DOI: 10.3303/CET2188066
Paper Received: 18 June 2021; Revised: 1 September 2021; Accepted: 11 October 2021
Please cite this article as: Kukulka D.J., Li W., Smith R., 2021, An Experimental Investigation to Determine the Effect of Tube Material on the
Tubeside Heat Transfer Performance of the Enhanced 1EHT Three Dimensional Heat Transfer Tube, Chemical Engineering Transactions, 88,
397-402 DOI:10.3303/CET2188066
CHEMICAL ENGINEERING TRANSACTIONS
VOL. 88, 2021
A publication of
The Italian Association
of Chemical Engineering
Online at www.cetjournal.it
Guest Editors: Petar S. Varbanov, Yee Van Fan, Jiří J. Klemeš
Copyright © 2021, AIDIC Servizi S.r.l.
ISBN 978-88-95608-86-0; ISSN 2283-9216
An Experimental Investigation to Determine the Effect of Tube
Material on the Tubeside Heat Transfer Performance of the
Enhanced 1EHT Three Dimensional Heat Transfer Tube
David John Kukulkaa,*, Wei Lib, Rick Smithc
a State University of New York College at Buffalo, 1300 Elmwood Avenue, Buffalo, New York, 14222 USA,
b Department of Energy Engineering, Zhejiang University, 866 Yuhangtang Road, Hangzhou 310027, PR China
c Vipertex, 658 Ohio Street, Buffalo New York. USA
kukulkdj@buffalostate.edu
Condensation heat transfer characteristics were experimentally investigated over a wide range of operating
conditions in order to determine the heat transfer performance inside horizontal, smooth and enhanced heat
transfer tubes; using R410A in tubes produced of copper and stainless steel. Experimental data was verified
and results were compared to the performance measured in a smooth tube. Results indicate that the
condensation heat transfer coefficient (HTC) enhancement ratio is in the range from 1.15 to 2.05 for the 1EHT
tube and for the HX tube it ranged from 1.18 to 1.69. Smooth tube heat transfer performance was slightly
affected by the thermal conductivity of the tube; however, larger enhancements are found in the enhanced
tubes.
Heat transfer coefficients increase with an increase of mass velocities. When the mass flux increases, the liquid
flow becomes more turbulent and the liquid film becomes thinner; this reduces the thermal resistance and
enhances the heat transfer. Heat transfer performance for low mass velocities rise slowly, showing only a small
difference in magnitude. Performance increase is larger at high mass flux rates than that those found at low
mass fluxes. The influence of thermal conductivity on the condensation heat transfer of the enhanced horizontal
tubes was discussed. Better heat transfer performance occurs in tubes produced of a higher thermal conductivity
material (copper) or in tubes with a smaller diameter.
1. Introduction
Heat transfer enhancement methods are efficient ways to conserve energy and various aspects have been
previously investigated. Surface enhancement is a heat transfer technique commonly employed in various
industries (i.e. air conditioning, aerospace, refrigeration, etc.) to enhance system performance. Condensation
heat transfer characteristics were experimentally investigated to determine the heat transfer performance inside
enhanced heat transfer tubes and smooth tubes; with tubes produced using different tube materials (copper and
stainless steel); using R410A; for a variety of operating conditions.
Li et al. (2020) evaluated heat transfer performance of several stainless-steel enhanced surface tubes with the
smooth, herringbone, helix micro-grooves, herringbone-dimple, and hydrophobic surfaces. Gu et al. (2020)
conducted experiments to study the condensation heat transfer characteristics of moist air outside three-
dimensional (3D), finned tubes. Zhang et al. (2018) experimentally studied the condensing heat transfer
characteristics (when using R410A) inside smooth and micro-fin tubes. Zhao et al. (2017) studied the influence
of surface structure and thermal conductivity on the condensation HTC inside enhanced 2D and 3D finned tubes
made of iron cupronickel and aluminium brass. Ji et al. (2014) discuss the material/conductivity differences of
enhanced tube materials and relate thermal conductivity differences to fin efficiency. Ali et al. (2013) reported
experimental data for condensation when using ethylene glycol and R-113 in three identical pairs of pin-fin tubes
made of copper, brass and bronze. Tang et al. (2020) observed and analyzed the flow patterns during
condensation (using R410A) inside three-dimensional enhanced tubes made of stainless steel and copper. Their
results show that the transition between annular flow and intermittent flow was shifted to a lower vapor quality
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for the enhanced tubes. Li et al. (2017), they evaluated the condensation heat transfer characteristics (using
R22 and R410A) in micro-fin tubes and compared the performance to smooth horizontal tubes, with outer
diameters of 5 mm and 9.52 mm. Kukulka et al. (2014) evaluated the inside condensation and evaporation heat
transfer of R410A, R22 and R32 that took place in a 12.7 mm horizontal cooper tube with low mass fluxes; they
found that tubeside evaporation and condensation HTC enhancement of the 1EHT tube is approximately 2 times
greater than smooth tubes. Kukulka et al. (2019) conducted an experimental investigation to explore the heat
transfer coefficient and the frictional pressure drop during condensation and evaporation using 1EHT tubes over
a limited range of conditions. As can be seen there is little published data regarding enhanced EHT tubes and
this apparent lack of information provides the motivation for this study. Little published data exist on: comparative
analysis of various enhanced dimensional tubes; evaluation of tube material; heat transfer performance
evaluation; heat transfer performance of different sized enhanced tubes. The work of Tang et al. (2020) and Li
et al. (2017) have been extended in this study and includes: a smooth tube, 2D helix micro-fin tube and a 3D
1EHT tube; with different tube thermal conductivities and diameters. An investigation of the condensation heat
transfer characteristics was performed using different tube thermal conductivities, tube diameters, and surface
structures.
