ABSTRACT 
 
 
 
Title of Dissertation: DESIGN AND PERFORMANCE 
CHARACTERIZATION OF AN 
ADDITIVELY-MANUFACTURED HEAT 
EXCHANGER FOR HIGH TEMPERATURE 
APPLICATIONS 
  
 Xiang Zhang, Doctor of Philosophy, 2018 
  
Dissertation directed by: Professor Michael Ohadi,  
Mechanical Engineering 
 
In its early stages of development, additive manufacturing was used chiefly for 
prototyping, but over the last decade, its use has evolved to include mass production of 
certain products for numerous industries in general, and speciality industries such as 
biomedical and aerospace industries in particular. Additive manufacturing can be used 
to fabricate unconventional/complex designs that are difficult and time-consuming 
through conventional fabrication methods, but offer significant performance advantage 
over state of the art. One such example is high temperature heat exchangers with 
complex novel geometries that can help improve the heat transfer density and provide 
better flow distribution, resulting in more compact and efficient designs and thereby 
also reducing materials costs considering fabrication of these heat exchangers from the 
suitable super alloys with the conventional manufacturing techniques is very difficult 
and laborious.  
  
This dissertation presents the results of the first high-temperature gas-to-gas manifold-
microchannel heat exchanger successfully fabricated using additive manufacturing. 
Although the application selected for this dissertation focuses on an aerospace pre-
cooling heat exchanger application, the results of this study can still directly and 
indirectly benefit other industrial sectors as heat exchangers are key components of 
most power conversion systems. In this work, optimization and numerical modelling 
were performed to obtain the optimal design, which show 30% weight reduction 
compared to the design baseline. Thereafter, the heat exchanger was scaled down to 66 
× 74 × 27 mm3 and fabricated as a single piece using direct metal laser sintering 
(DMLS). A minimum microchannel fin thickness of 165 μm was achieved. Next, the 
additively manufactured headers were welded to the heat exchanger core and the 
conventionally manufactured flanges. A high-temperature experimental loop was next 
built, and the additively manufactured heat exchanger was successfully tested at 600°C 
with ~ 450 kPa inlet pressure. A maximum heat duty of 2.78 kW and a heat transfer 
density close to 10 kW/kg were achieved with cold-side inlet temperature of 38°C during 
the experiments. A good agreement between the experimental and numerical results 
demonstrates the validity of the numerical models used for heat transfer and pressure drop 
predictions of the additively manufactured heat exchanger. Compared to conventional 
plate-fin heat exchangers, up to 25% improvement in heat transfer density was achieved. 
This work shows that additive manufacturing can be used to fabricate compact and 
lightweight high temperature heat exchangers, which benefit applications where space and 
weight are constrained. 
  
  
 
 
 
 
 
 
 
 
DESIGN AND PERFORMANCE CHARACTERIZATION OF AN 
ADDITIVELY-MANUFACTURED HEAT EXCHANGER FOR HIGH 
TEMPERATURE APPLICATIONS 
 
 
 
By 
 
Xiang Zhang 
 
 
 
 
 
Dissertation submitted to the Faculty of the Graduate School of the  
University of Maryland, College Park, in partial fulfillment 
of the requirements for the degree of 
Doctor of Philosophy 
2018 
 
 
 
 
 
 
 
 
 
Advisory Committee: 
Professor Michael Ohadi, Chair 
Professor Amir Shooshtari 
Professor Bao Yang  
Professor Christopher Cadou  
Professor Marino diMarzo  
Professor Peter Chung  
  
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
© Copyright by 
Xiang Zhang 
2018 
 
 
 
 
 
 
 
 
 
 
 
 
 
  
 
Dedication 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
To my parents, my wife, and my son 
who have helped, encouraged, and supported me always 
 
ii 
 
 
 
Acknowledgements 
First and foremost, I would like to thank my advisor Dr. Michael Ohadi and my co-
advisors Dr. Amir Shooshtari and Dr. Serguei Dessiatoun for their guidance, help, 
encouragement, and support for my research. I will carry what I learned from them with 
me throughout the rest of my life. 
I would like to express my gratitude to Dr. Hal Strumpf from Honeywell Aerospace for 
providing specifications and technical advices which were very helpful for the 
completion of this research project. I also wish to thank Mr. Howard Grossenbacher 
from the machine shop of the Aerospace Engineering Department at the University of 
Maryland for his continued help on the high-temperature experimental setup’s 
preparation. I wish to thank Ms.  Becky Lorenz for her help in welding of the test 
section as well.  
I am grateful to my friends at Smart and Small Thermal Systems Laboratory (S2TS), 
especially Farah Singer, Martinus Arie, Fabio Battaglia, Stefan Bangerth, David 
Hymas, Raphael Mandel, David Deisenroth, Daniel Bae, and Ratnesht Tiwari for 
providing a constructive and peaceful environment and also their helpful tips and 
suggestions regarding my research. 
iii 
 
 
 
Table of Contents 
Dedication ..................................................................................................................... ii 
Acknowledgements ....................................................................................................... iii 
Table of Contents ......................................................................................................... iv 
List of Tables ............................................................................................................... vii 
List of Figures ............................................................................................................ viii 
Nomenclature .............................................................................................................. xii 
Chapter 1: Introduction ................................................................................................ 1 
1.1. Motivation and Background .............................................................................. 1 
1.2. Dissertation Objective ........................................................................................ 6 
1.3. Approach ............................................................................................................ 6 
1.4. Major Contributions ......................................................................................... 10 
1.5. Dissertation Layout .......................................................................................... 10 
Chapter 2: Literature Review ..................................................................................... 12 
2.1.  Introduction ..................................................................................................... 12 
2.2. Types of High Temperature Heat Exchangers ................................................. 14 
2.2.1. Plate-fin heat exchanger (PFHX) .............................................................. 15 
2.2.2. Plate-and-frame heat exchanger ................................................................ 16 
2.2.3. Shell-and-tube heat exchanger .................................................................. 17 
2.3. Examples of HTHX Demonstrations ............................................................... 18 
2.3.1. Recuperator for gas turbine power generation .......................................... 18 
2.3.2. Intermediate heat exchanger for nuclear heat utilization .......................... 21 
2.3.3. High temperature prime heat exchanger in externally fired systems ........ 22 
2.3.4. Pre-cooler for aircraft environmental control system applications ........... 22 
2.3.5. High temperature industrial waste heat recovery ...................................... 23 
2.4. HTHX Fabrication using Advanced Manufacturing Techniques .................... 23 
2.4.1. Additive manufacturing ............................................................................ 24 
2.4.2. Photo-chemical etching ............................................................................. 27 
2.5. Comparison of Selected HTHXs ..................................................................... 29 
2.6. Summary .......................................................................................................... 30 
Chapter 3: Conceptual Design ................................................................................... 31 
3.1.  Introduction ..................................................................................................... 31 
iv 
 
 
 
3.2. Design Requirements ....................................................................................... 31 
3.3. Manifold-microchannel Technology ............................................................... 32 
3.4. Gas-to-gas Cross-flow Heat Exchanger Conceptual Design ........................... 39 
3.5. Summary .......................................................................................................... 41 
Chapter 4: Fabrication Alternatives ........................................................................... 42 
4.1. Introduction ...................................................................................................... 42 
4.2. Direct Metal Laser Sintering ............................................................................ 42 
4.3. Photo-chemical Etching ................................................................................... 45 
4.4. Laser Micro-machining .................................................................................... 47 
4.5. Diffusion Bonding ........................................................................................... 48 
4.6. Fabrication Alternatives Risk Assessment ....................................................... 50 
4.7. Summary .......................................................................................................... 52 
Chapter 5: Optimization & Design Selection ............................................................. 53 
5.1. Introduction ...................................................................................................... 53 
5.2. Single Objective Optimization ......................................................................... 53 
5.3. Modified Hybrid Method ................................................................................. 57 
5.4. Optimization Results & Design Selection ....................................................... 63 
5.5. Scaled-down Model ......................................................................................... 68 
5.6. Summary .......................................................................................................... 70 
Chapter 6: Test Print Coupon & Subscale Unit Fabrication ..................................... 71 
6.1. Introduction ...................................................................................................... 71 
6.2. Test Print Coupon Design ................................................................................ 71 
6.3. Test Print Coupon Fabrication & Evaluation ................................................... 74 
6.4. Pressure Containment Coupon ......................................................................... 77 
6.5. Subscale Unit Design ....................................................................................... 80 
6.6. Subscale Unit Fabrication & Evaluation .......................................................... 84 
6.7. Summary .......................................................................................................... 86 
Chapter 7: Low-temperature Performance Characterization .................................... 87 
7.1. Introduction ...................................................................................................... 87 
7.2. Single-Manifold-Multi-Microchannels Model ................................................ 87 
7.3. Low-temperature Experimental Setup ............................................................. 92 
7.4. Low-temperature Experiment Method ............................................................. 97 
v 
 
 
 
7.5. Low-temperature Experiment Results ........................................................... 101 
7.6. Summary ........................................................................................................ 107 
Chapter 8: High-temperature Performance Characterization ................................. 108 
8.1. Introduction .................................................................................................... 108 
8.2. Headers Design for High-temperature Experiment ....................................... 108 
8.3. Flow Simulation with Metal Header .............................................................. 114 
8.4. Metal Headers Fabrication & Evaluation ...................................................... 123 
8.5. Assembled Test Section ................................................................................. 126 
8.6. High-temperature Experimental Setup .......................................................... 131 
8.7. High-temperature Experiment Method .......................................................... 141 
8.8. High-temperature Experiment Results ........................................................... 145 
8.9. Summary ........................................................................................................ 155 
Chapter 9: Advanced Manifold-microchannel Heat Exchanger .............................. 156 
9.1. Introduction .................................................................................................... 156 
9.2. Design Requirement for Advanced Manifold-microchannel Heat Exchanger
............................................................................................................................... 156 
9.3. Advanced Manifold-microchannel Heat Exchanger Design ......................... 158 
9.4. Medium-scale Unit Design ............................................................................ 160 
9.5. Summary ........................................................................................................ 164 
Chapter 10: Conclusions and Proposed Future Work ............................................. 165 
10.1. Conclusions .................................................................................................. 165 
10.2. Proposed Future Work ................................................................................. 171 
10.2.1. Numerical modeling.............................................................................. 171 
10.2.2. Additive manufacturing ........................................................................ 173 
10.2.3. Project Continuity ................................................................................. 174 
Appendices ................................................................................................................ 177 
References ................................................................................................................. 186 
  
vi 
 
 
 
List of Tables 
Table 2-1 State-of-the-art heat exchanger comparison ............................................... 29 
Table 3-1 Physical specifications ................................................................................ 31 
Table 3-2 Operation specifications for full-scale heat exchanger .............................. 32 
Table 4-1 DMLS limitations and tolerances ............................................................... 44 
Table 4-2 Fabrication alternatives risk assessment ..................................................... 51 
Table 5-1 Optimization constraints ............................................................................. 54 
Table 5-2 Optimization results .................................................................................... 64 
Table 5-3 Summary of fabrication process and updated optimized results ................ 68 
Table 5-4 Comparison between full-scale design and subscale unit .......................... 70 
Table 6-1 Test print coupon designs’ minimum feature size dimensions .................. 74 
Table 6-2 Test print coupons measurement ................................................................ 76 
Table 6-3 Geometrical parameters of the subscale unit .............................................. 83 
Table 6-4 Percentage deviation of the fabricated dimensions and mass from the 
corresponding design values ....................................................................................... 85 
Table 7-1 Low-temperature experiment conditions .................................................... 98 
Table 7-2 List of low-temperature measurement equipment and its accuracy ......... 100 
Table 8-1 Inconel 718 spec sheet [110] .................................................................... 113 
Table 8-2 Estimated welding safety factor (at 650°C) ............................................. 114 
Table 8-3 Flow simulation results of the hot-side outlet mixing chamber ............... 135 
Table 8-4 High-temperature test hazard and operability analysis ............................ 140 
Table 8-5 High-temperature experimental test conditions ........................................ 141 
Table 8-6 Adiabatic experimental test conditions .................................................... 142 
Table 8-7 List of high-temperature measurement equipment and its accuracy ........ 144 
Table 8-8 Key geometrical specifications of plate-fin heat exchanger [114] ............. 153 
Table 9-1 Physical specifications for full-scale advanced heat exchanger ............... 157 
Table 9-2 Operating specifications for full-scale advanced heat exchanger ............ 157 
Table 9-3 Optimized advanced manifold-microchannel heat exchanger design ...... 158 
Table 9-4 Comparison between 3-D printed heat exchanger and advanced design . 159 
Table 9-5 Comparison between advanced full-size and medium-scale design ........ 160 
Table 9-6 Comparison between subscale unit and medium-scale design ................. 161 
 
 
 
 
 
 
 
 
 
 
 
 
vii 
 
 
 
List of Figures 
 
Figure 1-1: Heat exchanger in aircraft environmental control system ......................... 2 
Figure 1-2: Plate-fin concept (left) vs. manifold-microchannel concept (right) ........... 3 
Figure 1-3: Multiple layers of manifolds and microchannel surfaces .......................... 4 
Figure 1-4: Flow chart of the project approach ............................................................ 7 
Figure 2-1: Temperature ranges for heat exchanger materials (adapted from reference 
[45])............................................................................................................................. 13 
Figure 2-2: Illustration of main components of PFHX ............................................... 15 
Figure 2-3: Plate-and-frame heat exchanger structure concept .................................. 17 
Figure 2-4: Shell-and-tube heat exchanger structure concept ..................................... 18 
Figure 2-5: DMLS / SLM concept [80] ...................................................................... 25 
Figure 3-1: Manifold-microchannel concept .............................................................. 33 
Figure 3-2: Microchannel computational domain by Copeland et al. [85] ................. 34 
Figure 3-3: Single manifold-microchannel segment computational domain by Ryu et 
al. [89] ......................................................................................................................... 35 
Figure 3-4: Manifold-microchannel computational domain by Boteler et al. [93] ..... 37 
Figure 3-5: Air-to-water manifold-microchannel design concept by Arie [80] ......... 38 
Figure 3-6: Advanced gas-to-gas heat exchanger conceptual design (a); multi-layer 
manifold-microchannel heat exchanger concept (b) ................................................... 40 
Figure 4-1: Manifold design layout ............................................................................ 44 
Figure 4-2: Photo-chemical etching process (adapted from reference [99]) .............. 46 
Figure 4-3: Microchannel surface design layout ........................................................ 47 
Figure 4-4: Laser micro-machining process concept [101] ........................................ 48 
Figure 4-5: Diffusion bonding process concept [103] ................................................ 49 
Figure 4-6: Exploded view of assembly design for one stack of cross-flow manifold-
microchannel through diffusion bonding .................................................................... 50 
Figure 5-1: Manifold-microchannel design layout ..................................................... 55 
Figure 5-2: Single objective optimization flow chart ................................................. 57 
Figure 5-3: Computational domain and boundary conditions of a single manifold-
microchannel by Arie [106]: (1) manifold channel inlet surface with mass flow inlet 
and constant temperature, (2) manifold channel outlet surface with constant pressure 
outlet, (3) microchannel base surface with constant surface temperature, (4) symmetry 
planes .......................................................................................................................... 58 
Figure 5-4: CAD model with 2 passes on both hot and cold sides by Arie [50]: 
combined model (a), subdivided model (b) ................................................................ 60 
Figure 5-5: Unit-cell pattern from cross-flow manifold-microchannel heat exchanger 
[107] ............................................................................................................................ 63 
Figure 5-6: Build crash caused by overhangs [108] ................................................... 65 
Figure 5-7: Overhang issue with cross-flow manifold-microchannel design: 
overhangs at the manifold base and microchannel fins (a); overhangs at the manifold 
walls and microchannel fins (b) .................................................................................. 66 
Figure 5-8: Subscale unit design (the yellow part was designed as connection 
interface for headers) .................................................................................................. 69 
viii 
 
 
 
Figure 6-1: Cross-flow microchannel design layout: both side vertical fins (left); 
inclined fins on one side (right) .................................................................................. 72 
Figure 6-2: Zoom-in section view of cross-flow manifold-microchannel design with 
inclined microchannel fins on hot side and inclined manifold walls on cold-side (a); 
CAD design of the test print coupon (b) ..................................................................... 73 
Figure 6-3: 3-D printed 20 x 16 x 9 mm3 Inconel 718 test print coupon with inclined 
fins and inclined manifold walls ................................................................................. 75 
Figure 6-4: Measurements of the printed Inconel 718 coupon 2: straight fins (top); 
inclined fins (bottom) .................................................................................................. 77 
Figure 6-5: Pressure containment test coupon’s CAD design (left); cross section view 
– right side (middle); cross section view – front side (right) ...................................... 78 
Figure 6-6: 3-D printed Inconel 718 pressure containment test coupon ..................... 79 
Figure 6-7: Pressure containment test coupon with base of 0.3 mm at 20 psig .......... 80 
Figure 6-8: A small sectional sample for measuring internal structure dimensions 
(left); subscale unit printing orientation (right) .......................................................... 81 
Figure 6-9: Hot-side tapered manifold channel with vertical manifold walls (top); 
cold-side tapered manifold channel with inclined manifold walls (bottom) .............. 82 
Figure 6-10: Cross-flow manifold-microchannel configuration ................................. 84 
Figure 6-11: Additively manufactured Inconel 718 subscale unit and the small 
sectional coupon.......................................................................................................... 85 
Figure 7-1: SMMM model for CFD simulation on the hot side ................................. 89 
Figure 7-2: Flow distribution along the manifold channel on hot side (y-axis: mass 
flow rate in each microchannel; x-axis: microchannel location from the beginning to 
the end of manifold channel; “mnd" stands for manifold) .......................................... 91 
Figure 7-3: Hot side pressure drop prediction with SMMM model (cold side mass 
flow rate is constant at 0.006 kg/s) ............................................................................. 92 
Figure 7-4: Schematic of low-temperature experimental setup .................................. 93 
Figure 7-5: Headers’ design for low-temperature experimental setup (a); 3-D printed 
plastic headers (b) ....................................................................................................... 95 
Figure 7-6: Low-temperature experimental setup ....................................................... 96 
Figure 7-7: Experimental and numerical results for constant cold-side mass flow rate 
at 0.006 kg/s and 30°C temperature difference between hot and cold inlet air: heat 
duty (Q) vs. hot-side mass flow rate (ṁhot) (a); hot-side pressure drop (ΔPhot) vs. hot-
side mass flow rate (ṁhot) (b); overall heat transfer coefficient (U) vs. hot-side mass 
flow rate (ṁhot) (c) ................................................................................................... 102 
Figure 7-8: Experimental and numerical results for constant hot-side mass flow rate at 
0.019 kg/s and 30°C temperature difference between hot and cold inlet air: heat duty 
(Q) vs. cold-side mass flow rate (ṁcold) (a); cold-side pressure drop (ΔPcold) vs. cold-
side mass flow rate (ṁcold) (b); overall heat transfer coefficient (U) vs. cold-side mass 
flow rate (ṁcold) (c) .................................................................................................. 104 
Figure 7-9: Cross section view of the headers’ design (a); hot & cold side of the 
subscale unit (b) ........................................................................................................ 106 
Figure 8-1: Cross section view of headers’ design attached to the subscale unit ..... 109 
Figure 8-2: Metal header’s connection interface design: for hot-side inlet header (a) & 
(b); for hot-side outlet and cold-side inlet/outlet headers (c) & (d) .......................... 112 
ix 
 
 
 
Figure 8-3: Solidworks flow simulation results: hot-side inlet header (a); hot-side 
outlet header (b) ........................................................................................................ 115 
Figure 8-4: Simplified single manifold-porous-medium model (top); hot-side full 
model with inlet/outlet headers (bottom) .................................................................. 118 
Figure 8-5: Porous medium calibration: hot-side core pressure drop vs hot-side mass 
flow rate (a); cold-side core pressure drop vs cold-side mass flow rate (b) ............. 120 
Figure 8-6: Cross-section view of manifold channels on hot side (5 full manifold 
channels as inlet channel, 2 half & 4 full manifold channels as outlet channel) ...... 121 
Figure 8-7: Flow distribution in each manifold channel (hot-side mass flow rate at 
0.02 kg/s, cold-side mass flow rate at 0.006 kg/s) .................................................... 123 
Figure 8-8: 3-D printing orientation of the metal headers ........................................ 124 
Figure 8-9: Hot-side inlet header (a); hot-side outlet header (b); cold-side inlet header 
(c); cold-side outlet header (d) .................................................................................. 126 
Figure 8-10: A section from the subscale unit design (a); Welded 3-D printed 
welding-trial section through edge welding (b) ........................................................ 127 
Figure 8-11: Modified Inconel 718 flange’s engineering drawing (unit: inch) (a); 
fabricated Inconel 718 flanges through conventional machining (b) ....................... 129 
Figure 8-12: Test section including core, headers, and flanges before welding ....... 129 
Figure 8-13: Welded test section including core, headers, and flanges .................... 130 
Figure 8-14: Powerex oil-free scroll compressor pump [112] (left); 3 HP scroll 
compressor system .................................................................................................... 131 
Figure 8-15: Schematic of the commercial heater .................................................... 132 
Figure 8-16: Original exhaust fitting ring design on the heater (left); Modified 
stainless steel 304 flange as replacement of the original exhaust fitting ring (right) 133 
Figure 8-17: Power switch & control panel for the electric heater ........................... 134 
Figure 8-18: Numerical model of hot-side outlet mixing chamber at mass flow rate of 
0.02 kg/s .................................................................................................................... 135 
Figure 8-19: Schematic diagram of the high-temperature experimental setup ......... 137 
Figure 8-20: High-temperature experimental setup .................................................. 138 
Figure 8-21: Test section & mixing chambers: before installing thermal insulation 
(top), after installing thermal insulation (bottom) ..................................................... 139 
Figure 8-22: Subscale unit’s two manifold-microchannel layers ............................. 143 
Figure 8-23: Experimental and numerical results for constant cold-side microchannel 
Reynolds number (Rechn,c  120) and 560°C temperature difference between hot inlet 
N2 and cold inlet air: (a) heat duty (Q) vs. hot-side microchannel Reynolds number 
(Rechn,h); (b) overall heat transfer coefficient (U) vs. hot-side microchannel Reynolds 
number (Rechn,h) ....................................................................................................... 146 
Figure 8-24: Experimental and numerical results for constant hot-side microchannel 
Reynolds number (Rechn,h  450) and 560 °C temperature difference between hot 
inlet N2 and cold inlet air: (a) heat duty (Q) vs. cold-side microchannel Reynolds 
number (Rechn,c); (b) overall heat transfer coefficient (U) vs. cold-side microchannel 
Reynolds number (Rechn,c) ....................................................................................... 148 
Figure 8-25: Experimental and numerical results: (a) hot-side pressure drop (ΔPhot) 
vs. hot-side microchannel Reynolds number (Rechn,h); (b) cold-side pressure drop 
(ΔPcold) vs. cold-side microchannel Reynolds number (Rechn,c) ............................. 150 
x 
 
 
 
Figure 8-26: Experimental and numerical results: cold-side pressure drop vs cold-side 
mass flow rate ........................................................................................................... 152 
Figure 8-27: Performance comparison between manifold-microchannel heat 
exchanger with conventional plate-fin heat exchangers ........................................... 154 
Figure 8-28: Comparison of overall heat transfer coefficient (U) of manifold-
microchannel heat exchanger and selected conventional plate-fin heat exchangers 154 
Figure 9-1: Subscale unit design vs. medium-scale unit design ............................... 161 
Figure 9-2: Medium-scale design with the welding interface frame (marked in gray)
................................................................................................................................... 162 
Figure 9-3: Preferred printing orientation for the medium-scale design (a); tapered 
manifold channel’s overhang concern with 45-degree printing orientation (b) ....... 163 
Figure 10-1: Tapered manifold channel design: large tapered angle (a); small tapered 
angle (b) .................................................................................................................... 175 
xi 
 
 
 
Nomenclature 
𝐴𝑏𝑎𝑠𝑒    area of the microchannel surface base [m
2] 
𝐴 2𝑐ℎ𝑛    microchannel cross section area [m ] 
𝐶𝑑  discharge coefficient [-] 
CF  configuration correction factor of both unmixed cross-flow [-] 
COP  coefficient of performance [-] 
𝐶𝑝  specific heat [J/kg-K] 
𝐷𝑐ℎ𝑛    microchannel cross section hydraulic diameter [m] 
𝐷𝑖  nozzle inlet cross section diameter [m] 
𝐷𝑜  nozzle throat diameter [m] 
𝐻   heat exchanger core overall height [m] 
𝐻𝑏𝑎𝑠𝑒   microchannel surface base height [m] 
ℎ𝑏𝑎𝑠𝑒    base conductance (ℎ
2
𝑏𝑎𝑠𝑒 = 𝑄/𝐴𝑏𝑎𝑠𝑒(𝑇𝑖𝑛 − 𝑇𝑏𝑎𝑠𝑒)) [W/m K] 
𝐿   heat exchanger core overall length [m] 
𝐿𝑀𝑇𝐷  log mean temperature difference [K] 
𝑚  mass [kg] 
?̇?   mass flow rate [kg/s] 
𝑁𝑙𝑎𝑦𝑒𝑟  total number of manifold-microchannels layers [-] 
𝑃   system pressure [Pa] 
𝛥𝑃   pressure drop [Pa] 
𝑄   heat duty [W] 
𝑅𝑒   Reynolds number [-] 
𝑇   temperature [C]  
𝛥𝑇   temperature difference [°C] 
𝑡𝑓𝑖𝑛   fin thickness [m] 
𝑈   overall heat transfer coefficient [W/m2K] 
𝑈𝑋  uncertainty of variable X [-] 
?̇?  volumetric flow rate [m3/s] 
𝑊   heat exchanger core overall width [m] 
𝑊𝑐ℎ𝑛   microchannel width [m] 
𝑊𝑎𝑙𝑙𝑚𝑛𝑑  manifold wall thickness [m]  
 
Greek Symbols  
𝜇  dynamic viscosity [kg/(m·s)] 
𝜌   density [kg/m3] 
 
Subscripts 
base             microchannel surface base 
xii 
 
 
 
cal  calibration 
chn   microchannel 
cold   cold side flow 
exp  experiment 
hot   hot side flow 
in   inlet  
mnd  manifold 
num   numerical 
out   outlet 
std  standard  
 
Abbreviations 
AM  additive manufacturing 
DMLS  direct metal laser sintering 
DOE   design of experiment  
EBM  electron beam melting 
ECS  environmental control system 
HTGR  high temperature gas-cooled reactor 
HTHX  high temperature heat exchanger 
HX  heat exchanger 
IHX  intermediate heat exchanger 
LOM   laminated object manufacturing 
LTHX  low temperature heat exchanger 
M2HX  manifold-microchannel heat exchanger 
PCHX  printed circuit heat exchanger 
PFHX  plate-fin heat exchanger 
sCO2  Super critical CO2 
SLM  selective laser melting 
 
 
xiii 
 
 
 
Chapter 1: Introduction 
1.1. Motivation and Background 
Energy saving through heat recovery has been an important objective in many 
industries. Researchers have identified the importance of high temperature heat 
exchangers with operating temperatures of 500-1000°C to recover heat from a hot 
exhaust stream and transfer it for heating cold inputs to the system [1]. It is well 
established that higher cycle efficiencies can be obtained when the turbine inlet 
temperature for a typical power generation cycle is increased. However, increases 
in turbine inlet temperature are limited by the high temperature heat exchangers 
which support the power generation cycle. One of such high temperature heat 
exchangers is used in aircraft environmental control system (ECS), in which heat 
exchangers are used to cool compressor bleed air using ram air as the heat sink, and 
the cold exhaust air on the hot side is used to condition the aircraft cabin as shown 
in Figure 1-1 [2]. High temperature heat exchangers can also be used to increase 
gas turbine efficiency where high temperature recuperators used for this purpose 
have a typical operating temperature of about 580°C [3, 4]. Another application of 
high temperature heat exchangers is in gasoline fueled vehicles to reduce fuel 
consumption by recovering waste heat from exhaust ranging from 200 to 850°C 
and converting it to mechanical or electrical energy [5, 6]. High temperature heat 
exchanger is also used in gas cooled nuclear reactors where an intermediate gas-
gas heat exchanger recovers nuclear heat (> 800°C) and transfers it to secondary 
systems to generate electricity or hydrogen [7-9]. 
1 
 
 
 
Figure 1-1: Heat exchanger in aircraft environmental control system 
Currently, the most widely commercialized type of compact heat exchanger used 
in high temperature gas-to-gas applications is the plate-fin heat exchanger (PFHX) 
[10-12]. To improve the performance over the conventional designs, novel heat 
transfer surfaces have been suggested in recent years. One such design is the 
manifold-microchannel technique first proposed by Harpole and Eninger [13] in 
1991. Since then, numerous studies have been reported in the literature discussing 
the superior performance of this technology in various applications compared to 
conventional technologies [14-26]. Unlike the plate-fin design, the manifold-
microchannel concept design utilizes a manifold placed on top of microchannels as 
shown in Figure 1-2. Then, through the manifold, the air is distributed into the 
microchannels and travels a short length in each microchannel before it is guided 
out. Due to the short flow length, the air flow in the microchannel typically remains 
in the developing regions, which results in substantially higher heat transfer 
coefficients. Studies by Arie et al. [15, 25] show that an air-cooled air-to-water heat 
exchanger utilizing manifold-microchannel has higher heat transfer density and 
2 
 
 
heat transfer coefficient than other state-of-the-art dry cooling heat exchangers used 
for power plants. Jha et al. [17, 18] investigated the performance of a tubular 
manifold-microchannel evaporator for waste heat recovery systems and reported 
high heat transfer coefficients. Researchers also demonstrated enhanced 
performance of plate heat exchangers with manifold-microchannel used for 
refrigeration/air conditioning [19, 24, 26]. 
 
