
(
(8)
)
(9)
Then, q q2 2 qplateau − 2c o qplateau − 2d o = 0 (10) Re ref Re ref The falling film factor Kff with partial dryout can be predicted with the liquid film Reynolds number Reref approximately, K ff,plateau K ff = Re threshold
Re ref
(11)
Similarly, falling film heat transfer coefficient αref can be solved by multiplying Kff, by αpb from Eq. (5) when there is partial dryout on tube surfaces. Therefore, the dry patches areas on evaporation tubes Adryout can be predicted by Eq. (12), Adryout = ndryoutπd o Δx (12)
In this equation, ndryout is the grid point number of falling film evaporation tubes on which the dryout occurred. Usually, the falling film evaporator is operated in semi-falling film and semi-flooded conditions in order to keep a well overall performance. The lower portion of the falling film evaporator shell is flooded with liquid refrigerant to wet the lower tubes while the tubes in the upper portion are wetted only by liquid refrigerant from the trickling distributor. Therefore, the percentage of dry patches areas in outside surface of all falling film tubes is, Pdryout=(Adryout/Afalling)×100 (13) where, Afalling is the outside surface areas of all falling film tubes in falling film evaporators. The usual internal heat transfer coefficient αcw can be expressed as below, α cw =
Ci λcw 1 3 μ cw 0.14 Re 0cw.8 Prcw ( ) d in μw
(14)
where Ci is a coefficient related to the internal tube surface structure [9]. The overall heat transfer coefficient based on the outside surface area of the tubes is, 1 (15) U= 1 do
α cw d in
+
do d d 1 ln or + + Rfi o 2λw d in α ref d in
where Rfi ( m2 ·℃·W-1) is the fouling resistance of surfaces inside the tube. Here, its order of magnitude is 10-5 m2· ℃·W-1 in terms of the ARI Standard 550/590-98[10]. In falling film evaporators, there is usually two-tube-pass on the side of chilled water. In this simulation, the bottom-to-top tube pass arrangement is adopted, as shown in Fig. 3. Turbo-EHP tube is employed with outer and inside diameters of 19.05 mm and 16.54 mm respectively, and its length is 3.97 m and the number of total evaporation tubes is 236. It is assumed that the refrigerant HFC-134a
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al./ Procedia Engineering 00 (2017) 000–000
2180
5
evaporates at 6 ℃, chilled water mass flow rate inside the tubes is 52.49 kg·s-1 and its inlet and outlet temperatures are 12 ℃ and 7 ℃, respectively. 3. Simulation and discussion 3.1. The effect of mass flow rate
1.60
1130
1.50
1110
1.40
1090
1.30
1070
1.20
1050
1.10
Total heat transfer rate Falling film factor
1030 1010
50.0
Percentage of dry patches areas(%)
1150
Falling film factor
Total heat transfer rate(kW)
Fig. 4 shows the effect of liquid refrigerant mass flow rate to the top row of the tube bundle on the overall heat transfer performance of evaporators. It can be observed that the total heat transfer rate at first increases and then almost keeps constant with the increase of the refrigerant flow rate. Thus, the falling film heat transfer performance can be enhanced by increasing the refrigerant flow rates when they are in low values. The falling film factor is smaller than 1.0 when the refrigerant flow rate is lower than 3.1 kg·s-1 approximately. Therefore, the falling film heat transfer performance is much greater than the flooded performance when the refrigerant flow rate is higher than a certain value, while the performance is worse than the flooded one when liquid refrigerant flow rate is lower than this value at the same heat flux. The variation of the percentage of dry patches areas in outside surface areas is displayed in Fig. 5. Dry patches areas on falling film tubes gradually decrease and then nearly maintain a value of zero with the increasing refrigerant flow rate in a certain tube bundle arrangement. The percentage of dry patches areas corresponds to the total heat transfer rate at the same refrigerant flow rates, and they present contrary trends with liquid refrigerant flow rate. It is indicated that the heat transfer performance of evaporators can be enhanced by increasing the refrigerant flow rate within a certain range due to the decreasing dry patches areas on falling film tubes. Once the dry patches decrease to the minimum value of zero, the evaporators just reach the best thermal performance.
