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Nucleate boiling of low GWP refrigerants on highly enhanced tube surface
International Journal of Heat and Mass Transfer 96 (2016) 660–666
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Nucleate boiling of low GWP refrigerants on highly enhanced tube surface Evraam Gorgy ⇑ Great Lakes Thermal Technologies, Inc., United States
a r t i c l e
i n f o
Article history: Received 2 June 2015 Received in revised form 20 January 2016 Accepted 24 January 2016 Available online 13 February 2016 keywords: Low GWP refrigerants Nucleate boiling Enhanced surfaces Flooded evaporators Modified Wilson Plot Internal flow correlation
a b s t r a c t This paper presents the experimental investigation of the heat transfer performance of alternative low GWP (Global Warming Potential) refrigerants during nucleate boiling. The fluids used in the study are R-1234ze, R-1233zd(E), and R-450A. The first two HFOs (hydro-fluoro-olefins) have a GWP index less than or equal to 1 while the zeotropic blend R-450a has a GWP index of 547. This study compared R-1234ze and R-450A to their targeted replacement, R-134a, and compared R-1233zd(E) to R-123 using a highly enhanced tube surface. Tests were conducted at a heat flux range of 10–110 kW/m2 and at a saturation temperature of 4.44 °C. Results show that the performance of R-1234ze is very similar to that of R-134a while R-450a shows performance degradation of 28% compared to R-134a. For the R-123 replacement, R-1233zd(E) demonstrates a noticeable 19% performance increase. The results section includes a power-law model to predict the refrigerant-side heat transfer coefficient for all five fluids. Meanwhile, the heat flux used in the model covers a range of 10–60 kW/m2, a typical range of interest for refrigerant evaporator design. Ó 2016 Elsevier Ltd. All rights reserved.
1. Introduction The present publication addresses the shell-side heat transfer performance of flooded evaporators. A refrigerant evaporator is a shell and tube heat exchanger in which liquid circulates inside the tube bundle and is cooled by refrigerant circulating over the tube bundle and in the shell. Cooling takes place through phase change (boiling) of the refrigerant. In flooded evaporators, refrigerant flows over the tube bundle from the bottom up where it usually enters the shell at a thermodynamic quality near 10%, due to the expansion device, and leaves at 100% quality (saturated) or superheated vapor before entering the compressor. This paper investigates the nucleate (a.k.a. pool) boiling characteristics of three low GWP refrigerants and compares them to their respective targeted replacement, either R-134a or R-123. The tube sample, with a 19 mm diameter and active heat transfer length of 91.44 cm, has a doubly-enhanced heat transfer surface where the outside and inside surfaces were built for boiling and single phase heat transfer, respectively. Tests were based on a heat flux range of 10–110 kW/m2 and a saturation temperature of 4.44 °C. This study is useful for refrigerant evaporators such as large tonnage refrigeration equipment/centrifugal chillers. The performance of Honeywell refrigerants R-1234ze(E) and R450A were compared to that of R-134a, as well as the perfor⇑ Tel.: +1 256 841 4453. http://dx.doi.org/10.1016/j.ijheatmasstransfer.2016.01.057 0017-9310/Ó 2016 Elsevier Ltd. All rights reserved.
mance of Honeywell R-1233zd(E) to that of R-123. R-450A is zeotropic blend (i.e. it has a different vapor and liquid phase composition at equilibrium) of R-134a and R-1234ze while R-1234ze and R-1233zd(E) are HFO (hydro-fluoro-olefin) pure fluids. The latter two fluids have a Global Warming Potential (GWP) less than or equal to 1 while R-450A has a GWP of nearly 547 (due to the R-134a presence). Meanwhile, R-134a has a GWP of 1300 while R-123 is classified as a Class-II ozone-depleting substance. The GWP value (Eq. (1)) of a fluid is an index representing the energy absorbed by or contained in a fluid when released into the atmosphere relative to that of CO2 over a period of time (usually 100 years) as presented in Lashof and Ahuja [12]. Additionally, according to the European Commission’s website [2], F-gases such as HFC refrigerants have a global warming effect as high as 23,000 times that of CO2
R t max
ðGWP i Þt max ¼ R t0max 0
ai ðtÞci ðtÞdt
aC ðtÞcC ðtÞdt
:
ð1Þ
2. Background Many researchers have investigated pool boiling of highly enhanced tube-surfaces for a number of years because it is particularly useful for predicting the performance of flooded-type and falling-film evaporators. The types of tubes used range from smooth to integral-fin to enhanced-surface. Recently, enhanced
E. Gorgy / International Journal of Heat and Mass Transfer 96 (2016) 660–666
661
Nomenclature Ai Ao ai ðtÞ ci ðtÞ Co Cp Di Do fD hi ho k L
outside surface area ðpDi LÞ outside surface area ðpDo LÞ instantaneous radioactive forcing for unit increase in the concentration of gas i fraction of gas i remaining at time t shell-side correlation constant specific heat at constant pressure tube inside diameter tube outside diameter enhanced surface friction factor water-side heat transfer coefficient refrigerant-side heat transfer coefficient water thermal conductivity tube length
tubes have been the focus of the HVAC&R industry due to their high efficiency and advantage in reducing the size of heat exchangers. Furthermore, enhanced tube technology, as revealed by the number of recent patents, has been on the rise as manufacturability and productivity techniques advance and because of the increased competition among tube manufacturers. Among the studies that provided pool boiling investigations for enhanced surfaces are those by Gorgy and Eckels [3,4], Webb and Pais [27], Tatara and Payvar [24], Kim and Choi [9], Robinson and Thome [18], and Ribatski and Thome [17]. Those studies initially assessed traditional refrigerants such as R-11, R-12, R-22, R-123, and R-134a; however, interest in alternative low GWP refrigerants is growing with the rising worry over global warming, which prompted research based on thermodynamic and heat transfer characteristics. Compared to the available articles on shell-side heat transfer characteristics of traditional refrigerant, not many studies exist about alternative refrigerants; however, technical committees at ASHRAE and other societies are supporting research of low GWP refrigerants. Nonetheless, the above mentioned studies will guide researchers and engineers through data verification and the general practices of this type of research. While researchers (as mentioned in Gorgy and Eckels [3]) concluded that boiling of refrigerants over enhanced surfaces is unique for each tuberefrigerant combination (also shown in the surface-liquid correction factor of Rohsenow’s correlation [20]), comparing two different refrigerants for one tube enhancement is very useful for helping researchers and chiller manufacturers select alternative refrigerants. Perhaps the study by Rooyen and Thome [19] is among the first on shell-side boiling of low GWP refrigerants on highly enhanced tubes. These researchers studied three fluids, R-134a, R-236fa, and R-1234ze(E), concluding that the performance of R-236fa was lower than that of R-134a while R-1234ze(E) performed similarly to R-134a. The paper also provided a new prediction model for the pool boiling performance at lower heat fluxes. Several other studies investigated boiling performance of low GWP fluids but for in-tube boiling. The following studies outline those efforts. For example, Saitoh et al. [21] studied the boiling performance of R-1234yf inside a 2 mm inner diameter tube. They concluded that, for the most part, R-134a and R-1234yf produced quite similar results. Also, Li et al. [13] conducted a similar experiment (flow boiling inside 2 mm inner diameter tube), but this study used a refrigerant mixture of R-1234yf and R-32. These researchers concluded that at 20% mass fraction of R-32, the heat transfer coefficient is less than for pure R-1234yf while at 50% mass fraction, the heat transfer coefficient is higher than that of
_ m Nu Pr q00 Re Rwall
q
T sat T in T out Uo u V
water mass flow rate Nusselt number Prandtl number heat flux Reynolds number tube wall thermal resistance water density saturation temperature water inlet temperature water outlet temperature overall heat transfer coefficient uncertainty water velocity
pure R-1234yf. For both mass fractions, the heat transfer coefficient is lower than that of pure R-32. Kedzierski and Park [8] studied the convective boiling of R-1234yf/R-134a mixture and R-1234ze(E) inside an enhanced tube focusing on the heat transfer coefficient change with thermodynamic quality. They concluded that R-134a performed better than the other two fluids. Additionally, at qualities less than 30%, the heat transfer coefficient of R-1234yf/R-134a mixture is within 5% of that of R-134a. AHRI low GWP AREP (alternative refrigerants evaluation program) Phase I [26] evaluated a number of low GWP fluids, specifically evaluating the drop-in of new refrigerants in HVAC&R systems. The report identified that the targeted replacements are R-22, R-134a, R-404a, and R-410a. The R-134a replacements include R-450a (previously known as N13) and R-1234ze. While the report didn’t focus on R-123 (or low pressure refrigerants) replacements, R-1233zd(E) is commonly known as a strong candidate for R-123 replacement. In addition to studies involving tubes, Moreno et al. [14] studied the pool boiling of R-1234yf and R-134a on a heated 1 cm2 flat plate. They reported that the performances of R-134a and R-1234yf are almost identical at lower heat flux with an advantage favoring R-134a at higher heat fluxes. They also used copper microporous coating for enhancing the surface, which yielded higher boiling performance and enhanced critical heat flux (CHF) value.