Figure 1-2: Plate-fin concept (left) vs. manifold-microchannel concept (right) 
Despite the superior performance of manifold-microchannel heat exchangers for 
various applications reported in the literature, their main drawback is the difficulty 
associated with their manufacturing. When the manifolds and the microchannel 
surface are conventionally manufactured separately, the heat exchanger with 
multiple layers of manifolds and microchannel surfaces (as shown in Figure 1-3)  
must be assembled through brazing, welding or diffusion bonding. In these 
processes, the bonding material may penetrate and obstruct the microchannels, 
adversely affecting the performance of the heat exchanger. Moreover, for large-
scale heat exchangers consisting of hundreds of manifold-microchannel layers, the 
assembly process is challenging and time consuming. 
3 
 
 
 
Figure 1-3: Multiple layers of manifolds and microchannel surfaces 
Additive manufacturing (AM) has evolved from prototyping to mass production in 
the last decade and is now used in many industries such as biomedical, automotive, 
and aerospace [27, 28]. AM can produce lightweight structures with high 
mechanical robustness and has capability to fabricate unconventional designs that 
are not possible using conventional fabrication methods. Heat exchangers are one 
such applications for which AM has recently gained traction. Additively 
manufactured heat exchangers can be fabricated with complex internal geometries 
and external shapes with lighter weight and higher compactness which are key 
factors for heat exchangers used in high end applications such as aerospace 
industry. The shorter lead time to produce the heat exchangers is another advantage 
4 
 
 
of AM. High temperature heat exchangers typically involve expensive materials. 
Additive manufacturing of these heat exchangers by incorporating unconventional 
designs can help improve the heat transfer density, and provide better flow 
distribution, resulting in compact and efficient heat exchangers at reduced cost.  
Additive manufacturing technology has shown great potential not only for the 
design and construction of complex geometries but also for simplifying the 
assembly process and reducing fabrication steps. Recently the direct metal laser 
sintering (DMLS) process has been used successfully to fabricate metallic fins as 
thin as 220 μm for an air-to-water manifold-microchannel heat exchanger produced 
as a single unit (manifold-microchannel was only used on the air side while water 
side had simple rectangular channels) [29-33]. In another recent study, Grestler and 
Erno’s [34] metallic heat exchanger, which is used as a turbine engine fuel-cooled 
oil cooler, was also fabricated through selective laser melting (SLM) without any 
assembly process.  
Although there have been a few studies on additively manufactured heat 
exchangers, no work is reported in the literature which addresses development of 
AM heat exchangers for high-temperature gas-to-gas applications such as those 
found for heat exchangers in aircraft environmental control systems, which must 
be able to operate up to temperatures of 650°C with bleed hot air [16, 34-37]. The 
current work demonstrates the successful design and fabrication through DMLS of 
a gas-to-gas manifold-microchannel heat exchanger using Inconel 718. In this 
work, the manifold-microchannel was used on both hot and cold sides. The heat 
exchanger was experimentally tested with N2 on the hot-side at 600°C and air on 
5 
 
 
the cold-side at 38°C. The experimental results were compared against the 
calculated performance of conventional plate-fin heat exchangers to assess the 
performance enhancement.  
1.2. Dissertation Objective 
The main objective of this dissertation is to design, built and test a gas-to-gas heat 
exchanger that can provide significant performance enhancement compared to 
conventional heat exchangers for aerospace application. This is to be accomplished 
by implementing advanced fabrication techniques, advanced heat transfer surfaces, 
and design optimization. Accomplishing this objective will be an important step 
towards development of economical gas-to-gas heat exchangers for aircrafts 
environmental control system. In addition, although this study focuses on gas-to-
gas pre-cooling for aircraft, the knowledge gained in this study can be applied to 
many other applications where air or gas is used as the heat transfer medium, such 
as HVAC, automotive (car radiator), and electronic cooling applications. 
1.3. Approach 
To accomplish the project objective, metallic heat exchangers were exhaustively 
studied. Metallic heat exchangers are the most common type of heat exchangers, 
especially for aerospace applications, due to their high structural strength. To 
demonstrate the technology, a full-scale air-to-air heat exchanger design was 
created based on the operation conditions through an optimization process. Then, a 
scaled-down unit was designed, fabricated, and experimentally tested. This was 
motivated by the notion that a full-scale aerospace pre-cooling heat exchanger is 
6 
 
 
too risky and expensive to build. The scaled-down unit also does not require a 
sizeable test setup compared to the full-scale unit. The heat exchanger can be then 
scaled up for the given application based on the experimentally validated model. 
The scaled-up design was compared with the conventional design to calculate how 
much improvement the advance design could yield. A flow chart showing the major 
steps performed to accomplish the dissertation objective is shown in Figure 1-4.  
 
Figure 1-4: Flow chart of the project approach  
A summary of each step is provided in the following. 
1. Literature review of high temperature heat exchangers 
A detailed literature review of high temperature heat exchangers was 
performed. The study included identifying common types of high 
7 
 
 
temperature heat exchangers (HTHXs) and their corresponding applications 
in the literature.  
2. Define the design requirements & create conceptual design 
The design requirements for the heat exchanger, such as heat exchanger 
capacity, pressure drops on both hot and cold sides, inlet temperatures, 
system pressures, and mass flow rates were defined based on specifications 
of a typical heat exchanger on aircraft. A conceptual design was then created 
based on the manifold-microchannel technology and the design 
requirements. 
3. Identify fabrication alternatives & process constraints  
A study was performed on the advanced manufacturing techniques that can 
be used to fabricate the metallic high temperature heat exchanger. The study 
included fabrication techniques and fabrication limitations such as 
minimum feature size and maximum fabrication size.  
4. Create the heat exchanger design  
Based on the chosen fabrication approach, the fabrication limitations and 
constraints identified in the previous step were implemented for the design 
selection and sizing. A full-scale heat exchanger design was then created 
through the optimization process. 
5. Fabricate test print coupon & update fabrication process constraints 
To save time and confirm the actual fabrication constraints, several 
sectional test coupons were fabricated by the additive manufacturing 
8 
 
 
technique. A detailed evaluation was performed to check the quality of the 
fabricated coupons. If the measured parameter (for example, fin thickness) 
turned out to be significantly different from the design value, the design 
process constraints would be updated, and a new design would be obtained. 
Then this step would be repeated until all the manufacturing constraints are 
satisfied.  
6. Fabricate & test scale-down unit  
Once the test print coupon’s quality was confirmed from the previous step, 
a subscale unit was created by scaling down the full-scale unit to meet the 
AM fabrication size limitation for the identified AM printer. The subscale 
unit was then fabricated through additive manufacturing and tested under 
high temperature conditions, which were chosen based on the scaling-down 
process, initial design requirements, and the capability of the test setup.  
7. Compare test results to numerical prediction & update numerical model 
By comparing the test results of the subscale unit to the numerical 
predictions, if the difference was too large, the numerical model would be 
examined and improved, and some or all previous steps would be repeated. 
If the heat exchanger design requirements and fabrication 
constraints/limitations were updated, then another full-scale as well as 
subscale design would be created. The new subscale unit would be tested to 
check the revised numerical model. 
9 
 
 
1.4. Major Contributions  
The major contributions of this study are as follows: 
1. The successful use of additive manufacturing is demonstrated for the first 
time to print a metal gas-to-gas cross-flow manifold-microchannel heat 
exchanger for high-temperature applications.  
2. A computational model was developed, and the model was verified by 
comparison of the results with those of the experiments, thus verifying the 
model that can be used as a design tool. 
3. The performance of an additively manufactured manifold-microchannel 
heat exchanger is characterized for the first time at temperatures up to 600ºC 
and pressures up to 4.5 bar through a new high-temperature experimental 
setup. This setup can be used to test different high temperature heat 
exchangers under desired conditions.   
Details of the work are documented in respective chapters of this dissertation. 
1.5. Dissertation Layout 
The layout of this dissertation is as follows. First a literature study of conventional 
and state-of-the-art high temperature heat exchangers is presented in chapter 2. The 
design requirements are defined in chapter 3. In addition, the manifold-
microchannel concept is explained. Subsequently, the fabrication alternatives are 
presented in chapter 4. The fabrication process and constraints are also discussed 
in detail. Next, the numerical method used to predict the performance is explained 
10 
 
 
in chapter 5. The optimal full-scale air-to-air heat exchanger design was also 
selected based on the optimization results with different fabrication approaches. 
Then, the full-scale design’s scaling-down process is presented. The test print 
coupon study is discussed in chapter 6, and the subscale unit’s design and 
fabrication are discussed as well. The experimental work on the scaled-down heat 
exchanger is discussed in later chapters. The experimental results are then 
compared with numerical predictions. Considering the decent agreement between 
the experimental and numerical data, an advanced design was created and discussed 
in chapter 9. Lastly the conclusion and future work are presented in the last chapter. 
  
11 
 
 
Chapter 2: Literature Review 
2.1.  Introduction 
High temperature heat exchangers (HTHXs) are described in the literature as 
operating above an arbitrary value of 500°C, and early development of HTHXs was 
motivated by the first oil shock in 1973, which created a demand for energy saving 
and more effective use of energy [38]. Since then, various types of HTHXs have 
been developed, and many have been used as recuperators for gas turbine engines, 
aerospace, hydrogen production, waste heat recovery, supercritical CO2 power 
cycles, and high temperature fuel cell systems [39-43]. Heat exchangers operating 
at such high temperatures are subject to unique material challenges such as creep, 
reduced strength at higher temperatures, oxidation of material, corrosion, and 
thermal shock.   
Most materials used at lower temperatures, such as steel, copper, and aluminum, 
lose their thermo-mechanical properties at elevated temperatures. For this reason, 
they cannot be used for HTHX applications. Rather, materials which can retain their 
properties at high temperatures, such as high-grade steels and Ni-based alloys, are 
used for HTHXs. The strength of some material candidates for HTHXs application 
are shown in Figure 2-1. As shown there, most of these materials lose strength at 
temperatures exceeding 550°C and particularly above 600°C. Depending upon the 
operating temperature and pressure, high temperature materials can vary from steels 
to Ni-based alloys and ceramics. The prices for the material can vary from ~$2-
3/kg for steel to ~$70/kg for Ni-based alloys [44]. 
12 
 
 
 
Figure 2-1: Temperature ranges for heat exchanger materials (adapted from 
reference [45])  
As a result, expensive alloys that retain their strength at elevated temperatures are 
typically the material of choice. However, these alloys typically have low thermal 
conductivity, and difficult manufacturability [46, 47], and thus they present their 
own challenges in heat exchanger design and fabrication. Even these alloys lose 
strength at elevated temperatures. This low strength at higher operating pressures 
means that the walls must be thicker, requiring more material. Thick wall geometry 
13 
 
 
makes the HTHXs made of expensive alloys cost prohibitive in most power cycle 
applications. One way to offset these costs is to develop compact heat exchangers 
that have higher surface area to volume ratio and thus reduce the amount of material 
needed. As a result, in past couple of decades, there has been increased interest in 
development of compact and cost effective HTHXs [48, 49]. Advanced 
manufacturing techniques, mainly additive manufacturing of metals, have recently 
shown potentials in fabricating compact HTHXs with innovative heat transfer 
surface designs [50]. This chapter provides insights into recent developments in 
design and application areas of HTHXs. 
2.2. Types of High Temperature Heat Exchangers 
The cost of HTHX increases exponentially as the operating temperature increases, 
particularly above 600°C, mainly due to the material cost as well as manufacturing 
cost of the superalloy and ceramic heat exchangers [12]. Higher pressure 
applications such as those encountered in the power cycles complicates the issue 
even further. The majority of conventional heat exchanger designs used in low 
temperature applications prove to be uneconomical at high operating temperatures. 
The high cost of exchangers in power plant applications such as supercritical CO2 
Brayton cycles is a major stumbling block to making the cycle economical [10]. 
As a result, newer designs of heat exchangers that utilize the materials more 
efficiently, namely higher surface area to volume ratio designs, are being developed 
[10, 50]. These designs typically use microchannels as well as fin geometries to 
accomplish the higher area. An added advantage of smaller channel sizes is that the 
14 
 
 
heat transfer in such miniature geometries is much higher  [51, 52]. This section 
reviews various heat exchanger types and designs utilized in high temperature 
applications.  
2.2.1. Plate-fin heat exchanger (PFHX) 
Plate-fin heat exchangers (PFHXs) are one of the most commonly utilized HTHXs 
for diverse industrial sectors. They are used mainly for gas-to-gas heat transfer 
applications. The main components of a PFHX, including side bars, fins, and 
parting sheets, are shown in Figure 2-2. The fins are usually fabricated using a 
stamping process and are brazed together with the base plates. A brazed PFHX can 
withstand a maximum pressure of 90 bar, while diffusion bonded PFHXs can be 
used under pressures up to 200 bar [10].  
 
Figure 2-2: Illustration of main components of PFHX  
The fins in PFHXs can be easily rearranged, which allows the PFHXs to operate in 
any of the cross-flow, counter-flow, or cross-counter-flow configurations. The 
15 
 
 
main applications for PFHXs at high temperatures are gas turbine and power plants 
for hot gas heat recovery. Generally, PFHXs have a good heat transfer area to 
volume ratio and hence can be compact and economical for high-temperature 
applications [10]. 
2.2.2. Plate-and-frame heat exchanger 
Plate-and-frame heat exchangers are often used to transfer heat between two liquids 
or two gases. The fluid-separating plates of these heat exchangers are typically 
manufactured by compression processing of a thin metal sheet, and they come in 
several patterns such as wavy, chevron, washboard, herringbone, cross-corrugated, 
cross-undulated, or cross-wavy [49, 53-55]. Two such heat transfer plates are then 
stacked to produce a single cell, and this process is repeated to manufacture the 
required number of cells. The structural strength of the core is achieved through the 
connection of the end plates once all the cells are stacked, as shown in Figure 2-3 
[56]. For high pressure applications, the plates can be welded or brazed together to 
ensure operation up to 200 bar pressure and 815°C temperature [57]. Two layouts 
are typically used for plate-and-frame heat exchangers in recuperators in 
microturbine systems: 1) rectangular designs, which are installed behind the 
rotating machinery, and 2) annular designs, which are wrapped around the turbine 
[49, 58]. 
16 
 
 
 
Figure 2-3: Plate-and-frame heat exchanger structure concept  
2.2.3. Shell-and-tube heat exchanger 
Shell-and-tube heat exchangers (Figure 2-4) are the most common type used in 
industry. The tube diameters vary from 0.625” to 1.5” (~16 mm to 38 mm) in 
conventional heat exchangers. These heat exchangers have very low surface area 
to volume ratio and hence are generally not economical for high-temperature and 
high-pressure applications. Chordia et al. [10], however, developed a shell-and-
tube heat exchanger with a large number of tubes of diameter close to 1 mm to 
achieve small wall thickness and very high surface area to volume ratio. Since the 
channels are on the order of millimeters or less, the heat exchanger benefits from 
the high heat transfer. Since this type of heat exchanger can handle more severe 
conditions and higher pressures and temperatures, it can be used in gas turbine 
systems if there is no space limitation. It can also be used as a high temperature 
gas-cooled reactor for nuclear heat utilization.  
17 
 
 
 
Figure 2-4: Shell-and-tube heat exchanger structure concept 
2.3. Examples of HTHX Demonstrations   
High temperature heat exchangers have several applications ranging from 
aerospace, waste heat recovery, nuclear heat utilization, exhaust gas recuperators, 
and as primary heaters and recuperators in advanced high efficiency power cycles 
such as supercritical CO2 Brayton cycle. The present section reviews some of the 
applications of the HTHXs currently being used or developed. 
2.3.1. Recuperator for gas turbine power generation  
For over half a century, gas turbines have been widely used in power plants and 
aircraft propulsion, in which heat exchangers have always played an important role. 
The heat exchanger serves as a recuperator, which preheats compressed air before 
it enters the combustor by recovering heat from exhaust gas. Thus, the gas turbine 
18 
 
 
cycle efficiency can be increased and fuel consumption can be reduced [59]. For 
this application, the typical compressor exit gas temperature can be as high as 
725°C, which requires an HTHX as the recuperator [4].  Three major types of 
recuperators in gas turbine systems are used: PFHXs, plate-and-frame heat 
exchangers, and shell-and-tube heat exchangers [60]. 
For an optimized gas turbine cycle, whose inlet gas/air temperature is 825°C/ 521°C 
and pressure ratio is 24.3, Aquaro [7] proposed a PFHX fabricated from superalloy 
with increased fin thickness of 0.15 mm on the air side to withstand the pressure at 
these maximum operating temperatures. A PFHX designed by Toyo [61] was used 
as a recuperator with effectiveness of about 90% in intercooled and recuperated 
micro gas turbines at about 650°C under inlet pressure about 400 kPa. Similar to 
the external shape of the PFHX design from Toyo, Ingersoll-Rand [62] also 
developed a plate-fin recuperator, which adopts offset fins in the heat transfer area, 
operates at 700°C with effectiveness of about 90%, and has a cycle pressure ratio 
up to 14. AlliedSignal (Honeywell’s predecessor) [63, 64] produced a compact 
plate-fin ceramic recuperator for a cruise-missile propulsion application, and an 
industrial gas-turbine plate-fin recuperator (effectiveness of 84-89%) with offset 
plate-fin surfaces which operates at 510 – 575°C and has a pressure ratio of 10. 
A plate-and-frame type recuperator with effectiveness of about 90% and operating 
temperature of 650°C has been developed by Rekuperator Svenska AB (RSAB) 
[65] for application in a micro gas turbine power plant for combined electricity and 
heat generation. Another welded plate-and-frame counter-flow recuperator was 
developed by Honeywell [66]. A plate-and-frame recuperator with effectiveness of 
19 
 
 
90% was designed by Wilson et al. [67] using silicon carbide, which improved the 
overall thermal efficiency from 27% to over 40%, since the material can withstand 
inlet hot gas temperature of 955°C. 
Oak Ridge National Laboratory [58] developed an annular plate-and-frame 
recuperator which can withstand inlet gas temperatures as high as 850°C, inlet 
pressure of 690 kPa, and offers benefits from minimal ducting of the system. 
Another counter-flow annular plate-and-frame recuperator with effectiveness of 
90% and operating temperature of 725°C was designed by ACTE [68]. Rolls-Royce 
[69] also developed a spiral recuperator which operated with pressure ratio up to 
14 and effectiveness up to 92%. 
In addition to plate-and-frame heat exchangers and PFHXs, shell-and-tube heat 
exchangers are commonly used in gas turbine systems if size is not the limiting 
factor. Proe Power Systems [70] developed the Proe 90TM shell-and-tube 
recuperator, which operates with maximum temperature of 891°C and maximum 
inlet pressure of 621 kPa, and has a predicted heat duty of 30 kW, effectiveness of 
95%, and relatively low pressure drop (1-6 kPa). Schönenborn et al. [71] also 
developed a cross-counter flow recuperator, which operates with maximum 
temperature of 671°C and maximum inlet pressure of 3091 kPa. This recuperator 
was fabricated with two manifold tubes and a bundle of profile tubes using Iconel 
625. However, the estimated total weight of this shell-and-tube type recuperator 
was about 1,000 kg per engine, which clearly adds too much weight to the engine 
system on aircraft [4]. 
20 
 
 
2.3.2. Intermediate heat exchanger for nuclear heat utilization    
Another application of HTHXs is nuclear heat utilization. For nuclear power plants, 
the heat exchangers must be both economically competitive and meet stringent 
safety requirements.  Excellent safety characteristics can be achieved with high 
temperature gas-cooled reactors (HTGRs), since HTGRs have high heat capacity 
with their graphite core and high chemical stability with helium as the coolant [7]. 
However, to transfer the nuclear heat to the end user’s facility, such as a hydrogen 
production or steam reforming system, a high temperature intermediate heat 
exchanger (IHX), which may use He as the working fluid on both sides and must 
operate above 900°C, is needed [39]. In the 1980s, a shell-and-tube He/He IHX was 
designed and constructed in Germany with a heat duty of 10 MW, which 
successfully ran some tests for several months up to 950°C [64, 72]. Another 10 
MW He/He IHX based on the shell-and-tube design was designed in Japan to 
operate at temperatures above 900°C [9]. The IHX is a vertical helically coiled 
counter-flow heat exchanger. Primary He enters from the bottom at 950°C with 
inlet pressure of 4 MPa, while secondary He enters from top at 200°C with inlet 
pressure of 4.1 MPa. To minimize constraints of axial and radial thermal 
expansions on the tubes, a floating hot header with a combination of a central hot 
gas duct was adopted which passes through the central space inside the helix 
bundle. 
21 
 
 
2.3.3. High temperature prime heat exchanger in externally fired systems    
An externally fired combustion process takes place at atmospheric pressure outside 
the working fluid operating cycle. Generally, an externally fired system may have 
hot combustion temperatures of 1500°C and higher [73]. Therefore, an HTHX is 
required to transfer heat from combustion to the gas turbine working fluid, which 
may have pressure up to 17 bar for a gas turbine or 250 bar or higher for a 
supercritical CO2 power generation cycle. As the heat exchanger has to withstand 
the stresses imposed by the working conditions and the constituents in the 
combustion gases, a shell-and-tube heat exchanger may generally be the preferred 
candidate for externally fired systems [74]. 
One such HTHX is presented by K. A. Al-Attab [75]. In this biomass fuel powered 
system, a two-pass cross-flow heat exchanger with baffle shells and tubes was 
chosen to transfer the combustion heat to the compressed air. Due to the selected 
material, stainless steel, the maximum turbine inlet temperature was limited to 
694°C. If a nickel-based superalloy had been used for the heat exchanger’s 
fabrication, a maximum turbine inlet temperature of 800 – 825°C may have been 
reached [74]. 
2.3.4. Pre-cooler for aircraft environmental control system applications     
The environmental control system is used in aircrafts to maintain a comfortable 
closed environment by keeping temperature within acceptable limits. Hot bleed air 
from the engine compressor, which has temperature of 500 to 750°C is precooled 
using HTHX, so that it can circulate within the various systems of the aircraft for 
22 
 
 
other usage [76-78]. The pre-coolers are typically compact PFHXs that have 
relatively high heat transfer surface area per volume ratios. Since weight is a critical 
design point for components on aircraft, a plate-fin carbon-carbon heat exchanger 
operating at 650°C was designed by Stevenson et al. [2]. This HTHX achieved 
about 40% mass reduction over a common metallic HTHX. However, multi-
component coatings were needed for this carbon-carbon heat exchanger to avoid 
the oxidation issue at this operating temperature.  
2.3.5. High temperature industrial waste heat recovery     
Industrial waste heat refers to the energy generated as a byproduct in industrial 
processes that otherwise is not put to useful practical use. Examples of high 
temperature waste heat sources with temperatures of more than 500°C include 
metal refining/heating furnaces, hydrogen plants, and glass melting furnaces. For 
example, exhaust gases leaving the reverberatory furnace, which are frequently 
used in aluminum melting operations, usually have temperatures from 600 to 
1300°C, and 40-60% of furnace energy input can be carried way with these exhaust 
gases. HTHXs can be utilized to recover most or a good portion of this waste heat 
and improve the total cycle efficiency. 
2.4. HTHX Fabrication using Advanced Manufacturing Techniques 
Processing, machining, welding and brazing of superalloys is difficult due to their 
higher toughness, low thermal conductivity, tendency to crack during welding, or 
the unavailability of suitable brazing materials [46, 47, 79]. Specialized equipment 
and highly trained operators are often needed for processing such materials. 
23 
 
 
However, new developments in advanced manufacturing techniques such as 3-D 
printing can address the challenges faced by conventional manufacturing. This 
section discusses HTHXs fabricated using various advanced manufacturing 
techniques such as additive manufacturing and photo-chemical etching processes. 
2.4.1. Additive manufacturing 
Additive manufacturing (AM), also known as 3-D printing, emerged nearly three 
decades ago and was initially used mainly for quick prototyping and production of 
specialized parts. However, due to its high degree of freedom, especially for 
fabrication of complex parts, there has been significant development of industrial-
scale 3-D printers since then. 3-D printers now can print various metals, ceramics, 
and other tough-to-machine materials. As a result, AM has evolved from 
prototyping purposes to production of complex parts used in industries such as 
aerospace, biomedical, oil and gas. 
There are three major metal-based AM techniques: selective laser melting (SLM), 
direct metal laser sintering (DMLS), and electron beam melting (EBM). SLM and 
DMLS printers usually consist of two platforms. The first platform is a powder 
dispenser platform that houses the metal powder. The second platform is the build 
platform on which the 3-D structure is built. After a layer of the 3-D structure is 
built, the powder dispenser platform rises while the build platform lowers, so a new 
layer of powder can be distributed on top of the existing layer. A re-coater arm is 
used to uniformly distribute the metal powder, as shown in Figure 2-5. In the case 
of DMLS, the metal powder is sintered using a laser just below its melting 
24 
 
 
temperature. For the case of SLM, the powder is completely melted. Lenses focus 
the laser beam while a scanning mirror controls the beam or spot location based on 
a slicing of the 3-D structure CAD file. The process is repeated until the entire 3-D 
structure is built. The EBM process is similar to that of SLM. The major difference 
is that an electron beam is used instead of a laser to melt the powder. Due to the use 
of the high-power electron beam, parts fabricated using EBM will have better 
mechanical strength than parts fabricated using the SLM or DMLS process. 
 
Figure 2-5: DMLS / SLM concept [80] 
Laminated object manufacturing (LOM) is another AM technique capable of 
fabricating metal heat exchangers. In the LOM process, a 3-D structure is built layer 
by layer by cutting a sheet of material using a laser. A heated roller bonds the build 
part in the current layer onto the previous layer. The process is repeated until the 
entire 3-D structure is built. Compared to DMLS and SLM, the LOM process is 
25 
 
 
simpler and cheaper. In addition, LOM can be used to fabricate ceramic heat 
exchangers. However, the finish quality and accuracy of LOM are not as good as 
DMLS or SLM.   
Additive manufacturing can help fabricate the complex and compact geometries of 
HTHXs which are otherwise almost impossible to fabricate. Several superalloys 
such as Inconel 718 and Inconel 625 are compatible with AM. It also allows 
fabrication of the heat exchanger as a single component, which eliminates the need 
to weld or braze different parts. There have been a couple of successful attempts to 
fabricate HTHXs using AM, as summarized below. 
1. Multi-furcating heat exchanger  
A 3-D printed HTHX was developed by Gerstler and Erno [34] from 
General Electric Global Research. They successfully fabricated multi-
furcating heat exchangers using SLM for fuel-cooled oil cooler 
applications. The heat exchanger surfaces were fabricated using four 
different materials: aluminum, titanium alloy (Ti64), cobalt chrome, and 
Inconel 718. The test results showed that the heat exchangers met the 
pressure drop and heat transfer design requirement with 66% lower weight 
and 50% lower volume than the conventional heat exchangers.  
2. Advanced fin-and-tube heat exchanger  
In another study at University of Maryland, in collaboration with Oak Ridge 
National Laboratory (ORNL) [36], a novel fin-and-tube heat exchanger was 
26 
 
 
successfully fabricated out of titanium using DMLS. The heat exchanger 
has a complex fin geometry which cannot be economically fabricated using 
conventional manufacturing methods. 
3. Novel air-to-water heat exchanger 
Arie et al. [16] reported a novel air-to-water heat exchanger design 
fabricated through the DMLS process using different materials such as 
aluminium alloy (AlSi10Mg), stainless steel (SS17-4), and titanium alloy 
(Ti64). A fin thickness of 220 μm on the air side was also claimed as the 
minimum feature inside the heat exchanger fabricated using DMLS. Arie et 
al. [35] also developed another air-to-water polymer heat exchanger made 
from thin polyethylene sheets using layer-by-layer line welding and showed 
superior air-side performance over a plane plate fin surface. 
2.4.2. Photo-chemical etching  
Photo-chemical etching is a fabrication process that utilizes a photoresist and 
etchants to machine away a certain area of a metal plate. The process was initially 
developed for fabrication of printed circuit boards. However, due to its ability to 
etch various metals, including titanium, nickel superalloy, and copper superalloy, 
and its high accuracy, the process has been used to fabricate printed circuit heat 
exchangers. 
During the etching process, first, a photoresist layer is deposited on the metal 
surface. Then, the photoresist is exposed to UV light via a photo-tool. Afterward, 
27 
 
 
the exposed metal is dissolved via an etching process to form semi-circular 
channels with typical channel width of 0.5-2 mm [81]. Lastly, the photoresist is 
removed using a solution like alkaline. If the photo-tool specifies the area that needs 
to be dissolved by the UV light, the process is called positive-working photoresist. 
If the photo-tool specifies the area that is left, the process is called negative-working 
photoresist. 
Photo-chemical etching has been extensively used to fabricate printed circuit heat 
exchangers (PCHEs). The design consists of stacked plates with fine grooves 
etched into each plate. The stacked plates are usually bonded using a diffusion 
bonding technique. The common flow configurations of PCHE include counter-
flow, cross-flow, and cross-counter-flow. Some PCHEs are claimed to withstand a 
maximum design pressure of 900 bars and temperatures up to 980°C [82]. Heatric 
[83] has successfully marketed a printed circuit heat exchanger fabricated using 
photo-chemical etching and diffusion bonding. 
PCHEs achieve high thermal performance and are compact due to the smaller 
channel geometries. However, they incur larger pressure drops due to the long 
straight microchannels. The chemical etching process itself is expensive for the 
high temperature materials, as they are tough to etch. This makes these heat 
exchangers relatively expensive, especially for higher temperature applications. 
The low mass-based heat transfer density (kW/kg) and high cost of such heat 
exchangers has been a challenge for their adaptation in high temperature (>600°C) 
applications. 
28 
 
 
2.5. Comparison of Selected HTHXs  
Table 2-1 compares state-of-the art HTHXs. As seen in the table, for the listed heat 
exchangers, the shell-and-tube heat exchangers offer the highest maximum 
temperature and pressure. Shell-and-tube heat exchangers also benefit from being 
well researched and widely used in a diverse range of applications with good 
reliability. However, they are usually bulky and heavy due to their low heat transfer 
density. Plate-and-frame heat exchangers and plate-fin heat exchangers offer higher 
heat transfer density than shell-and-tube heat exchangers, although they need to 
operate at lower temperature and pressure. In addition, recent development of 
printed circuit heat exchangers has resulted in compact heat exchangers with very 
high heat transfer density. However, the high pressure drop of the printed circuit 
heat exchanger is the major limitation for many applications where pumping power 
is constrained. 
Table 2-1 State-of-the-art heat exchanger comparison 
 𝑨𝒉𝒆𝒂𝒕/𝑽  𝑻𝒎𝒂𝒙  𝑷𝒓𝒆𝒔𝒔𝒖𝒓𝒆𝒎𝒂𝒙  𝑹𝒆𝒇 
(𝒎𝟐/𝒎𝟑) (℃) (𝒃𝒂𝒓) 
Shell-and-tube 
50 – 100 1100  1000  [57] 
heat exchanger 
Plate-and-frame 
120 – 660 815 200 [57] 
heat exchanger 
Plate-fin heat 
800 – 1500 800  200 [10, 11] 
exchanger 
Printed circuit 
200 – 2500 980  900  [10, 82] 
heat exchanger 
29 
 
 
2.6. Summary   
Cost effective high temperature heat exchangers are key to the success of emerging 
high-temperature, high-efficiency modular power cycles for diverse applications of 
energy conversion, power generation and energy/waste heat recovery applications. 
Most common heat exchangers currently available in the market are the plate-fin, 
plate-and-frame, and shell-and-tube types. The ideal high temperature heat 
exchanger would offer an optimum balance among heat transfer effectiveness, 
pressure drop, size and weight (indirectly controlling the cost) of the heat 
exchanger, while meeting longevity and reliability requirements. For elevated 
temperatures, most heat-resistant superalloys that can withstand the required 
temperature suffer from low thermal conductivity and high cost. Therefore, 
innovative design and manufacturing techniques are crucial to successful 
development of such heat exchangers. While additive manufacturing of superalloys 
for high temperature applications faces numerous challenges, it is believed that 
additional research support in this field can overcome many of the challenges. 
Significant progress has already been made in the past few years and demonstrates 
the potential of AM for fabrication of high temperature heat exchangers. In the next 
chapter, a high temperature gas-to-gas heat exchanger’s conceptual design, which 
incorporates the manifold-microchannel technology and has the potential to be 
fabricated through AM, will be discussed.  
  
30 
 
 
Chapter 3: Conceptual Design 
3.1.  Introduction  
This chapter discusses the design requirements for the high temperature gas-to-gas 
heat exchanger. The concepts behind the design of the metallic heat exchanger 
surfaces (manifold-microchannel technology) are also explained. For the heat 
exchanger design, a gas-to-gas cross-flow heat exchanger with manifold-
microchannel on both hot and cold sides was proposed. 
3.2. Design Requirements  
The design requirements were provided by an aerospace company for an aerospace 
turbofan pre-cooler. The current conventional design is based on a compact plate-
fin heat exchanger that incorporates the microchannel technology.  The detailed 
design specifications including overall volume, build material, mass flow rates, 
inlet temperatures, system pressures, and pressure drops on both hot- and cold-sides 
are shown in Table 3-1 and Table 3-2. 
Table 3-1 Physical specifications  
Overall Volume ≤ 125 in3 (0.002 m3) 
Cold-flow Length ≤ 5 in (0.127 m) 
No-flow Length ≤ 5 in (0.127 m) 
Weight ≤ 10 lb (4.54 kg) 
Material  Inconel 718 
 
31 
 
 
 Table 3-2 Operation specifications for full-scale heat exchanger  
Parameter Unit Nominal Condition 
Flow Rate lb/min (kg/s) 10 (0.076) 
Temperature °F (°C) 1200 (649) 
Hot Side 
Pressure psia (kPa) 64.7 (448) 
Pressure Drop psid (kPa) 2.6 (18) 
Flow Rate lb/min (kg/s) 10 (0.076) 
Temperature °F (°C) 100 (38) 
Cold Side 
Pressure psia (kPa) 14.7 (101) 
Pressure Drop psid (kPa) 0.3 (2) 
 
3.3. Manifold-microchannel Technology  
The manifold-microchannel technology was first proposed by Harpole and Eninger 
[13] in 1991. The main advantage of this technology is that a high heat thermal 
performance can be achieved without significant increase in pressure drop as 
reported in [84]. 
The manifold-microchannel concept involves a flow guiding system called 
manifold positioned over microchannels as shown in Figure 3-1. The purpose of 
the manifold is to distribute and deliver the flow into multiple short microchannels. 
As shown in the figure, the flow first enters through the manifold channels, then is 
distributed into the microchannels where travels a short length before it is guided 
out. Due to the short flow length in the microchannels, the flow is typically in the 
32 
 
 
thermally developing regions, resulting in substantially higher heat transfer 
coefficients. In addition, a short flow length in microchannels can also reduce 
pressure drop, which in turn, reduces the pumping power requirement. The study 
from Cetegen [84] shows that for the same heat transfer rate, pressure drop in a 
channel can be reduced by a factor of N2, where N is the number of channel 
segments. 
 