1.00 0.90
2.0
3.0
4.0 5.0 6.0
7.0
30.0
20.0
10.0
0.0
8.0 9.0 10.0 11.0 12.0 13.0
2.0
Refrigerant mass flow rate(kg/s)
Fig. 4 Variations of total heat transfer rate and average falling film factor with refrigerant mass flow rate
40.0
3.0
4.0
5.0
6.0
7.0
8.0
9.0 10.0 11.0 12.0 13.0
Refrigerant mass flow rate(kg/s)
Fig. 5 Percentage of dry patches areas in the outside surface areas of all falling film tubes versus refrigerant mass flow rate
3.2. The effect of liquid maldistribution on the tube bundle The actual maldistribution of liquid refrigerant on the tube bundle usually leads to locally uneven among the tubes in the falling film evaporator. The distributor plays an important role in liquid refrigerant distribution, and its effect on heat transfer efficiency of the evaporator is analyzed in the following.
a
b
Fig. 6 Schematic drawing of liquid maldistribution on the undersurface of distributor (a) along the width direction, and (b) along the length direction
Fig. 6 shows the schematic drawing of liquid maldistribution along the width direction (i.e. k direction) and length direction (i.e. i direction) on the undersurface of distributor, where the filled grids denote the local zone of the decrease of liquid refrigerant flow rate compared with that in the uniform distribution conditions while the blank grids denote the local zone of the increase of liquid refrigerant flow rate due to the liquid maldistribution.
6
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al. / Procedia Engineering 00 (2017) 000–000
2181
Liquid maldistribution along the tube bundle width For analyzing the flow maldistribution, a parameter ΦW is defined to designate liquid refrigerant maldistribution on the tube bundle along its width. ΦW =
mmean − mmiddle- W 2mmean
(16)
where, mmiddle-W is the local mass flow rate of liquid refrigerant on the top of the middle tube array along the tube bundle width direction under non-uniform liquid flow conditions, and it is uniformly distributed onto the local middle tube array as shown in Fig. 6(a), and mmean is the mass flow rate of liquid refrigerant flowing onto the top of the same middle tube array as mmiddle-W in the tube bundle under uniform liquid flow conditions. Liquid flow maldistribution has an important effect on the overall characteristics of evaporators. The variations of the heat transfer capacity and average falling film factor of the evaporator are shown in Fig. 7. The heat transfer capacity and falling film factor decrease with the increasing ΦW. Furthermore, the falling film evaporation performance is worse than the flooded evaporation at the same heat flux when ΦW is larger than 0.35. This is mainly due to the increasing dry patches on falling film tubes as shown in Fig. 8. 1150
1070
1.3
1050 1030
1.2
1010
1.1
990
1.0
970 950
Total heat transfer rate
930
Falling film factor
Falling film factor K ff
1.4
1090
0.9
Percentage of dry patches areas (%)
1.5
1110 Total heat transfer rate (kW)
60
1.6
1130
50
40
30
20
10
0.8
910 890
0
0.7 0
10
20
Φ W (%)
30
40
0
50
10
20
30
40
50
Φ W (%)
Fig. 7 Total heat transfer rate and average falling film factor profiles versus ΦW Fig. 8 Percentage of dry patches areas in outside surface areas of all falling film tubes versus ΦW
tube(2,14,1) tube(2,14,4)
tube(1,6,1)
tube(1,6,4)
Fig. 9 Locations of the four tubes chosen in the tube bundle
In addition, liquid flow maldistribution greatly influences the heat transfer performance of tubes at different locations in tube bundles. Four tubes are chosen in the tube bundle as shown in Fig. 9. Fig. 10(a) displays the influence of flow maldistribution on the chilled water temperature in the tube (1, 6, 4). It can be observed that the chilled water temperature gradually increases with ΦW ranging from 0 to 0.45. Furthermore, the temperature of the chilled water almost keeps constant in this tube when ΦW is larger than 0.35. Therefore, the more the ΦW, the more dry patches appear on this tube, it makes the heat transfer performance drop. However, the chilled water temperature drops with the ΦW increasing in the tube (1, 6, 1) as shown in Fig. 10(b). It is indicated that the dry patches areas are slightly reduced, and the heat transfer capacity is raised between chilled water and liquid refrigerant with the increase of ΦW. It is due to the flow maldistribution that the refrigerant flow rate is slightly increased on this tube surface compared with uniform flow distribution. In addition, in uniform flow distribution the partial dry patches are formed on the tube (1, 6, 4) and tube (1, 6, 1) surfaces.
