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Enhancement of convective quenching heat transfer by coated tubes and intermittent cryogenic pulse flows
International Journal of Heat and Mass Transfer 141 (2019) 256–264
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International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
Enhancement of convective quenching heat transfer by coated tubes and intermittent cryogenic pulse flows J.N. Chung a,⇑, Jun Dong a, Hao Wang a, S.R. Darr a, J.W. Hartwig b a b
Cryogenics Heat Transfer Laboratory, Department of Mechanical and Aerospace Engineering, University of Florida, Gainesville, FL 32611-6300, USA NASA Glenn Research Center, Cleveland, OH 44135, USA
a r t i c l e
i n f o
Article history: Received 9 February 2019 Received in revised form 30 May 2019 Accepted 23 June 2019 Available online 27 June 2019 Keywords: Cryogenic fluid Quenching heat transfer Pulse flow Two-phase flow Film boiling
a b s t r a c t This paper reports heat transfer enhancement techniques for the cryogenic quenching process. An experiment was performed to evaluate the enhancement of quenching heat transfer by two techniques: 1. Using intermittent pulse flows and 2. Coating the inner surface of a test tube with layers of low thermal conductivity films. Pulsed liquid nitrogen flows with various duty cycles and periods were applied in the quenching of a room-temperature metal tube coated with four layers of Teflon thin films on the inner surface. In general, the results obtained indicate that the quenching thermal efficiency that measures the effectiveness of heat transfer enhancement increases with decreasing duty cycle, however it is relatively independent of the period. Comparing with non-coated bare surface test tube, the low-conductivity coating substantially improved the thermal efficiency and reduced the total quenching time. Additionally, the thermal efficiency was found to increase with decreasing source inlet pressure or coolant mass flow rate. The savings on the amount of cryogen consumed follow the same trends as those for the thermal efficiency with pulse flows and tube coating. Ó 2019 Elsevier Ltd. All rights reserved.
1. Introduction 1.1. Importance of quenching heat transfer In many convective liquid-vapor phase change heat transfer applications, cryogenic fluids are widely used in industrial processes, spacecraft and cryosurgery systems, and so on. For example, cryogens are usually used as liquid fuels such as liquid hydrogen and oxygen in the rocket industry, liquid nitrogen (LN2) and helium are frequently used to cool superconducting magnetic device for medical applications. In these systems, proper transport, handling, and storage of cryogenic fluids are of great importance. When a cryogenic system is first started up, its walls and hardware must go through a transient chilldown period prior to reaching a steady operation. Therefore, chilldown (quenching) is the process of keeping the system adjusted to the low temperature scale which is usually several hundred degrees below the room temperature. The chilldown or quenching process is complicated, involving unsteady two-phase heat and mass transfer, and has not been fully understood.
⇑ Corresponding author. E-mail address: [email protected] (J.N. Chung). https://doi.org/10.1016/j.ijheatmasstransfer.2019.06.080 0017-9310/Ó 2019 Elsevier Ltd. All rights reserved.
A new breakthrough advancement in the quenching phasechange heat transfer processes by a low-thermal conductivity surface coating can lead to much higher quenching efficiencies, substantial energy and liquid cryogen consumption savings, and global reduction in greenhouse gas emissions. Especially for space propulsion systems, higher quenching efficiencies translate to more fuels available for making longer and deeper space missions a reality. This paper reports an advancement on quenching heat transfer. 1.2. Characteristics of quenching heat transfer A quenching (chilldown) process is a liquid-to-vapor phase change phenomenon that is governed by the ‘‘boiling curve”. This curve [1] as shown in Fig. 1 illustrates the heat transfer surface heat flux, q00 , plotted against the surface degree of superheating, T W T sat , where T W is the surface temperature and T sat is the saturation temperature corresponding to the boiling fluid bulk pressure. In boiling, if the heating source is externally supplied to the heater surface such as the embedded electrical resistance heating element, the process is heat-flux controlled and it follows the route of A ? B ? D. In contrast, during quenching the warm wall where the heat comes out does not have a source of heat supply, therefore, the heat transferring out of the wall can only come internally from the thermal capacity (stored energy) of the wall. The only
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Fig. 1. A typical boiling curve including different boiling regimes and corresponding flow patterns.
way to remove heat from the wall is by lowering the inner wall surface temperature using a cooling flow. So quenching is a wall surface temperature-controlled process. Thus, a quenching process follows the route D ? C ? B ? A. During quenching, film boiling is always the first mode of heat transfer encountered due to a relatively very hot surface. Owing to its relatively low heat fluxes at high wall temperatures, film boiling usually dominates the quenching time and cannot be avoided. As a result, in traditional quenching processes, the thermal energy efficiency is extremely low. According to Shaeffer et al. [2], the average quenching efficiency that is defined as the ratio of the amount of thermal energy removed from the wall versus the required cooling capability of the cryogen spent in a quenching process is about 8%, highlighting the tremendous need to improve the quenching efficiency for many applications that require cryogens as the working fluid. 1.3. Convective quenching heat transfer in tubes and pipes Compared to the heated wall convective boiling case, the amount of literature on convective quenching heat transfer is far less. For convective quenching in tubes and pipes, most studies used water and normal room temperature fluids [3–5]. Yuan and Chung [6] and Yuan et al. [7] reported the pioneering work on liquid nitrogen (LN2) cryogenic chilldown in conventional stainless tubes under low flow rates in both terrestrial and simulated drop-tower microgravity conditions. Both papers [6,7] established the fundamental physical understanding and benchmark data for future convective cryogenic chilldown heat transfer research. As follow-ups, LN2 chilldown experiments of a short stainless steel tube have been carried out at low mass flux both at constant mass flux [8] and in a pulsed on/off flow [2]. Only recently has there been a thorough experimental investigation of the entire chilldown process in 1-g that reports the measured heat transfer coefficients for all boiling regimes over a large range of mass flow and thermodynamic conditions [9,10]. Another recent work [11] by the same group provided a wide ranging dataset of reduced gravity cryogenic pipe chilldown data from a parabolic flight experiment using LN2 onboard a C9 aircraft. Heat transfer data were obtained from
flowing LN2 through a short, stainless steel (SS) pipe in microgravity for developing heat transfer correlations. This experiment was an extension of 1-g experiments reported in Darr et al. [9,10]. 1.4. Effects of flow pulsation on heat transfer in tubes and pipes For single-phase liquid or gas convective heat transfer, the pulsation in the forms of a wave with a fixed amplitude and frequency such as a sinusoidal wave, results by Ishino et al. [12] pointed out that the heat transfer (Nusselt number) decreases initially as the amplitude increases, but then recovers gradually, and finally becomes much greater than the initial values. In general, most experimental results [13–16] showed that pulsation could either decrease or increase the heat transfer. Some analytical investigations [17–20] reported that for pulsating flows, the Nusselt numbers fluctuate periodically around value of steady flow for constant wall temperatures and constant wall heat fluxes in fully developed region. Some results [21–23] predicted by numerical simulations of pulsed flow heat transfer showed that the pulsing does not affect the time-averaged heat transfer in a pipe. The results also indicated the Nusselt number distribution varies with time in the near-entry region. In summary, for single-phase pulsating flows, usually the results among many investigators showed contradictory trends due to different definitions of the Nusselt number. For heat transfer with phase-change, pulse flow has not received much attention. We only found two reports [2,24] that are all dealing with cryogenic quenching heat transfer and are discussed below. Shaeffer et al. [2] performed a comparison between continuous flow and intermittent on/off pulse flow patterns on quenching heat transfer in a stainless steel bare surface tube without any coating. Liquid nitrogen was used as the working fluid and upward flows with Reynolds numbers ranging from 2500 to 7000 in a round tube were examined. To gain a better understanding of the effect of flow pulsation on the chilldown process, three different pulse flows (4-s, 10-s and 20-s periods and all with 50% duty cycle) were used to compare with the corresponding continuous flows. It is noted that
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the duty cycle, DC, which is the ratio of valve open time (t open ) to its period (topen þ t closed ), where t closed is the valve closed time, therefore, DC ¼ t open =ðtopen þ t closed Þ. For the chilldown of the entire process from film boiling to nucleate boiling regimes or just only the film boiling regime, the continuous flow pattern is favored for relatively low Reynolds flows. Specifically, the continuous flow prevailed over pulse flows if the Reynolds number was less than 4000. At higher Reynolds number flows, pulse flow patterns are more advantageous and shorter periods are more ideal than longer ones. Hartwig and Vera [24] investigated liquid hydrogen quenching of a tube using both continuous and intermittent on/off pulse flows. They used 15-s period with a 67% duty cycle and 20-s period with a 50% duty cycle for the pulse flows. Only limited data were reported and there was no specific comparison. For the same flow rates and identical liquid hydrogen initial temperatures, the pulse flow required more time to chilldown the tube to liquid temperature. 1.5. Theoretical basis of current experiment 1.5.1. Flow pulsing A mathematical model was developed based on the theory of transient one-dimensional heat conduction at the interface between two semi-infinite solids [25] to help explain the higher heat transfer enhancement by a pulse flow. A complete solution for this unsteady and conjugate heat transfer in a convectiveconductive system can be found in Mathie [26]. Since the quenching of the hot tube wall by cold cryogenic coolant flow is a transient and unsteady heat transfer process that takes place at the interface between two materials, therefore by neglecting the curvature effects, the metal tube wall can be approximated as the semi-infinite slab A and the coolant fluid is assumed as the semi-infinite slab B. If the two free surfaces of the slabs were to be placed in contact at time, t = 0, the heat flux, q00A!B , at the interface that flows from A to B can be given as follows [25].