Figure 3-1: Manifold-microchannel concept 
By studying the behavior of the flow along the microchannel in a U-shape path 
when it enters and exits, Copeland et al. [85] created a 3-D CFD simulation for the 
microchannel section which is part of a manifold-microchannel heat sink as shown 
in Figure 3-2. The model used water as the working fluid and silicon as solid 
material for fins and base. It was also assumed uniform laminar flow distribution 
in all channels. With different combinations of manifold channel pitches, 
microchannel depth, microchannel widths, and channel flow velocity (inlet velocity 
ranging from 0.1 m/s to 1.0 m/s), pressure drops in the range of 47 – 12,260 Pa and 
thermal resistance of 0.25 – 1.85 ºC/W were observed. Another 3-D CFD 
33 
 
 
simulation work by Ng and Poh [86, 87] also considered only the microchannel 
section, and assumed uniform flow in all microchannels. By using water as the 
working fluid, they observed pressure drops in the range of 45 – 10,209 Pa and 
thermal resistance of 0.53 – 6.66 ºC/W with inlet velocity of 0.1 – 1.0 m/s. 
 
Figure 3-2: Microchannel computational domain by Copeland et al. [85] 
In another study, Kim et al. [88] conducted experimental tests on manifold-
microchannel for forced air cooling with a design consisting of manifold and 
multiple microchannels. The heat sink was made of silicon, and air was used as the 
working fluid. The experiment results revealed that compared to microchannel heat 
sink, 35% reduction on thermal resistance was possible with using manifold-
microchannel. 
A 3-D CFD simulation study of a manifold-microchannel heat sink was reported 
by Ryu et al. [89]. The model included a single manifold-microchannel segment 
with a portion of the inlet and outlet manifold channels as shown in Figure 3-3. 
34 
 
 
Copper was chosen as the material of the solid part, and water was used as the 
working fluid. The results showed that compared to traditional microchannel heat 
sinks, manifold-microchannel technique can reduce thermal resistance by half, and 
improve temperature uniformity along the channel. 
 
Figure 3-3: Single manifold-microchannel segment computational domain by Ryu 
et al. [89] 
A small manifold-microchannel heat sink was designed by Wang et al. [90], and a 
3-D CFD modeling was performed. The heat exchanger design had only 30 
microchannels, and copper was used as the solid material. By using water as the 
working fluid, the simulation results showed that manifold-microchannel design 
could deliver better temperature distribution compared to microchannel heat sink 
and an increase in heat transfer by 75% was obtained. 
35 
 
 
A combined numerical and experimental study on a manifold-microchannel heat 
sink was performed by Escher et al. [91]. Water was used as the working fluid and 
silicon was selected as the fin and base material. To simplify the numerical model, 
porous media was used to replace the microchannel section. Anisotropic 
permeability was applied to account for the resistance of the microchannels. The 
simulation results matched with the experimental results to some degree. The 
results also showed that a thermal resistance of 0.09 cm2K/W with a corresponding 
pressure drop of 0.22 bar was possible for a 2 x 2 cm2 system. 
Cetegen [84] created a multi-objective optimization for a manifold-microchannel 
heat sink design. The numerical model only simulated a single manifold-
microchannel segment, and copper was used as the solid material and water as the 
working fluid. The simulation results showed that for the same pumping power, the 
manifold-microchannel heat sink may provide 72% higher heat transfer coefficient 
than a traditional microchannel heat sink, or 306% higher heat transfer coefficient 
than a jet impingement heat sink. Cetegen also conducted several single-phase and 
multi-phase heat transfer experiments using a manifold-microchannel heat sink 
which showed significant performance improvement compared to other 
conventional heat sink technologies. 
Boteler et al. [92] carried out a numerical investigation on manifold-microchannel 
heat sink which included both the manifold and microchannel section. Different 
models with 2 to 20 microchannels were studied. A model with 10 microchannels 
is shown in Figure 3-4. The numerical results showed that the manifold-
36 
 
 
microchannel heat sink yielded a 97% reduction on pressure drop and could provide 
better temperature distribution compared to a conventional microchannel heat sink. 
 
Figure 3-4: Manifold-microchannel computational domain by Boteler et al. [93] 
Mandel et al. studied the implementation of manifold-microchannel for high heat 
flux electronic cooling applications [94]. Several manifold-microchannel heat sinks 
were fabricated and tested using R245fa as the refrigerant. The experimental results 
showed that high heat transfer coefficient of 40 kW/m2K-140 kW/m2K at pressure 
drops of 10 kPa-50 kPa could be achieved by forcing thin film evaporation on the 
microchannel in two-phase regime.  
Another study regarding manifold-microchannel’s implementation for low heat 
flux applications was performed by Andhare et al. [26]. A water-to-water manifold-
microchannel plate heat exchanger was successfully fabricated and tested. The 
experimental results showed that with a pressure drop per length value of 5.85 
37 
 
 
bar/m, an overall heat transfer coefficient close to 20 kW/m2K was achieved for a 
flow rate of 20 g/s.  
Arie [80] successfully developed a hybrid numerical method to predict the 
performance of air-to-water manifold-microchannel heat exchanger, and fabricated 
a heat exchanger where  manifold-microchannel concept was implemented on the 
air-side (the design concept is shown in Figure 3-5). The numerical and 
experimental results showed that manifold-microchannel plate heat exchanger 
could yield significant performance improvement compared to conventional plate 
heat exchangers. 
 
Figure 3-5: Air-to-water manifold-microchannel design concept by Arie [80] 
38 
 
 
Although many studies have been reported regarding the performance 
improvement using manifold-microchannel technology, they were mostly focused 
on heat sink design used for chip cooling application. Even one study tried to 
implement manifold-microchannels in the air-side of an air-to-water heat 
exchanger, the heat exchanger was only used for drying cooling application under 
low-temperature (~ 40ºC on air-side & ~ 67.5ºC on water-side) operation 
conditions.  
3.4. Gas-to-gas Cross-flow Heat Exchanger Conceptual Design  
Since manifold-microchannel technology has not been implemented in a gas-to-gas 
application such as air-to-air heat exchangers used on aircraft where weight 
reduction is as important as performance enhancement, the proposed high 
temperature gas-to-gas heat exchanger’s conceptual design consists of manifold-
microchannels on both hot and cold sides in a cross-flow configuration as shown in 
Figure 3-6 (a). As discussed in the previous section, manifold-microchannel 
technology involves adding system of manifolds on top of the microchannels to 
reduce the flow length along each microchannel. Pumping power can be then 
reduced since reduced flow length would result in reduced pressure drop across the 
heat exchanger. At the same time, increased heat transfer performance can be 
expected as the flow is still in the developing region due to the short flow length. 
Since manifold-microchannel can improve the heat transfer performance and 
reduce the pumping power, by applying manifold-microchannels on both gas sides 
for the conceptual design, a weight reduction compared to conventional heat 
exchanger with the same performance can be expected. The complete heat 
39 
 
 
exchanger may be composed of several layers stacked vertically as shown in Figure 
3-6 (b).  
 
(a) 
Manifolds Microchannel 
surfaces 
 
(b) 
Figure 3-6: Advanced gas-to-gas heat exchanger conceptual design (a); multi-layer 
manifold-microchannel heat exchanger concept (b) 
40 
 
 
3.5. Summary  
In this chapter the design requirements for the high temperature heat exchanger 
were presented based on the operating and physical specifications provided by an 
aerospace company for an aerospace turbofan pre-cooler prototype which is 
conventionally based on compact plate-fin heat exchangers that incorporate 
microchannel technology. In addition, the rationality behind adoption of the 
manifold-microchannel technology were explained. With reduced flow length, 
higher heat transfer coefficients and lower pressure drops can be achieved with the 
same flow rate compared to conventional heat transfer geometries such as plate-fin 
or microchannel technique. Considering the proposed gas-to-gas high temperature 
heat exchanger’s conceptual design consisted of manifold-microchannels on both 
the hot and cold side in a cross-flow configuration, the fabrication alternatives for 
such complex gas-to-gas high temperature heat exchangers will be discussed in the 
next chapter.   
41 
 
 
Chapter 4: Fabrication Alternatives  
4.1. Introduction  
This chapter discusses the fabrication alternatives for the gas-to-gas cross-flow 
manifold-microchannel heat exchanger such as direct metal laser sintering, photo-
chemical etching, laser micro-machining, and diffusion bonding. The fabrication 
principles and the limitations of the techniques are discussed in detail. The 
presented information in this chapter aims to provide the rationality for adopting 
additive manufacturing as the fabrication method of the choice.  
4.2. Direct Metal Laser Sintering  
Direct metal laser sintering (DMLS) is a metal powder based additive 
manufacturing process.  A computer-controlled laser is used to sinter the metal 
powder based on a pre-programmed geometry (usually a STL file of the CAD 
model) to build a 3-D structure layer by layer. DMLS was first developed by EOS 
GmbH of Munich, Germany [95], and now is one of the most common additive 
manufacturing methods for metallic structures in United States. A DMLS machine 
generally consists of two platforms: a powder dispenser platform to house the metal 
powder, and a build platform on which the 3-D structure is being fabricated. When 
printing of a layer of the 3-D structure is completed, the powder dispenser platform 
will rise, and the build platform will lower down at the same time. By doing so, a 
new layer of metal powder can be evenly distributed on top of the previous layer 
through a re-coater arm. Based on the pre-programmed geometry, a scanning mirror 
42 
 
 
controls the laser to sinter the powder and build a cross section of 3-D structure on 
the new layer. The process is repeated until the entire 3-D structure is fabricated. 
Since it emerged as a new manufacturing technique, DMLS process has been 
improved significantly over the past decade. The minimum feature size that can be 
currently fabricated through DMLS is about 200 – 300 µm with a normal-resolution 
machine or 150 µm with a high-resolution machine. A list of the fabrication 
limitations and tolerances is given in Table 4-1 based on three top DMLS-service-
provider companies in United States (Proto Labs, Stratasys, and 3D Systems). 
Considering DMLS’s good compatibility with common heat exchanger materials 
(such as Aluminum, Stainless Steel, and Inconel) and its capability to fabricate 
complex geometry which may be difficult or impossible through conventional 
manufacturing technique, the fabrication method shows promise to be implemented 
for both manifold and microchannel surface’s fabrication.  
Based on the communications with all three DMLS-service-provider companies 
(Proto Labs, Stratasys, and 3D Systems), it was concluded that to avoid any 
warpage on the manifold wall (shown in Figure 4-1) and also thermal short circuit 
between incoming and outgoing flows, the minimum manifold wall thickness needs 
to be 500 µm. A minimum manifold base thickness of 1 mm is also needed if post-
process will be applied to meet certain flatness requirement for assembly.  
Moreover, the minimum microchannel fin thickness needs to be 150 µm due to the 
minimum feature size constraint. 
 
43 
 
 
Table 4-1 DMLS limitations and tolerances  
 Proto Labs [96] Stratasys [97] 3D Systems [98] 
Normal resolution:  
Layer  30 µm Normal resolution:  
N/A 
thickness High resolution:  40 µm 
20 µm 
Normal resolution:  
Minimum 380 µm Normal resolution:  Normal resolution:  
feature size High resolution:  300 µm 200 µm 
150 µm 
First inch: ± 127 µm 
Tolerance ± 76 µm ± 50 µm 
Thereafter: ± 50 µm 
Normal resolution:  
Maximum  246 × 246 × 274 mm3 Normal resolution:  
N/A 
build size High resolution:  275 × 275 × 420 mm3 
89 × 89 × 74 mm3 
 
𝑊𝑎𝑙𝑙𝑚𝑎𝑛𝑖𝑓𝑜𝑙𝑑 
𝐵𝑎𝑠𝑒𝑚𝑎𝑛𝑖𝑓𝑜𝑙𝑑 
 
Figure 4-1: Manifold design layout 
44 
 
 
4.3. Photo-chemical Etching  
Photo-chemical etching is a fabrication process which utilizes a photoresist mask 
and a liquid etchant to remove certain regions of a metal part. The process was 
initially developed for fabrication of printed circuit boards. However, due to its 
ability to etch various metals, the process is also commonly used to create 
microchannels in metal surfaces for printed circuit heat exchangers. As shown in 
Figure 4-2, during photo-chemical etching process, first the shape of the part is 
printed onto a photographic film. Then, photo-tool which is a sheet of photographic 
film showing negative images of the parts is prepared. Next, the metal sheet is 
laminated with a UV-sensitive photoresist. The coated metal is placed under the 
photo-tool.  The plate is then exposed to UV light. After the plate is "developed", 
the unexposed resist is washed away. During etching, etchant reacts with the 
unprotected metal, and essentially corrodes it away. After neutralizing and rinsing, 
the remaining resist is removed, and the finished metal sheet is cleaned and dried. 
Photo-chemical etching is a process by which a concentrated acid is used to remove 
material to shape desired structure. However, since etching on metal materials is 
homogeneous, it is hard to fabricate microchannels with rectangular cross sections.  
45 
 
 
 
Figure 4-2: Photo-chemical etching process (adapted from reference [99]) 
Based on the communications with the photo-chemical etching company 
(ChemTec - A Division of ChemArt), to fabricate a double-side cross-flow 
microchannel surface through the etching process, the aspect ratio of the 
microchannel width to the microchannel height would be 1.1. The minimum 
microchannel fin thickness and the minimum microchannel width would be 80 µm.  
The total microchannel surface height, as shown in Figure 4-3, is greater than 0.013 
mm and less than 2.032 mm. However, considerable effort will be needed if the 
etched double-side microchannel surface has to meet certain flatness requirement 
for assembly through post processing. 
46 
 
 
𝑇ℎ𝑖𝑐𝑘𝑛𝑒𝑠𝑠𝑓𝑖𝑛 
𝐻𝑒𝑖𝑔ℎ𝑡 𝑊𝑖𝑑𝑡ℎ 𝑐ℎ𝑛 𝑐ℎ𝑛
𝑇𝑜𝑡𝑎𝑙 𝑚𝑖𝑐𝑟𝑜𝑐ℎ𝑎𝑛𝑛𝑒𝑙 
𝑠𝑢𝑟𝑓𝑎𝑐𝑒 ℎ𝑒𝑖𝑔ℎ𝑡 
 
Figure 4-3: Microchannel surface design layout 
4.4. Laser Micro-machining  
Laser micro-machining is another approach to fabricate the double-side cross-flow 
microchannel surface. By applying a high-power laser beam, energy of the laser is 
converted to heat which vaporizes or melts the metal material to form the 
microchannel structure as shown in Figure 4-4 [100]. An assist gas or fluid is often 
utilized to improve the cut quality, expel molten metal, and minimize the heat affect 
zone. However, laser micro-machining is a very time-consuming process as it may 
take over 2 days to fabricate one 389 × 132 × 1.3 mm3 size microchannel surface. 
Also, the maximum channel height would be 500 µm according to the 
communications with the laser micro-machining company (Potomac Laser). 
47 
 
 
 
Figure 4-4: Laser micro-machining process concept [101] 
4.5. Diffusion Bonding   
As shown in Figure 4-5, diffusion bonding is a solid-state joining process by which 
two nominally flat surfaces are joined at an elevated temperature using an applied 
interfacial pressure for a period ranging from a few minutes to a few hours [102]. 
If manifold and microchannel surface are fabricated separately, diffusion bonding 
would be the best approach to assemble the parts. However, a thin bonding aid layer 
such as Ni-layer may be required for certain materials which may add additional 
thermal interfacial resistance to the assembled core. Due to the small size of the 
microchannels, improper bonding between the manifold and microchannels can 
48 
 
 
also cause clogging in the microchannels which can reduce the heat transfer 
performance and increase the pressure drop significantly. If diffusion bonding is 
used, in order to meet the parallel and flatness requirements for the bonding 
surfaces, considerable effort is needed to prepare the microchannel surfaces and 
manifolds.   
 
Figure 4-5: Diffusion bonding process concept [103] 
Based on the communication with the diffusion bonding company (Vacuum 
Process Engineering), in order to ensure the bonding quality, the bonding surface’s 
parallel needs to be 0.03 mm or better, and the surface’s flatness needs to be 0.127 
mm or better. Furthermore, microchannel surface needs to add 1 mm thick side bars 
on all 4 sides, and manifold also needs to add 1 mm thick side bars as shown in  
Figure 4-6 to provide enough bonding surface area. 
49 
 
 
Cold side manifold’s side 
bar 
Hot side manifold’s side 
Microchannel 
bar 
surface’s side bar 
 
Figure 4-6: Exploded view of assembly design for one stack of cross-flow 
manifold-microchannel through diffusion bonding 
4.6. Fabrication Alternatives Risk Assessment  
As shown in Table 4-2, in order to choose the best fabrication approach for the gas-
to-gas cross-flow manifold-microchannel heat exchanger, the limitations of the 
fabrication alternatives are summarised, and the risks of each fabrication alternative 
are assessed. Based on the risk assessment, it was concluded that if the heat 
exchanger is 3-D printed as a single unit, the assembly process can be avoided and 
the fabricating lead time can also be significantly shortened. 
  
50 
 
 
Table 4-2 Fabrication alternatives risk assessment  
 
Process Limitations Risk Level 
• Metal powder may be remained in the microchannels due to small 
channel width and sintering heat from the fins 
DMLS Moderate 
• Minimum fin thickness is more than 150 µm 
• Minimum separating sheet thickness is about 500 µm 
• Microchannel has a tolerance of 80 µm or more 
Microchannel  • Microchannel surface’s flatness after etching is generally 10 mm, but 
Surface Photochemical etching Moderate flatness of 0.127 mm is possible with post process 
• Cannot get rectangular cross section for microchannel  
• Cost is very expensive  
Laser micromachining • Very slow process which may take over two days to fabricate a single Moderate 
389 × 132 × 1.3 mm3 microchannel surface 
• Probably going to be some warp on the manifold base (if base thickness 
is 300 µm) due to the heat on the thin wall structure 
Manifold DMLS Moderate 
• Manifold base can be made thicker and be machined to meet the flatness 
requirement  
• Parallel and flatness required for both manifold base and microchannel 
Diffusion bonding surface are 0.03 mm and 0.127 mm respectively High 
Assembly • May cause clogging in the microchannels 
• If both manifold and microchannel surface are fabricated through DMLS, 
DMLS Low 
then it can be 3D printed as a single unit without assembly process 
51 
 
 
4.7. Summary  
This chapter discusses several fabrication options for the gas-to-gas cross-flow 
manifold-microchannel heat exchanger including direct metal laser sintering, 
photo-chemical etching, laser micro-machining, and diffusion bonding. The 
procedure for each technique is described and the associated limitations technique 
are discussed. A risk analysis was performed which indicated the additive 
manufacturing would be a better choice for the current design.  In the next chapter, 
the optimization process which was performed to obtain the optimal heat exchanger 
design is discussed, and the identified manufacturing constraints are included to 
predict the performance of the heat exchangers produced through different 
fabrication methods.  
52 
 
 
Chapter 5: Optimization & Design Selection  
5.1. Introduction  
This chapter discusses about a single objective optimization process performed for 
the gas-to-gas cross-flow manifold-microchannel heat exchanger to find the 
optimal design which meets the design requirements as listed in chapter 3. The first 
goal of this optimization was to compare the predicted performance of the 
manifold-microchannel heat exchangers produced through different fabrication 
approaches. The second goal was for design selection. An optimal design, which 
meets all the design requirements and shows the best performance with the chosen 
fabrication approach, was identified and scaled down to a lab-scale unit so that it 
could be fabricated and tested.  
5.2. Single Objective Optimization   
A single objective optimization was performed to calculate the highest heat transfer 
performance of the manifold-microchannel heat exchanger for the given physical 
and operating conditions listed in chapter 3. The goal of the optimization is to 
maximize the heat duty (which is the same as to maximize the effectiveness) of the 
cross-flow heat exchanger. Based on the fabrication alternatives discussed in the 
previous chapter, the optimization was performed for three manufacturing process 
constraints: 3-D metal printing (DMLS), photo-chemical etching, and laser micro-
machining. 
53 
 
 
The geometrical and performance constraints for all three manufacturing methods 
are shown in Table 5-1 for both hot and cold sides. The geometrical constraints 
were based on the manufacturing limitations as recommended by the 
manufacturers. These manufacturing limitations were also discussed in the previous 
chapter. 
Table 5-1 Optimization constraints  
 3-D Metal Photo-chemical Laser Micro-
Printing Etching machining 
Geometrical Constraints (hot and cold sides) 
𝑅𝑒𝑚𝑛𝑑 100-16,000 
𝐻𝑐h𝑛/𝐿𝑓𝑟𝑜𝑛𝑡 0.00236-0.0394 𝑊𝑐h𝑛/1.1 ≤0.00394 
𝑊𝑐h𝑛/𝑡𝑓𝑖𝑛 0.5-5 
𝑊𝑐h𝑛/𝐿𝑓𝑟𝑜𝑛𝑡 0.00157-0.00787 
𝑊𝑚𝑛𝑑−𝑐h𝑛/𝐿𝑓𝑟𝑜𝑛𝑡 0.0315-0.0787 
𝐻𝑚𝑛𝑑,𝑡𝑜𝑝/𝐿𝑓𝑟𝑜𝑛𝑡 0.00787-0.0787  
𝑛 30-1000 
𝑊𝑖𝑛/𝑊𝑚𝑛𝑑−𝑐h𝑛 0.1-0.9 
𝑡𝑓𝑖𝑛/𝐿𝑓𝑟𝑜𝑛𝑡 ≥0.00118 ≥0.000630 ≥0.000394  
𝐻𝑚𝑛𝑑,𝑏𝑜𝑡𝑡𝑜𝑚/𝐿𝑓𝑟𝑜𝑛𝑡 0.00236 
𝑊𝑚𝑛𝑑/𝐿𝑓𝑟𝑜𝑛𝑡 0.00394 
𝐻𝑏𝑎𝑠𝑒/𝐿𝑓𝑟𝑜𝑛𝑡 0.00236 
(𝐻𝑐ℎ𝑛,ℎ𝑜𝑡
+ 𝐻𝑐ℎ𝑛,𝑐𝑜𝑙𝑑 N/A 0.000102-0.016 N/A 
+ 𝐻𝑏𝑎𝑠𝑒)/𝐿𝑓𝑟𝑜𝑛𝑡 
Performance Constraints 
𝑇𝑖𝑛,ℎ𝑜𝑡 38 C 
𝑇𝑖𝑛,𝑐𝑜𝑙𝑑 649C 
∆𝑝ℎ𝑜𝑡 ≤17.9 kPa 
∆𝑝𝑐𝑜𝑙𝑑 ≤2.07 kPa 
𝑅𝑒𝑓𝑟𝑜𝑛𝑡,ℎ𝑜𝑡 18,000 
𝑅𝑒𝑓𝑟𝑜𝑛𝑡,𝑐𝑜𝑙𝑑 18,000 
𝑉𝑡𝑜𝑡 ≤0.00205 m
3 
 
54 
 
 
 
Figure 5-1: Manifold-microchannel design layout 
As shown in Figure 5-1, the geometrical parameters are defined as manifold wall 
thickness (𝑊𝑚𝑛𝑑), manifold channel width (𝑊𝑚𝑛𝑑,𝑐ℎ𝑛), manifold channel height 
(𝐻𝑚𝑛𝑑,𝑡𝑜𝑝), manifold base thickness (𝐻𝑚𝑛𝑑,𝑏𝑜𝑡𝑡𝑜𝑚), manifold-to-microchannel inlet 
width (𝑊𝑖𝑛, note: this is the width of the opening where either the flow enters from 
the manifold channel to the  microchannels or the flow exits from the microchannels 
to the manifold channel), microchannel fin thickness (𝑡𝑓𝑖𝑛), microchannel width 
(𝑊𝑐ℎ𝑛), microchannel height (𝐻𝑐ℎ𝑛), microchannel base thickness (𝐻𝑏𝑎𝑠𝑒, note: this 
is only for one side, which means the separating sheet thickness is 2𝐻𝑏𝑎𝑠𝑒), and 
microchannel numbers on one side (n). The geometrical constraints were 
normalized by the frontal area length scale (𝐿𝑓𝑟𝑜𝑛𝑡) of 0.127 m. The frontal area 
3
length scale was calculated as 𝐿𝑓𝑟𝑜𝑛𝑡 =  √𝑉𝑡𝑜𝑡  based on the maximum possible 
volume of the heat exchanger. The maximum heat exchanger volume (𝑉𝑡𝑜𝑡 ) is 
55 
 
 
0.00205 m3 which was listed as the heat exchanger physical requirement in chapter 
3. In addition to the geometrical constraints, the performance constraints based on 
the operating conditions were also included in Table 5-1, where 𝑅𝑒𝑓𝑟𝑜𝑛𝑡 is Reynold 
number calculated based on the frontal area and the mass flow rate as: 
𝑅𝑒𝑓𝑟𝑜𝑛𝑡 = 𝐿𝑓𝑟𝑜𝑛𝑡?̇?𝑓𝑟𝑜𝑛𝑡/(𝐴𝑓𝑟𝑜𝑛𝑡𝑎𝑙𝜇) (1) 
where 𝐴 = 𝐿 2𝑓𝑟𝑜𝑛𝑡𝑎𝑙 𝑓𝑟𝑜𝑛𝑡 , and 𝜇 is the dynamic viscosity based on the average of 
hot and cold inlet temperatures. 
Approximation-assisted optimization was employed to calculate the maximum heat 
capacity and the dimensions that meet all the constraints. The main advantage of 
such method is that it can reduce computational time significantly by using 
metamodel to predict the performance of the heat exchanger using the known 
information of the system. The optimization process was divided into 4 stages. The 
first stage is Design of Experiment (DOE) where initial sampling was performed 
using space filling method developed by Aute et al. [104]. The performance of the 
sampling points was then evaluated individually. The second stage is creating a 
metamodel using sampling points as input. Dace, a Kriging-based metamodel 
toolbox developed by Lophaven et al. [105], was used for this purpose.  The third 
stage is the optimization process which utilized genetic algorithm method. The 
basic principle of genetic algorithm is to use concept like natural selection to 
eliminate inferior solutions and retain high performance solutions. The last stage is 
to validate the metamodel by calculating the performance of the heat exchanger 
using the modified hybrid method developed by Arie [80] and compare it to the one 
56 
 
 
predicted using metamodel. If the error between the prediction and actual values is 
small, the process concludes. Otherwise, the metamodel is recreated by using the 
validation points as additional sampling points. The process is repeated until 
sufficiently low metamodel error is obtained, and the process flow chart is shown 
in Figure 5-2. 
 
Figure 5-2: Single objective optimization flow chart 
5.3. Modified Hybrid Method  
The pressure drop and heat transfer performance of the manifold-microchannel heat 
exchanger were evaluated using modified hybrid method developed by Arie et al. 
[80]. This method involves using a single manifold-microchannel CFD simulation 
(the model is shown in Figure 5-3) to calculate the heat transfer coefficient and 
57 
 
 
pressure drop in the microchannel. The manifold pressure drop, on the other hand, 
was calculated by solving a 1-D differential equation of mass and momentum 
balance in the manifold channel. The detailed description of the method can be 
found in Ref. [106]. 
 
Figure 5-3: Computational domain and boundary conditions of a single manifold-
microchannel by Arie [106]: (1) manifold channel inlet surface with mass flow inlet 
and constant temperature, (2) manifold channel outlet surface with constant 
pressure outlet, (3) microchannel base surface with constant surface temperature, 
(4) symmetry planes  
 
58 
 
 
The combined heat capacity (𝑄𝑡𝑜𝑡) of the cross-flow heat exchanger was evaluated 
by first calculating the heat capacity in each stack(𝑄1 𝑠𝑡𝑎𝑐𝑘). To do the calculation, 
first a single stack heat exchanger needs to be divided into 𝑁ℎ𝑜𝑡 sections on the hot 
flow direction and 𝑁𝑐𝑜𝑙𝑑 sections on the cold flow direction, each consisting of a 
single manifold-microchannel pass. An example is shown in Figure 5-4 for a single 
stack cross-flow manifold-microchannel heat exchanger with 2 passes on both the 
hot and cold sides (𝑁ℎ𝑜𝑡 = 2 and 𝑁𝑐𝑜𝑙𝑑 = 2), which is subdivided into 4 components 
with a single manifold-microchannel pass of each.  
 
(a) 
59 
 
 
 
(b) 
Figure 5-4: CAD model with 2 passes on both hot and cold sides by Arie [50]: 
combined model (a), subdivided model (b)  
The heat duty for each component can then be calculated by performing an energy 
balance on that section. The detailed steps for 𝑄𝑡𝑜𝑡 calculation are as follows: 
1. Set the temperature boundary condition: 
𝑇ℎ𝑜𝑡,𝑖𝑛,(𝑖,1) = 𝑇ℎ𝑜𝑡,𝑖𝑛  for 𝑖 = 1 to 𝑁𝑐𝑜𝑙𝑑 
Tcold,in,(1,j) = Tcold,in for j = 1 to Nhot 
2. Set i=1 and j=1 
60 
 
 
3. Perform an energy balance on the section i, j by solving Eqs. (2) – (5)  to solve 
for 4 unknown variables: heat capacity Qi,j, cold-side exit temperature Tcold,out,(i,j), 
hot-side exit temperature Thot,out,(j,i), and base temperature 𝑇b,(i,j), where hhot  and 
hcold are the base conductance obtained from the CFD simulations. Abase is the 
total base area of the heat exchanger evaluated as: Abase = Ltot,hot/× Ltot,cold. 
ṁ
     Q hot (2) i,j = C (T − TN p,hot hot,in,(i,j) hot,out,(i,j)
) 
cold
ṁ
     Q = cold C (3) i,j N p,cold
(Tcold,in,(i,j) − Tcold,out,(i,j)) 
hot
A
     Q = h basei,j hot (Thot,in,(i,j) − T ) (4) N N b,(i,j)hot cold
A
     Q basei,j = hcold (Tb,(i,j) − T (5) N N cold,in,(i,j)
) 
hot cold
4. Set Thot,in,(i,j+1) = Thot,out,(i,j) 
and Tcold,in,(i+1,j) = Tcold,out,(i,j) 
5. Update i value: i = i + 1 
6. Repeat step 3-5 until i = N_cold 
7. Update 𝑗 value: j = j + 1 and re-set i=1 
8. Repeat step 3-7 until j = N_hot 
9. Calculate the total capacity of each stack as:  
Nhot Ncold
𝑄1 𝑠𝑡𝑎𝑐𝑘 = ∑ ∑ Qi,j 
j=1 i=1
61 
 
 
10. Calculate the total capacity of the heat exchanger as: Qtot = NstackQ1 stack 
where Nstack is total stack number of the heat exchanger. 
Two main assumptions were considered for the modified hybrid method. One is 
that the flow distribution among the microchannels is uniform, which means the 
mass flow rate in each microchannel is the same. The second assumption is for each 
computational domain (i, j) the base temperature remains constant. This is a fair 
assumption because for the case of single manifold-microchannel pass, the same 
inlet temperature in all microchannels is expected as the flow comes from the top 
of the microchannels through the manifold channel, and heat transfer only occurs 
in the microchannel section (no heat transfer occurs in the manifold channel). 
Because both the hot and cold flows have the same inlet temperatures, and uniform 
flow is fed across microchannels, each layer of the heat exchanger can be divided 
into hundreds of repeated unit cells which have the same hot-inlet and cold-inlet 
temperatures as shown in Figure 5-5. For each unit cell, due to the short flow length, 
a constant base temperature can be assumed. 
 