3. Experimental apparatus The test section was a circular vessel 24.5 cm in diameter and 91.44 cm long accommodating three tube samples at the bottom. Three corresponding high performing condensation tubes were at the top of the test section to allow for the energy exchange between boiling and condensation to take place in the same vessel. Condensate droplets were collected on a tray and returned to the pool at the side of the test section, thus maintaining a constant liquid height. The tray was hung from the condensation tubes and sloped to the side of the test section to ensure a distant condensate-return from where the boiling action took place. Thus, ultimately, the amount of refrigerant charge allows for a rule-ofthumb of 2.5 cm of liquid height above the tube surface. The tube sample used in this research was of the doubly enhanced finned surface. The structured outer surface creates the boiling surface whereas the inside surface includes helical ridges (dimensions are presented in Table 1). The boiling surface was originally optimized for medium pressure refrigerants (such as R-134a). The tube was water heated while a low temperature
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E. Gorgy / International Journal of Heat and Mass Transfer 96 (2016) 660–666
1 1 1 ¼ þ Rwall þ : U o Ao hi Ai ho Ao
Table 1 Enhanced tube dimensions. Nominal finned dia. mm (in.)
Nominal wall mm (in.)
Inside dia. mm (in.)
Fin root dia. mm(in.)
19.05 (0.75)
0.635 (0.025)
16.68 (0.657)
17.95 (0.707)
ð2Þ
Multiplying by Ao and rearranging to solve for ho (based on the nominal outside diameter) yields,
1 1 Ao Rwall Ao Uo hi Ai
ho ¼ glycol, used as the cooling source, circulated inside the condensation tubes. The water circulated in the following order: test section, filter, flow meter, electric heater, reservoir, and pump as shown in Fig. 1. The refrigerant temperature was measured in four locations within the pool (farthest corners) to monitor any temperature distribution in the refrigerant pool; however, no noticeable change was detected. Additionally, two pressure transducers were placed in the refrigerant loop, low and high range transducers (for low and high pressure refrigerants) for increased accuracy. Calculated saturation pressure at the measured refrigerant pool temperature was compared to the measured pressure to confirm proper testing and any existence of non-condensable gases (especially for low pressure fluids). Prior to charging the test section, the test rig was brought to 65 Pa (<500 microns of mercury) using vacuum pump and monitored for at least three hours for any noticeable increase in pressure. The water temperature was measured using thermistors while the flow rate was measured using a Coriolistype flow meter. Finally, instrumentation accuracy is addressed in the uncertainty section.
1 :
ð3Þ
The water-side heat transfer coefficient hi is presented in the following section, and the overall heat transfer coefficient is determined using the LMTD method as,
_ C p ln ððT in T sat Þ=ðT out T sat ÞÞ=Ao : Uo ¼ m 4.1. Water-side heat transfer coefficient
For convective single phase heat transfer inside a tube, the following expression yields the heat transfer coefficient,
hi ¼
Nu kwall ; Di
ð5Þ
where all the water properties are evaluated at the average inlet and outlet temperatures. For this study, the water side Reynolds number varied from 10,000 to 75,000. Since the flow is exclusively turbulent, the Nusselt number expression could be predicted by a number of correlations available in literature. First, it manifests in Gnielinski [7] expression as,
ð2=3Þ # Di 1þ NuG ¼ : 1=2 L 1 þ 12:7ðf =8Þ ðPr2=3 1Þ ðf =8Þ ðRe 1000Þ Pr
4. Data reduction Data is reduced to solve for the refrigerant-side heat transfer coefficient (a.k.a. shell-side performance) while formulating the problem this way: heat is transferred by convection in the water, conduction in the tube wall, and convection in the refrigerant. Assuming that heat is carried only in the radial direction and ignoring fouling effects, the problem can be modeled as 1-D heat transfer. By solving the first law of thermodynamics, Newton’s law of cooling for convection heat transfer, and Fourier’s model for conduction heat transfer, the following thermal resistance equation could be written as:
"
NuP ¼
ðf =8Þ Re Pr 1:07 þ 12:7ðf =8Þ
1=2
ðPr2=3 1Þ
:
ð7Þ
Both correlations use the friction factor expression (mostly associated with smooth tubes), 2
f ¼ ð0:79 lnðReÞ 1:64Þ :
ð8Þ
Supply To Chiller Return Refrigerant Reservoir
Actuated Globe Valve
Charge/Reclaim
Condensaon Tubes Test Tubes Test Secon
Water Filter Flow Meter
Water Tank
ð6Þ
Moreover, the above correlation was a modification of the original Petukhov [16] correlation that reads,
Y Strainer
Water Heater
ð4Þ
Water Pump
Fig. 1. Schematic of refrigerant circuit.