q00A!B ¼
kA ðT A;i T s Þ ðpaA tÞ
1=2
¼
kB ðT B;i T s Þ ðpaB tÞ1=2
ð1Þ
where kA and kB are thermal conductivities of A and B, respectively. T A;i and T B;i are initial temperatures of A and B before contact, respectively. aA and aB are thermal diffusivities of A and B, respectively. T s and t are the common interface temperature and elapsed time, respectively. Eq. (1) above is introduced simply to provide a physical explanation for the concept that pulsing the flow enhances the heat transfer. Eq. (1) basically indicates that when we place two materials A and B, each with a different temperature in contact, the transient heat fluxes q00A!B across the interface from A to B would be proportional to t 1=2 (t is the elapsed time after the initial contact of the two materials, so t = 0 is the time when the two are placed in contact) as the rest of the parameters (thermal properties) in the equation are all constants. The physical meaning is that initially the heat transfer would be very high and the heat flux decreases quickly as time passes. In other words, Eq. (1) is the transient solution that corresponds to a step change (jump) in the surface temperature. For the pulse flow process, the tube inner surface gets a jump in temperature and a corresponding spike in the heat flux every time when intermittent pulse flow is turned on after the off period in a duty cycle that results in higher heat transfer than the continuous flow case. Please note that the above simplified 1-D semi-infinite slab model is only used to explain the transient intermittent heat transfer nature due to pulsing and is not intended for the modeling of the entire pulsed flow quenching process. In addition to the pulse
flows, we also used the test tubes coated with low thermal conductivity thin films on the inner wall surface to shorten the time for the heater surface to reach the Leidenfrost temperature as a second enhancement technique that is explained next. 1.5.2. Low-thermal conductivity coating The enhancement concept of coated tube is based on that there are two heat transfer mechanisms involved that are opposite in nature to each other. The first mechanism is the thermal insulating effect due to the low-thermal conductivity thin layer that facilities a fast drop of the tube inner surface temperature by restricting the heat flow from the bulk of the tube wall to the tube inner surface. The lower surface temperature allows the quenching process to quickly move from film boiling regime to the Leidenfrost point to initiate the higher heat transfer transition and nucleate boiling regimes. The second mechanism is the conduction of heat from the bulk of the tube wall to the cooling fluid through the inner surface during transition and nucleate boiling regimes that requires the thermal conductivity of the bulk wall material to be as high as possible such as metals to expedite the wall cooling process. There should exist an optimal thickness of the low-thermal conductivity layer that balances these two competing mechanisms such that the coating is just thick enough to quickly lower the tube surface temperature to that of the Leidenfrost point while it is still relatively very thin to not substantially reduce the heat flow from the bulk to the cooling fluid. 1.6. Research objective The primary objective of the current set of experiments is to evaluate the effects of intermittent ON/OFF pulse flow on transient temperature history curves of metal tubes, or chilldown curves, from room temperature to LN2 saturation temperatures during a quenching process. The test tubes are horizontally-aligned commercially available 304 stainless steel tubes without coating and with coating using low thermal conductivity thin Teflon layers. Therefore, the secondary objective is to investigate the effects of coating with pulse flows. Tests were carried out with a set of pulse flow conditions with 0.5–5-s periods and 80%, 60%, 40%, and 20% duty cycles over a wide range of test section inlet pressure levels and corresponding mass flow rates. The effectiveness of the pulse flow was evaluated by a comparison of chilldown curves with pulsations to those with constant and continuous flows. 2. Experimental methods 2.1. Experimental apparatus and procedure The current LN2 experimental system was built with small modifications to our original rig used by Darr et al. [9,10] and Darr et al. [11]. A flow and component schematic of the experiment is shown in Fig. 2. To avoid duplication, the readers are referred to Darr et al. [11] for a detailed description of the flow network, and the function and operation of each component. Specifically, it is important to note that a subcooler was placed before the entrance of the test section. The function of the subcooler was to insulate the tubing upstream of the test section so that to make sure that the fluid entering the test section was vapor free, subcooled liquid that provides a precise inlet boundary condition for the test tube during chilldown. A key modification in the current experiment is the installation of a solenoid valve (SV) that facilitates the intermittent ON/OFF pulse flow. The solenoid valve was placed after the relieve valve RV5 and before the subcooler. As shown in Fig. 3, the test section was a 57.2 cm long, 1.270 cm OD, 1.168 cm ID 304SS tube. A length of 2.54 cm of the test section
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Fig. 2. Fluid system schematic. The valves and important components of the fluid network. Relief valve settings, the burst disk setting, and pressure regulator settings are also included. BD, burst disk; BV, ball valve; CV, check valve; DAQ, data acquisition unit; FM, flow meter; GN2, gaseous nitrogen; GV, globe valve; LN2, liquid nitrogen;; NV, needle valve; PC, pre-cooler; PG, pressure gauge; PR, pressure regulator; PT, pressure transducer; RV, relief valve; SV, solenoid valve; TC, thermocouple; Vap, vaporizer; 3 V, threeway valve.
Fig. 3. Test section dimensions and details.
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tube protruded out of each side of the vacuum chamber. Six TCs were soldered to the outside of the tube, with three placed an axial distance of 14.9 cm from the test section inlet (upstream TC station), and three placed 40.1 cm from the inlet (downstream TC station). As detailed in Fig. 3, for each station, the TCs were spaced out radially in 90° increments such that each station had a top, side, and bottom TC. Two cryogenic rated PTs were connected to two short pieces of 304SS tube protruding perpendicularly from the test section, one near each TC station. These tubes were welded to the test section. As shown in Fig. 3, the test section was a 57.2 cm long, 1.270 cm OD, 1.168 cm ID 304SS tube. Six TCs were soldered to the outside of the tube, with three placed an axial distance of 14.9 cm from the test section inlet (upstream TC station), and three placed 40.1 cm from the inlet (downstream TC station). As detailed in Fig. 3, at each station, the TCs were spaced out radially in 90° increments such that each station had a top, side, and bottom TC. Two cryogenic rated PTs were connected to two short pieces of 304SS tube protruding perpendicularly from the test section, one near each TC station. These tubes were welded to the test section. The 316SS vacuum chamber that housed the test section was the same as the original experiment described in Darr et al. [11]. The purpose of the vacuum chamber was to reduce parasitic heat leak that reduced the uncertainty in the calculation of chilldown heat flux. A diaphragm pump and molecular turbopump were used to bring the air pressure inside the chamber down to approximately 1 Pa. This reduced the parasitic heat leak due to conduction between the test section and the air inside the vacuum chamber to less than 10% of the lowest measured convective heat flux, which occurred during film boiling at the lowest flow rate that was tested. Parasitic heat leak was less than 1% of the measured value for most of the data points. Except for setting the duty cycle and period for the pulse flow case, the rest of the experimental procedure is identical to that followed in our previous chilldown experiment. So, the readers are again referred to Darr et al. [11] for the details of the experimental procedure. 2.2. Coating of tube inner surface In order to evaluate the second enhancing technique in the current experiment, two different tubes were used. One is a bare surface stainless steel test tube with no coating. The other is a stainless steel test tube that was coated with low-thermal conductivity thin Teflon layers on the inner surface. Specifically, the coating material was made of Fluorinated Ethylene Propylene (FEP) by DuPont and classified by DuPont as Teflon 959G-203 that is a black color paint and has a thermal conductivity of 0.195 W/m K (DuPont publication [27]). The coating was put on by the Matrix Coating Corporation (West Palm Beach, FL 33404, USA) using the pour and drain process. The final thickness of the coating material on the tube depends on the number of layers processed. After each pour and drain, the fresh film layer was cured in a furnace through a standard sintering procedure before adding another layer by the same pour and drain procedure. As a result, the final thickness of coated layer depends on the total number of layers processed. The coated test tube used in the experiment is a four-layer (4L) one. The 4L coating went through the pour and drain process four separate times. To determine the coating layer thickness, high resolution computer tomography (CT) X-ray scans of the tube cross sections were obtained using a Phoenix v|tome|x M (GE’s Measurement & Control business, Boston, USA) system in the Nano Research Facility at the University of Florida. Scanning was carried out using a 240 kV X-ray tube and a tungsten on beryllium target, with the following settings: 200 kV, 50 milliamps, and 0.5 mm Tin filter. Images were collected from 1600 pixels horizontal, 2024
pixels vertical, 0.5 s detector exposure, averaging of 4 images per rotation position with a one-exposure skip and a total of 2200 rotational positions. The coating thickness for this 4L tube was measured at 64.8 ± 0.7 mm. 2.3. Experimental uncertainties As mentioned above, the current experimental system is virtually identical to that employed in Darr et al. [9,10], the readers are referred to Darr et al. [9,10] for uncertainties on those independent, directly measured quantities, such as temperatures and pressures.