62 
 
 
 
Figure 5-5: Unit-cell pattern from cross-flow manifold-microchannel heat 
exchanger [107] 
5.4. Optimization Results & Design Selection   
The optimization results of effectiveness (ε) and heat exchanger solid volume ratio 
(Vsolid/Vtot) based on different manufacturing constraints are shown in Table 5-2. 
By comparing the results, the heat exchanger with microchannel surface and 
manifold both fabricated by 3-D metal printing (DMLS) provides the highest 
effectiveness and lowest heat exchanger solid volume ratio. 18% effectiveness 
improvement and 34% solid volume ratio reduction are possible compared to the 
heat exchanger manufactured by photo-chemical etching, and 8% effectiveness 
improvement and 17% solid volume ratio reduction are possible compared to a heat 
exchanger manufactured by laser micro-machining. Since the build material for all 
three heat exchangers is assumed to be the same as Inconel 718, and the overall 
63 
 
 
volume of all three design is the same as 0.002 m3, a lower solid volume ratio also 
implies a lower mass of the heat exchanger design.  
Table 5-2 Optimization results 
Microchannel Surface  Manifold  
Optimized Results 
Manufacturing Process Manufacturing Process 
3-D Metal Printing 3-D Metal Printing ε = 0.569 
(DMLS) (DMLS) Vsolid/Vtot = 0.164 
3-D Metal Printing ε = 0.481 
Photo-chemical Etching 
(DMLS) Vsolid/Vtot = 0.249 
3-D Metal Printing ε = 0.528 
Laser Micro-machining 
(DMLS) Vsolid/Vtot = 0.197 
Besides higher effectiveness and lower heat exchanger solid volume ratio, there are 
also other benefits for heat exchanger manufactured by 3-D metal printing. First, 3-
D printing the entire heat exchanger core as a single piece eliminates the necessity 
of subsequent steps. Secondly, 3-D metal printing technique is quite efficient 
thereby saves fabrication time and reduces material wastage as compared to other 
techniques. Thirdly, additive manufacturing is highly effective in fabricating 
extremely complicated and compact geometries, which can aid in developing 
minimum feature size that will potentially yield higher performance using a variety 
of materials as compared to any of the conventional fabricating techniques. 
However, if 3-D metal printing (DMLS) is chosen as the fabrication approach based 
on the optimization results, the overhang issue must be considered. During the 3-D 
printing process, if the next layer is larger than the previous layer, an overhang issue 
will occur and cause build crash on the part as shown in Figure 5-6. DMLS has a 
64 
 
 
small allowance for unsupported overhangs (0.5mm), and structure with self-
supporting angle (≥ 45°) can avoid overhangs as well. 
 
Figure 5-6: Build crash caused by overhangs [108] 
For the cross-flow manifold-microchannel heat exchanger, depending on the 
printing orientation, the overhang issue will either occur at the manifold base and 
microchannel fins as shown in Figure 5-7 (a), or at the manifold walls and 
microchannel fins as shown in Figure 5-7 (b). Based on the communication with 
the manufacturer, it would be very challenging to print the part if the design was 
like the first case. And for the second case, if inclined fins and inclined manifold 
walls (which will be discussed in detail in the next chapter) can be implemented 
into the design, the printing of the part could still be realized.  
65 
 
 
Printing 
orientation 
Overhang 
 
(a) 
Printing 
orientation 
Overhang 
 
(b) 
Figure 5-7: Overhang issue with cross-flow manifold-microchannel design: 
overhangs at the manifold base and microchannel fins (a); overhangs at the 
manifold walls and microchannel fins (b) 
66 
 
 
Since heat does not significantly transfer along the manifold channels, and the 
pressure drop across these channels will not be affected much if the manifold wall 
is modified from vertical wall to inclined wall, another optimization which added 
the effect of the inclined microchannel fins was performed. The updated optimized 
result from the 3-D metal printing process was again compared with the other two 
manufacturing approaches as shown in Table 5-3. The results showed that by 
implementing the inclined microchannel fins, the heat duty of the heat exchanger 
using DMLS process was reduced less than the one using laser micro-machining. 
However, the optimal design based on 3-D metal printing still has the lowest weight 
compared to the other two. For the design using photo-chemical etching, the reason 
that it has more layers might because the aspect ratio of microchannel width to 
height was fixed. For the design using laser micro-machining, since the 
microchannel height could not be more than 0.5 mm, more fins were needed to 
enlarge the heat transfer area. Therefore, it also has the largest base area (389 x 132 
mm2). As the weight of the manifold-microchannel heat exchanger design is mainly 
from the microchannel surfaces, more layers or larger base area will result in 
significant increase of the heat exchanger’s total weight. Considering the higher 
heat transfer density (Q/m) which means less weight as an important aspect for the 
aerospace applications, and the elimination of assembly process, the heat exchanger 
design based on the 3-D metal printing (DMLS) approach was finally chosen. 
 
 
67 
 
 
Table 5-3 Summary of fabrication process and updated optimized results 
Microchannel Manifold Assembly Main Constraint Optimized Results 
Q = 23.8 kW  
3-D Metal  3-D Metal  m = 3.11 kg 
Fin thickness  
Printing Printing None Q/m = 7.65 kW/kg 
≥ 0.15 mm 
(DMLS) (DMLS) 368 x 115 x 48 mm3 
(Vtot = 0.00205 m3) 
Q = 23.4 kW  
Photo- 3-D Metal  Channel width  m = 4.29 kg 
Diffusion 
chemical Printing Bonding = 1.1 x channel 
Q/m = 5.46 kW/kg 
Etching (DMLS) height 341 x 113 x 52 mm3 
(V  = 0.00200 m3tot ) 
Q = 25.6 kW  
3-D Metal  m = 3.62 kg 
Laser Micro- Diffusion Channel height 
Printing Q/m = 7.08 kW/kg machining Bonding ≤ 0.5 mm 
(DMLS) 389 x 132 x 40 mm3 
(Vtot = 0.00205 m3) 
5.5. Scaled-down Model   
In order to fabricate and test the optimized heat exchanger design, the heat 
exchanger had to be scaled down considering both the manufacturer’s fabrication 
limitation and the testing capability in the lab. The dimensions of the subscale unit 
were determined by downscaling the full-scale 23.8 kW gas-to-gas cross-flow 
manifold-microchannel heat exchanger design which was obtained through the 
optimization process in the previous section. This downscaling process aims to 
reduce the overall volume of the full-scale heat exchanger to match the selected 
manufacturer’s fabrication capability while keeping both hot and cold side 
microchannels’ geometrical parameters (microchannel fin thickness, microchannel 
width, and microchannel height) unchanged. 
68 
 
 
To ensure the subscale unit’s overall volume was less than the manufacturer’s 
maximum building volume, a new optimization was performed. The effort involved 
additional geometrical constraints to ensure that hot side flow length (L) was less 
than 66 mm, cold side flow length (W) was less than 76 mm, and no-flow length 
(H) was less than 76 mm (see Figure 5-8). To keep similar performance 
characteristics, mass flow rate per microchannel on both hot/cold sides and COP 
on both hot/cold sides were kept the same as the full-scale design. The comparison 
between full-scale design and subscale unit is shown in Table 5-4. As can be seen, 
the subscale unit is about 1/10 of the full-scale unit in terms of heat transfer 
capacity.  
 
Figure 5-8: Subscale unit design (the yellow part was designed as connection 
interface for headers) 
69 
 
 
Table 5-4 Comparison between full-scale design and subscale unit  
 Full-scale Design Subscale Unit 
Capacity (𝑸) 23.8 kW 2.0 kW 
Hot flow length (𝑳) 368 mm 59 mm 
Cold flow length (𝑾) 115 mm 70 mm 
No-flow length (𝑯) 48 mm 26 mm 
Number of Layers 4 2 
5.6. Summary  
In summary, single-objective optimization for the gas-to-gas cross-flow manifold-
microchannel heat exchanger with different fabrication approaches was described 
in this chapter. The optimization constraints and process were discussed in detail. 
The optimized results showed that even by implementing inclined microchannel 
fins to avoid the overhang issue, the heat exchanger design using 3-D metal printing 
(DMLS) process could still provide the highest heat transfer density (Q/m) and 
eliminate the assembly process to significantly reduce the fabrication lead time. 
Lastly, the optimized full-scale gas-to-gas cross-flow manifold-microchannel heat 
exchanger was scaled down for fabrication and experimental testing. 
  
70 
 
 
Chapter 6: Test Print Coupon & Subscale Unit Fabrication  
6.1. Introduction  
This chapter first discusses the test print coupons which were fabricated to evaluate 
the build quality using the selected DMLS process. The coupons’ geometry was 
based on the scaled-down design proposed in chapter 5 with some design 
modifications according to the recommendations from the manufacturer (Proto 
Labs). After fabrication, the test print units were analyzed to evaluate the 
fabrication quality such as the printable minimum microchannel fin thickness and 
the minimum microchannel base thickness (which can also be called as separating 
sheet) that can prevent leakage from hot to cold side (as the design requirement has 
a higher system pressure on the hot side). Based on the test print units’ evaluation, 
the subscale unit’s geometry was updated to ensure the best fabrication quality, and 
the unit was successfully 3-D printed which matched very well to the final CAD 
design.    
6.2. Test Print Coupon Design    
Based on the communication with the manufacturer, DMLS only has a small 
allowance of 0.5 mm for unsupported structure [108]. Considering that the 
microchannel height on both sides of the design would be greater than 1 mm, either 
the hot side or cold side microchannel fins were suggested by the manufacturer to 
be printed at an inclined angle of 45 degrees to avoid overhangs during the 
fabrication process as shown in Figure 6-1. Based on the optimization study 
71 
 
 
discussed in the previous chapter, it was found that the heat exchanger’s heat duty 
would be higher if inclined microchannel fins were implemented on the hot side.  
Overhang Inclined fins 
Printing 
orientation 
              
Figure 6-1: Cross-flow microchannel design layout: both side vertical fins (left); 
inclined fins on one side (right) 
Since it was decided that hot side microchannel fins need to be inclined, to avoid 
overhang issue on the cold side, the manifold wall on cold side also had to be 
modified from vertical wall to inclined wall as shown in Figure 6-2 (a). With 
inclined microchannel fins on hot side and inclined manifold walls on cold side, a 
20 × 16 × 9 mm3 test print coupon design based on the scaled-down design’s 
geometry dimensions was created as shown in  Figure 6-2 (b). 
72 
 
 
 
(a) 
20  
mm 
Printing  
orientation 
16 mm 9 mm 
 
(b) 
Figure 6-2: Zoom-in section view of cross-flow manifold-microchannel design 
with inclined microchannel fins on hot side and inclined manifold walls on cold-
side (a); CAD design of the test print coupon (b) 
 
73 
 
 
Based on the initial scaled-down design obtained from the optimization discussed 
in the previous chapter, two test print coupon designs were prepared to check the 
manufacturer’s minimum printable feature size such as the minimum microchannel 
fin thickness and manifold wall thickness. As shown in Table 6-1, coupon 1 has 
0.200 mm vertical fins and 0.172 mm inclined fins as recommended by the 
manufacturer, and coupon 2 has the same dimensions as the initial scaled-down 
design except thinner inclined fins and manifold walls on the hot side. The same 
printing orientation were applied for both coupons as shown in Figure 6-2 (b). 
Table 6-1 Test print coupon designs’ minimum feature size dimensions  
 Coupon 1 Coupon 2 
 Hot side Cold side Hot side Cold side 
 Microchannel fin thickness (tfin) 0.172 mm 0.200 mm 0.106 mm 0.150 mm 
 Microchannel channel width (Wchn) 0.280 mm 0.280 mm 0.346 mm 0.330 mm 
 Manifold wall thickness (Wallmnd) 0.500 mm 0.500 mm 0.300 mm 0.500  
 
6.3. Test Print Coupon Fabrication & Evaluation   
Both test print coupons were successfully fabricated using DMLS by the 
manufacturer (Proto Labs) with Inconel 718 as shown in Figure 6-3. According to 
the weight measurement, both coupons showed very good agreement with the 
design value (< 2.3%) based on the material density provided by the manufacturer 
and the solid volume obtained from the CAD design.  
74 
 
 
 
Figure 6-3: 3-D printed 20 x 16 x 9 mm3 Inconel 718 test print coupon with inclined 
fins and inclined manifold walls  
By measuring the small feature dimensions with a digital microscope, the 
measurements showed a good match as shown in Table 6-2. However, it was found 
that the fabricated minimum inclined fin thickness was 180 µm (± 12 µm) even if 
the CAD design value was less than it. And if the designed inclined fin thickness 
was less than 100 µm, this feature would not be printed according to the 
75 
 
 
manufacturer. For the vertical fins, it was concluded that the minimum printable 
thickness was 165 µm (± 7 µm). 
Table 6-2 Test print coupons measurement 
 Coupon 1 Coupon 2 
Avg. Avg. 
 Design Design Measure diff % Measure diff % 
(mm) (mm) 
(mm) (mm) 
Hot side 
20.75 20.93 0.9% 20.50 20.66 0.8% 
length (L) 
Cold side 
15.80 15.76 -0.3% 15.80 15.79 -0.1% 
length (W) 
No-flow 
8.93 8.84 -1.0% 8.93 8.67 -2.9% 
length (H) 
Hman_cold 3.82 3.79 -0.8% 3.82 3.8 -0.5% 
Hman_hot 2.80 2.78 -0.8% 2.80 2.74 -2.1% 
Wallman_cold 0.50 0.51 2.0% 0.50 0.51 2.0% 
Wallman_hot 0.50 0.53 6.0% 0.30 0.32 6.7% 
tfin_cold 0.20 0.22 7.7% 0.15 0.17 10.6% 
tfin_hot 0.17 0.18 4.8% 0.11 0.19 75.8% 
 
No defects (such as broken microchannels or warped manifold walls) were 
identified based on the evaluation of the fabricated coupons. As shown in Figure 
6-4, both vertical fins and inclined fins showed good uniformity, and the 
manufacturer claimed that the surface roughness is 5 to 10 µm Ra. Except the 
different feature sizes of the coupons’ design, both fabricated coupons also 
demonstrated a good repeatability with DMLS process. 
76 
 
 
 
Figure 6-4: Measurements of the printed Inconel 718 coupon 2: straight fins (top); 
inclined fins (bottom) 
6.4. Pressure Containment Coupon    
Except the minimum feature size, the microchannel base thickness (which also can 
be call as separating sheet) was another important factor for the manifold-
microchannel heat exchanger design. Considering the weight of the microchannel 
base is usually about 40-50% of the total manifold-microchannel heat exchanger’s 
mass, a thinner microchannel base can significantly reduce the total weight of the 
77 
 
 
heat exchanger. However, it was noticed that the part fabricated by DMLS is not as 
solid as the one fabricated through conventional techniques. According to the 
DMLS manufacturer, the density of part is generally 99-99.5% of the raw material 
since it is a sintering process with micron size metal powder (which means there 
are some voids within the part). As a result, there might be a leaking issue if the 
microchannel base was too thin. In order to find the minimum microchannel base 
thickness which could prevent leakage under the requested system pressure (64.7 
psia), two pressure containment test coupon designs were created with the same 
dimensions as test print coupon 2 as shown in Figure 6-5.  
         
Figure 6-5: Pressure containment test coupon’s CAD design (left); cross section 
view – right side (middle); cross section view – front side (right) 
The difference between the pressure containment test coupon designs was that one 
has a microchannel base thickness of 0.5 mm, and the other’s is 0.3 mm. The 
pressure containment test coupons were successfully fabricated using DMLS as 
shown in Figure 6-6. Considering the small tube opening with an inner diameter of 
78 
 
 
4 mm and the inclined microchannels with a channel width of 0.346 mm, there 
might still be some power left in the enclosure area on the hot side which is shown 
in red in Figure 6-5, which could accidentally seal the voids and reduce the leakage. 
To ensure there was no powder remaining in the inclined microchannels, the 3-D 
printed coupons were post-processed using an ultrasonic cleaner. After the 
cleaning, both coupons showed very good agreements to the design values (within 
2%) based on the weight measurements. 
         
Figure 6-6: 3-D printed Inconel 718 pressure containment test coupon 
To perform the pressure containment test, the coupon was directly connected to a 
Nitrogen gas tank, and the applied pressure was controlled by the pressure 
regulator. By putting the coupon under water, if there was any leakage, it would be 
visually identified, and the leaking point could be located based on where the 
bubbles showed up. As shown in Figure 6-7, one leaking point was identified on 
the coupon with base thickness of 0.3 mm, and no leakage was found on the one 
with 0.5 mm thick microchannel base. To check the maximum pressure that can be 
held by the 3-D printed coupon with base thickness of 0.5mm, another high-
pressure test was performed. By connecting the coupon to a water-based pressure 
79 
 
 
pump, it was found that the coupon could withstand up to 3000 psig without any 
leakage.    
 
Figure 6-7: Pressure containment test coupon with base of 0.3 mm at 20 psig 
6.5. Subscale Unit Design   
Since the fabrication and evaluation of the test print coupons indicated the 
feasibility of the selected fabrication approach, a subscale unit design was created 
based on the DMLS process. The design has the same inclined microchannel fins 
on the hot side, and inclined manifold walls on the cold side. The printing 
orientation of the subscale unit is shown in Figure 6-8. A 10 × 10 × 5 mm3 small 
sectional sample design was also created for measuring internal structure 
dimensions considering it would be very difficult to directly measure the 
microchannel’s dimensions inside the subscale unit. 
80 
 
 
66 
mm 
Printing 
orientation 
74 mm 
27 mm 
 
Figure 6-8: A small sectional sample for measuring internal structure dimensions 
(left); subscale unit printing orientation (right) 
It was noticed by Arie [80] that straight manifold channel may induces a large 
maldistribution when the number of microchannels is high (n > 45), which makes 
the assumption of uniform flow distribution invalid for the numerical model 
discussed in the previous chapter. Therefore, the manifold channels on both hot and 
cold sides were designed as tapered channels for the subscale unit to better control 
the flow distribution. As shown in Figure 6-9, the flow area in the inlet manifold 
channels decreases longitudinally as more flow is fed into microchannels. The 
opposite process occurs for the outlet manifold channel where the flow area 
increases as more flow leaves the microchannels. 
81 
 
 
 
 
Figure 6-9: Hot-side tapered manifold channel with vertical manifold walls (top); 
cold-side tapered manifold channel with inclined manifold walls (bottom) 
82 
 
 
By implementing the fabricated minimum feature dimensions such as the minimum 
microchannel fin thickness of the test  print coupons and minimum microchannel 
base thickness of the pressure containment test coupons, the subscale unit’s design 
was updated from the initial scaled-down model, and the geometrical parameters of 
both the hot and cold sides are shown in Table 6-3 (see Figure 6-10  for the 
definitions of these parameters). Considering the subscale unit might need to be 
connected to upper-level manifolds (which can also be called as headers) through 
some conventional assembly processes, a connection interface was added into the 
design which is marked in yellow as shown in Figure 6-8.  
Table 6-3 Geometrical parameters of the subscale unit  
  Hot Side  
𝑡𝑓𝑖𝑛 0.180 mm 
𝑊𝑐ℎ𝑛 0.390 mm 
𝐻𝑏𝑎𝑠𝑒 0.250 mm 
Number of microchannels  72 
𝑊𝑎𝑙𝑙𝑚𝑛𝑑 0.500 mm 
  Cold Side  
𝑡𝑓𝑖𝑛 0.165 mm 
𝑊𝑐ℎ𝑛 0.340 mm 
𝐻𝑏𝑎𝑠𝑒 0.250 mm 
Number of microchannels  138 
𝑊𝑎𝑙𝑙𝑚𝑛𝑑 0.500 mm 
  Overall Size  
L 66.0 mm 
W 27.0 mm 
H 74.0 mm 
Mass 211 g 
83 
 
 
 
Figure 6-10: Cross-flow manifold-microchannel configuration 
6.6. Subscale Unit Fabrication & Evaluation  
Based on the subscale unit’s final design, the heat exchanger core and the small 
sectional coupon were successfully fabricated simultaneously in the same build 
through DMLS as shown in Figure 6-11. The subscale unit was built within 36 
hours which was much less than the quoted lead time of any other conventional 
manufacturing techniques. To ensure there was no powder remaining in the 
fabricated subscale unit, it was cleaned multiple times using an ultrasonic cleaner. 
After cleaning, the fabricated minimum feature dimensions were determined based 
on the measurements of the small sectional coupon, and the overall size of the 
subscale unit was measured as well. The variations between the fabricated and 
designed dimensions are shown in Table 6-4 and demonstrate fabrication fidelity 
within 14% of design values (except fin thickness on both sides, all other 
dimensions were within 5%). A pressure containment test was also performed on 
84 
 
 
the 3-D printed subscale unit. The result showed that the heat exchanger core could 
withstand a system pressure of 110 psig without any leakage from hot to cold sides. 
 
Figure 6-11: Additively manufactured Inconel 718 subscale unit and the small 
sectional coupon  
Table 6-4 Percentage deviation of the fabricated dimensions and mass from the 
corresponding design values  
  Hot Side  
𝑡𝑓𝑖𝑛 12.20 % 
𝑊𝑐ℎ𝑛 - 3.50 % 
  Cold Side  
𝑡𝑓𝑖𝑛 13.44 % 
𝑊𝑐ℎ𝑛 - 4.15 % 
  Overall Size  
L 0.14 % 
W 0.18 % 
H  0.03 % 
  Mass 1.0 % 
85 
 
 
6.7. Summary   
In summary, two test print coupons and two pressure containment test coupons 
were successfully fabricated through DMLS process using Inconel 718 as the 
material of choice. Inspection of the fabricated coupons showed that the minimum 
printable microchannel fin thickness was 165 µm for the vertical fins and 180 µm 
for the inclined fins. It was also concluded that the minimum microchannel base 
thickness needs to be 500 µm to prevent leakage from high-pressure (64.7 psia) to 
low-pressure (14.7 psia) side. By implementing the dimensions measured from the 
test print coupons in the initial scaled-down model, a subscale unit design was 
created, and it was successfully fabricated through DMLS using Inconel 718.  
  
86 
 
 
Chapter 7: Low-temperature Performance Characterization  
7.1. Introduction  
This chapter discusses the performance characterization of the 3-D printed subscale 
gas-to-gas cross-flow manifold-microchannel heat exchanger under low-
temperature conditions. The single-manifold-multi-microchannels model is 
discussed since the subscale unit was modified to meet the new fabrication 
constraints, and the effect of the tapered manifold channel need to be considered. 
Experimental tests under low-temperature conditions were performed to evaluate 
the heat exchanger’s performances. Lastly, the experimental results were compared 
with the numerical predictions. 
7.2. Single-Manifold-Multi-Microchannels Model     
The numerical model discussed in chapter 5 is based on straight manifold channel 
geometry and the assumption of uniform flow distribution into the microchannels 
along the manifold channel. The study from Arie [80] showed that if the flow 
maldistribution is less than 30%, the modified hybrid method with the single 
manifold-microchannel CFD model can still predict the performance accurately. 
The flow maldistribution, F, is defined as the ratio of the standard deviation of mass 
flow rate among all microchannels to the mean microchannel mass flow rate as  
 
87 
 
 
√1 ∑𝑛 ̅̅ ̅̅ ̅̅ ̅ 2
𝑛 𝑖=1(?̇?𝑐ℎ𝑛,𝑖 − ?̇?𝑐ℎ𝑛) (6) 
𝐹 =   
?̇?̅̅ ̅̅ ̅̅ ̅𝑐ℎ𝑛
where n is the number of microchannels along the manifold channel, and a smaller 
F value means more uniform flow distribution.   
However, since the subscale unit was modified based on the new fabrication 
constraints, the manifold channel was changed from straight channel to tapered 
channel as well. In order to confirm the flow maldistribution with the tapered 
manifold channel design is less than 30% so that the modified hybrid method can 
still be used to predict the performance, two single-manifold-multi-microchannels 
(SMMM) CFD models were created for both hot and cold sides based on the final 
CAD design and measured dimensions. The SMMM model was created with a 
mesh generation software Gambit 2.4.6, and solved numerically using the 
commercially available CFD software Fluent. Figure 7-1 shows an example of the 
SMMM model for the hot side. The goal of this model was to obtain the 
conductance ℎ𝑏𝑎𝑠𝑒 at the microchannel base area instead of the single manifold-
microchannel model (as shown in Figure 5-3), and the pressure drop ∆P𝑐𝑜𝑟𝑒 from 
the manifold inlet to manifold outlet (from surface A to surface B). 
88 
 
 
 
Figure 7-1: SMMM model for CFD simulation on the hot side 
To achieve the goal, two assumptions were made for the SMMM model:   
1. Mass flow rate in all manifold channels is equal. 
2. Symmetric condition exists for both sides and top planes of the SMMM model. 
The inlet surface of the manifold channel inlet extension was set to constant mass 
flow rate boundary condition and constant inlet temperature. The outlet surface of 
the manifold channel outlet extension was set as the pressure outlet boundary 
condition. The boundary condition of the microchannel base was set as a constant 
value calculated initially as the average of hot and cold inlet temperatures. This is 
based on the assumption that uniform flow was fed into each microchannel, and 
each layer of the subscale heat exchanger can be divided into hundreds of repeated 
89 
 
 
unit cells which have the same hot-inlet and cold-inlet temperatures. As a result of 
the short flow length of each unit cell, a constant base temperature can be assumed. 
To reduce the computational domain, only a half-section of the manifold was 
modeled, and symmetry boundary condition was applied on the top plane. A grid 
independency study was performed to find the minimum computational element 
number that provides less than 3% error while limiting the computational time. It 
was concluded that about 4,000,000 computational elements were sufficient for this 
purpose. 
The heat duty (Q), outlet temperature on the hot and cold sides (𝑇ℎ𝑜𝑡,𝑜𝑢𝑡  and 
𝑇𝑐𝑜𝑙𝑑,𝑜𝑢𝑡), and temperature at the microchannel base (𝑇𝑏𝑎𝑠𝑒) were calculated based 
on Equations (7) – (11) as there was only one pass on each side. 
Q𝑖 = ?̇?ℎ𝑜𝑡𝐶𝑝,ℎ𝑜𝑡(𝑇 (7) ℎ𝑜𝑡,𝑖𝑛 − 𝑇ℎ𝑜𝑡,𝑜𝑢𝑡) 
Q𝑖 = ?̇? (8) 𝑐𝑜𝑙𝑑𝐶𝑝,𝑐𝑜𝑙𝑑(−𝑇𝑐𝑜𝑙𝑑,𝑖𝑛 + 𝑇𝑐𝑜𝑙𝑑,𝑜𝑢𝑡) 
Q = ℎ 𝐴 (𝑇 − 𝑇 ) (9) 𝑖 ℎ𝑜𝑡,𝑏𝑎𝑠𝑒 𝑏𝑎𝑠𝑒 ℎ𝑜𝑡,𝑖𝑛 𝑏𝑎𝑠𝑒
Q𝑖 = ℎ (10) 𝑐𝑜𝑙𝑑,𝑏𝑎𝑠𝑒𝐴𝑏𝑎𝑠𝑒(−𝑇𝑐𝑜𝑙𝑑,𝑖𝑛 + 𝑇𝑏𝑎𝑠𝑒) 
𝑄 = 𝑁 × Q  (11) 𝑙𝑎𝑦𝑒𝑟 𝑖
where 𝑇ℎ𝑜𝑡,𝑖𝑛and 𝑇𝑐𝑜𝑙𝑑,𝑖𝑛are the inlet temperatures on the hot and cold sides, Qi is 
the heat duty for each manifold-microchannel layer, ℎℎ𝑜𝑡,𝑏𝑎𝑠𝑒 and ℎ𝑐𝑜𝑙𝑑,𝑏𝑎𝑠𝑒 are the 
conductance obtained from the CFD simulations, Abase is the microchannel base 
90 
 
 
area, and Nlayer is the number of manifold-microchannel layers. Five unknown 
variables (𝑇ℎ𝑜𝑡,𝑜𝑢𝑡, 𝑇𝑐𝑜𝑙𝑑,𝑜𝑢𝑡, 𝑇𝑏𝑎𝑠𝑒, Q𝑖, and 𝑄) can be solved. 
Similar to the one shown in Figure 7-1, another SMMM model was created for the 
case with straight manifold channel on the hot side. Based on the CFD simulation 
results, it showed that the flow maldistribution is 11.9% with tapered manifold 
channel, and 24.7% with straight manifold channel as shown in Figure 7-2. This 
means a more uniform flow distribution could be achieved for the subscale unit 
with the tapered manifold channel design. As a result, using the SMMM model with 
tapered manifold channel, the modified hybrid method would provide a more 
accurate prediction.  
 
Figure 7-2: Flow distribution along the manifold channel on hot side (y-axis: mass 
flow rate in each microchannel; x-axis: microchannel location from the beginning 
to the end of manifold channel; “mnd" stands for manifold) 
91 
 
 
It was also noticed that the heat transfer performance predictions with both straight 
and tapered manifold channel SMMM models were close to each other. This is 
because the geometry was optimized by setting the upper limit constraint on flow 
maldistribution as 30%, and most of the heat transfer occurs along the 
microchannels, which means the heat transfer performance might not be affected 
by the straight or tapered manifold channel design (if the maldistribution of each 
design is less than 30%). However, the predicted pressure drop with tapered 
manifold channel was less than the one with straight manifold channel due to a 
smaller flow maldistribution as shown in Figure 7-3.  
6000
TTapapeereredd m mnndd 
5000 sStrtaraigihght tm mnndd 
4000
3000
2000
1000
0
0.0065 0.0105 0.0145 0.0185
ṁhot (kg/s)
 
Figure 7-3: Hot side pressure drop prediction with SMMM model (cold side mass 
flow rate is constant at 0.006 kg/s) 
7.3. Low-temperature Experimental Setup   
To characterize the performance of the fabricated subscale unit, experiments under 
low temperature conditions (hot inlet temperature was ~ 43°C and cold inlet 
92 
 
dPhot (Pa)
 
temperature was ~ 13°C) were performed. A schematic diagram of the low-
temperature experimental setup is shown in Figure 7-4. An air-to-water heat 
exchanger was added to adjust the inlet air temperature to the blower, and a variable 
speed controller was connected to the blower to control the total air flow rate by 
adjusting the power frequency supplied to the blower. The air flow then was 
distributed through the T-joint, and the flow rates on hot and cold side were adjusted 
by the gate valve on cold side. The hot side air was heated by the electric heater to 
reach the desired temperature. The cold side air temperature was adjusted through 
another air-to-water heat exchanger which has circulated low-temperature (~ -5 °C) 
working fluid (ethanol-based water solution) supplied by a chiller. The subscale 
unit was installed on the end of both hot and cold side as shown in Figure 7-4. 
 