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al./ Procedia Engineering 00 (2017) 000–000
2182
b
a 12.5 12.0
The temperature of chilled water (℃)
The temperature of chilled water (℃)
Φw=10% Φw=25% Φw=40%
Φw=15% Φw=30% Φw=45%
11.0
11.0
10.5
10.5
10.0
10.0 9.5 9.0 8.5
Φw=0 Φw=20% Φw=35%
8.0 7.5 0
1
Φw=10% Φw=25% Φw=40%
Φw=15% Φw=30% Φw=45%
2 3 4 5 6 Grid point in the direction of tube length
9.5 9.0 8.5 8.0 7.5 0
7
d
9.5
1
2 3 4 5 6 Grid point in the direction of tube length
7
9.5 9.0
The temperature of chilled water (℃)
9.0 The temperature of chilled water (℃)
Φw=0 Φw=20% Φw=35%
12.0 11.5
11.5
c
12.5
7
8.5 8.0 7.5 7.0 Φw=0 Φw=20% Φw=35%
6.5
Φw=10% Φw=25% Φw=40%
Φw=15% Φw=30% Φw=45%
8.5 8.0 7.5 7.0 Φw=0 Φw=20% Φw=35%
6.5
Φw=10% Φw=25% Φw=40%
Φw=15% Φw=30% Φw=45%
6.0
6.0 0
1
2 3 4 5 6 Grid point in the direction of tube length
7
0
1
2 3 4 5 6 Grid point in the direction of tube length
7
Fig. 10 Chilled water temperature distribution in different tubes (a) in the tube (1, 6, 4), (b) in the tube (1, 6, 1), (c) in the tube (2, 14, 4), and (d) in the tube (2, 14, 1)
Figs. 10(c) and (d) show the effect of flow maldistribution on the temperature of chilled water in two tubes in the second tube pass. The flow maldistribution makes against the thermal performance, and moreover, the deterioration of the thermal performance depends critically on the tube location in the whole tube row. The effect on the tube (2, 14, 4) is greater than that on the tube (2, 14, 1). By comparison of the calculated results as shown in Figs. 10 (a) and (c), the dry patches areas on the tube (1, 6, 4) are larger than that on the tube (2, 14, 4) in flow maldistribution conditions, and the deterioration of heat transfer on the tube (1, 6, 4) is greater than that on the tube (2, 14, 4) as well. It is indicated that the dry patches areas on the lower tubes are larger than the upper ones in tube bundles when the refrigerant flows unequally over tube bundle in real operation. Therefore, the flow maldistribution influence on the lower tubes is greater, and the heat transfer deterioration on them is greater than the upper ones as well. Liquid maldistribution along the tube bundle length The flow uniformity greatly influences the performance of evaporation on the tube bundle along its length as well. Comparing to the flow maldistribution along the tube bundle width, a similar factor ΦL is introduced to analyze flow maldistribution along the evaporator axis direction, m − mmiddle-L (17) ΦL = mean 2mmean where, mmiddle-L is the local mass flow rate of liquid refrigerant flowing at the top and middle along the tube length in the whole tube bundles, and it is uniformly distributed onto the local tube bundle location, that is, mmiddle-L is the mass flow rate of the liquid refrigerant flowing through the filled grids as shown in Fig. 6(b) for non-uniform liquid flow, and mmean is the mass flow rate of liquid refrigerant flowing at the top of the same tube length as mmiddle-L in the tube bundle for uniform refrigerant flow conditions. In the calculated results, comparing to the liquid maldistribution along the tube bundle width, a similar thermal performance can be obtained to the effects of flow maldistribution ΦL, as shown in Fig. 11. The total heat transfer rate decreases with ΦL increasing due to the increasing dryout areas in flow maldistribution conditions along the evaporator axis direction.