3. Results and discussion As mentioned in Section 2.2, two different test tubes were used in the current experiment. One is a tube with 4-layer (4L) coating and the other is a bare surface tube used as a reference case for evaluating the coating effects with pulse flows. Three inlet pressures of 5.5 105 Pa (Gauge) (80 psig), 4.1 105 Pa (Gauge) (60 psig), and 2.1 Pa (Gauge) (30 psig) were applied as the flow driving force. For the pulse flow settings, it is reminded that the duty cycle, DC, is defined as the valve open time (t open ) to its period (topen þ t closed ), where tclosed is the valve closed time, therefore, DC ¼ t open =ðt open þ tclosed Þ. Six different periods of 0.5 s, 1 s, 2 s, 3 s, 4 s and 5 s were used with the duty cycles set at 20%, 40%, 60% and 80%, respectively. We have found that duty cycles with settings less than 20% would induce flow instabilities that prevented steady conditions for experimentation. 3.1. Typical chilldown curves Before we present and discuss the effects of pulsed flows on the quenching heat transfer of a metal tube, the physics and characteristics of the tube quenching process must be clearly explained first. Fig. 4 shows a typical chilldown (quenching) curve recorded in the current experiment for the case where the inlet pressure, duty cycle and period were set at 80 psig, 20% and 1 s, respectively. The inner surface of the test tube was bare with no coating. The six chilldown curves provide the tube wall outer surface temperature histories during chilldown at both upstream and downstream locations from the tube inlet. As shown in Fig. 4, six chilldown curves were plotted using the temperature data registered by the six thermocouples (TCs) shown in Fig. 3(b), where TCs 1(top), 2(side) and 3(bottom) are located in the upstream station and TCs 4(top), 5(side) and 6(bottom) are located in the downstream station. In general, all six chilldown curves display similar trends where the heat transfer process during cooling of the tube went through film, transition and nucleate boiling regimes, sequentially as explained by the boiling curve (Fig. 1). The effects of the pulse flow are clearly reflected by the zigzag pattern of the chilldown curves. At a specific TC station, the chilldown time is the shortest for the bottom location, second shortest for the side, and longest for the top. For example, at the upstream TC station, the chilldown time is the shortest for the bottom location (TC3), second shortest for the side (TC2), and longest for the top (TC1). However, for the same circumferential location at a TC station, the downstream one lagged behind that in the upstream station due to the fact it takes more time for the cooling fluid to reach the downstream location. However, it should be pointed out that the downstream ones actually started to lower their temperatures at the same time as the upstream ones but at slower rates due to cooling by the axial conduction heat transfer in the metal tube wall.
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Fig. 4. A typical chilldown curve.
3.2. The effects of pulse flow on the enhancement of chilldown thermal efficiency The performance of a cryogenic chilldown system solely depends on the effectiveness of the heat transfer from the tube wall to the cryogenic coolant. Therefore, to measure the performance of a chilldown system, the chilldown thermal efficiency, gCD , as defined below is introduced. The chilldown thermal efficiency represents the percent of total required quenching capacity of the cryogenic coolant heat sink that is actually utilized in removing the thermal energy from the tube wall and part of a globe valve from their initial room temperature to that of the liquid nitrogen. The chilldown thermal efficiency in %, gCD , is therefore defined as,
Q removed 100% Q required
gCD ¼
ð2Þ
In the above, Q removed is the total thermal energy removed from the tube wall and part of a globe valve by the cooling fluid during the entire chilldown period and is defined as,
Q removed ¼ ðM tube þ 0:3M v alv e Þcp; SS T initial T final
ð3Þ
where Mtube and Mv alv e are the mass of the tube section and that of the three-way globe bypass valve (shown in Fig. 3), respectively. A factor of 0.3 is applied to the valve mass, as it is estimated that only 30% of the valve mass needs to be chilldown when the valve is opened and the chilldown process starts, since part of the valve is in contact with the LN2 before the opening of the valve. Therefore, the valve is partially (about 70%) chilled down during precooling and reheating of the test section where the LN2 fills the path all the way to the valve. cP; SS is the stainless steel specific heat for both the tube and valve materials. T Initial and T Final are the initial room temperature and the final temperature when the chilldown is completed, respectively. The end of chilldown temperature, T Final , is the LN2 saturation temperature corresponding to the local pressure. Q required is the total quenching capacity required during the chilldown process. It is defined as,
Q required ¼ mhfg
ð4Þ
where m is the total mass of coolant required or consumed during chilldown and it is defined in Eq. (5) below.
Z m¼
tEnd
_ mðtÞdt
mass, that means Q required is the total quenching capacity of the cryogen of mass, m, required during the chilldown process. Before presenting the chilldown efficiency results, the relative uncertainties for mass flow rate and the chilldown efficiency, respectively, must be provided first. The relative uncertainties for the mass flow rates, were estimated in the range of 3% to 7%. Next, the relative uncertainties for the chilldown efficiencies are estimated to be of the order of 15%. The chilldown thermal efficiencies for source pressure of 80 psig and test tube with 4L coating are given in Table 1 for various duty cycles and periods. To evaluate the effects of pulse flows, the reference case for comparison is the one that has all the same conditions except that the coolant flow is continuous without pulsing. The efficiency of the reference case is 46.72%. In general, the measured efficiencies cover a range between 41.1% and 81.8% as shown in Table 3. There are two very clear trends. First, the chilldown efficiency is very much independent of the period for a given duty cycle. Table 3 indicates that for each of the four duty cycles, the six efficiencies at various periods fluctuated around a mean with a small standard deviation that is within the uncertainty range. Second, Consistently for each given period, the chilldown efficiency increases with decreasing DC. The relative independence of the efficiency on the period may be justified based on the simple model discussed in Section 1.5. Since the heat transfer spikes that take place every time when the tube surface is freshly wetted usually last only a very small time period that is shorter than the smallest period of 0.5 s adopted in the current experiment. As a result, the length of the period was not seen as a factor. To estimate the quantitative enhancement on chilldown efficiency due to pulse flows, based on the mean efficiency for each DC, a definition for the percent increase in mean efficiency over the corresponding continuous flow reference case is given in Eq. (6). The results show that the 80% DC only produced the minimum 1.41% increase in chilldown efficiency over the reference case while the maximum 65.77% increase is found for the 20% DC case. For the 60% DC and 40% DC, the enhancements are 21.21% and 38.70%, respectively.
Percentage Increase in Efficiency ¼
gmean gref 100% gref
ð6Þ
ð5Þ
0
_ where mðtÞ is the recorded time-dependent coolant mass flow rate during the chilldown and t End is the time at the end of chilldown. Since, m is the total mass of coolant consumed in the entire chilldown process and hfg is the latent heat of vaporization per unit
3.3. The pulse flow effects on the savings of chilldown cryogen consumption In engineering applications such as those in spacecraft thermal management systems, the savings in cryogen consumptions when
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Table 1 Chilldown efficiencies, gCD (%). DC/Period
0.5 s
1s
2s
3s
4s
5s
MEAN
STD
80% 60% 40% 20%
48.54 53.76 59.24 78.85
54.28 57.92 63.67 78.81
49.41 66.22 66.44 74.35
41.08 55.10 63.37 76.70
48.59 58.46 70.54 81.83
42.39 48.39 65.53 74.17
47.38 56.64 64.80 77.45
4.88 5.92 3.75 2.96
Reference case: Continuous flow without pulsing, gref = 46.72%.
coolant consumed, we would measure the pulsing effects by the ratio, m=m100% . Where m and m100% are the total mass of LN2 coolant consumed for a specific pulse flow case and that of the corresponding continuous flow case, respectively. The total coolant mass consumption is estimated using Eq. (5). Fig. 5 provides the results of m/m100% as function of the duty cycle for various periods. The following observations on the effects of different duty cycles at various periods are made based on the experimental results presented in Fig. 5.
the supply on board is limited are of foremost concerns. Even though, the amount of required cryogen is directly proportional to the chilldown thermal efficiency, we thought that presenting the results of required cryogen mass consumption would make more sense for design engineers whose most important design criterion is cryogen conservation. Next, we examine quantitatively the reduction in cryogen consumption due to enhanced quenching heat transfer by pulse flows. 3.3.1. The general effects of pulsed flow The cryogen coolant consumptions calculated using Eq. (5) for a source pressure of 80 psig and test tube with 4L coating are given in Table 2 for various duty cycles and periods. To evaluate the effects of pulse flows, the same reference case for comparison is again that of the continuous flow without pulsing with 80 psig source pressure and 4L coating test tube. The coolant consumed of the reference case is 0.6057 kg. In general, the measured coolant consumption covers a range between 0.3486 kg and 0.6910 kg as shown in Table 2. As expected, the coolant consumption follows the same trends as those for chilldown efficiency. First, the coolant consumption is very much independent of the period. Table 2 again indicates that for each of the four duty cycles, the six coolant consumptions at various periods fluctuated around a mean with a small standard deviation that is within the uncertainty range. Second, for a given period, the coolant consumption decreases with decreasing DC. To estimate the quantitative effectiveness of coolant consumption reduction due to pulse flows, based on the mean coolant consumption for each DC, a definition for the percent decrease in mean coolant consumption relative to the corresponding continuous flow reference case is given in Eq. (7). The results indicate that the 80% DC only produced the minimum 0.83% decrease in coolant consumption relative to the reference case while the maximum 38.22% decrease was found for the 20% DC case. For the 60% DC and 40% DC, the reductions are 17.14% and 27.60%, respectively.