Figure 7-4: Schematic of low-temperature experimental setup 
93 
 
 
To connect the subscale unit to the low-temperature experiment setup, custom 
designed hot and cold side headers were created as shown in Figure 7-5 (a). The 
headers were then fabricated using PLA, which has a glass transition temperature 
about 60°C, through fused deposition modeling (FDM) process. For FDM, the parts 
are built by laying down the melted plastic filament through a heated nozzle to form 
the cross section of the object. Then the next layer of the cross section is formed by 
repeating the additive process. During the 3-D printing process of FDM, if the 
original design has an unsupported structure, a sacrificial structure will be added as 
supporting material to solve the overhang issue based on the printer’s control 
program settings. As shown in Figure 7-5 (b), the 3-D printed headers had been 
post-processed to remove the unneeded supporting parts, and were connected to the 
subscale unit with a press-fit design.     
 
(a) 
94 
 
 
 
(b) 
Figure 7-5: Headers’ design for low-temperature experimental setup (a); 3-D 
printed plastic headers (b) 
As shown in Figure 7-6, the headers were connected to the low-temperature 
experimental setup through either Aluminum tapes (hot side) or rubber couplings 
with steel duct clamps (cold side). To prevent any leakage from the connection area, 
silicon glue was applied to seal the gaps. On both hot and cold sides, four T-type 
thermocouples were inserted into the headers and placed as close as possible to the 
inlet surface of the subscale unit to measure the inlet temperatures. Another four 
thermocouples were placed at the subscale unit’s outlet surface for outlet 
temperature measurements. Since the low-temperature tests were under 
atmospheric pressure conditions, only the inlet surfaces on both hot and cold sides 
were connected to the pressure transducers to measure the pressure drops across the 
subscale unit.  
95 
 
 
 
Figure 7-6: Low-temperature experimental setup 
The total flow rate entering both hot and cold sides was measured through a 
rotameter, and a correction correlation was used to calculate the actual flow rate at 
the operation temperature as:    
𝑃 𝑇 𝑀𝑊
?̇? 𝑠𝑡𝑑
𝑒𝑥𝑝 𝑐𝑎𝑙 𝑐𝑎𝑙
𝑡𝑜𝑡 = 𝜌 × ?̇?𝑚𝑒𝑎𝑠𝑢𝑟𝑒√  (12) 𝑃𝑐𝑎𝑙𝑇𝑒𝑥𝑝𝑀𝑊𝑒𝑥𝑝
where std means standard volumetric flow rate, exp stands for experiment, cal 
stands for manufacturing calibration, P is the air pressure, T is the flow temperature 
96 
 
 
(which is the same as the heater’s inlet temperature for 𝑇𝑒𝑥𝑝 ), and MW is the 
molecular weight. 
To get a more accurate data, the flow rate on hot side was calculated as the average 
of the measurements from the nozzle and the electric heater. The flow rate through 
the nozzle was calculated based on the measured pressure drop across the nozzle 
as: 
2∆𝑃
?̇?𝑛𝑜𝑧𝑧𝑙𝑒 = 𝜌 × 𝐶𝑑 × 𝐴𝑜 × √  𝐷𝑜 (13) 
𝜌[1 − ( )4]
𝐷𝑖
where ρ is the air density based on the heater’s outlet temperature measurement, Cd 
is the discharge coefficient from the nozzle’s calibration provided by the 
manufacturer, Ao is the cross-section area of the nozzle throat, Do is the diameter 
of the nozzle throat, and Di is the diameter of the nozzle inlet cross-section area. 
For the electric heater, the flow rate was calculated based on: 
𝑄𝑒 = 𝐶𝑝 × ?̇?ℎ𝑒𝑎𝑡𝑒𝑟 × (𝑇ℎ𝑒𝑎𝑡𝑒𝑟,out − 𝑇ℎ𝑒𝑎𝑡𝑒𝑟,𝑖𝑛) (14) 
where Qe is the measured electricity energy input to the heater, and Cp is the 
specific heat based on the average of the heater’s inlet and outlet temperatures. The 
flow rate on cold side then was the subtraction of hot side flow rate from the total 
flow rate. 
7.4. Low-temperature Experiment Method    
The low-temperature experiments were performed under the conditions shown in 
Table 7-1. Due to the glass transition temperature of header’s material, the hot side 
97 
 
 
air temperature was set at 43°C to avoid any deformation of the header during the 
tests. In order to achieve a proper temperature difference between hot and cold inlet 
temperatures, the cold side air temperature was set at 13°C. The experiments were 
performed twice, the first test is to change hot side flow rate while keeping cold 
side flow rate constant at 0.006 kg/s. Then, the second one is to change cold side 
flow rate while keeping hot side flow rate constant at 0.019 kg/s. During the tests, 
hot side inlet temperature (𝑇ℎ𝑜𝑡,𝑖𝑛), hot side outlet temperature (𝑇ℎ𝑜𝑡,𝑜𝑢𝑡), hot side 
pressure drop ( 𝛥𝑝 ), cold side inlet temperature ( 𝑇
ℎ𝑜𝑡 𝑐𝑜𝑙𝑑,𝑖𝑛
), cold side outlet 
temperature (𝑇𝑐𝑜𝑙𝑑,𝑜𝑢𝑡), and cold side pressure drop (𝛥𝑝 ) were measured and 𝑐𝑜𝑙𝑑
recorded. 
Table 7-1 Low-temperature experiment conditions  
Temperature Condition 
𝑇ℎ𝑜𝑡,𝑖𝑛 43 °C 
𝑇𝑐𝑜𝑙𝑑.𝑖𝑛 13 °C 
Varying Hot Side Flow Rate 
?̇?ℎ𝑜𝑡 0.007-0.018 m
3/s 
?̇?𝑐𝑜𝑙𝑑 0.005 m
3/s 
Varying Cold Side Flow Rate 
?̇?ℎ𝑜𝑡 0.017 m
3/s 
?̇?𝑐𝑜𝑙𝑑 0.001-0.007 m
3/s 
 
98 
 
 
Based on the measurements, the heat transfer rate of the subscale unit was evaluated 
as: 
?̇?ℎ𝑜𝑡 = (?̇?𝑛𝑜𝑧𝑧𝑙𝑒 + ?̇?ℎ𝑒𝑎𝑡𝑒𝑟)/2 (15) 
?̇?𝑐𝑜𝑙𝑑 = ?̇?𝑡𝑜𝑡 − ?̇?ℎ𝑜𝑡 (16) 
𝑄ℎ𝑜𝑡 = 𝐶𝑝ℎ𝑜𝑡 × ?̇?ℎ𝑜𝑡 × (𝑇ℎ𝑜𝑡,𝑖𝑛 − 𝑇ℎ𝑜𝑡,𝑜𝑢𝑡) (17) 
𝑄𝑐𝑜𝑙𝑑 = 𝐶𝑝𝑐𝑜𝑙𝑑 × ?̇?𝑐𝑜𝑙𝑑 × (−𝑇𝑐𝑜𝑙𝑑,𝑖𝑛 + 𝑇𝑐𝑜𝑙𝑑,𝑜𝑢𝑡) (18) 
  𝑄 = (𝑄ℎ𝑜𝑡 + 𝑄𝑐𝑜𝑙𝑑)/2 (19) 
where Cp and 𝜌 are air properties for hot and cold side based on the average 
temperature between inlet and outlet measurements. A heat balance within 10% 
between hot and cold sides was recorded during the low-temperature tests. 
The overall heat transfer coefficient (U) was evaluated using the log mean 
temperature difference (LMTD) method for both unmixed cross-flow arrangements 
as given in Eqs. (20) - (23) as a function of inlet and outlet temperatures on both 
the hot and cold sides. 
𝑑𝑇𝐴 = 𝑇ℎ𝑜𝑡,𝑖𝑛 −  𝑇𝑐𝑜𝑙𝑑,𝑜𝑢𝑡 (20) 
𝑑𝑇𝐵 = 𝑇ℎ𝑜𝑡,𝑜𝑢𝑡 −  𝑇𝑐𝑜𝑙𝑑,𝑖𝑛 (21) 
𝑑𝑇𝐴 − 𝑑𝑇𝐵
𝐿𝑀𝑇𝐷 = 𝐶𝐹  (22) 
ln( 𝑑𝑇𝐴 /𝑑𝑇𝐵) 
𝑄
𝑈 =  (23) 
𝑁𝑙𝑎𝑦𝑒𝑟𝐴𝑏𝑎𝑠𝑒𝐿𝑀𝑇𝐷 
99 
 
 
where CF is the configuration correction factor of both unmixed cross-flow, 𝑁𝑙𝑎𝑦𝑒𝑟 
is the number of manifold-microchannel layers, and Abase is the microchannel base 
area (Abase = 0.059 m × 0.070 m). 
Uncertainty propagation analysis was performed to calculate the uncertainty in the 
subscale heat exchanger performance parameters (Q, U) due to inaccuracy of the 
measurements. A list of the measurement equipment is shown in Table 7-2 with the 
corresponding accuracies. The uncertainty of the calculated quantity Y (UY), which is 
a function of X1, X2, . . . , XN with uncertainty of UX1, UX2, . . . , UXN, can be calculated 
using the method explained in NIST Technical Note 1297 [109] as: 
𝜕𝑌
  𝑈 2 2𝑌 = √∑ ( ) 𝑈𝑋  (24) 𝜕𝑋 𝑖𝑖 𝑖
The uncertainties of the subscale heat exchanger performance parameters (Q, U, ∆P) 
were included as error bars in the results presented in the next section. 
Table 7-2 List of low-temperature measurement equipment and its accuracy 
  Equipment function Equipment name Accuracy 
Fischer Porter F 
Total flow rate ± 2%  
Rotameter 
Pressure drop across Setra pressure 
± 0.14% FS of 1.2 kPa 
nozzle transducer 
± 1.0°C or 0.75% of 
Temperature T type thermocouple  
measurement 
Setra pressure 
Hot-side pressure drop ± 0.14% FS of 7.5 kPa 
transducer 
Setra pressure 
Cold-side pressure drop ± 0.14% FS of 2.5 kPa 
transducer 
100 
 
 
7.5. Low-temperature Experiment Results   
The subscale unit’s performances are shown in Figure 7-7 for the case of varying 
hot side flow rate while keeping the cold side flow rate constant at 0.006 kg/s. The 
heat duty, pressure drop on the hot side, and overall heat transfer coefficient are 
plotted as a function of hot side flow rate. The capacity (Q) is based on a 
temperature difference of 30°C between hot and cold inlet air. Figure 7-7 (a) shows 
a maximum capacity of 90 W for the subscale unit under low-temperature 
conditions. During the low-temperature tests, the hot-side pressure drop (𝛥𝑃ℎ𝑜𝑡) 
reached 4000 Pa at maximum hot-side mass flow rate of about 0.019 kg/s as shown 
in Figure 7-7 (b). The overall heat transfer coefficient (U) of the heat exchanger was 
in the range of 900 – 1200 W/m2K as shown in Figure 7-7 (c).  
120
100
80
60
40
Exp.
20
Num.
0
0.0065 0.0105 0.0145 0.0185
ṁhot (kg/s)
 
(a) 
101 
 
Q (W)
 
5,000
Exp.
4,000 Num.
3,000
2,000
1,000
0
0.0065 0.0105 0.0145 0.0185
ṁhot (kg/s)
 
(b) 
2,000
Exp.
1,500 Num.
1,000
500
0
0.0065 0.0105 0.0145 0.0185
ṁhot (kg/s)
 
(c) 
Figure 7-7: Experimental and numerical results for constant cold-side mass flow 
rate at 0.006 kg/s and 30°C temperature difference between hot and cold inlet air: 
heat duty (Q) vs. hot-side mass flow rate (?̇?ℎ𝑜𝑡) (a); hot-side pressure drop (𝛥𝑃ℎ𝑜𝑡) 
vs. hot-side mass flow rate (?̇?ℎ𝑜𝑡) (b); overall heat transfer coefficient (𝑈) vs. hot-
side mass flow rate (?̇?ℎ𝑜𝑡) (c) 
102 
 
U (W/m2K) ∆Phot (Pa)
 
The results showed a decent agreement between the numerical predictions and 
experimental results. Analysing the overall trend, heat duty, hot-side pressure drop, 
and overall heat transfer coefficient all increase as hot-side mass flow rate 
increases. This trend is as expected since the flow in microchannels is in the 
developing region, increasing flow rate leads in an increase of heat transfer 
performance as well as pressure drop.  
Figure 7-8 is showing the case of varying cold side flow rate while keeping the hot 
side flow rate constant at 0.019 kg/s. The heat duty, pressure drop on the cold side, 
and overall heat transfer coefficient are plotted as a function of cold side flow rate. 
Figure 7-8 (a) shows a maximum capacity of 100 W for the subscale unit with a 
30°C temperature difference between both sides’ inlet air. A maximum cold-side 
pressure drop (𝛥𝑃𝑐𝑜𝑙𝑑) of 1360 Pa was recorded when cold-side mass flow rate 
reached 0.008 kg/s as shown in Figure 7-8 (b). The overall heat transfer coefficient 
(U) of the heat exchanger was in the range of 800 – 1500 W/m2K as shown in Figure 
7-8 (c). 
120
100
80
60
40
Exp.
20
Num.
0
0.002 0.004 0.006 0.008
ṁcold (kg/s)  
(a) 
103 
 
Q (W)
 
5,000
Exp.
4,000 Num.
3,000
2,000
1,000
0
0.002 0.004 0.006 0.008
ṁcold (kg/s)
 
(b) 
2,000
Exp.
Num.
1,500
1,000
500
0
0.002 0.004 0.006 0.008
ṁcold (kg/s)
 
(c) 
Figure 7-8: Experimental and numerical results for constant hot-side mass flow rate 
at 0.019 kg/s and 30°C temperature difference between hot and cold inlet air: heat 
duty (Q) vs. cold-side mass flow rate (?̇?𝑐𝑜𝑙𝑑) (a); cold-side pressure drop (𝛥𝑃𝑐𝑜𝑙𝑑) 
vs. cold-side mass flow rate (?̇?𝑐𝑜𝑙𝑑) (b); overall heat transfer coefficient (𝑈) vs. cold-
side mass flow rate (?̇?𝑐𝑜𝑙𝑑) (c) 
104 
 
U (W/m2K) ∆Pcold (Pa)
 
Analyzing the overall trend, it also shows that heat duty, cold-side pressure drop, 
and overall heat transfer coefficient all increase as cold-side mass flow rate 
increases. Comparing the trend between Figure 7-7 and Figure 7-8, it shows that 
cold-side flow rate has more impact on the heat exchanger’s thermal performance. 
This is due to the fact that the cold-side flow rate is much lower (which is almost 50% 
of the hot-side flow rate). Therefore, the thermal resistance on the cold side is higher 
than the hot side, and any improvement in cold-side flow rate can positively impact 
the overall heat transfer performance. 
However, it was noticed that the measured cold-side pressure drop was about 40-
50% higher than the numerical prediction, while the hot-side was within 10%. Since 
both side’s pressure drops’ numerical predictions were based on the SMMM 
models with tapered manifold channel, the higher pressure drop on cold side might 
be due to that the cold side header’s design induced a non-uniform flow distribution 
when the flow enters the cold-side manifold channels, which means the mass flow 
rate in each cold-side manifold channel was not equal. As shown in Figure 7-9, the 
hot-side header design has several air deflectors to guide the air from the nozzle to 
the inlet of each hot-side manifold channel. But the cold-side header’s design only 
has two flat channels to separate the inlet air so that it enters either top or bottom 
manifold channels on the cold side. As a result, the flow entering the cold side 
might not be as uniform as the flow entering the hot side. 
105 
 
 
 
(a) 
Hot air 
Cold air 
 
(b) 
Figure 7-9: Cross section view of the headers’ design (a); hot & cold side of the 
subscale unit (b) 
106 
 
 
7.6. Summary   
In summary, the single-manifold-multi-microchannels model, which can be used to 
replace the single manifold-microchannel model in the modified hybrid method, is 
discussed. It shows that even the flow maldistribution from the straight manifold 
channel design is less than 30%, a more uniform flow distribution can be obtained 
with the tapered manifold channel design. The low-temperature experimental setup 
for performance characterization as well as the test conditions are presented.  Based 
on the low-temperature experimental results, it shows that the measured pressure 
drop on cold side is 40-50% higher than the numerical prediction, which might be 
due to the non-uniform flow distribution at the subscale unit’s cold-side inlet 
surface induced by the header’s design. 
  
107 
 
 
Chapter 8: High-temperature Performance Characterization  
8.1. Introduction  
Since the gas-to-gas cross-flow manifold-microchannel heat exchanger was 
designed for high temperature applications, to ensure the fabricated subscale unit 
can work as expected, this chapter discusses the performance characterization of 
the 3-D printed subscale unit under high-temperature (600°C) conditions. 
Considering most material cannot withstand high temperatures, and the press-fit 
connection does not work well with metal-to-metal interface, custom designed 
metal headers were created and skillfully welded to the subscale unit. A full CFD 
model including both inlet/outlet headers and heat exchanger core is discussed since 
the effect of the headers (such as flow distribution at the inlet surface of the subscale 
unit and pressure drop induced by the inlet/outlet headers) need to be considered. 
Experimental tests under high-temperature conditions were performed to evaluate 
the heat exchanger’s performances. Lastly, the experimental results were compared 
with the numerical predictions and conventional high temperature heat exchanger’s 
performance. 
8.2. Headers Design for High-temperature Experiment 
For high-temperature experiment, since test temperature is around 600°C, 
inlet/outlet headers on both sides have to be fabricated with high-temperature-
resistant material. Considering the subscale unit was fabricated using Inconel 718 
which has an operating temperature up to 750°C, and to save fabrication time for 
108 
 
 
complex designs which could not be easily built through conventional techniques, 
the high-temperature headers were designed based on 3-D metal printing (DMLS) 
using Inconel 718 as well. Based on the low-temperature experimental results, it 
was found that non-uniform flow induced by the header might increase the 
measured pressure drop significantly. To improve the uniformity of the flow 
entering the subscale unit, flow deflectors have been added into each header’s 
design as shown in Figure 8-1.  
Flange 
Hot inlet Header 
Cold outlet 
Hot outlet 
Cold inlet 
 
Figure 8-1: Cross section view of headers’ design attached to the subscale unit 
109 
 
 
Large surface area can easily cause very high heat loss when 600°C gas flows inside 
the headers, which in turn may induce a high heat balance value between hot and 
cold sides. Compared to the low-temperature headers as shown in Figure 7-5, the 
high-temperature headers were designed as compact as possible to reduce the 
surface area. The inlet/outlet of each header was designed as a 1 in OD tube so that 
the 1 in flanges, which was used to connect the test section to the high-temperature 
experimental setup, could be welded to the headers. Considering press-fit 
connection does not work well with metal-to-metal interface under high-
temperature conditions, the metal headers were decided to be welded to the subscale 
unit, and the connection interface of metal headers were modified as shown in 
Figure 8-2. Due to the interface design of the subscale unit, the hot-side inlet header 
has more contacting surface area compared to the other three. 
Interface for welding 
  
(a) 
110 
 
 
Edge Welding 
Edge welding 
 
(b) 
Interface for welding 
 
(c) 
111 
 
 
Welding at the corner 
Welding at 
the corner 
 
(d) 
Figure 8-2: Metal header’s connection interface design: for hot-side inlet header 
(a) & (b); for hot-side outlet and cold-side inlet/outlet headers (c) & (d) 
Since metal material’s strength reduces at high operating temperature, a welding 
safety factor calculation was performed considering the required system pressure 
112 
 
 
on hot side is 64.7 psia and 14.7 psia on cold side. Based on the welding joint 
design, the force (𝐹𝑤) it can withstand was calculated as  
𝐹𝑤 = 𝑃 × 𝑊 × 𝑌𝑆 (25) 
where P is the total length of the weld (which is also the perimeter of the welding 
edge), W is the width of the weld which is about 0.5 mm, and YS is the yield 
strength of the as-welded Inconel 718 based on the material spec sheet as shown in 
Table 8-1. 
Table 8-1 Inconel 718 spec sheet [110] 
 
Note: for high temperature condition, assuming the yield strength reduction ratio of 
the welded material is similar to parent material, the yield strength of the as-welded 
material can be about 68 ksi (469 MPa) at 1200°F (650°C).   
For the header under system pressure, the force applied to the header (𝐹𝑝) was 
calculated as 
𝐹𝑝 = 𝐴 × (𝑁 × 𝑃𝑠𝑦𝑠) (26) 
where A is the projected area of the header, N is the design coefficient for system 
pressure which means the applied pressure is 1.5 times the required value, and Psys 
is the required system pressure. 
113 
 
 
Then, the safety factor SF was obtained as 
𝐹𝑤
𝑆𝐹 =  (27) 
𝐹𝑝
Based on the calculations for both hot- and cold- inlet/outlet headers, the safety 
factors are included in Table 8-2. As all safety factors are more than 10, it was 
concluded that connecting the metal headers to the subscale unit through welding 
would be a safe and reliable approach to prepare the test section (which includes 
headers, flanges, and the heat exchanger core as shown in Figure 8-1) for the high-
temperature experiments.    
Table 8-2 Estimated welding safety factor (at 650°C) 
Header P W YS A N Psys SF 
Name  (mm) (mm) (MPa) (mm2) (-) (MPa) (-) 
Hot side 
186.1 0.5 469 1410 1.5 0.448 46 
inlet 
Hot side 
194.3 0.5 469 1,811 1.5 0.448 37 
outlet 
Cold side 
171.7 0.5 469 1,606 1.5 0.101 165 
inlet/outlet 
 
8.3. Flow Simulation with Metal Header    
Even though it was considered to add flow deflectors into each header’s design to 
improve the uniformity of the flow entering the subscale unit, no empirical 
correlations could be used to determine the length, location, and shape of the 
deflectors.  Therefore, considering the computational time, Solidworks flow 
simulations have been performed as a preliminary method to check whether 
uniform flow could be obtained through the custom designed flow deflectors inside 
114 
 
 
each header. By trial and error, a preliminary header design could be created with 
a flow maldistribution less than 5% based on the flow simulation results using 
Solidworks. 
 
(a) 
 
(b) 
Figure 8-3: Solidworks flow simulation results: hot-side inlet header (a); hot-side 
outlet header (b) 
115 
 
 
As shown in Figure 8-3, the flow simulation using Solidworks includes the solid 
body of the header design. For hot-side inlet header, the 1 in tube surface at inlet 
was set as mass flow rate inlet with constant inlet temperature. The five uniform 
manifold channel inlet areas of the subscale unit on the hot side were set as pressure 
outlet. No symmetry boundary was used as the Solidworks flow simulation includes 
the entire flow region within the header. The cross-section view in Figure 8-3 is 
only for the purpose of showing the structure inside header. 
For hot-side outlet header, the manifold channel outlet areas of the subscale unit on 
the hot side were set as mass flow rate inlet with constant temperature. A 1 in tube 
extension was added at the outlet surface of the header design to avoid re-
circulation at the header’s outlet surface. The extension’s outlet surface was then 
set as pressure outlet. One thing needs to be noticed is that half manifold channels 
were observed at the hot-side outlet of the subscale unit as well as cold-side 
inlet/outlet due to the flow behavior inside the manifold-microchannel heat 
exchanger. Therefore, the flow maldistribution of hot-side outlet header as well as 
cold-side inlet/outlet headers was evaluated as  
2 ?̇?∑ (?̇?  −  𝑡𝑜𝑡 )2𝑖=1 𝑖  ∑𝑛
?̇?
(?̇?  −  𝑡𝑜𝑡 22 × (𝑛 − 1) 𝑖=3 𝑖
𝐹 =  √ + 𝑛 − 1
)  
 (28) 
?̇?𝑡𝑜𝑡 2 ?̇?2 × ( ) (𝑛 − 2) × ( 𝑡𝑜𝑡 )2
2(𝑛 − 1) 𝑛 − 1
where area 1 and 2 are half size manifold channels, area 3 – 6 are full size manifold 
channels as shown in Figure 8-3 (b). 
116 
 
 
For cold-side inlet/outlet headers, similar boundary conditions and simulation 
approach were used. Based on the Solidworks flow simulation results, it showed 
that the original header design without any flow deflector usually has a flow 
maldistribution of ~ 20 – 40% with different inclined angles, which is the angle 
between the contacting surface of the subscale unit and the 1 in tube surface of the 
header.  By adding deflectors inside headers, after multiple design iterations 
through trial and error, the best flow maldistribution results obtained from 
Solidworks flow simulation were indicating ~ 4% maldistribution for hot-side inlet 
header, ~ 8% for hot-side outlet header, ~6% for cold-side inlet header, and ~ 8% 
for cold-side outlet header. 
However, due to the complexity of the subscale unit deign and the small microchannel 
size, the Solidworks flow simulation cannot include the subscale unit, which means 
the simulation result may not be accurate since the flow resistance from the heat 
exchanger core is not considered. To ensure the assumption of uniform flow 
distribution from headers among all manifold channels, and to get a more accurate 
and reliable flow simulation result, another Fluent flow simulation (including both 
inlet/outlet headers and heat exchanger core) was performed to determine whether the 
mass flow rate in each manifold channel is equal with the header design obtained 
based on the Solidworks flow simulation results.  
117 
 
 
 
Figure 8-4: Simplified single manifold-porous-medium model (top); hot-side full 
model with inlet/outlet headers (bottom) 
 
118 
 
 
Since the measured pressure drop of the test section includes pressure drop of the inlet 
and outlet headers, this Fluent simulation can also predict the total pressure drop from 
the inlet header to the outlet header. As shown in Figure 8-4, in order to reduce 
computational time, microchannels were replaced with a porous medium in the 
simplified model. For the simplified single manifold-porous-medium model, the same 
boundary conditions of the SMMM model were applied. Based on a grid 
independency study, it was concluded that about 1,300,000 computational elements 
were sufficient for the simplified model. Starting with the simplified single manifold-
porous-medium model, the porous medium parameters were calibrated by matching 
the pressure drop obtained from the SMMM model for a range of the same mass flow 
rates as shown in Figure 8-5. Then, the calibrated porous medium parameters were 
used in the full model to predict the total pressure drop, which includes pressure drops 
of both the inlet and outlet headers.  
 
(a) 
119 
 
 
 
(b) 
Figure 8-5: Porous medium calibration: hot-side core pressure drop vs hot-side 
mass flow rate (a); cold-side core pressure drop vs cold-side mass flow rate (b) 
For the full model, the inlet surface of the inlet header extension was set to mass flow 
rate boundary condition. The outlet surface of the outlet header extension was set as 
pressure outlet boundary condition. The symmetry boundary condition was applied 
on the plane where the headers and manifold channels are cut in half. The result of a 
grid independency study showed that about 8,500,000 computational elements were 
sufficient for the full model.  
120 
 
 
Full manifold channel 
Half manifold channel 
 
Figure 8-6: Cross-section view of manifold channels on hot side (5 full manifold 
channels as inlet channel, 2 half & 4 full manifold channels as outlet channel) 
As discussed before, due to the characteristics of manifold-microchannel heat 
exchanger’s flow path, half manifold channels were included in the subscale unit 
design as shown in Figure 8-6. Therefore, a uniform flow distribution means that mass 
flow rate at the half manifold channel needs to be half of the full manifold channel. 
Based on the full model’s simulation results, the mass flow rate at each manifold 
channel could be obtained, and it indicated that flow maldistribution is 6 – 8% for hot-
side inlet/outlet headers, and 3 – 5% for cold-side inlet/outlet headers as shown in 
Figure 8-7. 
121 
 
 
Hot inlet
25.00%
20.00%
15.00%
10.00%
5.00%
0.00%
1 2 3 4 5
Manifold channel
  
Hot outlet
25.00%
20.00%
15.00%
10.00%
5.00%
0.00%
1 2 3 4 5 6
Manifold channel
 
Cold inlet
20%
15%
10%
5%
0%
1 2 3 4 5 6
Manifold channel
 
122 
 
 
Cold outlet
20%
15%
10%
5%
0%
1 2 3 4 5 6
Manifold channel
 
Figure 8-7: Flow distribution in each manifold channel (hot-side mass flow rate at 
0.02 kg/s, cold-side mass flow rate at 0.006 kg/s) 
8.4. Metal Headers Fabrication & Evaluation 
Since the metal header design does not have any small feature (< 0.5 mm) that 
requires a high-resolution 3-D metal printer, all four headers were fabricated based 
on the printing orientation shown in Figure 8-8 through the DMLS process with 
normal resolution (30 µm per layer). However, supporting structures were added 
by the manufacturer as the welding interface of all headers’ design has the overhang 
issue. After the 3-D printing process, the headers together with the supporting 
structures were removed through wire EDM. The supporting structures were then 
removed with a power rotary tool, and the headers were polished using sandpaper 
to improve the surface finish quality of the overhung welding interface area. 
123 
 
 
Printing 
orientation 
Overhang issue  
Figure 8-8: 3-D printing orientation of the metal headers 
As shown in Figure 8-9, the metal headers’ as-built surface finish was stated by the 
manufacturer as 5 – 10 µm Ra. The density of 3-D printed headers was evaluated 
by measuring the weight with a scale and measuring the volume with water in a test 
tube. The results showed the density was ~8.081 g/cm3, which is 99% of the raw 
material density for Inconel 718 provided by the manufacturer. It was also noticed 
that the measured mass of the 3-D printed headers was ~8% less than the design 
value. According to the manufacturer, this might be due to the volume loss during 
the post-processing such as that more welding interface was machined away with 
the power rotary tool when removing the supporting structures. 
124 
 
 
 
(a) 
 
(b) 
 
(c) 
125 
 
 
 
(d) 
Figure 8-9: Hot-side inlet header (a); hot-side outlet header (b); cold-side inlet 
header (c); cold-side outlet header (d) 
8.5. Assembled Test Section 
As previously discussed, the test section for the high-temperature experiment 
includes hot- / cold-side headers, flanges, and the heat exchanger core (the subscale 
unit). All parts were decided to be assembled through welding. However, since 
headers and the subscale unit were 3-D printed through DMLS process, it was 
unclear what quality could be obtained after welding these 3-D printed metal pieces 
together. To ensure a successful welding job, an experienced welder was invited, 
and a small welding-trial piece was prepared as shown in Figure 8-10 (a). The 
welder cut the welding-trial piece in half and tried to weld them back-to-back 
through edge welding as shown in Figure 8-10 (b). According to the welder, these 
3-D printed metal pieces were still weldable. However, due to the voids inside the 
parts (which might have trapped some noble gas such as Argon during the sintering 
126 
 
 
process), welding 3-D printed Inconel 718 is much more difficult than the raw 
material. 
27 mm 10 mm 
10 mm 
 
(a) 
 
(b) 
Figure 8-10: A section from the subscale unit design (a); Welded 3-D printed 
welding-trial section through edge welding (b) 
In order to connect the test section to the high-temperature experimental setup, 
flanges and clamps connection was selected. For tube OD of 1 inch, the metal 
127 
 
 
flanges’ burst pressure is rated as 1200 psia at 800°F which is more than 10 times 
the required system pressure [111]. However, to weld the flanges to the headers, 
flanges made of Inconel 718 were needed, which are not commonly available in the 
market. Based on flange’s design standard and the welder’s suggestion, a modified 
flange design for use of flat gasket was created (Figure 8-11 (a)). As shown in 
Figure 8-11 (b), four Inconel 718 flanges were fabricated through conventional 
machining process. To ensure there is no gas leakage from the contacting surface 
between two flanges under required system pressure of 64.7 psia, a 0.02-inch-thick 
Flexible Grafoil gasket was used which has a service temperature up to 815°C.   
 
(a) 
128 
 
 
 
(b) 
Figure 8-11: Modified Inconel 718 flange’s engineering drawing (unit: inch) (a); 
fabricated Inconel 718 flanges through conventional machining (b) 
Since a tight-fit was required by the welder, and the 3-D printed headers could not 
directly fit into the subscale unit as well as the flanges, to complete the assembly 
of the test section, headers and subscale unit were post-processed through a precise 
machining. After the post-process, all pieces of the test section were fitted into each 
other tightly before welding as shown in Figure 8-12. 
 