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E. Gorgy / International Journal of Heat and Mass Transfer 96 (2016) 660–666
However, since the tube used in this study is internally enhanced, and to increase the accuracy of the above correlations, the friction coefficient is determined via the inlet-outlet pressure drop measurement according to,
fD ¼
DP Di 2 : q L V2
ð9Þ
A third possible heat transfer correlation for the convective flow model is a modification of Sieder and Tate [23] where,
NuST ¼ ðconstantÞ Re0:8 Pr1=3 :
ð10Þ
Further, the Wilson Plot technique (as the following section explains) should be used in conjunction with a convective flow model to investigate the tube-inside enhancement and thereby produce a correction factor associated with the convective model. Accordingly, the correction factor becomes a leading coefficient, C i , to the water-side heat transfer coefficient where the Nui can be any of the three Nusselt number expressions listed above or achieved by substituting the leading ðconstantÞ in the NuST expression,
hi ¼ C i
Nui k : Di
ð11Þ
The modified Wilson Plot technique was introduced by Briggs and Young [1] and Shah [22] for shell-and-tube heat exchangers. This technique generates the heat transfer coefficients for both the shell and tube sides based on the flow rate and temperature measurement of the two streams where using local heat transfer analysis (or a known model) in either side of the heat exchanger is difficult. The Wilson Plot is constructed through first transforming the thermal resistance equation into a straight line equation as,
1 : slope ðt val95%CI;DegFrdm Þ ðStdErrslope Þ
ð15Þ
Accordingly, six total leading coefficients were determined with each associated uncertainty. Based on the results shown in Table 2, the Petukhov-type correlation returned the most consistent results at the two Pr numbers. Additionally, the NuP coefficient determined at Pr 9.0 (underline value) has slightly less uncertainty than that at Pr 6.5. Therefore, the final value for the water-side heat transfer coefficient is determined as,
hi ¼ 1:662
1 1 1 Rwall ¼ þ : U o Ao ho Ao hi Ai |fflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflffl} |ffl{zffl} |{z}
ð12Þ
ðf D =8Þ Re Pr 1=2
1:07 þ 12:7ðf D =8Þ
ðPr2=3 1Þ
k=Di :
ð16Þ
mx
b
However, for this straight line equation to work out, the heat transfer coefficient on the shell side must be constant while the inside heat transfer coefficient varies over a certain range (preferably the full operating range of the heat exchanger). In the case of shell-side boiling, and for the purposes of this study, the shell-side heat transfer coefficient is assumed to be a function of heat flux only, as shown in Eq. (13) next, and thus it can be held constant by controlling the heat flux, n
ho ¼ C o ðq00 Þ :
ð13Þ
Further, the inside heat transfer coefficient varies as evidenced by changing Reynolds numbers (or water flow rates). Ultimately, the thermal resistances equation can be rearranged to include Eqs. (13) and (11) such that the final expression reads,
1 1 1 1 Rwall ¼ þ : 00 n U o Ao C i NuðG=P=STÞ k=Di Ai C o ðq Þ Ao |fflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflffl} |fflfflfflfflfflffl{zfflfflfflfflfflffl} |{z} |fflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl} Y
One of the disadvantages of applying the Wilson Plot method during pool boiling is the limited range of the water Prandtl number as it is governed mainly by the refrigerant saturation temperature. The limited range of the Pr number will therefore affect the applicability of the tube-side flow correlation, so this study addressed two ranges of Pr numbers, 6.5 and 9.0. As it is expected for each leading coefficient to be slightly different at each Pr number, the uncertainty analysis is used to identify the best leading coefficient. Using a Monte-Carlo simulation, the uncertainty analysis of the slope running 40,000 iterations (i.e. 40,000 straight-line fits) is studied based on a normal distribution of the uncertainties in the temperature and flow rate as given in Table 3. The straight line coefficients are calculated at each iteration where the minimum and maximum values of the 40,000 data points were used as the upper and lower limits of the slope uncertainty. When compared to the slope uncertainty calculation according to Eq. (15), a 4% difference arose between the slope uncertainty method and the Monte-Carlo method where the Monte-Carlo simulation produced higher, yet conservative uncertainty values. The Monte-Carlo simulation was the final method for determining the slope uncertainty.
DC i ¼
4.2. The modified Wilson Plot technique
Y
4.3. Prandtl number effect
b
m
ð14Þ
x
Nine data points were collected covering a Reynolds number range between 10,000 and 75,000 with the correction coefficient to the turbulence model C i determined as the inverse of the slope m through a linear regression. The following section addresses the investigation of each turbulent model in addition to the associated uncertainty and the final tube-side expression.
4.4. Uncertainty analysis The uncertainty analysis for the refrigerant-side heat transfer coefficient was performed using the Kline and McClintock [10] second order law, and determining the final uncertainty in the heat transfer coefficient first required defining the uncertainty of the input variables as presented in Table 3. Those can be the measured variables such as temperature, pressure, and flow rate or the calculated variables such as the water-side heat transfer coefficient. For the water properties, RefProp 8.0 [11] was used and knowing that the temperature measurement is the dominant uncertainty, a simplified version of the uncertainty analysis is applied such that the effect of dimensions and water properties was ignored in comparison to the temperature uncertainty. Thus, the refrigerant heat transfer coefficient uncertainty was determined by applying propagation of error on Eq. (3) producing,
Table 2 Wilson Plot results. NuG
NuP
NuST
C i;Pr 9:0
1.759
0.0864
Uncertainty C i;Pr 6:5 Uncertainty
4.2% 1.712 4.4%
1.662 4.2% 1.675 4.4%
4.2% 0.0892 4%
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E. Gorgy / International Journal of Heat and Mass Transfer 96 (2016) 660–666
sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2 2 2 2 2 @ho @ho @ho @ho @ho 2 _ 2þ uho ¼ ðuðTÞÞ2 þ ðuðTÞÞ2 þ ðuðTÞÞ2 þ ðuðmÞÞ ðuðhi ÞÞ _ @T in @T out @T sat @m @hi
5. Results Tests were performed at a saturation temperature of 4.44 °C (40 °F) and the refrigerant side heat transfer coefficient against the heat flux trends were plotted to generate the pool boiling results. First, for evaluating the enhanced surface efficiency over smooth surface, the smooth tube experimental data provided in Gorgy and Eckels [3] (under R-134a and the same saturation temperature) as shown in Eq. (18) was used. Compared to the enhanced tube’s R-134a results, a noticeable 5.5 times the heat transfer coefficient is realized over the heat flux range 10–60 kW/m2. At the lowest heat flux point, the enhancement is almost 10 times that of smooth tube
ho;smooth tube ¼ 596:281 q000:612
ð18Þ
The enhancement type of the current surface relies on increasing the number of nucleation sites, a typical approach for increasing the boiling performance (unlike using extended surfaces or integral fins). Nucleation site density is increased through creating plentiful pores and cavities; as opposed to the smooth surface where the nucleation sites are naturally present due to, for example, surface imperfection, grain structure, and polish. Boiling cavities are created by cold working of the tube surface, resulting in what is widely known as structured surfaces, to provide active evaporation cavities where the liquid pool does not flood the cavities.
Table 3 Input uncertainties. uT (°C)
um (kg/s)
uhi (W/m2 °C)
0.02
0.15% reading
4.2% hi
ð17Þ
The boiling mechanism in the surface is an oscillation cycle between vapor and liquid such that the vapor escaping the pores is immediately replaced by liquid. The heat transfer performance is enhanced due to two processes (also presented in Nakayama et al. [15]). The first enhancement is due to latent heat (liquid evaporation) while the second is bubbles convection on the surface. Additionally, the heat transfer coefficient behavior versus heat flux can help demonstrate the effect of the liquid feed mechanism [5]. The liquid feed to the enhanced surface has a substantial effect on the tube performance. At low heat flux the film thickness in the cavities is considerable thus the low performance. The performance reaches it peak as the film thickness decreases while the cavity walls remain wetted. For shell-side boiling on enhanced surfaces, the heat flux is typically the dominant effect (as outlined in Gorgy and Eckels [6]). Figs. 2 and 3 compare respectively the performance results of the medium and low pressure refrigerants versus the targeted replacement. Meanwhile, Fig. 2 shows results for R-134a compared to those for R-1234ze and R-450a. Considering the uncertainty limits, R-134a and R-1234ze present nearly the same results, as was previously confirmed by Rooyen and Thome [19]. Therefore, R-1234ze proves to be a viable R-134a replacement for the specified types of application. R-450a, on the other hand, showed a significantly lower performance than R-134a on average, a performance drop of nearly 28%. While R-450a’s thermodynamic properties are slightly closer to those of R-134a than is so between R-1234ze and R-134a, the lower performance of R-450a can be explained as due to the temperature glide of the zeotropic blend R-450a compared to the pure refrigerant R-1234ze. The degradation of the zeotropic blend’s boiling performance as compared to pure refrigerants was also highlighted in Thome and Shock [25]. With respect to low pressure refrigerants, Fig. 3 compares R-123 and R-1233zd(E), the latter showing a performance improvement of 19%. While R-1233zd(E) shows a significant
Fig. 2. Medium pressure refrigerants comparisons.
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E. Gorgy / International Journal of Heat and Mass Transfer 96 (2016) 660–666
Fig. 3. Low pressure refrigerants comparison.