Percentage Decrease in Coolantright mref mmean Consumption ¼ 100% mref
1. First let us examine the effects of vaious period. As seen in Fig. 5 where we plotted the data for all six periods performed in the experiment. In general, it is very clear that the cryogen consumption is relatively independent of the period that is consistent with the finding in Section 3.3.1. However, it can be seen that there are fluctuations and scatters with the measured data for a given duty cycle that are basicaly due to measurement uncerttainties. The standard deviation of the fluctuations increases with increasing duty cycle that is reasonable as the uncertailty does get larger with the increase of mass flow rates. 2. On the effects of different duty cycles, we found a clear relationship. The dash line in Fig. 5 is a linear fit of the data that clearly demonstrates the linear relationship between m/m100% and the duty cycle. Please note that for 80% duty cycle, the top two points (3 s and 5 s data) were not included in the fitting calculation as their values are over 100% that is physically unrealistic. Another fact that suppotrts the physical validity of this curve fit is that the fitted curve predicts a 99.63% for m=m100% at 100% duty cycle (a continuous flow case). By definition, m=m100% is 100% for 100% duty cycle. The fitted linear curve is represented by Eq. (8) below,
m ð%Þ ¼ 0:48 Duty cycle ð%Þ þ 51:63 m100%
ð8Þ
3.3.3. The effects of tube coating with pulse flows Again, the effects of tube coating are assessed using the tube without coating (bare surface tube) as the reference base case for comparison. As shown in Fig. 6, with a common inlet pressure of 80 psig and a common 1 s period, the coating results are very consistent. On the average, with pulse flows the cryogen mass consumption, mconsumed, of a coated tube is about 57% of that for a bare surface tube (to be exact, 59% for 100% duty cycle, 54% for 60% duty cycle and 57% for 20% duty cycle). Also for each of the two tube types, the cryogen mass consumption increases relatively
ð7Þ
3.3.2. The specific effects of different duty cycles and periods Since the pulse flows are characterized by both the duty cycle and the period (or frequency), we first examine the specific effects of various duty cycles (100%, 80%, 60%, 40%, and 20%) with six different periods, a 80 psig inlet pressure, and a 4L coated tube. Since the main concern in chilldown is the total amount of cryogen or
Table 2 Total nitrogen consumed for chilldown process (kg). DC/Period
0.5 s
1s
2s
3s
4s
5s
MEAN
STD
80% 60% 40% 20%
0.5880 0.5291 0.4917 0.3648
0.5184 0.4819 0.4432 0.3667
0.5671 0.4259 0.4234 0.3915
0.6910 0.5088 0.4428 0.3787
0.5788 0.4783 0.3984 0.3486
0.6609 0.5871 0.4314 0.3948
0.6007 0.5019 0.4385 0.3742
0.0638 0.0544 0.0308 0.0176
Reference case: Continuous flow without pulsing, mref = 0.6057 kg.
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J.N. Chung et al. / International Journal of Heat and Mass Transfer 141 (2019) 256–264 Table 3 Comparison between 30psig and 80psig under 20% DC & 2 s period condition. Flow Condition
Efficiency (%)
Coolant consumed (kg)
30 psig, 20%, 2 s 60 psig, 20%, 3 s 80 psig, 20%, 2 s
90.90 83.84 74.35
0.2833 0.3180 0.3915
bare surface and 4L coated test tubes at various duty cycles and 1 s period with 80 psig inlet pressure. Fig. 7 also provides some perspectives on the chilldown time required for each of the six cases. Similar to the finding for continuous flows, the low thermal conductivity coating was also able to shorten the chilldown time for pulse flows by expediting the approach to Leindenfrost point in the transient quenching process that can lead to a much faster chilldown of the pipe wall to liquid temperatures, which results in significant increases in chilldown efficiency and savings in cryogen consumption. For example, the 4L coating shortened the chilldown time from 17.6 s to 11.5 s for 20% duty cycle condition. Fig. 5. A plot of m/m100% as a function of duty cycle for 80 psig inlet pressure, 4L coated tub and various periods.
3.3.4. The effects of inlet source pressure Finally the effects of different driving source pressures or mass flow rates are also examined. Table 3 lists the results. Three different source pressures of 30 psig, 60 psig and 80 psig were selected with 4L coated test tube and 20% DC. For the 30 psig and 80 psig cases, 2 s period was used, however 3 s period was used for 60 psig. It is apparent that the thermal efficiency increases with decreasing inlet source pressure or decreasing mass flow rate. As expected, the coolant consumption increases with increasing inlet source pressure. 4. Conclusion
Fig. 6. A plot of mconsumed as a function of the duty Cycle for different coating layers with 1 s period.
linearly with the duty cycle that is consistent with the results in Fig. 5. Another factor we addressed is the time to complete the chilldown process. Fig. 7 provides some additional information on the chilldown curves or tube outer surface temperature histories for
The experimental results obtained shows that the chilldown efficiencies with pulse flows cover a range between 41.1% and 81.8%. A general trend was found that indicates the chilldown efficiency increases with decreasing duty cycle and is relatively independent of the pulse flow period. Based on the mean efficiency over the six periods, the 80% DC only produced a 1.41% increase in chilldown efficiency over the reference case of continuous flow without pulsing, while a 65.77% increase is found for the 20% DC case. On the amount of cryogen consumed, a higher chilldown efficiency means less amount of coolant required. Again using the mean cryogen consumption over the six periods, we found that 80% DC only produced a 0.83% decrease in coolant consumption
Fig. 7. Chilldown curves for bare surface and 4L coated test tubes at various duty cycles and 1 s period with 80 psig inlet pressure.
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relative to the reference case, while a 38.22% decrease was obtained for the 20% DC case. Based on a curve fit, we found that m=m100% , where m and m100% are the total mass of LN2 coolant consumed for the pulse flow case and that for the corresponding continuous flow case, respectively, is a linear function of the duty cycle. On the average, with pulse flows, a common inlet pressure of 80 psig, and a common 1 s period, the cryogen mass consumption for a tube coated on the inner surface with 4-layer low-thermal conductivity films is about 57% of that for a bare surface tube (to be exact, 59% for 100% duty cycle, 54% for 60% duty cycle and 57% for 20% duty cycle). Also for each of the two tube types, the cryogen mass consumption increases relatively linearly with the duty cycle. Another factor we addressed is the time to complete the chilldown process. Similar to the finding reported for continuous flows, the low thermal conductivity coating was also able to shorten the chilldown time in film boiling with pulse flows by expediting the approach to the Leindenfrost point in the transient quenching process that can lead to a much faster chilldown of the pipe wall to liquid temperatures, which results in significant increases in chilldown efficiency and savings in cryogen consumption. For example, the 4L coating shortened the chilldown time from 17.6 s to 11.5 s for 20% duty cycle condition. For the effects of different inlet pressures, it is apparent that the chilldown thermal efficiency increases with decreasing inlet source pressure or decreasing mass flow rate. As expected, the coolant consumption increases with increasing inlet source pressure. Acknowledgments This work was partially supported by the Flight Opportunities Program at National Aeronautics and Space Administration (NASA) under the award number NNX16AP68G. This research was also partially supported by the Andrew H. Hines, Jr./Progress Energy Endowment Fund at the University of Florida. Declarations of Competing Interest None. Appendix A. Supplementary material Supplementary data to this article can be found online at https://doi.org/10.1016/j.ijheatmasstransfer.2019.06.080. References [1] V.P. Carey, Liquid-Vapor Phase-Change Phenomena, second ed., Taylor & Francis Group, LLC., New York, 2008. [2] R. Shaeffer, H. Hu, J.N. Chung, An experimental study on liquid nitrogen pipe chilldown and heat transfer with pulse flows, Int. J. Heat Mass Transf. 67 (2013) 955–966. [3] S.K. Sahu, P.K. Das, S. Bhattacharyya, An experimental investigation on the quenching of a hot vertical heater by water injection at high flow rate, Nucl. Eng. Des. 240 (2010) 1558.