Figure 8-12: Test section including core, headers, and flanges before welding 
129 
 
 
The test section was successfully welded together, however some leakages were 
identified after the first welding job. As it was leaking at several points from the 
welding line to the ambient when the test section was pressurized, more welding 
was applied to seal the identified leaking points. After multiple trials, the test 
section was eventually proved to be leak-free (no leakage from hot side to cold side 
through headers, and no leakage from the pressurized test section to the ambient) 
as shown in Figure 8-13. This assembly process also demonstrated the feasibility 
of welding additively manufactured and conventional machined Inconel 718 parts. 
 
Figure 8-13: Welded test section including core, headers, and flanges 
130 
 
 
8.6. High-temperature Experimental Setup   
A high-temperature experimental setup was built from scratch to characterize the 
performance of the test section under high-temperature (600°C) conditions. As the 
hot side has a system pressure of 64.7 psia based on the design requirements 
discussed in chapter 3, the setup’s hot side was initially designed as an open loop 
with a compressor to supply the pressure and flow rate. However, a compressor 
which can supply 0.02 kg/s air at 64.7 psia requires at least 15 – 20 HP power 
supply, has a giant size that required too much laboratory space, and would cause 
huge noises. Therefore, a close loop was designed for the hot side with a smaller 
size compressor. To supply the required flow rate within the close loop, a 3 HP oil-
free scroll enclosed compressor, which can use pressurized gas as inlet flow, was 
selected as shown in Figure 8-14.     
       
Figure 8-14: Powerex oil-free scroll compressor pump [112] (left); 3 HP scroll 
compressor system 
131 
 
 
Because the hot-side inlet temperature for the subscale unit was required as 650°C, 
a commercial electric heater which has a capacity of 13 kW with 3-phase power (@ 
208 V) was selected. Based on the specification, the heater which is made from 
Stainless Steel 304 could supply a maximum exhaust gas temperature of 704°C 
(1300°F) and withstand a maximum pressure of 120 psig. With an average Nitrogen 
temperature of 300°C at the given pressure of 64.7 psia, a pressure drop of ~16 psig 
is expected when mass flow rate inside the heater is 0.02 kg/s. This high pressure 
drop is due to the S-shape flow path inside the heater as shown in Figure 8-15, and 
can be compensated by the selected compressor.  
 
Figure 8-15: Schematic of the commercial heater 
However, as a commercial product for standard industrial applications, the heater 
did not have a connection option as the flange/clamp type used in the high-
temperature setup. Therefore, a modified flange design was created to replace the 
original exhaust fitting ring of the commercial electric heater (as shown in Figure 
8-16) so that it could be installed in the setup. The modified flange was then 
fabricated at University of Maryland and delivered to the heater’s manufacturer to 
be welded to the heater. 
132 
 
 
Weld 
Exhaust 
fitting ring 
          
Figure 8-16: Original exhaust fitting ring design on the heater (left); Modified 
stainless steel 304 flange as replacement of the original exhaust fitting ring (right) 
The electric heater was controlled by a PID controller provided by its manufacturer, 
however, fuses were not included in the control panel due to the large size of the 
high-current fuse and the limited space within the control box. As a result, custom 
designed fuse holder was 3-D printed using high-temperature-resistant plastic 
material (which can withstand a maximum temperature of 200°C), and three 50 
Amp fuses were installed inside the fuse holder. As shown in Figure 8-17. 
Considering most of the risks during the high-temperature experiments (which are 
discussed at the end of this section) could be related to the electric heater, a 3-phase 
power emergency switch, a flow rate relay (check hot-side minimum flow rate), 
and a temperature relay (check compressor inlet temperature) were added in the 
power system.     
133 
 
 
3-phase power 
plug for heater 
3-phase power 
emergency switch 
3-D printed 
fuse holder 
Electric heater 
control panel 
 
Figure 8-17: Power switch & control panel for the electric heater 
Since the subscale unit of the gas-to-gas manifold-microchannel heat exchanger has 
a cross-flow configuration, it has a chance to create non-uniform flow temperatures 
at the hot- and cold-outlets. In order to obtain a uniform flow temperature 
measurement, custom designed mixing chambers were created. To ensure the 
measured location inside the mixing chamber can represent the average 
temperature, Solidworks flow simulations were performed by assuming the 
temperature varies 50°C across the inlet cross-section of the mixing chamber. As 
shown in Figure 8-18 and Table 8-3, at the measured location, the simulation results 
for this extreme case show that the measured temperature is almost the same as the 
134 
 
 
average temperature at mixing chamber’s inlet for the range of mass flow rates. 
Similar results are obtained for other mixing chambers.   
Thermocouple 
measurement 
Outlet Inlet 
 
Figure 8-18: Numerical model of hot-side outlet mixing chamber at mass flow rate 
of 0.02 kg/s 
Table 8-3 Flow simulation results of the hot-side outlet mixing chamber 
?̇?hot  Tavg  Tavg ∆Tavg  ∆Tavg/Tavg  
(kg/s) @ inlet (°C)  @ TC location (°C) (°C) @ inlet (%) 
0.010 475.0 472.9 -2.1 -0.44% 
0.015 475.0 473.2 -1.8 -0.38% 
0.020 475.0 473.0 -2.0 -0.42% 
 
 
 
135 
 
 
The schematic diagram of the high-temperature experimental setup is shown in 
Figure 8-19. In summary, to reach the required system pressure, the hot side was 
built as a closed loop with N2 gas as the working fluid supplied via a N2 gas cylinder. 
A scroll compressor, controlled by a variable speed controller, was used to provide 
the desired flow rate. A 13-kW electric heater was installed on the hot side to heat 
the N2 gas stream to 600°C. A 20-kW chiller and an air-to-water heat exchanger 
(Heat Exchanger 1) were used on the hot side to cool the N2 exhaust from the outlet 
of the test section to ensure the gas temperature at the inlet of compressor was less 
than 20°C. Coriolis flow meters were used to measure the mass flow rates on both 
the hot and cold sides. The cold side was built as an open loop with atmosphere 
system pressure. On the cold side, air was used as the working fluid, and a blower 
controlled by a variable speed controller was used to provide the desired flow rate. 
During the high temperature test, the entire setup needs to be run for at least 30 
minutes for a proper steady temperature measurement. Because of long operation 
of the blower, the outlet air from the blower will increase to more than 45°C for an 
inlet blower temperature of 25°C. A pre-cooler (Heat Exchanger 2) was used on the 
cold-side to maintain a constant cold-side inlet temperature (~ 38°C). Pressure 
transducers and high temperature thermocouples were used to measure the pressure 
drops and temperatures. On both the hot and cold sides, custom designed mixing 
chambers ensured uniform temperature measurements at both the inlets and outlets 
of the test section.   
 
 
136 
 
 
 
Figure 8-19: Schematic diagram of the high-temperature experimental setup 
137 
 
 
3-phase power 
plug for heater AM fabricated 
subscale HX core 
Cold side outlet 
mixing chamber 
Cold side inlet mixing 
chamber 
Hot side outlet 
mixing chamber 
Hot side inlet mixing 
chamber 
Flexible tube 
system  
Electric heater 
Chiller 
Fuse holder 
Pressure transducers 
Electric heater 
control panel 
Cold side  
Tube-in-tube HX Coriolis flow meter 
(heat exchanger 1) 
 Hot side  
Coriolis flow meter 
Compressor 
Variable speed 
controller 
 
Figure 8-20: High-temperature experimental setup 
138 
 
 
The high-temperature experimental setup is shown in Figure 8-20, where the test 
section was connected to the setup through flanges/clamp connection with flexible 
graphite gasket to prevent leakage from contacting flanges’ surfaces. Three thermal 
insulation layers were also covered the entire test section and all mixing chambers 
to reduce heat loss during high temperature tests as shown in Figure 8-21. 
Cold side outlet Hot side outlet Cold side inlet 
mixing chamber mixing chamber mixing chamber 
AM fabricated 
core 
Hot side inlet  
mixing chamber 
 
 
Figure 8-21: Test section & mixing chambers: before installing thermal insulation 
(top), after installing thermal insulation (bottom) 
139 
 
 
Considering the risks involved with the high-temperature experiment setup (such 
as pressurized N2 gas was heated to 600°C or more through a 13-kW electric 
heater), a hazard and operability analysis was conducted before the high-
temperature tests as shown in Table 8-4. Even the high-temperature setup was well 
prepared and insulated, if blower or compressor stops working properly, certain 
risks still need to be considered and methods of mitigation have been concluded in 
the table as well. 
Table 8-4 High-temperature test hazard and operability analysis 
 
Condition Effect Risk Mitigation 
Compressor inlet Compressor Temperature relay off → 
No flow 
temperature too high damaged heater off 
Cold side 
Vent paint 
air flow Low flow Vent exhaust is hot Additional cooling fan 
damaged 
High flow N/A N/A N/A 
Heater minimum Heater Flow relay off  
No flow 
flow not reached damaged → heater off 
Flow relay &  
Hot side Heater minimum Heater 
Low flow Temperature relay off 
N2 flow flow not reached damaged 
→ heater off 
Compressor inlet Compressor Temperature relay off 
High flow 
temperature too high damaged → heater off 
Compressor inlet Compressor Temperature relay off 
No flow 
temperature too high damaged → heater off 
Chiller 
Compressor inlet Compressor  
water Temperature relay off 
Low flow temperature too high & Chiller 
flow → heater off 
& boiling water damaged 
High flow N/A N/A N/A 
Setup Pressure relief valve @ 
High Setup burst 
damaged 95 psia open 
System 
pressure Flow relay &  Heater minimum Heater 
Low Temperature relay off 
flow not reached damaged 
→ heater off 
 
140 
 
 
8.7. High-temperature Experiment Method     
The experiments were performed under high-temperature and adiabatic conditions as 
shown in Table 8-5 and Table 8-6. Four sets of experiments were conducted. For the 
high-temperature tests, first the hot-side flow rate was varied while keeping the cold-
side flow rate constant at 6 g/s. Then, in the second test, the cold-side flow rate was 
varied while keeping the hot-side flow rate constant at 20 g/s. For the adiabatic tests, 
which do not involve any heat transfer, first the hot-side flow rate was varied while 
allowing no flow on the cold-side (flow rate = 0 g/s). Then, the cold-side flow rate 
was varied while allowing no flow on the hot-side. During the tests, the hot-side inlet 
temperature (𝑇ℎ𝑜𝑡,𝑖𝑛), the hot-side outlet temperature (𝑇ℎ𝑜𝑡,𝑜𝑢𝑡), the hot-side pressure 
drop (𝛥𝑃ℎ𝑜𝑡), the hot-side system pressure (𝑃ℎ𝑜𝑡), the hot-side mass flow rate (?̇?ℎ𝑜𝑡), 
the cold-side inlet temperature (𝑇𝑐𝑜𝑙𝑑,𝑖𝑛), the cold-side outlet temperature (𝑇𝑐𝑜𝑙𝑑,𝑜𝑢𝑡), 
the cold-side pressure drop (𝛥𝑃𝑐𝑜𝑙𝑑), and the cold-side mass flow rate (?̇?𝑐𝑜𝑙𝑑) were 
measured and recorded.   
Table 8-5 High-temperature experimental test conditions 
 Hot Side Cold Side 
Temperature Condition 600 °C 38 °C 
System Pressure 448 kPa 101 kPa 
Varying Hot Side Flow Rate 10 – 20 g/s 6 g/s  
Varying Cold Side Flow Rate 20 g/s 4 – 11 g/s 
 
141 
 
 
Table 8-6 Adiabatic experimental test conditions 
 Hot Side Cold Side 
Temperature Condition 38 °C 38 °C 
System Pressure 448 kPa 101 kPa 
Varying Hot Side Flow Rate 0 – 20 g/s 0 g/s  
Varying Cold Side Flow Rate 0 g/s 0 – 14 g/s 
 
Based on the measurement, the heat duty (Q) of the subscale heat exchanger was 
evaluated as: 
𝑄ℎ𝑜𝑡 = 𝐶𝑝ℎ𝑜𝑡?̇?ℎ𝑜𝑡(𝑇ℎ𝑜𝑡,𝑖𝑛 − 𝑇ℎ𝑜𝑡,𝑜𝑢𝑡) (29) 
𝑄𝑐𝑜𝑙𝑑 = 𝐶𝑝𝑐𝑜𝑙𝑑?̇?𝑐𝑜𝑙𝑑(𝑇𝑐𝑜𝑙𝑑,𝑜𝑢𝑡 − 𝑇𝑐𝑜𝑙𝑑,𝑖𝑛) (30) 
 𝑄 = (𝑄ℎ𝑜𝑡 + 𝑄𝑐𝑜𝑙𝑑)/2 (31) 
where Cp is the N2 or air specific heat for the hot- and cold-sides calculated at the 
average temperature between the inlet and outlet measurements. The typical deviation 
observed between Qhot and Qcold was about 6 – 8% (it was about 20% before installing 
the thermal insulation).  
The overall heat transfer coefficient was evaluated using the log mean temperature 
difference (LMTD) method for both unmixed cross-flow arrangements as given in 
Eqs. (32) - (35) as a function of inlet and outlet temperatures on both hot and cold 
sides. 
142 
 
 
𝑑𝑇𝐴 = 𝑇ℎ𝑜𝑡,𝑖𝑛 −  𝑇𝑐𝑜𝑙𝑑,𝑜𝑢𝑡 (32) 
𝑑𝑇𝐵 = 𝑇ℎ𝑜𝑡,𝑜𝑢𝑡 −  𝑇𝑐𝑜𝑙𝑑,𝑖𝑛 (33) 
𝑑𝑇𝐴 − 𝑑𝑇𝐵
𝐿𝑀𝑇𝐷 = 𝐶𝐹  (34) 
ln (𝑑𝑇𝐴 /𝑑𝑇𝐵) 
Q𝑖
𝑈 =  (35) 
𝐴𝑏𝑎𝑠𝑒𝐿𝑀𝑇𝐷 
where CF is the configuration correction factor of both unmixed cross-flow (CF  
0.94 – 0.96), Abase is the microchannel base area (Abase = 0.059 m × 0.070 m), and Qi 
is the heat duty for each manifold-microchannel layer (as shown in Figure 8-22, the 
subscale unit design has two manifold-microchannel layers, and each layer consists 
of a full cross-flow microchannel surface, a half hot-side manifolds, and a half cold-
side manifolds). 
1st layer 
2nd layer 
 
Figure 8-22: Subscale unit’s two manifold-microchannel layers 
 
143 
 
 
Uncertainty propagation analysis was performed to calculate the uncertainty in the 
subscale heat exchanger performance parameters (Q, U) due to inaccuracy of the 
measurements. A list of the measurement equipment is shown in Table 8-7 with the 
corresponding accuracies. The uncertainty of the calculated quantity Y (UY), which is 
a function of X1, X2, . . . , XN with uncertainty of UX1, UX2, . . . , UXN, can be calculated 
using the method explained in NIST Technical Note 1297 [109] as: 
𝜕𝑌
  𝑈𝑌 = √∑ ( )
2𝑈2  (36) 
𝜕𝑋 𝑋𝑖𝑖 𝑖
Table 8-7 List of high-temperature measurement equipment and its accuracy 
  Equipment function Equipment name Accuracy 
Endress Hausser 
Hot-side flow rate Coriolis flow meter ± 0.5% of measurement 
(0.5-inch size) 
Endress Hausser 
Cold-side flow rate Coriolis flow meter  ± 3% of measurement 
(1-inch size) 
± 2.2°C or 0.75% of 
Temperature K type thermocouple  
measurement 
Validyne P55 pressure 
Hot-side pressure drop ± 0.25% FS of 5.5 kPa 
transducer 
Validyne P55 pressure 
Cold-side pressure drop ± 0.25% FS of 5.5 kPa 
transducer 
 
The uncertainties of the subscale heat exchanger performance parameters (Q, U) were 
calculated. The maximum uncertainties in Q, U, and ∆P were evaluated as ± 2.50%, 
± 5.12%, and ± 0.25% respectively, and the error bars are included in the results 
presented in the next section. 
144 
 
 
8.8. High-temperature Experiment Results  
The heat transfer performance of the subscale unit for a constant cold-side 
microchannel Reynolds number is shown in Figure 8-23. The results are plotted as 
functions of hot-side microchannel Reynolds number (Rechn, h) which is defined as: 
𝐷𝑐ℎ𝑛,ℎ?̇?𝑐ℎ𝑛,ℎ
  𝑅𝑒 =  (37) 𝑐ℎ𝑛,ℎ
𝐴𝑐ℎ𝑛,ℎ𝜇ℎ𝑜𝑡
where Achn, h is the hot-side microchannel cross section area (Achn, h = 0.00039 m × 
0.00128 m), Dchn, h is the hot-side microchannel hydraulic diameter (Dchn, h = 0.0006 
m), and ?̇?𝑐ℎ𝑛,ℎ is the mass flow rate per microchannel on the hot-side.  
 
(a) 
145 
 
 
 
(b) 
Figure 8-23: Experimental and numerical results for constant cold-side 
microchannel Reynolds number (𝑅𝑒𝑐ℎ𝑛,𝑐  120) and 560°C temperature difference 
between hot inlet N2 and cold inlet air: (a) heat duty (Q) vs. hot-side microchannel 
Reynolds number (𝑅𝑒𝑐ℎ𝑛,ℎ); (b) overall heat transfer coefficient (𝑈) vs. hot-side 
microchannel Reynolds number (𝑅𝑒𝑐ℎ𝑛,ℎ) 
Figure 8-23 (a) shows the heat duty variation with the hot-side microchannel 
Reynolds number for a temperature difference of 560°C between the hot- and cold-
sides. Up to 2140 W heat duty was achieved at the hot-side’s highest flow rate (when 
the cold-side microchannel Reynolds number is constant about 120). The overall heat 
transfer coefficient of the heat exchanger was in the range of 700 – 800 W/m2K as 
shown in Figure 8-23 (b). This number is higher than the overall heat transfer 
coefficient of a typical air-cooled heat exchanger with gases as the process fluid, 
146 
 
 
which is in the range of 50 – 300 W/ m2K [113].  Heat transfer coefficient in the 
present design is improved due to the two main facts: 1) high area to volume ratio of 
the heat exchanger, and 2) the enhancement related to the manifold-microchannel 
technology as reported in the literature [15, 26]. As expected, the heat duty and overall 
heat transfer coefficient increase with increase in hot-side microchannel Reynolds 
number for the test conditions.  
The heat transfer performance results for the subscale unit with varying cold-side 
microchannel Reynolds number (Rechn, c) are shown in Figure 8-24. Rechn, c is defined 
as: 
𝐷𝑐ℎ𝑛,𝑐?̇?𝑐ℎ𝑛,𝑐
𝑅𝑒𝑐ℎ𝑛,𝑐 =  (38) 𝐴𝑐ℎ𝑛,𝑐𝜇𝑐𝑜𝑙𝑑
where Achn, c is the cold-side microchannel cross section area (Achn, c = 0.00034 m × 
0.00101 m), Dchn, c is the cold-side microchannel hydraulic diameter (Dchn, c = 0.00051 
m), and ?̇?𝑐ℎ𝑛,𝑐 is the mass flow rate per microchannel on the cold-side. 
 
(a) 
147 
 
 
 
(b) 
Figure 8-24: Experimental and numerical results for constant hot-side 
microchannel Reynolds number (𝑅𝑒𝑐ℎ𝑛,ℎ  450) and 560 °C temperature difference 
between hot inlet N2 and cold inlet air: (a) heat duty (Q) vs. cold-side microchannel 
Reynolds number (𝑅𝑒𝑐ℎ𝑛,𝑐); (b) overall heat transfer coefficient (𝑈) vs. cold-side 
microchannel Reynolds number (𝑅𝑒𝑐ℎ𝑛,𝑐) 
Figure 8-24 (a) shows the heat duty variation with the change in cold-side flow rate 
for a constant hot-side microchannel Reynolds number of about 450.  A maximum of 
2780 W heat duty was achieved at cold-side microchannel Reynolds number of 200. 
The overall heat transfer coefficient of the heat exchanger was in the range of 700 – 
1000 W/m2K as shown in Figure 8-24 (b). The graphs also show that heat duty and 
overall heat transfer coefficient increase with increase in cold-side microchannel 
Reynolds number. Comparing the trend between Figure 8-23 and Figure 8-24, it 
shows that cold-side flow rate has more impact on the heat exchanger’s thermal 
148 
 
 
performance. This is due to the fact that the cold-side flow rate is much lower (which 
is almost 50% of the hot-side flow rate). Therefore, the thermal resistance on the cold-
side is higher than the hot-side, and any improvement in cold-side flow rate can 
positively impact the overall heat transfer performance. 
The pressure drop performance of the heat exchanger test section is shown in Figure 
8-25. The trend is as expected: pressure drop increases with increased Reynolds 
number. During the high temperature tests, the hot-side pressure drop reached 4000 
Pa at maximum hot-side microchannel Reynolds number of about 450.  Maximum 
cold-side pressure drop was measured as 5200 Pa at maximum cold-side 
microchannel Reynolds number of about 200. For the adiabatic tests, the hot-side 
maximum pressure drop was measured as 1410 Pa at maximum hot-side 
microchannel Reynolds number of about 900.  Cold-side maximum pressure drop was 
measured as 5210 Pa at cold-side microchannel Reynolds number of about 370.  
 
(a) 
149 
 
 
 
(b) 
Figure 8-25: Experimental and numerical results: (a) hot-side pressure drop (𝛥𝑃ℎ𝑜𝑡) 
vs. hot-side microchannel Reynolds number (𝑅𝑒𝑐ℎ𝑛,ℎ); (b) cold-side pressure drop 
(𝛥𝑃𝑐𝑜𝑙𝑑) vs. cold-side microchannel Reynolds number (𝑅𝑒𝑐ℎ𝑛,𝑐) 
From Figure 8-23 and Figure 8-24, it can be observed that the heat transfer 
performance results show a decent agreement between the numerical and 
experimental data. From Figure 8-25, for the high-temperature tests it can be seen that 
the measured pressure drops are about 10% higher than the numerical predictions on 
the hot-side, and about 25% higher than the numerical predictions on the cold-side. 
For the adiabatic tests, the pressure drop measurements are within 5% of the 
predictions. The deviation observed between the high-temperature tests and adiabatic 
tests is due to the fact that constant gas properties (calculated at average temperature 
between inlet and outlet) were assumed in the numerical model used for pressure drop 
prediction. For the adiabatic tests, since there was no heat transfer involved, the gas 
150 
 
 
properties did not change during the tests. However, for the high-temperature tests, 
the gas properties varied along the flow due to the temperature change. The higher 
deviation observed on the cold-side is due to the temperature difference between cold 
inlet and outlet, which was about 300°C, while the temperature difference between 
the hot inlet and outlet was relatively small (~ 100°C).  
As shown in Figure 8-26, if the constant air properties at cold inlet temperature 
(which is the same as the adiabatic test condition) is used in the calculations, the 
pressure drop prediction on cold side is much lower than the experimental result. If 
the constant air properties at hot inlet temperature is used, the pressure drop 
prediction is much higher than the experimental result. In fact, the experimental 
results fall between two modeling cases: one is that the average value of the cold-
side inlet and outlet temperatures is used and another one is that the cold-side outlet 
temperature is used. From these results, it can be concluded that the most of 
subscale unit on cold side effectively operates at higher temperatures than the inlet 
and outlet average temperature. This may not always be the case for other heat 
exchanger designs since it depends on temperature distribution inside the heat 
exchanger. However, it seems for this particular subscale unit the deviation 
observed is due to this temperature effect.     
 
151 
 
 
6000
5000 Experimental results
4000 Prediction using T =
(T_cold_in + T_cold_out)/2
3000 Prediction using T_cold_in
2000 Prediction using T_hot_in
1000
Prediction using T_cold_out
0
3 4 5 6 7 8 9 10 11 12
ṁcold (g/s)  
Figure 8-26: Experimental and numerical results: cold-side pressure drop vs cold-
side mass flow rate  
Considering plate-fin heat exchangers are the most commonly utilized compact high 
temperature heat exchangers for gas-to-gas heat transfer applications, the 
experimental results of the additively manufactured manifold-microchannel heat 
exchanger were compared with plate-fin heat exchangers to assess its superiority. 
Based on a database provided in Ref. [114] as shown in Table 8-8, the geometries of 
plate-fin heat exchangers were selected through an optimization using COP of 62 and 
effectiveness of 60% as the same design point of the subscale unit under the same 
operating conditions. The COP was evaluated as: 
𝑄
𝐶𝑂𝑃 =  
?̇?
(∆𝑃 × ℎ𝑜𝑡
?̇? (39) 
ℎ,𝑐𝑜𝑟𝑒 + ∆𝑃𝑐,𝑐𝑜𝑟𝑒 ×
𝑐𝑜𝑙𝑑)
𝜌ℎ𝑜𝑡 𝜌𝑐𝑜𝑙𝑑
where ∆𝑃ℎ,𝑐𝑜𝑟𝑒 and ∆𝑃𝑐,𝑐𝑜𝑟𝑒 are the pressure drop of the heat exchanger core on the 
ho- and cold side, 𝜌ℎ𝑜𝑡 and 𝜌𝑐𝑜𝑙𝑑 are the gas density on the hot and cold sides, ?̇?ℎ𝑜𝑡 
and ?̇?𝑐𝑜𝑙𝑑 are the mass flow rates, and Q is the heat duty. 
152 
 
∆Pcold (Pa)
 
Table 8-8 Key geometrical specifications of plate-fin heat exchanger [114] 
  Name Fin pitch (fin/cm) Plate spacing (cm) 
PFHX 1 7.82 0.64 
PFHX 2 5.94 1.06 
PFHX 3 5.82 0.84 
PFHX 4 4.37 0.64 
 
The plate-fin heat exchangers were optimized for a compact size, which was the same 
to minimize mass (m). The sizing was obtained by fixing the mass flow rates, 
effectiveness, and COP to match the subscale manifold-microchannel heat exchanger 
design. The results are presented in the form of heat transfer density (Q/m) versus 
COP in Figure 8-27. The heat transfer density is defined as the ratio of heat duty 
(Q) over the heat exchanger core’s mass (m). The subscale manifold-microchannel 
heat exchanger shows improvement of heat transfer density of up to 25% for the 
same COP compared to the best plate-fin heat exchanger design. 
Figure 8-28 shows the comparison among all heat exchangers in terms of the overall 
heat transfer coefficient. Note that for the conventional plate-fin heat exchangers, the 
base thickness ( 2Hbase ) is assumed to be 0.3 mm whereas for the manifold-
microchannel heat exchanger the fabricated base thickness was 0.5 mm. The result 
again demonstrates the superiority of manifold-microchannel heat exchanger in terms 
of thermal performance over the conventional plate-fin heat exchangers. 
153 
 
 
 
Figure 8-27: Performance comparison between manifold-microchannel heat 
exchanger with conventional plate-fin heat exchangers  
 
Figure 8-28: Comparison of overall heat transfer coefficient (𝑈 ) of manifold-
microchannel heat exchanger and selected conventional plate-fin heat exchangers 
154 
 
 
8.9. Summary  
In summary, the fabrication of the metal headers and the test section (including 
headers, flanges, and the subscale unit) demonstrated the feasibility of welding 
additively manufactured and conventionally machined Inconel 718 parts. A high-
temperature experimental setup was built so that high-temperature experiments (up to 
650°C with maximum system pressure of 655 kPa) for various high temperature heat 
exchangers can be conducted. This chapter also demonstrates that the additively 
manufactured Inconel 718 high temperature gas-to-gas subscale heat exchanger was 
successfully tested at 600°C with ~ 450 kPa inlet pressure. A maximum heat duty of 
2.78 kW and a heat transfer density close to 10 kW/kg was achieved with cold-side 
inlet temperature of 38°C. The decent agreement between the experimental and the 
numerical results demonstrates the validity of the updated numerical analysis models 
used for heat transfer and pressure drop predictions. The performance comparison 
shows that additive manufacturing can be used to achieve the compact and light-
weight high temperature heat exchangers, and benefit the applications where space 
and weight are constrained. 
 
  
155 
 
 
Chapter 9: Advanced Manifold-microchannel Heat 
Exchanger  
9.1. Introduction  
Considering the successful fabrication of the subscale unit through additive 
manufacturing and the promising experimental results compared to the numerical 
predictions, this chapter discusses an advanced gas-to-gas cross-flow manifold-
microchannel heat exchanger design which was created based on the new design 
requirement given by the sponsor. More constraints regarding the advanced heat 
exchanger’s overall size were added during the design stage. And a medium-scale 
unit design was created, which has a heat duty of 20 kW at the design point 
compared to the 2-kW heat duty of the subscale unit, to further evaluate the 
maximum building size and fabrication limits of the DMLS process.  
9.2. Design Requirement for Advanced Manifold-microchannel Heat Exchanger 
Since the fabrication of the subscale unit demonstrated the feasibility of the selected 
additive manufacturing approach (DMLS), and the experimental results showed 
that the 3-D printed heat exchanger can be leak-free and withstand required 
pressure at high temperatures, an advanced design with new operating and physical 
specifications was requested. As shown in Table 9-1 and Table 9-2, the advanced 
heat exchanger design must have a size which is equal to or smaller than the given 
size of the baseline heat exchanger with the same heat duty of 417 kW. 
 
156 
 
 
Table 9-1 Physical specifications for full-scale advanced heat exchanger 
Overall Volume ≤ 1161 in3 (0.019 m3) 
Hot-flow Length ≤ 10.5 in (0.27 m) 
Cold-flow Length ≤ 9.1 in (0.23 m) 
No-flow Length ≤ 12.15 in (0.31 m) 
Weight ≤ 82 lb (37.2 kg) 
Material  Inconel 625 
Table 9-2 Operating specifications for full-scale advanced heat exchanger 
Parameter Unit Nominal Condition 
Flow lb/min (kg/s) 212.2 (1.604) 
Temperature °F (°C)  908 (487) 
Hot Side Pressure psia (kPa) 49.2 (339.2) 
Pressure Drop psid (kPa) 5.43 (37.4)  
Effectiveness % 52.9 
Flow lb/min (kg/s) 229.9 (1.738) 
Temperature °F (°C)  74 (23) 
Cold Side Pressure psia (kPa) 9.52 (65.6) 
Pressure Drop psid (kPa) 1.9 (13.1) 
Effectiveness % 50.9 
 
157 
 
 
9.3. Advanced Manifold-microchannel Heat Exchanger Design    
A single objective optimization was performed to find the minimum weight of the 
advanced manifold-microchannel heat exchanger for the given physical and 
operating conditions listed in the previous section. As shown in Table 9-3, the 
optimal design achieved a 22% weight reduction with a 5% increase of the no-flow 
length and a 2% increase of the pressure drop on the cold side.  
Table 9-3 Optimized advanced manifold-microchannel heat exchanger design 
 𝑄 416 kW (– 0.3%) 
 𝐿ℎ𝑜𝑡 0.267 m  
 𝐿𝑐𝑜𝑙𝑑 0.231 m  
 𝐿𝑛𝑜−𝑓𝑙𝑜𝑤 0.323 m (+ 4.8%) 
 𝑁𝑙𝑎𝑦𝑒𝑟 54 
 M𝑎𝑠𝑠 28.9 kg (– 22.2%) 
 Volume 0.0199 m3 (+ 4.8%) 
 𝑑𝑃ℎ𝑜𝑡 𝑓𝑢𝑙𝑙 37.6 kPa (+ 0.5%) 
 𝑑𝑃𝑐𝑜𝑙𝑑 𝑓𝑢𝑙𝑙 13.4 kPa (+ 2.3%) 
 
Compared to the first additively manufactured gas-to-gas cross-flow manifold-
microchannel heat exchanger discussed before, the advanced design includes a 
much shorter manifold height on the hot side, a thinner microchannel base, thinner 
microchannel fins on both sides, and most importantly, it has vertical fins / manifold 
walls on both sides. As shown in Table 9-4, the advanced design’s hot-side 
manifold height for each layer is 0.95 mm, the microchannel base thickness is 
reduced to 0.15 mm (which means the separating sheet thickness is 0.3 mm), and 
the fin thickness on both sides is ~0.15 mm. 
   