Fig. 4. R-134a and R-1233zd(E) comparison.
improvement over R-123, it still comes 19% short of the performance of R-134a as Fig. 4 shows. In this case, R-1233zd(E) proves to be a better substitute for the targeted replacement R-123. Comparing the thermodynamic properties of R-1233zd(E) to R-123 shows the former has a lower vapor specific volume and a higher liquid specific volume, which could mean that for the same boiling surface enhancement (pore and tunnel size), the lower vapor specific volume allows for better pore activation (higher bubble convection), and the higher liquid specific volume allows for better liquid coverage of the surface enhancement (liquid latent heat), hence the higher heat transfer coefficient. The same logic applies when explaining the performance increase of R-134a over R-1234zd(E) where the former has lower vapor specific volume and higher liquid specific volume. Next, Table 4 presents the model coefficients associated with each fluid using the pool boiling heat transfer coefficient model (Eq. (19)). The applicable units are W/m2 °C and kW/m2 for the heat transfer coefficient and heat flux, respectively. For increased
Table 4 Model summary.
Co n % uho;max % uho;min
R-134a
R-1234ze
R-450a
R-123
R-1233zd (E)
20417.08 0.0851 24 6
23685.84 0.039 26 6
7260.53 0.286 18 5
5697.525 0.329 17 5
11677.83 0.184 22 5
model accuracy, a heat flux range of 10–60 kW/m2 (a typical operating range for refrigerant evaporators) was used. Within that range, the heat transfer coefficient experiences an exponentialfunction rise. Noticeably, the uncertainty range of each refrigerant correlates with the performance level. The high uncertainty is indicative of better fluid performance where a smaller temperature difference between the inlet and outlet temperatures, or low heat flux, is observed
666
hmodel ¼ C o q00n :
E. Gorgy / International Journal of Heat and Mass Transfer 96 (2016) 660–666
ð19Þ
6. Conclusion This paper presents the experimental investigation of three low GWP refrigerants: R-1234ze, R-1233zd(E), and R-450A. The first two HFOs have a GWP less than or equal to 1 while the zeotropic blend R-450a has a GWP index of 547. Nucleate boiling performance of R-1234ze and R-450a were compared to their targeted replacement, R-134a; and R-1233zd(E) to its targeted replacement, R-123 conducting the test using a highly enhanced tube surface. Moreover, data was collected at a saturation temperature of 4.44 °C and a heat flux range of 10-110 kW/m2. Also, the study included an investigation of the suitable tubeside turbulent flow correlation. A Monte-Carlo simulation for the tube-side uncertainty analysis helped show that the Petukhov correlation is consistent at Prandtl numbers of 6.5 and 9.0. Results show that the performance of R-1234ze is very similar to that of R-134a while R-450a demonstrates a significant 28% degradation. For the R-123 comparison, R-1233zd(E) demonstrates a noticeable 19% increase in performance. Finally, an exponential-function model was introduced for all five fluids where the applicable heat flux covered a range between 10 and 60 kW/m2, a typical range of interest for refrigerant evaporator design. Acknowledgements The efforts of Mark Spatz, Samuel Motta, and Ankit Sethi of Honeywell, who donated the refrigerant samples for this study, are greatly appreciated. Moreover, thanks to Joey Smith, Mark Owens, Justin Ryan, and Ben Turnbull for helping with samples manufacturing and testing. References [1] D.E. Briggs, E.H. Young, Modified Wilson plot techniques for obtaining heat transfer correlations for shell and tube heat exchangers, in: Chemical Engineering Progress Symposium Series, 92 (65), 1969, pp. 35–45. [2] Fluorinated greenhouse gases, F-gas facts,. [3] E.I. Gorgy, S. Eckels, Average heat transfer coefficient for pool boiling of R-134a and R-123 on smooth and enhanced tubes (RP-1316), HVAC&R 16 (5) (2010) 657–676. [4] E.I. Gorgy, S. Eckels, Local heat transfer coefficient for pool boiling of R-134a and R-123 on smooth and enhanced tubes, Int. J. Heat Mass Transfer 55 (11– 12) (2012) 2751–3326.
[5] E.I. Gorgy, Experimental evaluation of heat transfer impacts of tube pitch on highly enhanced surface tube bundle (Ph.D. Dissertation), Kansas State University, Chapter 8, 2011. [6] E.I. Gorgy, S. Eckels, Convective boiling of R-134a and R-123 on a standard pitch enhanced tube bundle, HVAC&R Res. J. 19 (2) (2013) 193–206. [7] V. Gnielinski, New equations for heat and mass transfer in turbulent pipe and channel flow, Int. Chem. Eng 16 (2) (1976) 359–368. [8] M. Kedzierski, K.J. Park, Horizontal convective boiling of R-134a, R-1234yf/R134a, and R-1234ze(E) within a micro-fin tube with extensive measurement and analysis details, Nist Technical Note 1807, 2013. [9] N.H. Kim, K.K. Choi, Nucleate pool boiling on structured enhanced tubes having pores with connecting gaps, Int. J. Heat and Mass Transfer 44 (1) (2001) 17–28. [10] S.J. Kline, F. McClintock, Describing uncertainties in single-sample experiments, Mech. Eng. 75 (1) (1953) 3–8. [11] E.W. Lemmon, M.L. Huber, M.O. McLinden, NIST thermodynamic and transport properties of refrigerants and refrigerant mixtures (REFPROP) version 8.0. NIST, 2007. [12] D.A. Lashof, D.R. Ahuja, Relative contribution of greenhouse gas emissions to global warming, Lett. Nat. 334 (6266) (1990) 529–531. [13] M. Li, C. Dang, E. Hihara, Flow boiling heat transfer of HFO 1234yf and R32 refrigerant mixtures in a smooth horizontal tube: part I. Experimental investigation, Int. J. Heat Mass Transfer 55 (13–14) (2012) 3437–3446. [14] G. Moreno, S. Narumanchi, C. King, Pool boiling heat transfer characteristics of HFO-1234yf on plain and microporous enhanced surfaces, J. Heat Transfer 135 (11) (2013) 111014-1–111014-10. [15] W. Nakayama, T. Daikoku, H. Kuwahara, T. Nakajima, Dynamic model of enhanced boiling heat transfer on porous surfaces part II: analytical modeling, J. Heat Transfer 102 (1980) 451–456. [16] B.S. Petukhov, Heat transfer and friction in turbulent pipe flow with variable physical properties, Adv. Heat Transfer 6 (1970) 503–565. [17] G. Ribatski, J.R. Thome, Two-phase flow and heat transfer across horizontal tube bundles – a review, Heat Transfer Eng. 28 (6) (2007) 508–524. [18] D.M. Robinson, J.R. Thome, Local bundle boiling heat transfer coefficients on a turbo-BII HP tube bundle (RP-1089), HVAC&R Res. 10 (4) (2004) 441–457. [19] E.V. Rooyen, J.R. Thome, Pool boiling data and prediction method for enhanced boiling tubes with R-134a, R-236fa, and R-1234ze(E), Int. J. Refrig. 36 (2) (2013) 447–455. [20] W.M. Rohsenow, A method of correlating heat transfer data for surface boiling of liquids, ASME Trans. 74 (1952) 969–976. [21] S. Saitoh, C. Dand, Y. Nakamura, E. Hihara, Boiling heat transfer of HFO-1234yf flowing in a smooth small-diameter horizontal tube, Int. J. Refrig. 34 (8) (2011) 1846–1853. [22] R.K. Shah, Assessment of modified Wilson plot techniques for obtaining heat exchanger design data, in: Heat Transfer, Proceedings of the International Heat Transfer Conference 5, 1990, pp. 51–56. [23] E.N. Sieder, G.E. Tate, Heat transfer and pressure drop of liquids in tubes, Ind. Eng. Chem. 28 (12) (1936) 1429–1435. [24] R.A. Tatara, P. Payvar, Pool boiling of pure R134a from a single Turbo-BII-HP tube, Int. J. Heat Mass Transfer 43 (12) (2000) 2233–2236. [25] J.R. Thome, R.A.W. Shock, Boiling of multicomponent mixtures, Adv. Heat Transfer 16 (1984) 59–156. [26] X. Wang, K. Amrane, AHRI low global warming potential alternative refrigerants evaluation program (Low-GWP AREP) – summary of phase I testing results, in: International Refrigeration and Air Conditioning Conference. Paper 1416, 2014. [27] R.L. Webb, C. Pais, Nucleate pool boiling data for five refrigerants on plain, integral-fin, and enhanced tube geometries, Int. J. Heat Mass Transfer 35 (8) (1992) 1893–1904.