[4] C. Agrawal, R. Kumar, A. Gupta, B. Chatterjee, Rewetting and maximum surface heat flux during quenching of hot surfac by round water jet impingement, Int. J. Heat Mass Transf. 55 (2012) 4772–4782. [5] S.A. Nada, M. Shoukri, A.F. El-Dib, Rewetting of hot vertical tubes by a falling liquid film with different directions of venting the geneated steam, Int. J. Thermal Sci. 85 (2014) 62–72. [6] K. Yuan, Y. Ji, J.N. Chung, Cryogenic chilldown process under low flow rates, Int. J. Heat Mass Transf. 50 (2007) 4011–4022. [7] K. Yuan, Y. Ji, J.N. Chung, Numerical modeling of cryogenic quenching process in terrestrial gravity and microgravity, Int. J. Heat Fluid Flow 30 (2009) 44–53. [8] H. Hu, J.N. Chung, An experimental study on flow patterns and heat transfer characteristics during cryogenic chilldown in a vertical pipe, Cryogenics 52 (2012) 268–277. [9] S.R. Darr, Hu. Hong, N. Glikin, J.W. Hartwig, A.K. Majumdar, A.C. Leclair, J.N. Chung, An experimental study on terrestrial cryogenic transfer line chilldown I. Effect of mass flux, equilibrium quality, and inlet subcooling, Int. J. Heat Mass Transf. 103 (2016) 1225–1242. [10] S.R. Darr, Hong Hu, N. Glikin, J.W. Hartwig, A.K. Majumdar, A.C. Leclair, J.N. Chung, An experimental study on terrestrial cryogenic transfer line chilldown, II. Effect of flow direction with respect to gravity and new correlation set, Int. J. Heat Mass Transf. 103 (2016) 1243–1260. [11] S.R. Darr, Jun Dong, N. Glikin, J.W. Hartwig, A.K. Majumdar, A.C. Leclair, J.N. Chung, The effect of reduced gravity on cryogenic nitrogen boiling and pipe chilldown, Nat. Partner J. (NPJ): Microgr. 2 (2016) 16033. [12] Y. Ishino, Suzuki, T. Abe, N. Ohiwa, S. Yamaguchi, Flow and heat transfer characteristics in pulsating pipe flows (effects of pulsation on internal heat transfer in a circular pipe flow), Heat Transf.-Japanese Res. 25 (1996) 323–336. [13] E.B. Denison, W.H. Stevenson, R.W. Fox, Pulsating laminar flow measurements with a directionally sensitive laser velocimeter, AIChE J. 17 (1971) 781–787, https://doi.org/10.1002/aic.690170405. [14] S.K. Gupta, T.R.D. Patel, R.C. Ackerberg, Wall heat/mass transfer in pulsatile flow, Chem. Eng. Sci. 37 (1982) 1727–1739, https://doi.org/10.1016/00092509(82)80045-6. [15] A.A. Al-Haddad, N. Al-Binally, Prediction of heat transfer coefficient in pulsating flow, Int. J. Heat. Fluid Flow 10 (1989) 131–133, https://doi.org/ 10.1016/0142-727X(89)90006-4. [16] M.A. Habib, A.M. Attya, A.I. Eid, A.Z. Aly, Convective heat transfer characteristics of laminar pulsating pipe air flow, Heat Mass Transf. 38 (2002) 221–232, https://doi.org/10.1007/s002310100206. [17] Y. Kita, T. Hayashi, K. Hirose, Heat transfer in pulsating laminar flow in a pipe: a constant wall temperature, Bull. JSME 25 (1982) 217–224, https://doi.org/ 10.1299/jsme1958.25.217. [18] J.C. Yu, Z.-X. Li, T.S. Zhao, An analytical study of pulsating laminar heat convection in a circular tube with constant heat flux, Int. J. Heat. Mass Transf. 47 (2004) 5297–5301, https://doi.org/10.1016/j. ijheatmasstransfer.2004.06.029. [19] D.A. Nield, A.V. Kuznetsov, Forced convection with laminar pulsating flow in a channel or tube, Int. J. Therm. Sci. 46 (2007) 551–560, https://doi.org/10.1016/ j.ijthermalsci.2006.07.011. [20] B.H. Yan, L. Yu, Y.H. Yang, Heat transfer with laminar pulsating flow in a channel or tube in rolling motion, Int. J. Therm. Sci. 49 (2010) 1003–1009, https://doi.org/10.1016/j.ijthermalsci.2010.01.011. [21] H. Chattopadhyay, F. Durst, S. Ray, Analysis of heat transfer in simultaneously developing pulsating laminar flow in a pipe with constant wall temperature, Int. Commun. Heat. Mass Transf. 33 (2006) 475–481, https://doi.org/10.1016/j. icheatmasstransfer.2005.12.008. [22] T.K. Nandi, H. Chattopadhyay, Numerical investigations of simultaneously developing flow in wavy microchannels under pulsating inlet flow condition, Int. Commun. Heat. Mass Transf. 47 (2013) 27–31, https://doi.org/10.1016/j. icheatmasstransfer.2013.06.008. [23] T.K. Nandi, H. Chattopadhyay, Numerical investigations of developing flow and heat transfer in raccoon type microchannels under inlet pulsation, Int. Commun. Heat. Mass Transf. 56 (2014) 37–41, https://doi.org/10.1016/j. icheatmasstransfer.2014.04.017. [24] J.W. Hartwig, J. Vera, Numerical modeling of transient chilldown of a cryogenic propellant transfer line, J. Thermophys. Heat Transfer 30 (2016) 403–408. [25] T.L. Bergman, A.S. Lavin, F.P. Incropera, D.P. Dewitt, Fundamentals of Heat and Mass Transfer, seventh ed., Wiley, New York, 2011. [26] R. Mathie, Unsteady and Conjugate Heat Transfer in Convective-Conductive Systems, Doctoral Thesis, Department of Chemical Engineering, Imperial College London, 2012. [27] DuPontTM Teflon FEP, Fluoroplastic Film (2013), www.teflon.com/Industrial.
Contents lists available at ScienceDirect
International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
Enhancement of convective quenching heat transfer by coated tubes and intermittent cryogenic pulse flows J.N. Chung a,⇑, Jun Dong a, Hao Wang a, S.R. Darr a, J.W. Hartwig b a b
Cryogenics Heat Transfer Laboratory, Department of Mechanical and Aerospace Engineering, University of Florida, Gainesville, FL 32611-6300, USA NASA Glenn Research Center, Cleveland, OH 44135, USA
a r t i c l e
i n f o
Article history: Received 9 February 2019 Received in revised form 30 May 2019 Accepted 23 June 2019 Available online 27 June 2019 Keywords: Cryogenic fluid Quenching heat transfer Pulse flow Two-phase flow Film boiling
a b s t r a c t This paper reports heat transfer enhancement techniques for the cryogenic quenching process. An experiment was performed to evaluate the enhancement of quenching heat transfer by two techniques: 1. Using intermittent pulse flows and 2. Coating the inner surface of a test tube with layers of low thermal conductivity films. Pulsed liquid nitrogen flows with various duty cycles and periods were applied in the quenching of a room-temperature metal tube coated with four layers of Teflon thin films on the inner surface. In general, the results obtained indicate that the quenching thermal efficiency that measures the effectiveness of heat transfer enhancement increases with decreasing duty cycle, however it is relatively independent of the period. Comparing with non-coated bare surface test tube, the low-conductivity coating substantially improved the thermal efficiency and reduced the total quenching time. Additionally, the thermal efficiency was found to increase with decreasing source inlet pressure or coolant mass flow rate. The savings on the amount of cryogen consumed follow the same trends as those for the thermal efficiency with pulse flows and tube coating. Ó 2019 Elsevier Ltd. All rights reserved.
1. Introduction 1.1. Importance of quenching heat transfer In many convective liquid-vapor phase change heat transfer applications, cryogenic fluids are widely used in industrial processes, spacecraft and cryosurgery systems, and so on. For example, cryogens are usually used as liquid fuels such as liquid hydrogen and oxygen in the rocket industry, liquid nitrogen (LN2) and helium are frequently used to cool superconducting magnetic device for medical applications. In these systems, proper transport, handling, and storage of cryogenic fluids are of great importance. When a cryogenic system is first started up, its walls and hardware must go through a transient chilldown period prior to reaching a steady operation. Therefore, chilldown (quenching) is the process of keeping the system adjusted to the low temperature scale which is usually several hundred degrees below the room temperature. The chilldown or quenching process is complicated, involving unsteady two-phase heat and mass transfer, and has not been fully understood.
⇑ Corresponding author. E-mail address: [email protected] (J.N. Chung). https://doi.org/10.1016/j.ijheatmasstransfer.2019.06.080 0017-9310/Ó 2019 Elsevier Ltd. All rights reserved.
A new breakthrough advancement in the quenching phasechange heat transfer processes by a low-thermal conductivity surface coating can lead to much higher quenching efficiencies, substantial energy and liquid cryogen consumption savings, and global reduction in greenhouse gas emissions. Especially for space propulsion systems, higher quenching efficiencies translate to more fuels available for making longer and deeper space missions a reality. This paper reports an advancement on quenching heat transfer. 1.2. Characteristics of quenching heat transfer A quenching (chilldown) process is a liquid-to-vapor phase change phenomenon that is governed by the ‘‘boiling curve”. This curve [1] as shown in Fig. 1 illustrates the heat transfer surface heat flux, q00 , plotted against the surface degree of superheating, T W T sat , where T W is the surface temperature and T sat is the saturation temperature corresponding to the boiling fluid bulk pressure. In boiling, if the heating source is externally supplied to the heater surface such as the embedded electrical resistance heating element, the process is heat-flux controlled and it follows the route of A ? B ? D. In contrast, during quenching the warm wall where the heat comes out does not have a source of heat supply, therefore, the heat transferring out of the wall can only come internally from the thermal capacity (stored energy) of the wall. The only
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Fig. 1. A typical boiling curve including different boiling regimes and corresponding flow patterns.