Li Yang et al. / Procedia Engineering 205 (2017) 2176–2183 Li Yang et al. / Procedia Engineering 00 (2017) 000–000 1160
55
1140
50 45
Total heat transfer rate(kW)
1120
40
1100
35
Total heat transfer rate
1080
30
Percentage of dry patches areas
25
1060
20
1040
15
1020
10
1000
2183
Percentage of dry patches areas (%)
8
5
980
0 0
10
20
Φ L (%)
30
40
50
Fig. 11 Total heat transfer rate and percentage of dry patches areas in the outside surface areas of all falling film tubes versus ΦL
4. Conclusions • According to the simulated results, the capacity of heat transfer has decreased by 12% while the dry patches have reached half heat transfer surface of falling film tubes in the tube bundle. The occurrence of dry patches is mainly because of the less liquid refrigerant flow rate and the nonuniform liquid distribution on the tube bundles in practical operation. • At low refrigerant flow rates, the overall heat transfer performance is enhanced by increasing the flow rate, and the dry patch percentage in the total falling film tube surfaces is decreased at the same time. While the refrigerant flow rate rises above a certain value, the heat transfer performance hardly varies with the flow rate due to dry patches being too small. • The refrigerant flow maldistribution can lead to local dryout, and affect the evaporation heat transfer of the falling film evaporator. The capacity of heat transfer deteriorates with the extending degree of the refrigerant nonuniform flow. Moreover, the liquid maldistribution has different effects on heat transfer performances of the tubes at different locations in the tube bundle, and tubes at lower rows present worse heat transfer performance than the upper ones. Acknowledgements The authors gratefully acknowledge the support from the Doctoral Fund Project of Shandong Jianzhu University (XNBS1224), and Wang Liqiu Overseas Taishan Scholar Research Group on Heat and Moisture Transfer.
References [1] B.X. Wang, J.T. Zhang, X.F. Peng, Experimental study on the dryout heat flux of falling liquid film. International Journal of Heat and Mass Transfer, 43 (2000) 1897-1903. [2] M.S. Bohn, S.H. Davis, Thermocapillary breakdown of falling liquid films at high Reynolds numbers. International Journal of Heat and Mass Transfer, 36 (1993) 1875-1881. [3] Tang Zhiwei, Yang Xiaoke, Jiang Zhangyan, Experimental study on surface wave and film breakdown of falling liquid film flow. Heat Mass Transfer, 45 (2009) 673-677. [4] Gherhardt Ribatski, John R. Thome, Experimental study on the onset of local dryout in an evaporating falling film on horizontal plain tubes. Experimental Thermal & Fluid Science, 31 (2007) 483-493. [5] J.F. Roques, Falling film evaporation on a single tube and on a tube bundle, Ph.D. thesis, École Polytechnique Fédérale de Lausanne, Switzerland, 2004. [6] J.F. Roques and J. R. Thome, Falling films on arrays of horizontal tubes with R-134a, Part II: flow visualization, onset of dryout, and heat transfer predictions. Heat Transfer Engineering 28(5) (2007), 415-434. [7] Webb R L, Pals C., Nucleate boiling data for five refrigerants on plain, integral-fin and enhanced tube geometries. Int. J. Heat and Mass Transfer, 35(8) (1992) 1893-1904. [8] Kutateladze, S. S. On the transition to film boiling under nature convection. Kotloturbostoenie, 3(1948) 10-15. [9] Bell K J, Muller A C. Wolverine engineering data book II [M]. Wolverine Tube Co, 2001. [10] ARI Standard 550/590-98. Standard for water chilling packages using the vapor compression cycle, 1998, Air-Conditioning and Refrigeration Institute, 4301 North Fairfax Drive, Suite 425, Arlington, Va. 22203, U.S.A.