Contents lists available at ScienceDirect
International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
Nucleate boiling of low GWP refrigerants on highly enhanced tube surface Evraam Gorgy ⇑ Great Lakes Thermal Technologies, Inc., United States
a r t i c l e
i n f o
Article history: Received 2 June 2015 Received in revised form 20 January 2016 Accepted 24 January 2016 Available online 13 February 2016 keywords: Low GWP refrigerants Nucleate boiling Enhanced surfaces Flooded evaporators Modified Wilson Plot Internal flow correlation
a b s t r a c t This paper presents the experimental investigation of the heat transfer performance of alternative low GWP (Global Warming Potential) refrigerants during nucleate boiling. The fluids used in the study are R-1234ze, R-1233zd(E), and R-450A. The first two HFOs (hydro-fluoro-olefins) have a GWP index less than or equal to 1 while the zeotropic blend R-450a has a GWP index of 547. This study compared R-1234ze and R-450A to their targeted replacement, R-134a, and compared R-1233zd(E) to R-123 using a highly enhanced tube surface. Tests were conducted at a heat flux range of 10–110 kW/m2 and at a saturation temperature of 4.44 °C. Results show that the performance of R-1234ze is very similar to that of R-134a while R-450a shows performance degradation of 28% compared to R-134a. For the R-123 replacement, R-1233zd(E) demonstrates a noticeable 19% performance increase. The results section includes a power-law model to predict the refrigerant-side heat transfer coefficient for all five fluids. Meanwhile, the heat flux used in the model covers a range of 10–60 kW/m2, a typical range of interest for refrigerant evaporator design. Ó 2016 Elsevier Ltd. All rights reserved.
1. Introduction The present publication addresses the shell-side heat transfer performance of flooded evaporators. A refrigerant evaporator is a shell and tube heat exchanger in which liquid circulates inside the tube bundle and is cooled by refrigerant circulating over the tube bundle and in the shell. Cooling takes place through phase change (boiling) of the refrigerant. In flooded evaporators, refrigerant flows over the tube bundle from the bottom up where it usually enters the shell at a thermodynamic quality near 10%, due to the expansion device, and leaves at 100% quality (saturated) or superheated vapor before entering the compressor. This paper investigates the nucleate (a.k.a. pool) boiling characteristics of three low GWP refrigerants and compares them to their respective targeted replacement, either R-134a or R-123. The tube sample, with a 19 mm diameter and active heat transfer length of 91.44 cm, has a doubly-enhanced heat transfer surface where the outside and inside surfaces were built for boiling and single phase heat transfer, respectively. Tests were based on a heat flux range of 10–110 kW/m2 and a saturation temperature of 4.44 °C. This study is useful for refrigerant evaporators such as large tonnage refrigeration equipment/centrifugal chillers. The performance of Honeywell refrigerants R-1234ze(E) and R450A were compared to that of R-134a, as well as the perfor⇑ Tel.: +1 256 841 4453. http://dx.doi.org/10.1016/j.ijheatmasstransfer.2016.01.057 0017-9310/Ó 2016 Elsevier Ltd. All rights reserved.
mance of Honeywell R-1233zd(E) to that of R-123. R-450A is zeotropic blend (i.e. it has a different vapor and liquid phase composition at equilibrium) of R-134a and R-1234ze while R-1234ze and R-1233zd(E) are HFO (hydro-fluoro-olefin) pure fluids. The latter two fluids have a Global Warming Potential (GWP) less than or equal to 1 while R-450A has a GWP of nearly 547 (due to the R-134a presence). Meanwhile, R-134a has a GWP of 1300 while R-123 is classified as a Class-II ozone-depleting substance. The GWP value (Eq. (1)) of a fluid is an index representing the energy absorbed by or contained in a fluid when released into the atmosphere relative to that of CO2 over a period of time (usually 100 years) as presented in Lashof and Ahuja [12]. Additionally, according to the European Commission’s website [2], F-gases such as HFC refrigerants have a global warming effect as high as 23,000 times that of CO2
R t max
ðGWP i Þt max ¼ R t0max 0
ai ðtÞci ðtÞdt
aC ðtÞcC ðtÞdt
:
ð1Þ
2. Background Many researchers have investigated pool boiling of highly enhanced tube-surfaces for a number of years because it is particularly useful for predicting the performance of flooded-type and falling-film evaporators. The types of tubes used range from smooth to integral-fin to enhanced-surface. Recently, enhanced
E. Gorgy / International Journal of Heat and Mass Transfer 96 (2016) 660–666
661
Nomenclature Ai Ao ai ðtÞ ci ðtÞ Co Cp Di Do fD hi ho k L
outside surface area ðpDi LÞ outside surface area ðpDo LÞ instantaneous radioactive forcing for unit increase in the concentration of gas i fraction of gas i remaining at time t shell-side correlation constant specific heat at constant pressure tube inside diameter tube outside diameter enhanced surface friction factor water-side heat transfer coefficient refrigerant-side heat transfer coefficient water thermal conductivity tube length
tubes have been the focus of the HVAC&R industry due to their high efficiency and advantage in reducing the size of heat exchangers. Furthermore, enhanced tube technology, as revealed by the number of recent patents, has been on the rise as manufacturability and productivity techniques advance and because of the increased competition among tube manufacturers. Among the studies that provided pool boiling investigations for enhanced surfaces are those by Gorgy and Eckels [3,4], Webb and Pais [27], Tatara and Payvar [24], Kim and Choi [9], Robinson and Thome [18], and Ribatski and Thome [17]. Those studies initially assessed traditional refrigerants such as R-11, R-12, R-22, R-123, and R-134a; however, interest in alternative low GWP refrigerants is growing with the rising worry over global warming, which prompted research based on thermodynamic and heat transfer characteristics. Compared to the available articles on shell-side heat transfer characteristics of traditional refrigerant, not many studies exist about alternative refrigerants; however, technical committees at ASHRAE and other societies are supporting research of low GWP refrigerants. Nonetheless, the above mentioned studies will guide researchers and engineers through data verification and the general practices of this type of research. While researchers (as mentioned in Gorgy and Eckels [3]) concluded that boiling of refrigerants over enhanced surfaces is unique for each tuberefrigerant combination (also shown in the surface-liquid correction factor of Rohsenow’s correlation [20]), comparing two different refrigerants for one tube enhancement is very useful for helping researchers and chiller manufacturers select alternative refrigerants. Perhaps the study by Rooyen and Thome [19] is among the first on shell-side boiling of low GWP refrigerants on highly enhanced tubes. These researchers studied three fluids, R-134a, R-236fa, and R-1234ze(E), concluding that the performance of R-236fa was lower than that of R-134a while R-1234ze(E) performed similarly to R-134a. The paper also provided a new prediction model for the pool boiling performance at lower heat fluxes. Several other studies investigated boiling performance of low GWP fluids but for in-tube boiling. The following studies outline those efforts. For example, Saitoh et al. [21] studied the boiling performance of R-1234yf inside a 2 mm inner diameter tube. They concluded that, for the most part, R-134a and R-1234yf produced quite similar results. Also, Li et al. [13] conducted a similar experiment (flow boiling inside 2 mm inner diameter tube), but this study used a refrigerant mixture of R-1234yf and R-32. These researchers concluded that at 20% mass fraction of R-32, the heat transfer coefficient is less than for pure R-1234yf while at 50% mass fraction, the heat transfer coefficient is higher than that of
_ m Nu Pr q00 Re Rwall
q
T sat T in T out Uo u V
water mass flow rate Nusselt number Prandtl number heat flux Reynolds number tube wall thermal resistance water density saturation temperature water inlet temperature water outlet temperature overall heat transfer coefficient uncertainty water velocity
pure R-1234yf. For both mass fractions, the heat transfer coefficient is lower than that of pure R-32. Kedzierski and Park [8] studied the convective boiling of R-1234yf/R-134a mixture and R-1234ze(E) inside an enhanced tube focusing on the heat transfer coefficient change with thermodynamic quality. They concluded that R-134a performed better than the other two fluids. Additionally, at qualities less than 30%, the heat transfer coefficient of R-1234yf/R-134a mixture is within 5% of that of R-134a. AHRI low GWP AREP (alternative refrigerants evaluation program) Phase I [26] evaluated a number of low GWP fluids, specifically evaluating the drop-in of new refrigerants in HVAC&R systems. The report identified that the targeted replacements are R-22, R-134a, R-404a, and R-410a. The R-134a replacements include R-450a (previously known as N13) and R-1234ze. While the report didn’t focus on R-123 (or low pressure refrigerants) replacements, R-1233zd(E) is commonly known as a strong candidate for R-123 replacement. In addition to studies involving tubes, Moreno et al. [14] studied the pool boiling of R-1234yf and R-134a on a heated 1 cm2 flat plate. They reported that the performances of R-134a and R-1234yf are almost identical at lower heat flux with an advantage favoring R-134a at higher heat fluxes. They also used copper microporous coating for enhancing the surface, which yielded higher boiling performance and enhanced critical heat flux (CHF) value.