way to remove heat from the wall is by lowering the inner wall surface temperature using a cooling flow. So quenching is a wall surface temperature-controlled process. Thus, a quenching process follows the route D ? C ? B ? A. During quenching, film boiling is always the first mode of heat transfer encountered due to a relatively very hot surface. Owing to its relatively low heat fluxes at high wall temperatures, film boiling usually dominates the quenching time and cannot be avoided. As a result, in traditional quenching processes, the thermal energy efficiency is extremely low. According to Shaeffer et al. [2], the average quenching efficiency that is defined as the ratio of the amount of thermal energy removed from the wall versus the required cooling capability of the cryogen spent in a quenching process is about 8%, highlighting the tremendous need to improve the quenching efficiency for many applications that require cryogens as the working fluid. 1.3. Convective quenching heat transfer in tubes and pipes Compared to the heated wall convective boiling case, the amount of literature on convective quenching heat transfer is far less. For convective quenching in tubes and pipes, most studies used water and normal room temperature fluids [3–5]. Yuan and Chung [6] and Yuan et al. [7] reported the pioneering work on liquid nitrogen (LN2) cryogenic chilldown in conventional stainless tubes under low flow rates in both terrestrial and simulated drop-tower microgravity conditions. Both papers [6,7] established the fundamental physical understanding and benchmark data for future convective cryogenic chilldown heat transfer research. As follow-ups, LN2 chilldown experiments of a short stainless steel tube have been carried out at low mass flux both at constant mass flux [8] and in a pulsed on/off flow [2]. Only recently has there been a thorough experimental investigation of the entire chilldown process in 1-g that reports the measured heat transfer coefficients for all boiling regimes over a large range of mass flow and thermodynamic conditions [9,10]. Another recent work [11] by the same group provided a wide ranging dataset of reduced gravity cryogenic pipe chilldown data from a parabolic flight experiment using LN2 onboard a C9 aircraft. Heat transfer data were obtained from
flowing LN2 through a short, stainless steel (SS) pipe in microgravity for developing heat transfer correlations. This experiment was an extension of 1-g experiments reported in Darr et al. [9,10]. 1.4. Effects of flow pulsation on heat transfer in tubes and pipes For single-phase liquid or gas convective heat transfer, the pulsation in the forms of a wave with a fixed amplitude and frequency such as a sinusoidal wave, results by Ishino et al. [12] pointed out that the heat transfer (Nusselt number) decreases initially as the amplitude increases, but then recovers gradually, and finally becomes much greater than the initial values. In general, most experimental results [13–16] showed that pulsation could either decrease or increase the heat transfer. Some analytical investigations [17–20] reported that for pulsating flows, the Nusselt numbers fluctuate periodically around value of steady flow for constant wall temperatures and constant wall heat fluxes in fully developed region. Some results [21–23] predicted by numerical simulations of pulsed flow heat transfer showed that the pulsing does not affect the time-averaged heat transfer in a pipe. The results also indicated the Nusselt number distribution varies with time in the near-entry region. In summary, for single-phase pulsating flows, usually the results among many investigators showed contradictory trends due to different definitions of the Nusselt number. For heat transfer with phase-change, pulse flow has not received much attention. We only found two reports [2,24] that are all dealing with cryogenic quenching heat transfer and are discussed below. Shaeffer et al. [2] performed a comparison between continuous flow and intermittent on/off pulse flow patterns on quenching heat transfer in a stainless steel bare surface tube without any coating. Liquid nitrogen was used as the working fluid and upward flows with Reynolds numbers ranging from 2500 to 7000 in a round tube were examined. To gain a better understanding of the effect of flow pulsation on the chilldown process, three different pulse flows (4-s, 10-s and 20-s periods and all with 50% duty cycle) were used to compare with the corresponding continuous flows. It is noted that
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the duty cycle, DC, which is the ratio of valve open time (t open ) to its period (topen þ t closed ), where t closed is the valve closed time, therefore, DC ¼ t open =ðtopen þ t closed Þ. For the chilldown of the entire process from film boiling to nucleate boiling regimes or just only the film boiling regime, the continuous flow pattern is favored for relatively low Reynolds flows. Specifically, the continuous flow prevailed over pulse flows if the Reynolds number was less than 4000. At higher Reynolds number flows, pulse flow patterns are more advantageous and shorter periods are more ideal than longer ones. Hartwig and Vera [24] investigated liquid hydrogen quenching of a tube using both continuous and intermittent on/off pulse flows. They used 15-s period with a 67% duty cycle and 20-s period with a 50% duty cycle for the pulse flows. Only limited data were reported and there was no specific comparison. For the same flow rates and identical liquid hydrogen initial temperatures, the pulse flow required more time to chilldown the tube to liquid temperature. 1.5. Theoretical basis of current experiment 1.5.1. Flow pulsing A mathematical model was developed based on the theory of transient one-dimensional heat conduction at the interface between two semi-infinite solids [25] to help explain the higher heat transfer enhancement by a pulse flow. A complete solution for this unsteady and conjugate heat transfer in a convectiveconductive system can be found in Mathie [26]. Since the quenching of the hot tube wall by cold cryogenic coolant flow is a transient and unsteady heat transfer process that takes place at the interface between two materials, therefore by neglecting the curvature effects, the metal tube wall can be approximated as the semi-infinite slab A and the coolant fluid is assumed as the semi-infinite slab B. If the two free surfaces of the slabs were to be placed in contact at time, t = 0, the heat flux, q00A!B , at the interface that flows from A to B can be given as follows [25].
q00A!B ¼
kA ðT A;i T s Þ ðpaA tÞ
1=2
¼
kB ðT B;i T s Þ ðpaB tÞ1=2
ð1Þ
where kA and kB are thermal conductivities of A and B, respectively. T A;i and T B;i are initial temperatures of A and B before contact, respectively. aA and aB are thermal diffusivities of A and B, respectively. T s and t are the common interface temperature and elapsed time, respectively. Eq. (1) above is introduced simply to provide a physical explanation for the concept that pulsing the flow enhances the heat transfer. Eq. (1) basically indicates that when we place two materials A and B, each with a different temperature in contact, the transient heat fluxes q00A!B across the interface from A to B would be proportional to t 1=2 (t is the elapsed time after the initial contact of the two materials, so t = 0 is the time when the two are placed in contact) as the rest of the parameters (thermal properties) in the equation are all constants. The physical meaning is that initially the heat transfer would be very high and the heat flux decreases quickly as time passes. In other words, Eq. (1) is the transient solution that corresponds to a step change (jump) in the surface temperature. For the pulse flow process, the tube inner surface gets a jump in temperature and a corresponding spike in the heat flux every time when intermittent pulse flow is turned on after the off period in a duty cycle that results in higher heat transfer than the continuous flow case. Please note that the above simplified 1-D semi-infinite slab model is only used to explain the transient intermittent heat transfer nature due to pulsing and is not intended for the modeling of the entire pulsed flow quenching process. In addition to the pulse
flows, we also used the test tubes coated with low thermal conductivity thin films on the inner wall surface to shorten the time for the heater surface to reach the Leidenfrost temperature as a second enhancement technique that is explained next. 1.5.2. Low-thermal conductivity coating The enhancement concept of coated tube is based on that there are two heat transfer mechanisms involved that are opposite in nature to each other. The first mechanism is the thermal insulating effect due to the low-thermal conductivity thin layer that facilities a fast drop of the tube inner surface temperature by restricting the heat flow from the bulk of the tube wall to the tube inner surface. The lower surface temperature allows the quenching process to quickly move from film boiling regime to the Leidenfrost point to initiate the higher heat transfer transition and nucleate boiling regimes. The second mechanism is the conduction of heat from the bulk of the tube wall to the cooling fluid through the inner surface during transition and nucleate boiling regimes that requires the thermal conductivity of the bulk wall material to be as high as possible such as metals to expedite the wall cooling process. There should exist an optimal thickness of the low-thermal conductivity layer that balances these two competing mechanisms such that the coating is just thick enough to quickly lower the tube surface temperature to that of the Leidenfrost point while it is still relatively very thin to not substantially reduce the heat flow from the bulk to the cooling fluid. 1.6. Research objective The primary objective of the current set of experiments is to evaluate the effects of intermittent ON/OFF pulse flow on transient temperature history curves of metal tubes, or chilldown curves, from room temperature to LN2 saturation temperatures during a quenching process. The test tubes are horizontally-aligned commercially available 304 stainless steel tubes without coating and with coating using low thermal conductivity thin Teflon layers. Therefore, the secondary objective is to investigate the effects of coating with pulse flows. Tests were carried out with a set of pulse flow conditions with 0.5–5-s periods and 80%, 60%, 40%, and 20% duty cycles over a wide range of test section inlet pressure levels and corresponding mass flow rates. The effectiveness of the pulse flow was evaluated by a comparison of chilldown curves with pulsations to those with constant and continuous flows. 2. Experimental methods 2.1. Experimental apparatus and procedure The current LN2 experimental system was built with small modifications to our original rig used by Darr et al. [9,10] and Darr et al. [11]. A flow and component schematic of the experiment is shown in Fig. 2. To avoid duplication, the readers are referred to Darr et al. [11] for a detailed description of the flow network, and the function and operation of each component. Specifically, it is important to note that a subcooler was placed before the entrance of the test section. The function of the subcooler was to insulate the tubing upstream of the test section so that to make sure that the fluid entering the test section was vapor free, subcooled liquid that provides a precise inlet boundary condition for the test tube during chilldown. A key modification in the current experiment is the installation of a solenoid valve (SV) that facilitates the intermittent ON/OFF pulse flow. The solenoid valve was placed after the relieve valve RV5 and before the subcooler. As shown in Fig. 3, the test section was a 57.2 cm long, 1.270 cm OD, 1.168 cm ID 304SS tube. A length of 2.54 cm of the test section
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Fig. 2. Fluid system schematic. The valves and important components of the fluid network. Relief valve settings, the burst disk setting, and pressure regulator settings are also included. BD, burst disk; BV, ball valve; CV, check valve; DAQ, data acquisition unit; FM, flow meter; GN2, gaseous nitrogen; GV, globe valve; LN2, liquid nitrogen;; NV, needle valve; PC, pre-cooler; PG, pressure gauge; PR, pressure regulator; PT, pressure transducer; RV, relief valve; SV, solenoid valve; TC, thermocouple; Vap, vaporizer; 3 V, threeway valve.
Fig. 3. Test section dimensions and details.