158 
 
 
Table 9-4 Comparison between 3-D printed heat exchanger and advanced design  
  3-D Printed HX Advanced Design 
𝐻𝑐h𝑛 1.010 mm 0.876 mm 
𝑡𝑓𝑖𝑛 0.165 mm 0.149 mm 
𝑊𝑐h𝑛 0.340 mm 0.220 mm 
𝐻𝑐ℎ𝑛,𝑏𝑎𝑠𝑒 0.250 mm 0.150 mm 
𝑊𝑚𝑛𝑑−𝑐h𝑛 4.440 mm 17.657 mm 
𝐻𝑚𝑛𝑑 5.250 mm 3.300 mm 
𝑊𝑖𝑛 3.490 mm 3.531 mm 
Cold side 
𝑊𝑎𝑙𝑙𝑚𝑛𝑑 0.350 mm 0.300 mm 
𝐻𝑚𝑛𝑑,𝑏𝑎𝑠𝑒 0.300 mm 0.150 mm 
𝑁𝑐ℎ𝑛 per in 50 69 
𝑁𝑚𝑛𝑑 11 14 
𝑁𝑐ℎ𝑛  138 626 
𝑁𝑝𝑎𝑠𝑠 1 1 
𝐻𝑚𝑛𝑑/𝐻𝑐h𝑛 5.2 3.8 
𝐻𝑐h𝑛 1.280 mm 0.259 mm 
𝑡𝑓𝑖𝑛 0.180 mm 0.153 mm 
𝑊𝑐h𝑛 0.390 mm 0.310 mm 
𝐻𝑐ℎ𝑛,𝑏𝑎𝑠𝑒 0.250 mm 0.150 mm 
𝑊𝑚𝑛𝑑−𝑐h𝑛 6.100 mm 12.252 mm 
𝐻𝑚𝑛𝑑 4.810 mm 0.950 mm 
𝑊𝑖𝑛 2.480 mm 0.196 mm 
Hot side 
𝑊𝑎𝑙𝑙𝑚𝑛𝑑 0.500 mm 0.300 mm 
𝐻𝑚𝑛𝑑,𝑏𝑎𝑠𝑒 0.300 mm 0.150 mm 
𝑁𝑐ℎ𝑛 per in 32 55 
𝑁𝑐ℎ𝑛  72 577 
𝑁𝑚𝑛𝑑 10 18 
𝑁𝑝𝑎𝑠𝑠 1 1 
𝐻𝑚𝑛𝑑/𝐻𝑐h𝑛 3.8 3.7 
159 
 
 
9.4. Medium-scale Unit Design 
Considering the current DMLS printer in the market does not have the capability to 
print the full-size advanced manifold-microchannel heat exchanger, a medium-
scale unit design was created by scaling down the full-size design. During the 
scaling-down process, channel height (𝐻𝑐h𝑛), fin thickness (𝑡𝑓𝑖𝑛), channel width 
(𝑊𝑐h𝑛), base thickness (𝐻𝑐ℎ𝑛,𝑏𝑎𝑠𝑒), and mass flow rate per microchannel were kept 
the same for the microchannel surface as the full-size design. For the manifold 
scaling, width of manifold channel ( 𝑊𝑚𝑛𝑑−𝑐h𝑛 ), manifold wall thickness 
( 𝑊𝑎𝑙𝑙𝑚𝑛𝑑 ), inlet width to microchannel ( 𝑊𝑖𝑛 ), manifold base thickness 
(𝐻𝑚𝑛𝑑,𝑏𝑎𝑠𝑒), and pressure drop in the manifold per manifold channel length (which 
is the same as flow length) were the same as full-size design. As shown in Table 
9-5, the medium-scale unit’s manifold height and number of microchannels on both 
sides were reduced compared to the full-size design.   
Table 9-5 Comparison between advanced full-size and medium-scale design 
 Cold side Hot side 
 Full-size Medium-Scale Full-size Medium-Scale 
𝐻𝑐h𝑛 0.876 mm 0.876 mm 0.259 mm 0.259 mm 
𝑡𝑓𝑖𝑛 0.149 mm 0.149 mm 0.172 mm 0.172 mm 
𝑊𝑐h𝑛 0.220 mm 0.220 mm 0.291 mm 0.291 mm 
𝐻𝑐ℎ𝑛,𝑏𝑎𝑠𝑒 0.175 mm 0.175 mm 0.175 mm 0.175 mm 
𝑊𝑚𝑛𝑑−𝑐h𝑛 17.657 mm 17.657 mm 12.252 mm 12.252 mm 
𝐻𝑚𝑛𝑑 3.300 mm 1.890 mm 0.950 mm 0.458 mm 
𝑊𝑎𝑙𝑙𝑚𝑛𝑑 0.300 mm 0.300 mm 0.300 mm 0.300 mm 
𝑊𝑖𝑛 3.531 mm 3.531 mm 0.250 mm 0.250 mm 
𝐻𝑚𝑛𝑑,𝑏𝑎𝑠𝑒 0.150 mm 0.150 mm 0.150 mm 0.150 mm 
𝑁𝑐ℎ𝑛 626 204 577 155 
 
160 
 
 
However, compared to the 3-D printed subscale unit, the medium-scale design is 
still ~ 4 times as much as the overall size and 10 times as much as the heat duty as 
shown in Table 9-6. It can also be noticed from Figure 9-1 that medium-scale 
design’s no-flow length is ~ 4 times the subscale design, and much more manifold-
microchannel layers are included in the medium-scale design due to the shorter 
manifold height on both sides.   
Table 9-6 Comparison between subscale unit and medium-scale design 
 3D Printed Subscale Core Medium-scale Core Design  
 𝐿ℎ𝑜𝑡 0.059 m  0.072 m  
 𝐿𝑐𝑜𝑙𝑑 0.070 m  0.075 m  
 𝐿𝑛𝑜−𝑓𝑙𝑜𝑤 0.027 m 0.114 m 
 𝑁𝑙𝑎𝑦𝑒𝑟 2 28 
 𝑄 2.0 kW * 20.3 kW 
𝐹𝑖𝑛 𝑡𝑦𝑝𝑒 Cold vertical / Hot inclined  Cold vertical / Hot vertical 
Note: Heat duty (Q) of 3-D printed subscale IN718 heat exchanger core is based on 
previous operating specifications. 
Medium-scale 
Cold Hot 
Subscale  
Cold Hot 
 
Figure 9-1: Subscale unit design vs. medium-scale unit design 
161 
 
 
Based on the lesson learnt from the welding of the test section discussed in the 
previous chapter, the medium-scale unit’s welding interface frame was created such 
that the welding edge has a thickness of 3 mm, and headers on both sides have 
separate welding edges (by extending the welding interface on all four sides as 
shown in Figure 9-2) to avoid any potential leakage directly from hot-side header 
to cold-side header.   
75 + 7×2 mm  
Hot Cold 
114 + 3×2 mm  72 + 7×2 mm  
 
Figure 9-2: Medium-scale design with the welding interface frame (marked in gray)  
Because the medium-scale design has vertical microchannel fins and vertical 
manifold walls on both sides, unlike the 3-D printed subscale unit, a 45-degree 
printing orientation has to be applied to the medium-scale design so that the cross-
flow microchannel fins can still be self-supported during the sintering process as 
shown in Figure 9-3 (a). However, considering the tapered manifold channel 
162 
 
 
design, when the medium-scale unit is printing in a 45-degree orientation, half of 
the manifold walls will raise an overhang concern since they will be printing in an 
angle less than 45 degrees as shown in Figure 9-3 (b).  
Printing 
orientation 
 
(a) 
Manifolds without 
overhang concern 
Manifolds with 
overhang concern 
Printing 
orientation 
Build plate 
 
(b) 
Figure 9-3: Preferred printing orientation for the medium-scale design (a); tapered 
manifold channel’s overhang concern with 45-degree printing orientation (b) 
163 
 
 
9.5. Summary  
In summary, an advanced manifold-microchannel heat exchanger design was created 
based on the new design specification which has a required minimum heat transfer 
density (Q/m) of 11 kW/kg compared to the previously required one of 6.7 kW/kg. 
To comply with the maximum build envelop size of currently available DMLS 
printers in the market, a medium-scale unit which has the same microchannel 
geometrical dimensions was created. By comparing to the 3-D printed subscale heat 
exchanger, the medium-scale design is 4 times as much as the overall size and 10 
times as much as the heat duty at design point. However, since the advanced cross-
flow design has vertical microchannels fins on both sides for the purpose of avoiding 
additional pressure drop included by the inclined fins, the medium-scale unit has to be 
fabricated with a 45-degree printing orientation. 
  
164 
 
 
Chapter 10: Conclusions and Proposed Future Work  
10.1. Conclusions 
In this dissertation, an additive manufacturing technique (DMLS) was used to 
fabricate a high-temperature gas-to-gas cross-flow manifold-microchannel heat 
exchanger. An advanced design optimization scheme was applied to meet the heat 
transfer density (Q/m) requirement for pre-cooling applications on the aircraft. 
Although this study focused on an aerospace pre-cooling heat exchanger 
application, the results can also be of significance to other industries that rely on 
compact gas-to-gas applications. These include air-cooled condensers (heat pump, 
refrigeration), air-cooled heat sinks (electronic cooling), air-to-air heat exchangers 
(economizers), and most importantly, gas-to-gas high-temperature heat exchangers 
(recuperators for gas turbine power generation). 
As the first additively manufactured high-temperature manifold-microchannel heat 
exchanger, the heat exchanger designed and fabricated for this study was composed 
of manifold-microchannels on both hot and cold sides in a cross-flow configuration. 
An optimization utilizing the modified hybrid method was performed to obtain the 
initial design, which had a minimum weight and met all the design requirements as 
well as the manufacturing constraints of the DMLS process. For comparison 
purposes, single-objective optimizations were also performed based on different 
manufacturing approaches including photo-chemical etching, laser micro-
machining, and diffusion bonding for the same performance constraints. The results 
showed that the 3-D printed heat exchanger offered the highest heat transfer density 
165 
 
 
(Q/m) compared to other manufacturing approaches (40% higher than photo-
chemical etching, and 8% higher than laser micro-machining). This approach 
allowed the heat exchanger to be fabricated from Inconel 718, which is a very tough 
material for conventional machining process, as a single piece. The selected 
approach (DMLS) also eliminated the time-consuming assembly process, which 
may cause unwanted additional pressure drops across the microchannel section due 
to clogging. The full-size heat exchanger design was scaled down to a subscale unit 
that could be fabricated by a 3-D metal printer currently on the market and 
experimentally tested under the existing constraints of the laboratory at University 
of Maryland. 
Several test print coupons were first fabricated through DMLS using Inconel 718 
to evaluate the quality of the 3-D metal printed parts. Different small-feature sizes 
such as microchannel fin thickness and manifold wall thickness were incorporated 
in the coupon design. By measuring the fabricated small features under a digital 
microscope, it was found that the minimum printed vertical microchannel fin 
thickness was 165 µm (± 7 µm), and the minimum printed inclined fin was 180 µm 
(± 12 µm) thick. For all other features whose size was greater than 300 µm, the 
fabricated dimensions were within 7% of the design values. Since the 3-D printed 
metal piece may have had some porosity issues due to the voids generated during 
the powder sintering process, two pressure containment test coupons were 
fabricated to identify the minimum microchannel base thickness (which is the same 
as separating sheet thickness) needed to prevent leakage from the high-pressure 
(64.7 psia) to low-pressure (14.7 psia) side. The pressure containment test results 
166 
 
 
showed a small leakage on the coupon with microchannel base thickness of 0.3 mm 
at 34.7 psia, and the coupon with 0.5 mm thick microchannel base could be leak-
free at the required pressure of 64.7 psia. Another water-based pressure 
containment test also showed that the leak-free coupon could withstand up to 3000 
psig pressure without any deformation.   
Based on the knowledge gained from the test print coupons, the subscale cross-flow 
manifold-microchannel heat exchanger design was modified according to the 
updated manufacturing constraints and successfully fabricated out of Inconel 718 
through the DMLS additive manufacturing technique. The entire unit was 3-D 
printed within 36 hours, which was much less than the quoted lead time of any other 
conventional manufacturing techniques. A detailed inspection of the fabricated 
subscale heat exchanger as well as the small sectional coupon (which was 3-D 
printed simultaneously in the same build as the subscale heat exchanger for 
measuring the internal feature size) demonstrated fabrication fidelity within 14% 
of design values. Except fin thickness on both sides, all other dimensions were 
within 5%. Since tapered manifold channels were incorporated in the subscale 
unit’s design, to achieve more accurate performance predictions, single-manifold-
multi-microchannels (SMMM) models were created to replace the single manifold-
microchannel models initially used in the modified hybrid method. The CFD 
simulation results of the SMMM model also showed that a more uniform flow 
distribution could be obtained with the tapered manifold channels compared to the 
straight channels. The flow maldistribution was ~ 12% with the tapered manifold 
channel and ~25% with the straight manifold channel.  
167 
 
 
The 3-D printed subscale heat exchanger was experimentally tested under both low-
temperature and high-temperature conditions to characterize its performance. The 
low-temperature test results showed a decent agreement for the heat transfer 
performance between the experimental data and the numerical prediction using the 
SMMM model. However, due to the non-uniform flow at the subscale unit’s cold-
side inlet induced by the low-temperature header design, the measured pressure 
drop on the cold side was ~ 40 – 50% higher than the numerical prediction. Since 
the subscale manifold-microchannel heat exchanger was designed for high-
temperature applications, a high-temperature experimental setup was built from 
scratch to characterize the heat exchanger’s performance at high temperatures. To 
address the non-uniform flow induced by the low-temperature header, flow 
deflectors were added into the high-temperature metal headers and flow simulations 
using both Solidworks and Fluent were performed to ensure the flow 
maldistribution from the metal header was less than 8%. Since the full CFD model 
included both headers and the heat exchanger core, a more accurate prediction 
including pressure drop across the inlet/outlet headers and pressure drop across the 
heat exchanger core could be obtained.  
A maximum heat duty of 2.78 kW and a heat transfer density close to 10 kW/kg was 
achieved with cold-side inlet temperature of 38°C during the high-temperature 
experiments. It was observed that the heat transfer performance results again showed 
a decent agreement between the numerical and experimental data. Since the high-
temperature experiments included two test conditions, for the high-temperature tests 
the measured pressure drops were about 10% higher than the numerical predictions 
168 
 
 
on the hot side, and about 25% higher than the numerical predictions on the cold side. 
For the adiabatic tests (no heat transfer involved), the pressure drop measurements on 
both sides were within 5% of the predictions. The deviation observed between the 
high-temperature tests and adiabatic tests was due to the fact that constant gas 
properties (calculated at average temperature between inlet and outlet) were assumed 
in the full CFD model used for pressure drop prediction. For the adiabatic tests, the 
gas properties did not change during the tests, as there was no heat transfer involved. 
However, for the high-temperature tests, the gas properties varied along the flow due 
to the temperature change. The higher deviation observed on the cold side was due to 
the temperature difference between cold inlet and outlet, which was about 300°C, 
while the temperature difference between the hot inlet and outlet was relatively small 
(~ 100°C). The performance of the subscale heat exchanger was also compared to 
selected conventional plate-fin heat exchangers (which are the most common compact 
high temperature heat exchangers for gas-to-gas applications) under the same 
operating conditions. The subscale manifold-microchannel heat exchanger showed 
improvement of heat transfer density of up to 25% for the same COP compared to the 
best plate-fin heat exchanger design.  
Considering the good agreement between experimental and numerical results, the 
method used to predict the additively manufactured manifold-microchannel heat 
exchanger was verified, and the process of the method is summarized as following: 
1. Obtain optimal heat exchanger design based on the operating specifications 
and manufacturing constraints through the optimization using modified hybrid 
method with single manifold-microchannel models. 
169 
 
 
2. Create single-manifold-multi-microchannels (SMMM) models to obtain more 
accurate performance predictions if tapered manifold channel and/or inclined 
microchannel fins are utilized in the final heat exchanger design. 
3. Design headers (with flow deflectors if needed) using Solidworks flow 
simulation as the preliminary approach to obtain the minimum flow 
maldistribution among the manifold channels at the inlet and outlet of the 
manifold-microchannel heat exchanger. 
4. Create simplified single manifold-porous-medium models in which the 
microchannel section is replaced with porous medium to reduce 
computational time, and calibrate the porous medium parameters by matching 
mass flow rate vs. pressure drop based on the predictions obtained from the 
single-manifold-multi-microchannels (SMMM) models. 
5. Conduct Fluent flow simulations including inlet/outlet headers and heat 
exchanger core (with the simplified single manifold-porous-medium models) 
to obtain more accurate flow distribution among inlet/outlet manifold 
channels and pressure drops (including ΔPheader, inlet, ΔPheader, outlet, and ΔPcore). 
Given the successful fabrication of the subscale heat exchanger through additive 
manufacturing and the promising experimental results compared to the numerical 
predictions, an advanced high-temperature gas-to-gas manifold-microchannel heat 
exchanger was designed based on the newly identified physical and operating 
specifications. As a more compact design, the advanced heat exchanger’s manifold 
height on both sides was significantly reduced, and vertical fins were incorporated 
170 
 
 
on both sides of the cross-flow microchannel surface to improve the overall 
performance. To further evaluate the maximum building size of currently available 
DMLS printers on the market, a medium-scale design which was 4 times larger 
than the overall size of the subscale heat exchanger was created and incorporated 
the recommended 45-degree printing orientation.  
10.2. Proposed Future Work 
Several directions for future work in the areas of numerical modeling of manifold-
microchannel heat exchangers, additive manufacturing enhancement, and other 
aspects associated with continuity of the project are proposed in this section. 
10.2.1. Numerical modeling 
The single-manifold-multi-microchannels (SMMM) model discussed in this 
dissertation did not consider the change of gas properties due to different 
temperatures along the flow path inside the manifold-microchannel heat exchanger. 
As mentioned in chapter 8, the deviation of the pressure drop prediction from the 
experimental results on the cold side shows that when the temperature difference 
between inlet and outlet is large, the pressure drop may be greatly underestimated 
with improper gas properties. To improve the SMMM model, a temperature-
dependent flow simulation should be studied, and the gas properties as a function 
of local temperature should be implemented in the SMMM model. 
During the design stage of the high-temperature manifold-microchannel heat 
exchanger, the heat exchanger’s material properties were considered at a constant 
171 
 
 
temperature of 600°C to ensure that the heat exchanger structure would be strong 
enough to withstand the required system pressure under high-temperature conditions. 
However, as the temperature inside the heat exchanger may change significantly from 
600°C to 40°C, some structures inside the heat exchanger could be designed with a 
much smaller thickness to further reduce the heat exchanger’s weight. Such design 
can only be created if a thermal-mechanical model is applied or developed.  
Currently, both the single manifold-microchannel model and the single-manifold-
multi-microchannels model used in the modified hybrid method to predict the 
manifold-microchannel heat exchanger’s performance are based on CFD 
simulations running with a commercially available software (ANSYS Fluent). If a 
universal meta-model that includes the most possible flow characteristics of 
different manifold-microchannel heat exchanger geometries could be created, the 
computational time of the optimization process for a manifold-microchannel heat 
exchanger design could be shortened considerably. However, such universal meta-
model may need thousands of different manifold-microchannel heat exchangers as 
the initial samplings, considering there are at least six variables on each side of the 
manifold-microchannel heat exchanger: microchannel height (𝐻𝑐h𝑛), microchannel 
fin thickness ( 𝑡𝑓𝑖𝑛 ), microchannel width ( 𝑊𝑐h𝑛 ), width of manifold channel 
(𝑊𝑚𝑛𝑑−𝑐h𝑛), height of manifold channel (𝐻𝑚𝑛𝑑), and inlet width to microchannel 
(𝑊𝑖𝑛). 
172 
 
 
10.2.2. Additive manufacturing 
In the field of additive manufacturing, more work is needed to improve the 
fabrication quality of DMLS. To obtain the highest heat transfer performance from 
the advanced manifold-microchannel design, the microchannel fin thickness on 
both sides should be ~ 150 µm. But since the minimum vertical fin thickness of the 
fabricated subscale heat exchanger was 165 µm, either a different DMLS printer 
(the one used to fabricate the subscale heat exchanger is built by Concept Laser, 
but there are other 3-D metal printers built by EOS or 3D Systems) or different 
printing settings are needed to achieve a cross-flow microchannel surface with fin 
thickness of 100 µm or even smaller in some designs.  
Because most of a manifold-microchannel heat exchanger’s weight is from the 
microchannel base, or separating sheet, a thinner microchannel base would not only 
improve the heat transfer by decreasing the thermal resistance between hot and cold 
streams, but also significantly reduce the mass of the heat exchanger. However, to 
prevent leakage directly from the high-pressure to low-pressure side, the printed 
minimum microchannel base was 0.5 mm. This minimum base thickness could be 
reduced to 0.3 mm if the porosity issue of the DMLS process could be resolved by 
either using smaller size metal powder or adjusting the laser power during the 
sintering process. Considerable trial-and-error testing will be needed to achieve a 
thinner leak-free separating sheet for all heat exchanger designs that plan to use 
additive manufacturing as the fabrication approach.  
173 
 
 
Even though many studies have shown that the manifold-microchannel technique 
can significantly improve heat transfer compared to conventional heat exchanger 
designs, and this dissertation has demonstrated that a gas-to-gas cross-flow 
manifold-microchannel heat exchanger can be successfully fabricated using 
additive manufacturing, there is no DMLS printer on the market that can print a 
full-scale manifold-microchannel heat exchanger with a minimum overall size of 
127 × 127 × 127 mm3. The overall size of the medium-scale design, which was 
obtained by scaling down the full-scale advanced heat exchanger, was 86 × 89 × 
120 mm3. This was the maximum build size among the DMLS printers that could 
print a minimum feature size of 150 µm. When a larger build size is desired, a major 
limitation of current 3-D metal printers is that the minimum feature size will be 
about 500 µm due to the fixed location of the laser head. Since industrial-size heat 
exchangers require both large build size and small feature size, current 3-D metal 
printers will need to be modified, such as by incorporating an adjustable laser 
header, if they are to be used to fabricate such heat exchangers.  
10.2.3. Project Continuity 
For continuity of the project, the medium-scale advanced manifold-microchannel 
heat exchanger should be fabricated. Based on the knowledge gained from the 
subscale unit’s fabrication, new test print coupons should be 3-D printed using the 
recommended 45-degree orientation to evaluate the fabrication quality. If the 
tapered manifold wall cannot be fabricated due to the overhang issue discussed in 
the previous chapter, either the tapered angle needs to be modified as shown in 
Figure 10-1 so that a smaller tapered angle can be used by reducing the inlet width 
174 
 
 
of manifold channel (𝑊𝑚𝑛𝑑−𝑐h𝑛−𝑖𝑛), or a straight manifold channel which may 
cause a high flow maldistribution needs to be re-incorporated. 
 
(a) 
 
(b) 
Figure 10-1: Tapered manifold channel design: large tapered angle (a); small 
tapered angle (b) 
175 
 
 
Based on the evaluation of the new test print coupons, the medium-scale design 
should be updated accordingly. The optimization can be re-run if considerable 
changes are identified regarding the manufacturing constraints such as minimum 
printable microchannel fin thickness (or minimum microchannel width which can 
help to reduce fin thickness). If a new numerical model which can consider the 
effect of different manifold channel’s tapered angles is developed, a new advanced 
manifold-microchannel design that can meet all the new design requirements (such 
as shorter manifold height so that no-flow length is also equal to or less than the 
required value) will be possible. 
Considering some high temperature heat exchangers need to operate under both 
high-temperature and high-pressure (≥600°C & ≥80 bars) conditions, a thermal-
mechanical model, which can predict the structural stress conditions at different 
locations and temperatures inside a heat exchanger, should be developed so that 
more reliable manifold-microchannel heat exchanger designs can be obtained. And 
for certain case when the operating temperature is more than 1000°C, using 
ceramics as the material of fabrication will be more demanding with the thermal-
mechanical model. 
  
176 
 
 
Appendices 
Pre-test Procedures for High Temperature Experimental Setup 
Instruments 
• Validyne differential pressure transmitter   
o Hot/cold side 
▪ Range: 0 – 5.5 kPa  
(this can be adjusted by replacing the diaphragm) 
▪ Uncertainty: +/- 0.25% FS 
 
• Sensotec absolute pressure transducer 
o Hot side 
▪ Range: 0 – 200 psia 
▪ Uncertainty: +/- 0.10% FS 
 
• Endress Hausser Coriolis flow meter 
o Hot side (flow meter 1, meter size DN 15) 
▪ Range: 0 – 105 g/s 
▪ Uncertainty: measured error less than 1% (when flow rate ≥ 
6 g/s) 
 
o Cold side option 1 (flow meter 2, meter size DN 50) 
▪ Range: 0 – 250 g/s 
▪ Uncertainty: measured error less than 5% (when flow rate ≥ 
20 g/s) 
 
o Cold side option 2 (flow meter 2, meter size DN 25) 
▪ Range: 0 – 150 g/s 
▪ Uncertainty: measured error less than 3% (when flow rate ≥ 
4 g/s) 
 
177 
 
 
• Omega high-temperature-glass K type thermocouple 
o Hot/cold side 
▪ Range: 0 – 704°C 
▪ Uncertainty: standard limits of error (above 0 °C) is 2.2°C or 
0.75% of measurement 
Experimental Instrumentation Check  
To ensure measurements are as accurate as the instruments can have, all instruments 
will be checked and calibrated individually before the tests 
 
• Pressure transmitter/transducer calibration 
o Differential pressure transmitter’s output voltage reading will be 
calibrated based on water column level reading or special 
differential pressure calibration tool 
o Absolute pressure transducer’s output voltage reading will be 
calibrated based on N2 gas tank’s pressure gauge reading 
 
• Coriolis flow meter 
o Two Coriolis flow meters will be tested in series with a fixed 
flowrate blower, both readings should be the same as the blower’s 
rated flowrate. Otherwise, the flow meters will be sent back to 
manufacturer for re-calibration  
 
• High temperature thermocouple 
o High temperature K type thermocouple will be tested in ice bath to 
check if temperature reading is 0°C  
 
Setup Preparation  
To ensure no leakage issue will be found during the tests (which may either cause 
inaccurate measurement or damage the setup), all individual components and the 
assembled test setup will be checked before the tests 
178 
 
 
• Individual Component Check 
o Electric heater 
▪ Pressure containment test will be conducted on the electric 
heater’s body, no bubble (leak point) should be noticed when 
the body is covered with soap 
 
o Tube-in-tube heat exchanger (heat exchanger 1) 
▪ Pressure containment test will be conducted, no bubble (leak 
point) should be noticed on both outer and inner tubes when 
the heat exchanger is submerged underwater 
 
o Mixing chambers 
▪ Pressure containment tests will be conducted on all four 
mixing chambers (hot/cold side inlet/outlet), no bubble (leak 
point) should be noticed when the mixing chambers are 
submerged underwater 
 
o Test Section  
▪ Pressure containment test will be conducted on the test 
section (both hot and cold side), no bubble (leak point) 
should be noticed when the test section was submerged 
underwater 
 
• Setup Assembly Check 
o Assembled setup 
▪ Pressure containment test will be conducted on the 
assembled setup’s hot side, no bubble (leak point) should be 
noticed along fittings (NPT fittings, compression fittings, 
flange/clamp connections) and pipes (welded flexible metal 
hose) when they are covered with soap 
 
179 
 
 
o Heat loss check 
▪ At the beginning of test, only hot side will be running 
▪ Once desired mass flow rate is reached, wait until flow meter 
reading is stable 
▪ Turn on the heater to heat the hot side gas to desired 
temperature, wait for 20 – 30 minutes until both inlet and 
outlet temperature is stable 
▪ Take thermal images of the test section and inlet/outlet 
mixing chambers to locate any existing hot spot  
▪ If big hot spot is found along the insulated area, and 
calculated heat loss is more than acceptable range (5% of 
desired capacity), turn off hot side 
▪ Wait until hot side is completely cooled, check the hot spot 
area, and add more insulation layers 
▪ Repeat previous steps until calculated heat loss is within 
acceptable range 
 
Experiment Tests 
▪ Adiabatic test 
o Before high temperature tests, in order to quickly assess the quality 
of the heat exchanger (mainly to check whether microchannels are 
clogged), only one side (hot or cold) will be running at room 
temperature (electric heater will be off), and only pressure drop data 
will compared to numerical prediction 
▪ Standard high temperature test procedures 
o Turn on the chiller and ensure chilling water is flowing through the 
tube-in-tube heat exchanger 
o Turn on cold side blower, check cold side flow rate and pressure 
drop readings 
o Pressurize hot side to desired system pressure, turn on hot side 
compressor, check hot side flow rate and pressure drop readings 
180 
 
 
o When hot side mass flow rate is more than 6 g/s, turn on the electric 
heater 
o Heat up hot side gas flow to desired temperature while ensure the 
exhaust gas is cooled down to room temperature through tube-in-
tube heat exchanger 
o Wait until both test section’s inlet and outlet temperature are stable, 
then record the measurements (including test section’s hot/cold side 
inlet/outlet temperatures, hot/cold side pressure drops, hot/cold side 
mass flow rates, and hot side system pressure) 
o Change flow rate on hot and/or cold side, adjust the electric heater 
to control test section hot side inlet temperature, wait until both test 
section’s inlet and outlet temperature are stable 
o Repeat the process based on the pre-determined experiment data sets 
o After all date sets are collected, turn off electric heater while keep 
both hot and cold side running with room temperature gas flow to 
cool down the setup 
o Turn off hot side compressor when all hot side temperature are 
stable as room temperature, then turn off cold side blower 
o Turn off the chiller, and depressurize hot side to atmosphere 
pressure 
 
• Heat balance check 
o During the test, if calculated heat balance is more than acceptable 
range (10%), take thermal images to check if any hot spot can be 
found along test section and mixing chambers 
o If no hot spot is found, turn off heater 
o Check flow meter to see if mass flow rate reading is consistent and 
correct based on previously calibrated data 
o When the test loop is cooled down (less than 50°C), use soap to 
check if there is any leakage along fittings and pipes 
181 
 
 
o If no issue can be found on flow meter, and no leakage can be 
identified along fittings and pipes, turn off hot side and take out the 
thermocouples 
o Test the thermocouples in ice bath and boiling water to check if any 
issue with thermocouples’ reading 
o If thermocouples’ issue is resolved, turn on hot side, and re-check if 
heat balance is within acceptable range 
  