3. Experimental apparatus The test section was a circular vessel 24.5 cm in diameter and 91.44 cm long accommodating three tube samples at the bottom. Three corresponding high performing condensation tubes were at the top of the test section to allow for the energy exchange between boiling and condensation to take place in the same vessel. Condensate droplets were collected on a tray and returned to the pool at the side of the test section, thus maintaining a constant liquid height. The tray was hung from the condensation tubes and sloped to the side of the test section to ensure a distant condensate-return from where the boiling action took place. Thus, ultimately, the amount of refrigerant charge allows for a rule-ofthumb of 2.5 cm of liquid height above the tube surface. The tube sample used in this research was of the doubly enhanced finned surface. The structured outer surface creates the boiling surface whereas the inside surface includes helical ridges (dimensions are presented in Table 1). The boiling surface was originally optimized for medium pressure refrigerants (such as R-134a). The tube was water heated while a low temperature
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E. Gorgy / International Journal of Heat and Mass Transfer 96 (2016) 660–666
1 1 1 ¼ þ Rwall þ : U o Ao hi Ai ho Ao
Table 1 Enhanced tube dimensions. Nominal finned dia. mm (in.)
Nominal wall mm (in.)
Inside dia. mm (in.)
Fin root dia. mm(in.)
19.05 (0.75)
0.635 (0.025)
16.68 (0.657)
17.95 (0.707)
ð2Þ
Multiplying by Ao and rearranging to solve for ho (based on the nominal outside diameter) yields,
1 1 Ao Rwall Ao Uo hi Ai
ho ¼ glycol, used as the cooling source, circulated inside the condensation tubes. The water circulated in the following order: test section, filter, flow meter, electric heater, reservoir, and pump as shown in Fig. 1. The refrigerant temperature was measured in four locations within the pool (farthest corners) to monitor any temperature distribution in the refrigerant pool; however, no noticeable change was detected. Additionally, two pressure transducers were placed in the refrigerant loop, low and high range transducers (for low and high pressure refrigerants) for increased accuracy. Calculated saturation pressure at the measured refrigerant pool temperature was compared to the measured pressure to confirm proper testing and any existence of non-condensable gases (especially for low pressure fluids). Prior to charging the test section, the test rig was brought to 65 Pa (<500 microns of mercury) using vacuum pump and monitored for at least three hours for any noticeable increase in pressure. The water temperature was measured using thermistors while the flow rate was measured using a Coriolistype flow meter. Finally, instrumentation accuracy is addressed in the uncertainty section.
1 :
ð3Þ
The water-side heat transfer coefficient hi is presented in the following section, and the overall heat transfer coefficient is determined using the LMTD method as,
_ C p ln ððT in T sat Þ=ðT out T sat ÞÞ=Ao : Uo ¼ m 4.1. Water-side heat transfer coefficient
For convective single phase heat transfer inside a tube, the following expression yields the heat transfer coefficient,
hi ¼
Nu kwall ; Di
ð5Þ
where all the water properties are evaluated at the average inlet and outlet temperatures. For this study, the water side Reynolds number varied from 10,000 to 75,000. Since the flow is exclusively turbulent, the Nusselt number expression could be predicted by a number of correlations available in literature. First, it manifests in Gnielinski [7] expression as,
ð2=3Þ # Di 1þ NuG ¼ : 1=2 L 1 þ 12:7ðf =8Þ ðPr2=3 1Þ ðf =8Þ ðRe 1000Þ Pr
4. Data reduction Data is reduced to solve for the refrigerant-side heat transfer coefficient (a.k.a. shell-side performance) while formulating the problem this way: heat is transferred by convection in the water, conduction in the tube wall, and convection in the refrigerant. Assuming that heat is carried only in the radial direction and ignoring fouling effects, the problem can be modeled as 1-D heat transfer. By solving the first law of thermodynamics, Newton’s law of cooling for convection heat transfer, and Fourier’s model for conduction heat transfer, the following thermal resistance equation could be written as:
"
NuP ¼
ðf =8Þ Re Pr 1:07 þ 12:7ðf =8Þ
1=2
ðPr2=3 1Þ
:
ð7Þ
Both correlations use the friction factor expression (mostly associated with smooth tubes), 2
f ¼ ð0:79 lnðReÞ 1:64Þ :
ð8Þ
Supply To Chiller Return Refrigerant Reservoir
Actuated Globe Valve
Charge/Reclaim
Condensaon Tubes Test Tubes Test Secon
Water Filter Flow Meter
Water Tank
ð6Þ
Moreover, the above correlation was a modification of the original Petukhov [16] correlation that reads,
Y Strainer
Water Heater
ð4Þ
Water Pump
Fig. 1. Schematic of refrigerant circuit.