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tube protruded out of each side of the vacuum chamber. Six TCs were soldered to the outside of the tube, with three placed an axial distance of 14.9 cm from the test section inlet (upstream TC station), and three placed 40.1 cm from the inlet (downstream TC station). As detailed in Fig. 3, for each station, the TCs were spaced out radially in 90° increments such that each station had a top, side, and bottom TC. Two cryogenic rated PTs were connected to two short pieces of 304SS tube protruding perpendicularly from the test section, one near each TC station. These tubes were welded to the test section. As shown in Fig. 3, the test section was a 57.2 cm long, 1.270 cm OD, 1.168 cm ID 304SS tube. Six TCs were soldered to the outside of the tube, with three placed an axial distance of 14.9 cm from the test section inlet (upstream TC station), and three placed 40.1 cm from the inlet (downstream TC station). As detailed in Fig. 3, at each station, the TCs were spaced out radially in 90° increments such that each station had a top, side, and bottom TC. Two cryogenic rated PTs were connected to two short pieces of 304SS tube protruding perpendicularly from the test section, one near each TC station. These tubes were welded to the test section. The 316SS vacuum chamber that housed the test section was the same as the original experiment described in Darr et al. [11]. The purpose of the vacuum chamber was to reduce parasitic heat leak that reduced the uncertainty in the calculation of chilldown heat flux. A diaphragm pump and molecular turbopump were used to bring the air pressure inside the chamber down to approximately 1 Pa. This reduced the parasitic heat leak due to conduction between the test section and the air inside the vacuum chamber to less than 10% of the lowest measured convective heat flux, which occurred during film boiling at the lowest flow rate that was tested. Parasitic heat leak was less than 1% of the measured value for most of the data points. Except for setting the duty cycle and period for the pulse flow case, the rest of the experimental procedure is identical to that followed in our previous chilldown experiment. So, the readers are again referred to Darr et al. [11] for the details of the experimental procedure. 2.2. Coating of tube inner surface In order to evaluate the second enhancing technique in the current experiment, two different tubes were used. One is a bare surface stainless steel test tube with no coating. The other is a stainless steel test tube that was coated with low-thermal conductivity thin Teflon layers on the inner surface. Specifically, the coating material was made of Fluorinated Ethylene Propylene (FEP) by DuPont and classified by DuPont as Teflon 959G-203 that is a black color paint and has a thermal conductivity of 0.195 W/m K (DuPont publication [27]). The coating was put on by the Matrix Coating Corporation (West Palm Beach, FL 33404, USA) using the pour and drain process. The final thickness of the coating material on the tube depends on the number of layers processed. After each pour and drain, the fresh film layer was cured in a furnace through a standard sintering procedure before adding another layer by the same pour and drain procedure. As a result, the final thickness of coated layer depends on the total number of layers processed. The coated test tube used in the experiment is a four-layer (4L) one. The 4L coating went through the pour and drain process four separate times. To determine the coating layer thickness, high resolution computer tomography (CT) X-ray scans of the tube cross sections were obtained using a Phoenix v|tome|x M (GE’s Measurement & Control business, Boston, USA) system in the Nano Research Facility at the University of Florida. Scanning was carried out using a 240 kV X-ray tube and a tungsten on beryllium target, with the following settings: 200 kV, 50 milliamps, and 0.5 mm Tin filter. Images were collected from 1600 pixels horizontal, 2024
pixels vertical, 0.5 s detector exposure, averaging of 4 images per rotation position with a one-exposure skip and a total of 2200 rotational positions. The coating thickness for this 4L tube was measured at 64.8 ± 0.7 mm. 2.3. Experimental uncertainties As mentioned above, the current experimental system is virtually identical to that employed in Darr et al. [9,10], the readers are referred to Darr et al. [9,10] for uncertainties on those independent, directly measured quantities, such as temperatures and pressures.
3. Results and discussion As mentioned in Section 2.2, two different test tubes were used in the current experiment. One is a tube with 4-layer (4L) coating and the other is a bare surface tube used as a reference case for evaluating the coating effects with pulse flows. Three inlet pressures of 5.5 105 Pa (Gauge) (80 psig), 4.1 105 Pa (Gauge) (60 psig), and 2.1 Pa (Gauge) (30 psig) were applied as the flow driving force. For the pulse flow settings, it is reminded that the duty cycle, DC, is defined as the valve open time (t open ) to its period (topen þ t closed ), where tclosed is the valve closed time, therefore, DC ¼ t open =ðt open þ tclosed Þ. Six different periods of 0.5 s, 1 s, 2 s, 3 s, 4 s and 5 s were used with the duty cycles set at 20%, 40%, 60% and 80%, respectively. We have found that duty cycles with settings less than 20% would induce flow instabilities that prevented steady conditions for experimentation. 3.1. Typical chilldown curves Before we present and discuss the effects of pulsed flows on the quenching heat transfer of a metal tube, the physics and characteristics of the tube quenching process must be clearly explained first. Fig. 4 shows a typical chilldown (quenching) curve recorded in the current experiment for the case where the inlet pressure, duty cycle and period were set at 80 psig, 20% and 1 s, respectively. The inner surface of the test tube was bare with no coating. The six chilldown curves provide the tube wall outer surface temperature histories during chilldown at both upstream and downstream locations from the tube inlet. As shown in Fig. 4, six chilldown curves were plotted using the temperature data registered by the six thermocouples (TCs) shown in Fig. 3(b), where TCs 1(top), 2(side) and 3(bottom) are located in the upstream station and TCs 4(top), 5(side) and 6(bottom) are located in the downstream station. In general, all six chilldown curves display similar trends where the heat transfer process during cooling of the tube went through film, transition and nucleate boiling regimes, sequentially as explained by the boiling curve (Fig. 1). The effects of the pulse flow are clearly reflected by the zigzag pattern of the chilldown curves. At a specific TC station, the chilldown time is the shortest for the bottom location, second shortest for the side, and longest for the top. For example, at the upstream TC station, the chilldown time is the shortest for the bottom location (TC3), second shortest for the side (TC2), and longest for the top (TC1). However, for the same circumferential location at a TC station, the downstream one lagged behind that in the upstream station due to the fact it takes more time for the cooling fluid to reach the downstream location. However, it should be pointed out that the downstream ones actually started to lower their temperatures at the same time as the upstream ones but at slower rates due to cooling by the axial conduction heat transfer in the metal tube wall.
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Fig. 4. A typical chilldown curve.
3.2. The effects of pulse flow on the enhancement of chilldown thermal efficiency The performance of a cryogenic chilldown system solely depends on the effectiveness of the heat transfer from the tube wall to the cryogenic coolant. Therefore, to measure the performance of a chilldown system, the chilldown thermal efficiency, gCD , as defined below is introduced. The chilldown thermal efficiency represents the percent of total required quenching capacity of the cryogenic coolant heat sink that is actually utilized in removing the thermal energy from the tube wall and part of a globe valve from their initial room temperature to that of the liquid nitrogen. The chilldown thermal efficiency in %, gCD , is therefore defined as,
Q removed 100% Q required
gCD ¼
ð2Þ
In the above, Q removed is the total thermal energy removed from the tube wall and part of a globe valve by the cooling fluid during the entire chilldown period and is defined as,
Q removed ¼ ðM tube þ 0:3M v alv e Þcp; SS T initial T final
ð3Þ
where Mtube and Mv alv e are the mass of the tube section and that of the three-way globe bypass valve (shown in Fig. 3), respectively. A factor of 0.3 is applied to the valve mass, as it is estimated that only 30% of the valve mass needs to be chilldown when the valve is opened and the chilldown process starts, since part of the valve is in contact with the LN2 before the opening of the valve. Therefore, the valve is partially (about 70%) chilled down during precooling and reheating of the test section where the LN2 fills the path all the way to the valve. cP; SS is the stainless steel specific heat for both the tube and valve materials. T Initial and T Final are the initial room temperature and the final temperature when the chilldown is completed, respectively. The end of chilldown temperature, T Final , is the LN2 saturation temperature corresponding to the local pressure. Q required is the total quenching capacity required during the chilldown process. It is defined as,
Q required ¼ mhfg
ð4Þ
where m is the total mass of coolant required or consumed during chilldown and it is defined in Eq. (5) below.
Z m¼
tEnd
_ mðtÞdt
mass, that means Q required is the total quenching capacity of the cryogen of mass, m, required during the chilldown process. Before presenting the chilldown efficiency results, the relative uncertainties for mass flow rate and the chilldown efficiency, respectively, must be provided first. The relative uncertainties for the mass flow rates, were estimated in the range of 3% to 7%. Next, the relative uncertainties for the chilldown efficiencies are estimated to be of the order of 15%. The chilldown thermal efficiencies for source pressure of 80 psig and test tube with 4L coating are given in Table 1 for various duty cycles and periods. To evaluate the effects of pulse flows, the reference case for comparison is the one that has all the same conditions except that the coolant flow is continuous without pulsing. The efficiency of the reference case is 46.72%. In general, the measured efficiencies cover a range between 41.1% and 81.8% as shown in Table 3. There are two very clear trends. First, the chilldown efficiency is very much independent of the period for a given duty cycle. Table 3 indicates that for each of the four duty cycles, the six efficiencies at various periods fluctuated around a mean with a small standard deviation that is within the uncertainty range. Second, Consistently for each given period, the chilldown efficiency increases with decreasing DC. The relative independence of the efficiency on the period may be justified based on the simple model discussed in Section 1.5. Since the heat transfer spikes that take place every time when the tube surface is freshly wetted usually last only a very small time period that is shorter than the smallest period of 0.5 s adopted in the current experiment. As a result, the length of the period was not seen as a factor. To estimate the quantitative enhancement on chilldown efficiency due to pulse flows, based on the mean efficiency for each DC, a definition for the percent increase in mean efficiency over the corresponding continuous flow reference case is given in Eq. (6). The results show that the 80% DC only produced the minimum 1.41% increase in chilldown efficiency over the reference case while the maximum 65.77% increase is found for the 20% DC case. For the 60% DC and 40% DC, the enhancements are 21.21% and 38.70%, respectively.