2. Experimental Details
Figure 1 is the experimental apparatus used in this study. Table 1 provides the main parameters of the tubes
tested; pictures of the enhanced surfaces are presented in Figure 2.
Figure 1 Schematic diagram of the experimental setup
Table 1: Geometric parameters of the tested tubes
Parameter Smooth tube 1EHT tube HX tube
Material Cu/SS Cu/SS Cu/SS
Outer diameter (mm) 9.52/12.7 9.52/12.7 9.52/12.7
Thickness (mm) 0.61 0.61 0.61
Length (mm) 2 2 2
Dimple/fin height (mm) - 0.19/1.71 0.25
Dimple/fin width (mm) - 0.35/1.34 0.31
Dimple/fin pitch (mm) - 4 0.8
Helix angle (°) - 60 21
Surface Area enhancement ratio (1) 1 1.44 1.34
Deionized water is the working medium used in the constant temperature water bath in the experiment section;
industrial alcohol is used to run at a lower temperature in the supercooling section. As shown in Figure 1, the
test section is a typical horizontal, tube-in-tube heat exchanger with a heated length of 2.0 m. Tubes evaluated
are utilized as the inner tube, a copper tube with an outer diameter of 17.0 mm is utilized as the outer tube.
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Experimental conditions to investigate the effects of tube diameter and material include: saturation temperature
of 45 ºC; mass flux values in the range of 75 to 400 kg m-2 s-1; with an inlet vapor quality of 0.8 and outlet vapor
quality of 0.2. Maximum relative error of the HTC is calculated to be ± 11.32 %; Table 2 presents the maximum
relative errors of the measurement and calculation parameters.
(a) 1EHT tube (b) HX tube
Figure 2: Pictures of enhanced surfaces: (a) 1EHT (b) HX
Table 2: Maximum relative errors of measurement and calculation parameters
Measurement Parameters Uncertainty
Diameter (mm) ± 0.05
Length (mm) ± 0.2
Temperature (K) ± 0.05
Pressure (range: 0 - 40 bar) ± 0.2 % of full scale
Differential pressure (range: 0 - 100 kPa) ± 0.05 % of reading
Water flow rate (range: 0 - 12 L min-1) ± 0.35 % of reading
Refrigerant flow rate (range: 0 - 60 kg h-1) ± 0.2 % of reading
Calculation Parameters Uncertainty
Mass flux (kg m-2s-1) ± 3.25 %
Heat flux (W m-2) ± 4.71 %
Vapor quality ± 6.30 %
Heat transfer coefficient (W m-2 k-1) ± 11.32 %
3. Results
Data were evaluated using the necessary fluid properties from Lemmon et al. (2010). Heat transfer, 𝑄𝑡𝑒, in the
test section is determined using the heat flux of the annular waterside, 𝑄𝑤,𝑡𝑒, as shown in Eq(1)：
𝑄𝑡𝑒 = 𝑄𝑤,𝑡𝑒 = c𝑤 𝑚𝑤,𝑡𝑒 (𝑡𝑤,𝑡𝑒,𝑜𝑢𝑡 − 𝑡𝑤,𝑡𝑒,𝑖𝑛) (1)
where c𝑤 is the specific heat capacity of water; 𝑚𝑤,𝑡𝑒, water mass flux in the test section; 𝑡𝑤,𝑡𝑒,𝑜𝑢𝑡 and 𝑡𝑤,𝑡𝑒,𝑖𝑛