182 
 
 
High Temperature Setup Operation Checklist 
❖ Please read this checklist carefully 
❖ Always request a second person when running the high temperature test 
❖ Be aware if you see or smell anything that looks like burning/melting, turn 
off the heater immediately by turning off the control panel 
 
o Before test 
o Ensure all thermocouples, pressure transducers, and flowmeters are 
connected to the data acquisition system 
o Ensure all power plugs (including dP transducers, temperature switch, 
scroll compressor, blower, flowmeters, and data acquisition system) are 
safely connected                          
o Turn on the chiller and ensure chilling water is flowing through the pipe 
o Set the cooling temperature on the chiller to 10-15°C (push the bottom 
while changing cooling temperature set point) 
o Open the valve on the N2 gas cylinder, ensure there are enough N2 gas 
for the following test (A full cylinder with 2500 psi reading can last 
about 10 hours) 
o Open the hot side charging valve (as shown below) to start filling the 
hot side loop with N2 gas 
                               
o Start the data acquisition system, ensure all the readings make sense for 
non-operation conditions 
o Ensure the bypass valve for hot side dP transducer is open, then slowly 
increase the hot side system pressure through the pressure regulator, and 
183 
 
 
ensure hot side dP reading does not exceed the max reading of the dP 
transducer 
o Turn on the 3-phase power switch for electric heater’s control panel (at 
this point, be careful when standing around the test setup, do not touch 
any 3-phase power cable) 
o During test 
o Turn on the blower by push the start bottom (the green one), slowly 
increase the flow rate by increasing the frequency showing on the 
variable frequency controller’s monitor 
o Check the mass flow rate based on the cold side flowmeter’s reading, 
and ensure the cold side dP reading does not exceed the max reading of 
the dP transducer 
o Close the bypass valve for hot side dP transducer 
o Turn on the scroll compressor by push the start bottom (the green one), 
slowly increase the flow rate by increasing the frequency showing on 
the variable frequency controller’s monitor 
o Check the mass flow rate based on the hot side flowmeter’s reading, and 
ensure the hot side dP reading does not exceed the max reading of the 
dP transducer 
o When hot side mass flow rate is more than 6 g/s, turn on the electric 
heater’s control panel  
▪ Note: the heater has a minimum mass flow rate requirement of 5.5 g/s 
o When hot side and cold side mass flow rate reach the desired point, 
change the heater’s control panel set point to 70°C, then start the heater 
by click “heat on/off” 
o Slowing increase the heating set point to desired temperature (e.g. wait 
2-3 min then increase the set point by 50-80°C each time, and at the 
beginning, it may take 20-30 min for all the temperature measurement 
to be stable) 
▪ Note: remember to check the mass flow rate reading since temperature 
change may cause mass flow rate change as well 
184 
 
 
o For the heater’s control panel, in the setup mode, turn on the “auto tune” 
function if a stable exhaust temperature (± 1-2°C) is desired (for more 
information regarding how to set up the heater’s control panel and the 
alarm system, please read through the heater’s control panel operation 
manual) 
o After test 
o Turn off the electric heater by click “heat on/off”  
o When the process temperature showing on the heater’s control panel 
monitor is less than 40°C, turn off the heater’s control panel 
o Turn off the 3-phase power switch for electric heater’s control panel 
o Slowly decrease the mass flow rate on cold side to about 3 g/s, and 
slowly decrease the mass flow rate on hot side to 6 g/s 
o Let the system cool down for 10-15 min 
o Open the bypass valve for hot side dP transducer 
o Close the valve on the N2 gas cylinder 
▪ Note: when hot side system pressure is reducing, the hot side mass 
flow rate is reducing as well, no need to increase it or keep it 
constant at 6 g/s anymore 
o When the system pressure reduced to atmosphere pressure (when the 
pressure gauge of the pressure regulator shows 0 psi), release the 
pressure regulator by rotate the regulator valve through counter-clock 
direction 
o Close the hot side charging valve 
o After ensuring all the temperature measurement reading is less than 
30°C, turn off the chiller, wait for 5 min, close the yellow valve 
o Decrease the frequency on both scroll compressor and blower’s variable 
frequency controller to 0 Hz, then click the red button 
o Disconnect the power plugs (including dP transducers, temperature 
switch, scroll compressor, blower, flowmeters, and data acquisition 
system) 
  
185 
 
 
References 
 
[1] W. A. Whittenberger, G. W. Brunson, and J. H. Becker, "High-temperature 
heat exchanger," U.S. Patent US8047272, 2011. 
[2] D. Stevenson, D. L. Vrable, and R. J. Watts, "Development of an 
intermediate temperature carbon-carbon heat exchanger for aircraft 
applications," presented at the 44th International SAMPE Symposium, 
1999.  
[3] C. F. McDonald and D. G. Wilson, "The utilization of recuperated and 
regenerated engine cycles for high-efficiency gas turbines in the 21st 
century," Applied Thermal Engineering, vol. 16, no. 8-9, pp. 635-653, 1996. 
[4] J. K. Min, J. H. Jeong, M. Y. Ha, and K. S. Kim, "High temperature heat 
exchanger studies for applications to gas turbines," Heat and Mass 
Transfer, vol. 46, no. 2, pp. 175-186, 2009. 
[5] X. Liu, Y. Deng, K. Zhang, M. Xu, Y. Xu, and C. Su, "Experiments and 
simulations on heat exchangers in thermoelectric generator for automotive 
application," Applied Thermal Engineering, vol. 71, no. 1, pp. 364-370, 
2014. 
[6] T. Wang, Y. Zhang, J. Zhang, G. Shu, and Z. Peng, "Analysis of recoverable 
exhaust energy from a light-duty gasoline engine," Applied Thermal 
Engineering, vol. 53, no. 2, pp. 414-419, 2013. 
[7] D. Aquaro and M. Pieve, "High temperature heat exchangers for power 
plants: Performance of advanced metallic recuperators," Applied Thermal 
Engineering, vol. 27, no. 2-3, pp. 389-400, 2007. 
[8] R. E. Mizia, "Next generation nuclear plant intermediate heat exchanger 
acquisition strategy," Idaho National Laboratory (INL), Idaho Falls, ID 
(United States), 2008. 
[9] K. Hada, I. Nishiguchi, Y. Muto, and H. Tsuji, "Developments of metallic 
materials and a high-temperature structural design code for the HTTR," 
Nuclear engineering and design, vol. 132, no. 1, pp. 1-11, 1991. 
[10] L. Chordia, M. A. Portnoff, and E. Green, "High temperature heat 
exchanger design and fabrication for systems with large pressure 
differentials," Thar Energy, LLC, Pittsburgh, PA (United States), 2017. 
[11] Q. Li, G. Flamant, X. Yuan, P. Neveu, and L. Luo, "Compact heat 
exchangers: A review and future applications for a new generation of high 
temperature solar receivers," Renewable and Sustainable Energy Reviews, 
vol. 15, no. 9, pp. 4855-4875, 2011. 
[12] X. Zhang, H. Keramati, M. Arie, F. Singer, R. Tiwari, A. Shooshtari, and 
M. Ohadi, "Recent developments in high temperature heat exchangers: a 
review," Frontiers in Heat and Mass Transfer, vol. 11, no. 18, 2018. 
[13] G. M. Harpole and J. E. Eninger, "Micro-channel heat exchanger 
optimization," in Semiconductor Thermal Measurement and Management 
Symposium, 1991. SEMI-THERM VII. Proceedings., Seventh Annual IEEE, 
1991, pp. 59-63: IEEE. 
[14] M. Arie, A. Shooshtari, S. Dessiatoun, E. Al-Hajri, and M. Ohadi, 
"Numerical modeling and thermal optimization of a single-phase flow 
186 
 
 
manifold-microchannel plate heat exchanger," International Journal of 
Heat and Mass Transfer, vol. 81, pp. 478-489, 2015. 
[15] M. A. Arie, A. Shooshtari, S. Dessiatoun, and M. Ohadi, "Thermal 
optimization of an air-cooling heat exchanger utilizing manifold-
microchannels," in Thermal and Thermomechanical Phenomena in 
Electronic Systems (ITherm), 2014 IEEE Intersociety Conference on, 2014, 
pp. 807-815: IEEE. 
[16] M. A. Arie, A. H. Shooshtari, and M. M. Ohadi, "Experimental 
characterization of an additively manufactured heat exchanger for dry 
cooling of power plants," Applied Thermal Engineering, vol. 129, pp. 187-
198, 2018. 
[17] V. Jha, S. Dessiatoun, M. Ohadi, A. Shooshtari, and E. Al-Hajri, "High 
performance micro-grooved evaporative heat transfer surface for low grade 
waste heat recovery applications," in ASME 2011 Pacific Rim Technical 
Conference and Exhibition on Packaging and Integration of Electronic and 
Photonic Systems, 2011, pp. 277-283: American Society of Mechanical 
Engineers. 
[18] V. Jha, S. Dessiatoun, A. Shooshtari, E. S. Al-Hajri, and M. M. Ohadi, 
"Experimental characterization of a nickel alloy-based manifold-
microgroove evaporator," Heat Transfer Engineering, vol. 36, no. 1, pp. 33-
42, 2015. 
[19] D. Boyea, A. Shooshtari, S. V. Dessiatoun, and M. M. Ohadi, "Heat transfer 
and pressure drop characteristics of a liquid cooled manifold-microgroove 
condenser," in ASME 2013 Heat Transfer Summer Conference collocated 
with the ASME 2013 7th International Conference on Energy Sustainability 
and the ASME 2013 11th International Conference on Fuel Cell Science, 
Engineering and Technology, 2013: American Society of Mechanical 
Engineers. 
[20] M. Ohadi, K. Choo, S. Dessiatoun, and E. Cetegen, Next generation 
microchannel heat exchangers. Springer, 2013. 
[21] E. Kermani, S. Dessiatoun, A. Shooshtari, and M. M. Ohadi, "Experimental 
investigation of heat transfer performance of a manifold microchannel heat 
sink for cooling of concentrated solar cells," in Electronic Components and 
Technology Conference, 2009. ECTC 2009. 59th, 2009, pp. 453-459: IEEE. 
[22] E. Cetegen, "Force fed microchannel high heat flux cooling utilizing 
microgrooved surfaces," Ph.D thesis, University of Maryland, College Park, 
2010. 
[23] M. Arie, A. Shooshtari, S. Dessiatoun, M. Ohadi, and E. Al Hajri, 
"Simulation and thermal optimization of a manifold microchannel flat plate 
heat exchanger," in ASME 2012 International Mechanical Engineering 
Congress and Exposition, 2012, pp. 209-220: American Society of 
Mechanical Engineers. 
[24] L. Boteler, N. Jankowski, P. McCluskey, and B. Morgan, "Numerical 
investigation and sensitivity analysis of manifold microchannel coolers," 
International Journal of Heat and Mass Transfer, vol. 55, no. 25-26, pp. 
7698-7708, 2012. 
187 
 
 
[25] M. A. Arie, A. Shooshtari, and M. Ohadi, "Air side enhancement of heat 
transfer in an additively manufactured 1 kW heat exchanger for dry cooling 
applications," in Thermal and Thermomechanical Phenomena in Electronic 
Systems (ITherm), 2017 16th IEEE Intersociety Conference on, 2017, pp. 1-
8: IEEE. 
[26] R. S. Andhare, A. Shooshtari, S. V. Dessiatoun, and M. M. Ohadi, "Heat 
transfer and pressure drop characteristics of a flat plate manifold 
microchannel heat exchanger in counter flow configuration," Applied 
Thermal Engineering, vol. 96, pp. 178-189, 2016. 
[27] D. L. Bourell, J. J. Beaman, M. C. Leu, and D. W. Rosen, "A brief history 
of additive manufacturing and the 2009 roadmap for additive 
manufacturing: looking back and looking ahead," Proceedings of 
RapidTech, pp. 24-25, 2009. 
[28] SPI Lasers. Metal Additive Manufacturing. Available: 
http://www.spilasers.com/whitepapers/metal-additive-manufacturing 
[Accessed Jan 2, 2018]. 
[29] C. Hutter, D. Büchi, V. Zuber, and P. Rudolf von Rohr, "Heat transfer in 
metal foams and designed porous media," Chemical Engineering Science, 
vol. 66, no. 17, pp. 3806-3814, 2011. 
[30] M. Wong, I. Owen, C. J. Sutcliffe, and A. Puri, "Convective heat transfer 
and pressure losses across novel heat sinks fabricated by Selective Laser 
Melting," International Journal of Heat and Mass Transfer, vol. 52, no. 1, 
pp. 281-288, 2009. 
[31] S. Tsopanos, C. Sutcliffe, and I. Owen, "The manufacture of micro cross-
flow heat exchangers by selective laser melting," in Enhanced, Compact 
and Ultra-Compact Heat Exchangers: Science, Engineering and 
Technology, 2005: ECI Symposium Series. 
[32] S. M. Thompson, Z. S. Aspin, N. Shamsaei, A. Elwany, and L. Bian, 
"Additive manufacturing of heat exchangers: a case study on a multi-
layered Ti–6Al–4V oscillating heat pipe," Additive Manufacturing, vol. 8, 
pp. 163-174, 2015. 
[33] M. Arie, A. Shooshtari, S. Dessiatoun, and M. Ohadi, "Performance 
characterization of an additively manufactured titanium (Ti64) heat 
exchanger for an air-water cooling application," in ASME 2016 Heat 
Transfer Summer Conference collocated with the ASME 2016 Fluids 
Engineering Division Summer Meeting and the ASME 2016 14th 
International Conference on Nanochannels, Microchannels, and 
Minichannels, 2016: American Society of Mechanical Engineers. 
[34] W. D. Gerstler and D. Erno, "Introduction of an additively manufactured 
multi-furcating heat exchanger," in Thermal and Thermomechanical 
Phenomena in Electronic Systems (ITherm), 2017 16th IEEE Intersociety 
Conference on, 2017, pp. 624-633: IEEE. 
[35] M. A. Arie, A. H. Shooshtari, R. Tiwari, S. V. Dessiatoun, M. M. Ohadi, 
and J. M. Pearce, "Experimental characterization of heat transfer in an 
additively manufactured polymer heat exchanger," Applied Thermal 
Engineering, vol. 113, pp. 575-584, 2017. 
188 
 
 
[36] L. Leslie. (2016). Metal Additive Manufacturing Enables Lean, Green Heat 
Exchanger. Available: http://www.makepartsfast.com/metal-additive-
manufacturing-enables-lean-green-heat-exchanger/ [Accessed March 31, 
2018]. 
[37] Q. Wang, X. Han, A. Sommers, Y. Park, C. T'Joen, and A. Jacobi, "A 
review on application of carbonaceous materials and carbon matrix 
composites for heat exchangers and heat sinks," International Journal of 
refrigeration, vol. 35, no. 1, pp. 7-26, 2012. 
[38] Y. Mori, A. E. Sheĭndlin, and N. H. Afgan, High temperature heat 
exchangers (Proceedings of the International Centre for Heat and Mass 
Transfer; 22.). Washington, D.C.: Hemisphere Pub. Corp., 1986. 
[39] L. M. Crosbie and D. Chapin, "Hydrogen production by nuclear heat," 
GENES4/ANP2003, Kyoto, Japan, 2003. 
[40] C. Mansilla, J. Sigurvinsson, A. Bontemps, A. Maréchal, and F. Werkoff, 
"Heat management for hydrogen production by high temperature steam 
electrolysis," Energy, vol. 32, no. 4, pp. 423-430, 2007. 
[41] L. Magistri, A. Traverso, A. Massardo, and R. Shah, "Heat exchangers for 
fuel cell and hybrid system applications," Journal of fuel cell science and 
technology, vol. 3, no. 2, pp. 111-118, 2006. 
[42] G. Musgrove, R. LePierres, and Nash J., "Heat exchangers for supercritical 
CO2 power cycle applications," in The 4th International Symposium on 
Supercritical CO2 Power Cycles, Pittsburgh Pennsylvania, 2014. 
[43] I. Johnson, W. T. Choate, and A. Davidson, "Waste heat recovery. 
Technology and opportunities in US industry," BCS, Inc., Laurel, MD 
(United States), 2008. 
[44] K. Brun, P. Friedman, and R. Dennis, Fundamentals and applications of 
supercritical carbon dioxide (sCO2) based power cycles. Woodhead 
Publishing, 2017. 
[45] C. F. McDonald, "The role of the ceramic heat exchanger in energy and 
resource conservation," Journal of Engineering for Power, vol. 102, no. 2, 
pp. 303-315, 1980. 
[46] Marlin Steel. (2016). 4 Inconel Machining Tips You Need to Know. 
Available: https://www.marlinwire.com/blog/inconel-machining-tips 
[Accessed March 31,  2018]. 
[47] Continental Steel & Tube Company. (2015). Considerations when 
Machining Nickel Alloys. Available: 
https://continentalsteel.com/blog/considerations-when-machining-nickel-
alloys/ [Accessed March 31, 2018]. 
[48] M. M. Ohadi and S. G. Buckley, "High temperature heat exchangers and 
microscale combustion systems: applications to thermal system 
miniaturization," Experimental Thermal and Fluid Science, vol. 25, no. 5, 
pp. 207-217, 2001. 
[49] C. F. McDonald, "Low-cost compact primary surface recuperator concept 
for microturbines," Applied Thermal Engineering, vol. 20, no. 5, pp. 471-
497, 2000. 
189 
 
 
[50] X. Zhang, M. Arie, D. Deisenroth, A. Shooshtari, S. Dessiatoun, and M. 
Ohadi, "Impact of additive manufacturing on performance enhancement of 
heat exchangers: a case study on an air-to-air heat exchanger for high 
temperature applications," in IX Minsk International Seminar on Heat 
Pipes, Heat Pumps, Refrigerators, Power Sources, Minsk, Belarus, 2015. 
[51] M. A. Arie, A. H. Shooshtari, V. V. Rao, S. V. Dessiatoun, and M. M. 
Ohadi, "Air-side heat transfer enhancement utilizing design optimization 
and an additive manufacturing technique," Journal of Heat Transfer, vol. 
139, no. 3, 2017. 
[52] S. Kandlikar, S. Garimella, D. Li, S. Colin, and M. R. King, Heat transfer 
and fluid flow in minichannels and microchannels. elsevier, 2005. 
[53] E. Utriainen and B. Sundén, "Evaluation of the cross corrugated and some 
other candidate heat transfer surfaces for microturbine recuperators," 
Journal of Engineering for Gas Turbines and Power, vol. 124, no. 3, pp. 
550-560, 2002. 
[54] J. Stasiek, "Experimental studies of heat transfer and fluid flow across 
corrugated-undulated heat exchanger surfaces," International Journal of 
Heat and Mass Transfer, vol. 41, no. 6-7, pp. 899-914, 1998. 
[55] S. Foerster and M. Kleemann, "Compact metallic and ceramic recuperators 
for gas turbines," in ASME 1978 International Gas Turbine Conference and 
Products Show, 1978: American Society of Mechanical Engineers. 
[56] J.-W. Seo, C. Cho, S. Lee, and Y.-D. Choi, "Thermal characteristics of a 
primary surface heat exchanger with corrugated channels," Entropy, vol. 18, 
no. 12, 2015. 
[57] R. Shah and D. Sekulic, "Heat exchangers," in Handbook of Heat Transfer, 
Warren M.   Rohsenow, James P. Hartnett, and Y. I. Cho, Eds. 3rd ed. New 
York: McGraw-Hill, 1998. 
[58] E. Lara-Curzio, P. J. Maziasz, B. A. Pint, M. Stewart, D. Hamrin, N. 
Lipovich, and D. DeMore, "Test facility for screening and evaluating 
candidate materials for advanced microturbine recuperators," presented at 
the Volume 1: Turbo Expo 2002, 2002.  
[59] A. Franco and C. Casarosa, "Thermoeconomic evaluation of the feasibility 
of highly efficient combined cycle power plants," Energy, vol. 29, no. 12-
15, pp. 1963-1982, 2004. 
[60] G. Xiao, T. Yang, H. Liu, D. Ni, M. L. Ferrari, M. Li, Z. Luo, K. Cen, and 
M. Ni, "Recuperators for micro gas turbines: a review," Applied Energy, 
vol. 197, pp. 83-99, 2017. 
[61] K. Takase, H. Furukawa, and K. Nakano, "A preliminary study of an inter-
cooled and recuperative microgasturbine below 300 kW," in ASME Turbo 
Expo 2002: Power for Land, Sea, and Air, 2002, pp. 1009-1016: American 
Society of Mechanical Engineers. 
[62] J. Kesseli, T. Wolf, J. Nash, and S. Freedman, "Micro, industrial, and 
advanced gas turbines employing recuperators," in ASME Turbo Expo 2003, 
collocated with the 2003 International Joint Power Generation Conference, 
2003, pp. 789-794: American Society of Mechanical Engineers. 
190 
 
 
[63] C. M. Parker K, "Development in plate-fin heat exchangers," ASME 
Publication HTD, vol. 10, pp. 171–179, 1979. 
[64] C. F. McDonald, "Compact buffer zone plate-fin IHX—the key component 
for high-temperature nuclear process heat realization with advanced MHR," 
Applied Thermal Engineering, vol. 16, no. 1, pp. 3-32, 1996. 
[65] G. Lagerström and M. Xie, "High performance and cost effective 
recuperator for micro-gas turbines," no. 36061, pp. 1003-1007, 2002. 
[66] A. Muley and B. Sundén, "Advances in recuperator technology for gas 
turbine systems," in ASME 2003 International Mechanical Engineering 
Congress and Exposition, 2003, pp. 161-169: American Society of 
Mechanical Engineers. 
[67] M. A. Wilson, K. Recknagle, and K. Brooks, "Design and development of 
a low-cost, high temperature silicon carbide micro-channel recuperator," in 
ASME Turbo Expo 2005: Power for Land, Sea, and Air, 2005, pp. 1029-
1034: American Society of Mechanical Engineers. 
[68] H. Antoine and L. Prieels, "The ACTE spiral recuperator for gas turbine 
engines," in ASME Turbo Expo 2002: Power for Land, Sea, and Air, 2002, 
pp. 1025-1031: American Society of Mechanical Engineers. 
[69] J. I. Oswald, D. A. Dawson, and L. A. Clawley, "A new durable gas turbine 
recuperator," in ASME 1999 International Gas Turbine and Aeroengine 
Congress and Exhibition, 1999: American Society of Mechanical 
Engineers. 
[70] R. A. Proeschel, "Proe 90™ recuperator for microturbine applications," in 
ASME Turbo Expo 2002: Power for Land, Sea, and Air, 2002, pp. 1033-
1044: American Society of Mechanical Engineers. 
[71] H. Schönenborn, E. Ebert, B. Simon, and P. Storm, "Thermomechanical 
design of a heat exchanger for a recuperative aero engine," in ASME Turbo 
Expo 2004: Power for Land, Sea, and Air, 2004, pp. 187-193: American 
Society of Mechanical Engineers. 
[72] R. Cook, R. Exner, and L. Graham, "Post-service examination of a 10 MW 
helium-helium heat exchanger and comparison with long term behaviour in 
laboratory tests," High temperature metallic materials for gas-cooled 
reactors, p. 129, 1989. 
[73] M. E. Collings, B. A. Dockter, D. R. Hajicek, A. K. Henderson, J. P. Hurley, 
P. L. Kleven, and G. F. Weber, "High-temperature heat exchanger testing 
in a pilot-scale slagging furnace system,"; University of North Dakota (US), 
1999, Available: http://www.osti.gov/scitech/servlets/purl/824978. 
[74] M. Anheden, "Analysis of gas turbine systems for sustainable energy 
conversion," Ph.D thesis, Trita-KET, Kemiteknik, Stockholm, 2000. 
[75] K. A. Al-attab and Z. A. Zainal, "Performance of high-temperature heat 
exchangers in biomass fuel powered externally fired gas turbine systems," 
Renewable Energy, vol. 35, no. 5, pp. 913-920, 2010. 
[76] I. Martinez. Aircraft Environmental Control. Available: 
http://webserver.dmt.upm.es/~WoSdoro/tc3/Aircraft%20ECS.pdf 
[Accessed March 31, 2018]. 
191 
 
 
[77] Bombardier Inc., "System and method for operating a precooler in an 
aircraft," 2012. Available: https://patents.justia.com/patent/9624831. 
[78] H. Tsuji, High temperature air combustion : from energy conservation to 
pollution reduction (Environmental and energy engineering series; 
Environmental and energy engineering series.). Boca Raton, FL :: CRC 
Press, 2003. 
[79] S. A. David, J. A. Siefert, J. N. DuPont, and J. P. Shingledecker, 
"Weldability and weld performance of candidate nickel base superalloys for 
advanced ultrasupercritical fossil power plants part I: fundamentals.," 
Science and Technology of Welding and Joining, vol. 20, no. 7, pp. 532-
552., 2015. 
[80] M. A. Arie, "Air side heat transfer enhancement in heat exchangers utilizing 
innovative designs and the additive manufacturing technique," Ph.D thesis, 
Univerisity of Maryland, College Park, 2016. 
[81] S. K. Mylavarapu, X. Sun, R. N. Christensen, R. R. Unocic, R. E. Glosup, 
and M. W. Patterson, "Fabrication and design aspects of high-temperature 
compact diffusion bonded heat exchangers," Nuclear Engineering and 
Design, vol. 249, pp. 49-56, 2012. 
[82] Heatric. Typical characteristics of diffusion-bonded heat exchangers. 
Available: 
https://www.heatric.com/typical_characteristics_of_PCHEs.html 
[Accessed March 31, 2018]. 
[83] Heatric. The exceptional performance of Heatric PCHE heat exchangers. 
Available: https://www.heatric.com/heat_exchanger_performance.html 
[Accessed March 31, 2018]. 
[84] E. Cetegen, "Force fed microchannel high heat flux cooling utilizing 
microgrooved surface," Ph.D thesis, University of Maryland, College Park, 
2010. 
[85] D. Copeland, M. Behnia, and W. Nakayama, "Manifold microchannel heat 
sinks: isothermal analysis," IEEE Transactions on Components, Packaging, 
and Manufacturing Technology – Part A, vol. 20, no. 2, pp. 96-102, 1997. 
[86] S. T. Poh and E. Y. K. Ng, "Investigative study of manifold microchannel 
heat sinks for electronic cooling design," Journal of Electronics 
Manufacturing, vol. 9, no. 2, pp. 155-166, 1999. 
[87] S. T. Poh and E. Y. K. Ng, "Heat transfer and flow issues in manifold 
microchannel heat sinks: a CFD approach," presented at the 2nd Electronics 
Packaging Technology Conference, Singapore, 1998.  
[88] Y. H. Kim, W. C. Chun, J. T. Kim, B. C. Pak, and B. J. Baek, "Forced air 
cooling by using manifold microchannel heat sinks," Journal of Mechanical 
Science and Technology, vol. 12, no. 4, pp. 709-718, 1998. 
[89] J. H. Ryu, D. H. Choi, and S. J. Kim, "Three-dimensional numerical 
optimization of a manifold microchannel heat sink," International Journal 
of Heat andMass Transfer, vol. 46, no. 9, pp. 1553-1562, 2003. 
[90] Y. Wang and G. F. Ding, "Numerical analysis of heat transfer in a manifold 
microchannel heat sink with high efficient copper heat spreader," 
Microsystem Technologies, vol. 14, no. 3, pp. 389-395, 2008. 
192 
 
 
[91] W. Escher, T. Brunschwiler, B. Michel, and D. Poulikakos, "Experimental 
investigation of an ultrathin manifold microchannel heat sink for liquid-
cooled chips," Journal of Heat Transfer, vol. 132, 2010. 
[92] L. M. Boteler, "Microfabrication and analysis of manifold microchannel 
coolers for power electronics," Ph.D thesis, University of Maryland, 
College Park, 2011. 
[93] L. Boteler, N. Jankowski, P. McCluskey, and B. Morgan, "Numerical 
investigation and sensitivity analysis of manifold microchannel coolers," 
International Journal of Heat and Mass Transfer, vol. 55, no. 25, pp. 7698-
7708, 2012. 
[94] R. Mandel, S. Dessiatoun, P. McCluskey., and M. Ohadi, "Embedded two-
phase cooling of high flux electronics via micro-enabled surface and fluid 
delivery system (FEEDS)," presented at the Proceedings of the ASME 2015 
International Technical Conference and Exhibition on Packaging and 
Integration of Electronic and Photonic Microsystems, San Francisco, 
California, USA, 2015.  
[95] A. Lohner, "Laser sintering ushers in new route to PM parts," Metal Powder 
Report (MPR), vol. 52, no. 2, pp. 24-30, 1997. 
[96] ProtoLabs. (2018). Design Guidelines: Direct Metal Laser Sintering. 
Available: https://www.protolabs.com/services/3d-printing/direct-metal-
laser-sintering/design-guidelines/ [Accessed September 20, 2018]. 
[97] A. Carter, Stratasys Ltd., private communication ed, 2014. 
[98] 3D Systems. (2018). Direct Metal Printing. Available: 
https://www.3dsystems.com/on-demand-manufacturing/direct-metal-
printing/materials [Accessed Septermber 20, 2018]. 
[99] M.  Eesa and G.   Marriott. Photochemical Etching of Heat Exchanger 
Plates.Available: 
http://www.heatexdesign.com/_resources/assets/attachment/full/0/301720
5.pdf [Accessed March 31, 2018]. 
[100] Potomac-laser. (2018). Laser Micromachining Metals. Available: 
http://www.potomac-laser.com/services/core/laser-micromachining 
[Accessed September 20, 2018]. 
[101] N. Muhammad, B. D. Rogers, and L. Li, "Understanding the behaviour of 
pulsed laser dry and wet micromachining processes by multi-phase 
smoothed particle hydrodynamics (SPH) modelling," Journal of Physics D: 
Applied Physics, vol. 46, 2013. 
[102] M. G. Nicholas, Joining processes. In: Introduction to Brazing and 
Diffusion Bonding. Kluwer Academic, 1998. 
[103] K. Kitazono, K. Susumu, and K. Yuki, "Mechanical properties of titanium 
particles dispersed magnesium matrix composite produced through 
accumulative diffusion bonding process," Materials transactions vol. 52, 
no. 2, pp. 155-158, 2011. 
[104] V. Aute, O. Abdelaziz, S. Azarm, and R. Radermacher, "Cross-validation 
Based Single Response Adaptive Design of Experiments," presented at the 
12th AIAA/ISSMO Multidisciplinary Analysis and Optimization 
Conference, 2008.  
193 
 
 
[105] S. N. Lophaven, H. B. Nielsen, and J. Søndergaard, "DACE: A Matlab 
kriging toolbox, Citeseer," 2002. 
[106] M. A. Arie, A. H. Shooshtari, S. V. Dessiatoun, E. Al-Hajri, and M. M. 
Ohadi, "Numerical modeling and thermal optimization of a single-phase 
flow manifold-microchannel plate heat exchanger," International Journal 
of Heat and Mass Transfer, vol. 81, pp. 478-489, 2015. 
[107] R. K. Mandel, M. A. Arie, A. H. Shooshtari, and M. M. Ohadi, "A heat 
spreading model for double-sided, cross-flow, manifold-microchannel heat 
exchangers," presented at the The 17th Intersociety Conference on Thermal 
and Thermomechanical Phenomena in Electronic Systems, San Diego, CA, 
2018.  
[108] B. David, "Designing for the DMLS Process," ed: Protolabs, 2014. 
[109] B. N. Taylor and C. E. Kuyatt, Guidelines for evaluating and expressing the 
uncertainty of NIST measurement results. Citeseer, 1994. 
[110] Special Metals. (2018). Inconel 718 spec sheet. Available: 
http://www.specialmetals.com/documents/Inconel%20alloy%20718.pdf 
[Accessed September 20, 2018] 
[111] Eaton. (2016). Industrial Clamp Products. Available: 
https://www.eaton.com/ecm/groups/public/@pub/@eaton/@aero/documen
ts/content/ct_194111.pdf [Accessed September 20, 2018] 
[112] Compressor World. (2018). Powerex 5 HP Air Compressor Oilless Scroll 
Pump | SLAE05E. Available: https://www.compressorworld.com/slae05e-
5-horsepower-oilless-scroll-air-compressor-replacement-pump-
manufactured-by-powerex.html [Accessed September 20, 2018] 
[113] Engineering Page. Typical Overall Heat Transfer Coefficients. Available: 
http://www.engineeringpage.com/technology/thermal/transfer.html 
[Accessed April 2, 2018]. 
[114] W. M. Kays and A. L. London, Compact heat exchangers, third ed ed. 
Florida: Krieger, 1984. 
 
194