Percentage Increase in Efficiency ¼
gmean gref 100% gref
ð6Þ
ð5Þ
0
_ where mðtÞ is the recorded time-dependent coolant mass flow rate during the chilldown and t End is the time at the end of chilldown. Since, m is the total mass of coolant consumed in the entire chilldown process and hfg is the latent heat of vaporization per unit
3.3. The pulse flow effects on the savings of chilldown cryogen consumption In engineering applications such as those in spacecraft thermal management systems, the savings in cryogen consumptions when
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Table 1 Chilldown efficiencies, gCD (%). DC/Period
0.5 s
1s
2s
3s
4s
5s
MEAN
STD
80% 60% 40% 20%
48.54 53.76 59.24 78.85
54.28 57.92 63.67 78.81
49.41 66.22 66.44 74.35
41.08 55.10 63.37 76.70
48.59 58.46 70.54 81.83
42.39 48.39 65.53 74.17
47.38 56.64 64.80 77.45
4.88 5.92 3.75 2.96
Reference case: Continuous flow without pulsing, gref = 46.72%.
coolant consumed, we would measure the pulsing effects by the ratio, m=m100% . Where m and m100% are the total mass of LN2 coolant consumed for a specific pulse flow case and that of the corresponding continuous flow case, respectively. The total coolant mass consumption is estimated using Eq. (5). Fig. 5 provides the results of m/m100% as function of the duty cycle for various periods. The following observations on the effects of different duty cycles at various periods are made based on the experimental results presented in Fig. 5.
the supply on board is limited are of foremost concerns. Even though, the amount of required cryogen is directly proportional to the chilldown thermal efficiency, we thought that presenting the results of required cryogen mass consumption would make more sense for design engineers whose most important design criterion is cryogen conservation. Next, we examine quantitatively the reduction in cryogen consumption due to enhanced quenching heat transfer by pulse flows. 3.3.1. The general effects of pulsed flow The cryogen coolant consumptions calculated using Eq. (5) for a source pressure of 80 psig and test tube with 4L coating are given in Table 2 for various duty cycles and periods. To evaluate the effects of pulse flows, the same reference case for comparison is again that of the continuous flow without pulsing with 80 psig source pressure and 4L coating test tube. The coolant consumed of the reference case is 0.6057 kg. In general, the measured coolant consumption covers a range between 0.3486 kg and 0.6910 kg as shown in Table 2. As expected, the coolant consumption follows the same trends as those for chilldown efficiency. First, the coolant consumption is very much independent of the period. Table 2 again indicates that for each of the four duty cycles, the six coolant consumptions at various periods fluctuated around a mean with a small standard deviation that is within the uncertainty range. Second, for a given period, the coolant consumption decreases with decreasing DC. To estimate the quantitative effectiveness of coolant consumption reduction due to pulse flows, based on the mean coolant consumption for each DC, a definition for the percent decrease in mean coolant consumption relative to the corresponding continuous flow reference case is given in Eq. (7). The results indicate that the 80% DC only produced the minimum 0.83% decrease in coolant consumption relative to the reference case while the maximum 38.22% decrease was found for the 20% DC case. For the 60% DC and 40% DC, the reductions are 17.14% and 27.60%, respectively.
Percentage Decrease in Coolantright mref mmean Consumption ¼ 100% mref
1. First let us examine the effects of vaious period. As seen in Fig. 5 where we plotted the data for all six periods performed in the experiment. In general, it is very clear that the cryogen consumption is relatively independent of the period that is consistent with the finding in Section 3.3.1. However, it can be seen that there are fluctuations and scatters with the measured data for a given duty cycle that are basicaly due to measurement uncerttainties. The standard deviation of the fluctuations increases with increasing duty cycle that is reasonable as the uncertailty does get larger with the increase of mass flow rates. 2. On the effects of different duty cycles, we found a clear relationship. The dash line in Fig. 5 is a linear fit of the data that clearly demonstrates the linear relationship between m/m100% and the duty cycle. Please note that for 80% duty cycle, the top two points (3 s and 5 s data) were not included in the fitting calculation as their values are over 100% that is physically unrealistic. Another fact that suppotrts the physical validity of this curve fit is that the fitted curve predicts a 99.63% for m=m100% at 100% duty cycle (a continuous flow case). By definition, m=m100% is 100% for 100% duty cycle. The fitted linear curve is represented by Eq. (8) below,
m ð%Þ ¼ 0:48 Duty cycle ð%Þ þ 51:63 m100%
ð8Þ
3.3.3. The effects of tube coating with pulse flows Again, the effects of tube coating are assessed using the tube without coating (bare surface tube) as the reference base case for comparison. As shown in Fig. 6, with a common inlet pressure of 80 psig and a common 1 s period, the coating results are very consistent. On the average, with pulse flows the cryogen mass consumption, mconsumed, of a coated tube is about 57% of that for a bare surface tube (to be exact, 59% for 100% duty cycle, 54% for 60% duty cycle and 57% for 20% duty cycle). Also for each of the two tube types, the cryogen mass consumption increases relatively
ð7Þ
3.3.2. The specific effects of different duty cycles and periods Since the pulse flows are characterized by both the duty cycle and the period (or frequency), we first examine the specific effects of various duty cycles (100%, 80%, 60%, 40%, and 20%) with six different periods, a 80 psig inlet pressure, and a 4L coated tube. Since the main concern in chilldown is the total amount of cryogen or
Table 2 Total nitrogen consumed for chilldown process (kg). DC/Period
0.5 s
1s
2s
3s
4s
5s
MEAN
STD
80% 60% 40% 20%
0.5880 0.5291 0.4917 0.3648
0.5184 0.4819 0.4432 0.3667
0.5671 0.4259 0.4234 0.3915
0.6910 0.5088 0.4428 0.3787
0.5788 0.4783 0.3984 0.3486
0.6609 0.5871 0.4314 0.3948
0.6007 0.5019 0.4385 0.3742
0.0638 0.0544 0.0308 0.0176
Reference case: Continuous flow without pulsing, mref = 0.6057 kg.
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Efficiency (%)
Coolant consumed (kg)
30 psig, 20%, 2 s 60 psig, 20%, 3 s 80 psig, 20%, 2 s
90.90 83.84 74.35
0.2833 0.3180 0.3915
bare surface and 4L coated test tubes at various duty cycles and 1 s period with 80 psig inlet pressure. Fig. 7 also provides some perspectives on the chilldown time required for each of the six cases. Similar to the finding for continuous flows, the low thermal conductivity coating was also able to shorten the chilldown time for pulse flows by expediting the approach to Leindenfrost point in the transient quenching process that can lead to a much faster chilldown of the pipe wall to liquid temperatures, which results in significant increases in chilldown efficiency and savings in cryogen consumption. For example, the 4L coating shortened the chilldown time from 17.6 s to 11.5 s for 20% duty cycle condition. Fig. 5. A plot of m/m100% as a function of duty cycle for 80 psig inlet pressure, 4L coated tub and various periods.
3.3.4. The effects of inlet source pressure Finally the effects of different driving source pressures or mass flow rates are also examined. Table 3 lists the results. Three different source pressures of 30 psig, 60 psig and 80 psig were selected with 4L coated test tube and 20% DC. For the 30 psig and 80 psig cases, 2 s period was used, however 3 s period was used for 60 psig. It is apparent that the thermal efficiency increases with decreasing inlet source pressure or decreasing mass flow rate. As expected, the coolant consumption increases with increasing inlet source pressure. 4. Conclusion
Fig. 6. A plot of mconsumed as a function of the duty Cycle for different coating layers with 1 s period.
linearly with the duty cycle that is consistent with the results in Fig. 5. Another factor we addressed is the time to complete the chilldown process. Fig. 7 provides some additional information on the chilldown curves or tube outer surface temperature histories for
The experimental results obtained shows that the chilldown efficiencies with pulse flows cover a range between 41.1% and 81.8%. A general trend was found that indicates the chilldown efficiency increases with decreasing duty cycle and is relatively independent of the pulse flow period. Based on the mean efficiency over the six periods, the 80% DC only produced a 1.41% increase in chilldown efficiency over the reference case of continuous flow without pulsing, while a 65.77% increase is found for the 20% DC case. On the amount of cryogen consumed, a higher chilldown efficiency means less amount of coolant required. Again using the mean cryogen consumption over the six periods, we found that 80% DC only produced a 0.83% decrease in coolant consumption
Fig. 7. Chilldown curves for bare surface and 4L coated test tubes at various duty cycles and 1 s period with 80 psig inlet pressure.
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relative to the reference case, while a 38.22% decrease was obtained for the 20% DC case. Based on a curve fit, we found that m=m100% , where m and m100% are the total mass of LN2 coolant consumed for the pulse flow case and that for the corresponding continuous flow case, respectively, is a linear function of the duty cycle. On the average, with pulse flows, a common inlet pressure of 80 psig, and a common 1 s period, the cryogen mass consumption for a tube coated on the inner surface with 4-layer low-thermal conductivity films is about 57% of that for a bare surface tube (to be exact, 59% for 100% duty cycle, 54% for 60% duty cycle and 57% for 20% duty cycle). Also for each of the two tube types, the cryogen mass consumption increases relatively linearly with the duty cycle. Another factor we addressed is the time to complete the chilldown process. Similar to the finding reported for continuous flows, the low thermal conductivity coating was also able to shorten the chilldown time in film boiling with pulse flows by expediting the approach to the Leindenfrost point in the transient quenching process that can lead to a much faster chilldown of the pipe wall to liquid temperatures, which results in significant increases in chilldown efficiency and savings in cryogen consumption. For example, the 4L coating shortened the chilldown time from 17.6 s to 11.5 s for 20% duty cycle condition. For the effects of different inlet pressures, it is apparent that the chilldown thermal efficiency increases with decreasing inlet source pressure or decreasing mass flow rate. As expected, the coolant consumption increases with increasing inlet source pressure. Acknowledgments This work was partially supported by the Flight Opportunities Program at National Aeronautics and Space Administration (NASA) under the award number NNX16AP68G. This research was also partially supported by the Andrew H. Hines, Jr./Progress Energy Endowment Fund at the University of Florida. Declarations of Competing Interest None. Appendix A. Supplementary material Supplementary data to this article can be found online at https://doi.org/10.1016/j.ijheatmasstransfer.2019.06.080. References [1] V.P. Carey, Liquid-Vapor Phase-Change Phenomena, second ed., Taylor & Francis Group, LLC., New York, 2008. [2] R. Shaeffer, H. Hu, J.N. Chung, An experimental study on liquid nitrogen pipe chilldown and heat transfer with pulse flows, Int. J. Heat Mass Transf. 67 (2013) 955–966. [3] S.K. Sahu, P.K. Das, S. Bhattacharyya, An experimental investigation on the quenching of a hot vertical heater by water injection at high flow rate, Nucl. Eng. Des. 240 (2010) 1558.
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