are the outlet and inlet water temperature in the test section. In order to perform a heat balance analysis, the
heat flux of the refrigerant, 𝑄𝑟𝑒,𝑡𝑒, is calculated using:
𝑄𝑟𝑒,𝑡𝑒 = 𝑚𝑟𝑒,𝑡𝑒 (𝑥𝑖𝑛𝑖𝑣,𝑖𝑛 + (1 − 𝑥𝑖𝑛)𝑖𝑙,𝑖𝑛 − [(1 − 𝑥𝑜𝑢𝑡 )𝑖𝑙,𝑜𝑢𝑡 + 𝑥𝑜𝑢𝑡𝑖𝑣,𝑜𝑢𝑡 ]) (2)
where 𝑚𝑟𝑒,𝑡𝑒 is the refrigerant mass flux in the test section; 𝑖𝑣,𝑖𝑛 and 𝑖𝑣,𝑜𝑢𝑡 are the gaseous enthalpy of the
refrigerant in the test section inlet and outlet; 𝑖𝑙,𝑖𝑛 and 𝑖𝑙,𝑜𝑢𝑡 are the liquid enthalpy of refrigerant in the test section
inlet and outlet. Inlet and outlet vapor quality of the refrigerant, 𝑥𝑖𝑛 and 𝑥𝑜𝑢𝑡 are determined from:
𝑥𝑖𝑛 =
𝑖𝑖𝑛 − 𝑖𝑠
𝑖𝑣
(3)
𝑥𝑜𝑢𝑡 =
𝑖𝑜𝑢𝑡 − 𝑖𝑠
𝑖𝑣
(4)
where 𝑖𝑖𝑛 is the enthalpy of the refrigerant at the inlet of the test section; 𝑖𝑜𝑢𝑡 is the enthalpy of the refrigerant at
the outlet; 𝑖𝑠, the enthalpy of the saturated liquid; and 𝑖𝑣, the vaporization enthalpy of the refrigerant at the
saturation temperature. Enthalpy at the outlet of the preheat exchanger can be computed from the enthalpy at
the inlet of the preheat exchanger, 𝑖𝑝𝑟,𝑖𝑛; heat flux in the preheat exchanger, 𝑄𝑝𝑟; and the refrigerant mass flux
in the preheat exchanger, 𝑚𝑟𝑒, as follows:
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𝑖𝑖𝑛 = 𝑖𝑝𝑟,𝑖𝑛 +
𝑄𝑝𝑟
𝑚𝑟𝑒
(5)
Similarly, the enthalpy of the refrigerant at the outlet of the test section is calculated from the enthalpy at the
test section inlet, 𝑖𝑡𝑒,𝑖𝑛; heat flux in the test section, 𝑄𝑡𝑒; and the refrigerant mass flux, 𝑚𝑟𝑒 , as follows:
𝑖𝑜𝑢𝑡 = 𝑖𝑖𝑛 +
𝑄𝑡𝑒
𝑚𝑟𝑒
(6)
The heat transfer coefficient is computed in Eq(7):
ℎ𝑟𝑒,𝑡𝑒,𝑖 =
1
𝐴𝑖 [
𝐿𝑀𝑇𝐷
𝑄𝑡𝑒
−
1
ℎ𝑤,𝑡𝑒,𝑜𝐴𝑜
−
𝑙𝑛(𝑑𝑜 /𝑑𝑖 )
2𝜋𝐿 · 𝑘
]
(7)
where 𝐴𝑖 and 𝐴𝑜 are the heat transfer areas of the refrigerant side and waterside; ℎ𝑤,𝑡𝑒,𝑜 , waterside heat
transfer coefficient; 𝐿 , tube length; 𝑘, thermal conductivity of the tested tube; 𝑑𝑖 , nominal inner diameter of the
tube evaluated; 𝑑𝑜 , outer diameter of the evaluated tube and LMTD, the logarithmic mean temperature
difference.
Effects of tube parameters (wall thermal conductivity and tube inner diameter -ID) on the inside tube HTC is
discussed. Results showing the condensation HTCs inside copper and stainless-steel tubes with a 12.7-mm
OD; as a function of mass flux are compared in Figure 3. Differences in enhanced tube performance are the
result of differences in tube material thermal conductivity. Smooth tube copper HTCs are slightly higher than
those found in the SS smooth tubes; while the HTCs inside the copper enhanced tubes are significantly higher
than enhanced SS tubes.