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E. Gorgy / International Journal of Heat and Mass Transfer 96 (2016) 660–666
However, since the tube used in this study is internally enhanced, and to increase the accuracy of the above correlations, the friction coefficient is determined via the inlet-outlet pressure drop measurement according to,
fD ¼
DP Di 2 : q L V2
ð9Þ
A third possible heat transfer correlation for the convective flow model is a modification of Sieder and Tate [23] where,
NuST ¼ ðconstantÞ Re0:8 Pr1=3 :
ð10Þ
Further, the Wilson Plot technique (as the following section explains) should be used in conjunction with a convective flow model to investigate the tube-inside enhancement and thereby produce a correction factor associated with the convective model. Accordingly, the correction factor becomes a leading coefficient, C i , to the water-side heat transfer coefficient where the Nui can be any of the three Nusselt number expressions listed above or achieved by substituting the leading ðconstantÞ in the NuST expression,
hi ¼ C i
Nui k : Di
ð11Þ
The modified Wilson Plot technique was introduced by Briggs and Young [1] and Shah [22] for shell-and-tube heat exchangers. This technique generates the heat transfer coefficients for both the shell and tube sides based on the flow rate and temperature measurement of the two streams where using local heat transfer analysis (or a known model) in either side of the heat exchanger is difficult. The Wilson Plot is constructed through first transforming the thermal resistance equation into a straight line equation as,
1 : slope ðt val95%CI;DegFrdm Þ ðStdErrslope Þ
ð15Þ
Accordingly, six total leading coefficients were determined with each associated uncertainty. Based on the results shown in Table 2, the Petukhov-type correlation returned the most consistent results at the two Pr numbers. Additionally, the NuP coefficient determined at Pr 9.0 (underline value) has slightly less uncertainty than that at Pr 6.5. Therefore, the final value for the water-side heat transfer coefficient is determined as,
hi ¼ 1:662
1 1 1 Rwall ¼ þ : U o Ao ho Ao hi Ai |fflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflffl} |ffl{zffl} |{z}
ð12Þ
ðf D =8Þ Re Pr 1=2
1:07 þ 12:7ðf D =8Þ
ðPr2=3 1Þ
k=Di :
ð16Þ
mx
b
However, for this straight line equation to work out, the heat transfer coefficient on the shell side must be constant while the inside heat transfer coefficient varies over a certain range (preferably the full operating range of the heat exchanger). In the case of shell-side boiling, and for the purposes of this study, the shell-side heat transfer coefficient is assumed to be a function of heat flux only, as shown in Eq. (13) next, and thus it can be held constant by controlling the heat flux, n
ho ¼ C o ðq00 Þ :
ð13Þ
Further, the inside heat transfer coefficient varies as evidenced by changing Reynolds numbers (or water flow rates). Ultimately, the thermal resistances equation can be rearranged to include Eqs. (13) and (11) such that the final expression reads,
1 1 1 1 Rwall ¼ þ : 00 n U o Ao C i NuðG=P=STÞ k=Di Ai C o ðq Þ Ao |fflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflffl} |fflfflfflfflfflffl{zfflfflfflfflfflffl} |{z} |fflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl} Y
One of the disadvantages of applying the Wilson Plot method during pool boiling is the limited range of the water Prandtl number as it is governed mainly by the refrigerant saturation temperature. The limited range of the Pr number will therefore affect the applicability of the tube-side flow correlation, so this study addressed two ranges of Pr numbers, 6.5 and 9.0. As it is expected for each leading coefficient to be slightly different at each Pr number, the uncertainty analysis is used to identify the best leading coefficient. Using a Monte-Carlo simulation, the uncertainty analysis of the slope running 40,000 iterations (i.e. 40,000 straight-line fits) is studied based on a normal distribution of the uncertainties in the temperature and flow rate as given in Table 3. The straight line coefficients are calculated at each iteration where the minimum and maximum values of the 40,000 data points were used as the upper and lower limits of the slope uncertainty. When compared to the slope uncertainty calculation according to Eq. (15), a 4% difference arose between the slope uncertainty method and the Monte-Carlo method where the Monte-Carlo simulation produced higher, yet conservative uncertainty values. The Monte-Carlo simulation was the final method for determining the slope uncertainty.
DC i ¼
4.2. The modified Wilson Plot technique
Y
4.3. Prandtl number effect
b
m
ð14Þ
x
Nine data points were collected covering a Reynolds number range between 10,000 and 75,000 with the correction coefficient to the turbulence model C i determined as the inverse of the slope m through a linear regression. The following section addresses the investigation of each turbulent model in addition to the associated uncertainty and the final tube-side expression.
4.4. Uncertainty analysis The uncertainty analysis for the refrigerant-side heat transfer coefficient was performed using the Kline and McClintock [10] second order law, and determining the final uncertainty in the heat transfer coefficient first required defining the uncertainty of the input variables as presented in Table 3. Those can be the measured variables such as temperature, pressure, and flow rate or the calculated variables such as the water-side heat transfer coefficient. For the water properties, RefProp 8.0 [11] was used and knowing that the temperature measurement is the dominant uncertainty, a simplified version of the uncertainty analysis is applied such that the effect of dimensions and water properties was ignored in comparison to the temperature uncertainty. Thus, the refrigerant heat transfer coefficient uncertainty was determined by applying propagation of error on Eq. (3) producing,
Table 2 Wilson Plot results. NuG
NuP
NuST
C i;Pr 9:0
1.759
0.0864
Uncertainty C i;Pr 6:5 Uncertainty
4.2% 1.712 4.4%
1.662 4.2% 1.675 4.4%
4.2% 0.0892 4%
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sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2 2 2 2 2 @ho @ho @ho @ho @ho 2 _ 2þ uho ¼ ðuðTÞÞ2 þ ðuðTÞÞ2 þ ðuðTÞÞ2 þ ðuðmÞÞ ðuðhi ÞÞ _ @T in @T out @T sat @m @hi
5. Results Tests were performed at a saturation temperature of 4.44 °C (40 °F) and the refrigerant side heat transfer coefficient against the heat flux trends were plotted to generate the pool boiling results. First, for evaluating the enhanced surface efficiency over smooth surface, the smooth tube experimental data provided in Gorgy and Eckels [3] (under R-134a and the same saturation temperature) as shown in Eq. (18) was used. Compared to the enhanced tube’s R-134a results, a noticeable 5.5 times the heat transfer coefficient is realized over the heat flux range 10–60 kW/m2. At the lowest heat flux point, the enhancement is almost 10 times that of smooth tube
ho;smooth tube ¼ 596:281 q000:612
ð18Þ
The enhancement type of the current surface relies on increasing the number of nucleation sites, a typical approach for increasing the boiling performance (unlike using extended surfaces or integral fins). Nucleation site density is increased through creating plentiful pores and cavities; as opposed to the smooth surface where the nucleation sites are naturally present due to, for example, surface imperfection, grain structure, and polish. Boiling cavities are created by cold working of the tube surface, resulting in what is widely known as structured surfaces, to provide active evaporation cavities where the liquid pool does not flood the cavities.
Table 3 Input uncertainties. uT (°C)
um (kg/s)
uhi (W/m2 °C)
0.02
0.15% reading
4.2% hi
ð17Þ
The boiling mechanism in the surface is an oscillation cycle between vapor and liquid such that the vapor escaping the pores is immediately replaced by liquid. The heat transfer performance is enhanced due to two processes (also presented in Nakayama et al. [15]). The first enhancement is due to latent heat (liquid evaporation) while the second is bubbles convection on the surface. Additionally, the heat transfer coefficient behavior versus heat flux can help demonstrate the effect of the liquid feed mechanism [5]. The liquid feed to the enhanced surface has a substantial effect on the tube performance. At low heat flux the film thickness in the cavities is considerable thus the low performance. The performance reaches it peak as the film thickness decreases while the cavity walls remain wetted. For shell-side boiling on enhanced surfaces, the heat flux is typically the dominant effect (as outlined in Gorgy and Eckels [6]). Figs. 2 and 3 compare respectively the performance results of the medium and low pressure refrigerants versus the targeted replacement. Meanwhile, Fig. 2 shows results for R-134a compared to those for R-1234ze and R-450a. Considering the uncertainty limits, R-134a and R-1234ze present nearly the same results, as was previously confirmed by Rooyen and Thome [19]. Therefore, R-1234ze proves to be a viable R-134a replacement for the specified types of application. R-450a, on the other hand, showed a significantly lower performance than R-134a on average, a performance drop of nearly 28%. While R-450a’s thermodynamic properties are slightly closer to those of R-134a than is so between R-1234ze and R-134a, the lower performance of R-450a can be explained as due to the temperature glide of the zeotropic blend R-450a compared to the pure refrigerant R-1234ze. The degradation of the zeotropic blend’s boiling performance as compared to pure refrigerants was also highlighted in Thome and Shock [25]. With respect to low pressure refrigerants, Fig. 3 compares R-123 and R-1233zd(E), the latter showing a performance improvement of 19%. While R-1233zd(E) shows a significant
Fig. 2. Medium pressure refrigerants comparisons.
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E. Gorgy / International Journal of Heat and Mass Transfer 96 (2016) 660–666
Fig. 3. Low pressure refrigerants comparison.