Effects of tube material thermal conductivity on the inside condensation heat transfer of enhanced tubes was
determined from the experimental results. Performance differences observed in tubes produced from different
tube material conductivity is explained because of temperature differences produced by the fin efficiency of the
enhanced surface. The structure of the enhanced surfaces produces different temperature gradients for different
thermal conductivity values; the actual temperature of the heat transfer surface is not the same as the fin root
temperature. For an enhanced surface with a lower thermal conductivity, a higher actual temperature of heat
transfer surface leads to a lower fin efficiency; this leads to poorer condensation heat transfer performance.
Meanwhile, the area enhancement ratio of the HX tube is 1.44, which is greater than that of the 1EHT tube
(1.34). Hence, the HX tube is more susceptible to thermal conductivity differences. Heat transfer enhancement
of the HX tube made of copper is 1.13 to 1.18 times that of the SS tube for the same operating conditions; while
enhancement ratio for the 1EHT tube is only 1.01 to 1.09. In conclusion, the tube thermal conductivity has a
more significant effect on the HTC for the HX tube than on the 1EHT tube.
For mass flux values larger than 100 kg m-2 s-1, the HTCs of all three tubes increase with an increase of the
mass flux. However, for the enhanced tubes, the HTCs decrease with an increase of the mass flux when the
refrigerant mass flux is less than 100 kg m-2 s-1; at those values the main flow pattern is a stratified-wave flow;
that is, the inner surface of the tube is divided into a submerged surface and unsubmerged surface. Gas-phase
refrigerant transfers heat directly with the wall surface on the unsubmerged surface; this has a greater heat
transfer efficiency. However, the latent heat released by the phase transition at the submerged surface can only
reach the wall surface through the condensate by thermal conduction. As the mass flux increases, the shear
force makes the condensate spread over the inner surface of the tube and causes a more serious inundation
problem; the enhanced surface improves the function of surface tension and in turn intensifies the phenomenon;
explaining why this trend appears only in the enhanced tubes and not in the smooth tube. Additional research
is needed in this area to fully understand this.
Figure 4 presents the thermal resistance for in tube condensation heat transfer; for a mass flux of 200 kg m-2s-
1. It can be seen that the thermal resistance ratios change with the thermal conductivity of the tube. Finally, only
a slight variation occurs with the surface structure. Thermal conduction resistance of the wall is a small portion
of the total resistance; the thermal resistance of the convection heat transfer on the refrigerant side dominates
the process. The external surface of the 1EHT tube is enhanced, resulting in a lower outside thermal resistance;
this is more evident in the SS tube. Total HTC of the 1EHT tube is the highest with both internal and external
enhancement; it is followed by the HX tube and finally the smooth (ST) tube. Figure 5 shows the effect of tube
diameter on the condensation HTC. For the most part, the HTCs of the 9.52-mm-OD tubes are higher than that
of 12.7-mm-OD tubes. This can partially be explained by the fact that as the tube diameter decreases, the shear
force and surface tension gradually take the place of gravity and they become the dominant forces; this is
beneficial to removing and thinning the liquid film at the bottom. Additionally, the tube surface with a small
diameter has a higher area density (ratio of surface area to volume); this leads to higher heat flux per unit
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volume. The trend of HTCs initially rising and then falling also appeared in the 9.52-mm-OD enhanced tubes;
however, the turning point is delayed.
Figure 3: Comparison of condensation heat transfer coefficients inside 12.7-mm-OD copper and stainless steel
tubes with various mass fluxes
(a) (b)
Figure 4: Comparison of the thermal resistance for (a) copper tubes and (b) stainless steel tubes
4. Conclusions
An experimental investigation of tubeside condensation heat transfer characteristics was performed using
R410A in horizontal tubes that were enhanced and smooth. Effects of tube diameter and tube conductivity on
the tubeside condensation heat transfer were discussed. The following conclusions can be drawn:
1. Smooth tube thermal-hydraulic performance is slightly affected by the thermal conductivity of the tube;
however, the HTC of the enhanced tubes increase significantly with higher thermal conductivity tube
material.
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Figure 5: Condensation HTCs inside 9.52-mm-OD and 12.7-mm-OD copper tubes with various mass fluxes
2. Smaller tube diameters can achieve a better heat transfer performance; enhanced heat transfer is
more obvious in the enhanced tubes.
3. It can also be concluded that tubes produced of higher thermal conductivity material or using a smaller
diameter of the tube will lead to better heat transfer performance. They should be considered for high
performance heat transfer systems.
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