Fig. 4. R-134a and R-1233zd(E) comparison.
improvement over R-123, it still comes 19% short of the performance of R-134a as Fig. 4 shows. In this case, R-1233zd(E) proves to be a better substitute for the targeted replacement R-123. Comparing the thermodynamic properties of R-1233zd(E) to R-123 shows the former has a lower vapor specific volume and a higher liquid specific volume, which could mean that for the same boiling surface enhancement (pore and tunnel size), the lower vapor specific volume allows for better pore activation (higher bubble convection), and the higher liquid specific volume allows for better liquid coverage of the surface enhancement (liquid latent heat), hence the higher heat transfer coefficient. The same logic applies when explaining the performance increase of R-134a over R-1234zd(E) where the former has lower vapor specific volume and higher liquid specific volume. Next, Table 4 presents the model coefficients associated with each fluid using the pool boiling heat transfer coefficient model (Eq. (19)). The applicable units are W/m2 °C and kW/m2 for the heat transfer coefficient and heat flux, respectively. For increased
Table 4 Model summary.
Co n % uho;max % uho;min
R-134a
R-1234ze
R-450a
R-123
R-1233zd (E)
20417.08 0.0851 24 6
23685.84 0.039 26 6
7260.53 0.286 18 5
5697.525 0.329 17 5
11677.83 0.184 22 5
model accuracy, a heat flux range of 10–60 kW/m2 (a typical operating range for refrigerant evaporators) was used. Within that range, the heat transfer coefficient experiences an exponentialfunction rise. Noticeably, the uncertainty range of each refrigerant correlates with the performance level. The high uncertainty is indicative of better fluid performance where a smaller temperature difference between the inlet and outlet temperatures, or low heat flux, is observed
666
hmodel ¼ C o q00n :
E. Gorgy / International Journal of Heat and Mass Transfer 96 (2016) 660–666
ð19Þ
6. Conclusion This paper presents the experimental investigation of three low GWP refrigerants: R-1234ze, R-1233zd(E), and R-450A. The first two HFOs have a GWP less than or equal to 1 while the zeotropic blend R-450a has a GWP index of 547. Nucleate boiling performance of R-1234ze and R-450a were compared to their targeted replacement, R-134a; and R-1233zd(E) to its targeted replacement, R-123 conducting the test using a highly enhanced tube surface. Moreover, data was collected at a saturation temperature of 4.44 °C and a heat flux range of 10-110 kW/m2. Also, the study included an investigation of the suitable tubeside turbulent flow correlation. A Monte-Carlo simulation for the tube-side uncertainty analysis helped show that the Petukhov correlation is consistent at Prandtl numbers of 6.5 and 9.0. Results show that the performance of R-1234ze is very similar to that of R-134a while R-450a demonstrates a significant 28% degradation. For the R-123 comparison, R-1233zd(E) demonstrates a noticeable 19% increase in performance. Finally, an exponential-function model was introduced for all five fluids where the applicable heat flux covered a range between 10 and 60 kW/m2, a typical range of interest for refrigerant evaporator design. Acknowledgements The efforts of Mark Spatz, Samuel Motta, and Ankit Sethi of Honeywell, who donated the refrigerant samples for this study, are greatly appreciated. Moreover, thanks to Joey Smith, Mark Owens, Justin Ryan, and Ben Turnbull for helping with samples manufacturing and testing. References [1] D.E. Briggs, E.H. Young, Modified Wilson plot techniques for obtaining heat transfer correlations for shell and tube heat exchangers, in: Chemical Engineering Progress Symposium Series, 92 (65), 1969, pp. 35–45. [2] Fluorinated greenhouse gases, F-gas facts,
[5] E.I. Gorgy, Experimental evaluation of heat transfer impacts of tube pitch on highly enhanced surface tube bundle (Ph.D. Dissertation), Kansas State University, Chapter 8, 2011. [6] E.I. Gorgy, S. Eckels, Convective boiling of R-134a and R-123 on a standard pitch enhanced tube bundle, HVAC&R Res. J. 19 (2) (2013) 193–206. [7] V. Gnielinski, New equations for heat and mass transfer in turbulent pipe and channel flow, Int. Chem. Eng 16 (2) (1976) 359–368. [8] M. Kedzierski, K.J. Park, Horizontal convective boiling of R-134a, R-1234yf/R134a, and R-1234ze(E) within a micro-fin tube with extensive measurement and analysis details, Nist Technical Note 1807, 2013. [9] N.H. Kim, K.K. Choi, Nucleate pool boiling on structured enhanced tubes having pores with connecting gaps, Int. J. Heat and Mass Transfer 44 (1) (2001) 17–28. [10] S.J. Kline, F. McClintock, Describing uncertainties in single-sample experiments, Mech. Eng. 75 (1) (1953) 3–8. [11] E.W. Lemmon, M.L. Huber, M.O. McLinden, NIST thermodynamic and transport properties of refrigerants and refrigerant mixtures (REFPROP) version 8.0. NIST, 2007. [12] D.A. Lashof, D.R. Ahuja, Relative contribution of greenhouse gas emissions to global warming, Lett. Nat. 334 (6266) (1990) 529–531. [13] M. Li, C. Dang, E. Hihara, Flow boiling heat transfer of HFO 1234yf and R32 refrigerant mixtures in a smooth horizontal tube: part I. Experimental investigation, Int. J. Heat Mass Transfer 55 (13–14) (2012) 3437–3446. [14] G. Moreno, S. Narumanchi, C. King, Pool boiling heat transfer characteristics of HFO-1234yf on plain and microporous enhanced surfaces, J. Heat Transfer 135 (11) (2013) 111014-1–111014-10. [15] W. Nakayama, T. Daikoku, H. Kuwahara, T. Nakajima, Dynamic model of enhanced boiling heat transfer on porous surfaces part II: analytical modeling, J. Heat Transfer 102 (1980) 451–456. [16] B.S. Petukhov, Heat transfer and friction in turbulent pipe flow with variable physical properties, Adv. Heat Transfer 6 (1970) 503–565. [17] G. Ribatski, J.R. Thome, Two-phase flow and heat transfer across horizontal tube bundles – a review, Heat Transfer Eng. 28 (6) (2007) 508–524. [18] D.M. Robinson, J.R. Thome, Local bundle boiling heat transfer coefficients on a turbo-BII HP tube bundle (RP-1089), HVAC&R Res. 10 (4) (2004) 441–457. [19] E.V. Rooyen, J.R. Thome, Pool boiling data and prediction method for enhanced boiling tubes with R-134a, R-236fa, and R-1234ze(E), Int. J. Refrig. 36 (2) (2013) 447–455. [20] W.M. Rohsenow, A method of correlating heat transfer data for surface boiling of liquids, ASME Trans. 74 (1952) 969–976. [21] S. Saitoh, C. Dand, Y. Nakamura, E. Hihara, Boiling heat transfer of HFO-1234yf flowing in a smooth small-diameter horizontal tube, Int. J. Refrig. 34 (8) (2011) 1846–1853. [22] R.K. Shah, Assessment of modified Wilson plot techniques for obtaining heat exchanger design data, in: Heat Transfer, Proceedings of the International Heat Transfer Conference 5, 1990, pp. 51–56. [23] E.N. Sieder, G.E. Tate, Heat transfer and pressure drop of liquids in tubes, Ind. Eng. Chem. 28 (12) (1936) 1429–1435. [24] R.A. Tatara, P. Payvar, Pool boiling of pure R134a from a single Turbo-BII-HP tube, Int. J. Heat Mass Transfer 43 (12) (2000) 2233–2236. [25] J.R. Thome, R.A.W. Shock, Boiling of multicomponent mixtures, Adv. Heat Transfer 16 (1984) 59–156. [26] X. Wang, K. Amrane, AHRI low global warming potential alternative refrigerants evaluation program (Low-GWP AREP) – summary of phase I testing results, in: International Refrigeration and Air Conditioning Conference. Paper 1416, 2014. [27] R.L. Webb, C. Pais, Nucleate pool boiling data for five refrigerants on plain, integral-fin, and enhanced tube geometries, Int. J. Heat Mass Transfer 35 (8) (1992) 1893–1904.